Pump-controlled hydraulic circuits for operating a differential hydraulic actuator

Abstract

Pump-controlled hydraulic circuits are more efficient than valve-controlled circuits, as they eliminate the energy losses due to flow throttling in valves and require less cooling effort. Presently existing pump-controlled solutions for single rod cylinders encounter an undesirable performance during certain operating conditions. Novel circuit designs employ use of different charge pressures on a pair of pilot-operated charging-control valves or different piston areas and/or spring constants on a shuttle-type charging control valve to shift a critical loading region in a load-force/actuator-velocity plane to a lower load force range, thereby reducing the undesired oscillations experienced in the response of the typical critical loading region. One or more specialized valves are controlled by fluid pressures to provide throttling in the circuit only within the critical loading region, thereby reducing the oscillatory amplitude while avoiding throttling-based energy losses outside the critical region over the majority of the circuit's operational overall operating area.

Claims

1. A pump-controlled hydraulic circuit for operating a differential hydraulic actuator, said circuit comprising: a reversible hydraulic pump; a first main fluid line connecting a first side of the reversible hydraulic pump to an extension side of the differential hydraulic actuator; a second main fluid line connecting a second side of the reversible hydraulic pump to a retraction side of the differential hydraulic actuator; a hydraulic charging system for supplying/releasing charging fluid to and from the first and second main fluid lines to compensate for differential flow on opposing sides of the differential hydraulic actuator, said charging system having two different outlets respectively providing higher and lower pressure supplies of charging fluid; first and second charging lines respectively connecting the charging system to the first and second main fluid lines, and each being connected to a different one of said two different outlets of the charging system; and at least one charging-control valve (32, or p.sub.crA & p.sub.crB) operably installed in the first and/or second charging lines and operable to switch between at least a first charging fluid supply/release state enabling flow through the first charging line between the first main fluid line and the charging circuit, and a second charging fluid supply/release state enabling flow through the second charging line between the second main fluid line and the charging circuit, thereby enabling supply and release of the charging fluid to and from the first and second main fluid lines, whereby the reversible hydraulic pump cooperates with the differential hydraulic cylinder via the main charging lines, the charging lines and the charging system to operate to provide a four quadrant mode operation including a first load-resistive actuator-extension quadrant, a second load-assistive actuator-extension quadrant, a third load-resistive actuator-retraction quadrant and a fourth load-assistive actuator-retraction quadrant; wherein said at least one charging-control valve comprises first and second charging control valves (p.sub.crA & p.sub.crB), at least one of which is further configured to also operate as pilot-operated vibration-damping valve (42a, 42b, 44a, 44b) configured to throttle flow in the hydraulic circuit in a critical loading zone of the four-quadrant mode of operation, while allowing unthrottled flow in the hydraulic circuit outside the critical loading zone.

2. The hydraulic circuit of claim 1 wherein a higher pressure one of said two different outlets of the charging system is connected to the second charging line to connect the higher pressure supply of charging fluid to the second main fluid line in the second charging fluid supply/release state of the at least one charging control valve.

3. A pump-controlled hydraulic circuit for operating a differential hydraulic actuator, said circuit comprising: a reversible hydraulic pump; a first main fluid line connecting a first side of the reversible hydraulic pump to an extension side of the differential hydraulic actuator; a second main fluid line connecting a second side of the reversible hydraulic pump to a retraction side of the differential hydraulic actuator; a hydraulic charging system for supplying/releasing charging fluid to and from the first and second main fluid lines to compensate for differential flow on opposing sides of the differential hydraulic actuator, said charging system having two different outlets respectively providing higher and lower pressure supplies of charging fluid; first and second charging lines respectively connecting the charging system to the first and second main fluid lines, and each being connected to a different one of said two different outlets of the charging system; and at least one charging-control valve (32, or p.sub.crA & p.sub.crB) operably installed in the first and/or second charging lines and operable to switch between at least a first charging fluid supply/release state enabling flow through the first charging line between the first main fluid line and the charging circuit, and a second charging fluid supply/release state enabling flow through the second charging line between the second main fluid line and the charging circuit, thereby enabling supply and release of the charging fluid to and from the first and second main fluid lines, whereby the reversible hydraulic pump cooperates with the differential hydraulic cylinder via the main charging lines, the charging lines and the charging system to operate to provide a four quadrant mode operation including a first load-resistive actuator-extension quadrant, a second load-assistive actuator-extension quadrant, a third load-resistive actuator-retraction quadrant and a fourth load-assistive actuator-retraction quadrant; wherein the hydraulic charging system comprises a charging pump, and a pressure reducer connected between the charging pump and the first fluid charging line to define a lower pressure one of said two different outputs of the charging system, the first charging line being connected to said lower pressure one of said two different outputs to connect the lower pressure supply of charging fluid to the first main fluid line in the first charging fluid supply/release state of the first and second charging control valves.

4. A pump-controlled hydraulic circuit for operating a differential hydraulic actuator, said circuit comprising: a reversible hydraulic pump; a first main fluid line connecting a first side of the reversible hydraulic pump to an extension side of the differential hydraulic actuator; a second main fluid line connecting a second side of the reversible hydraulic pump to a retraction side of the differential hydraulic actuator; a hydraulic charging system for supplying/releasing charging fluid to and from the first and second main fluid lines to compensate for differential flow on opposing sides of the differential hydraulic actuator; a first charging line connecting the charging circuit to the first main fluid line; a second charging line connecting the charging circuit to the second main fluid line; and at least one charging-control valve operably installed in the first and/or second charging lines and operable to switch between at least a first charging fluid supply/release state enabling flow through the first circuit-charging line between the first main fluid line and the charging circuit, and a second charging fluid supply/release state enabling flow through the second circuit-charging line between the second main fluid line and the charging circuit, thereby enabling supply and release of the charging fluid to and from the first and second main fluid lines, whereby the reversible hydraulic pump cooperates with the differential hydraulic cylinder via the main charging lines, the charging lines and the charging system to operate to provide a four quadrant mode operation including a first load-resistive actuator-extension quadrant, a second load-assistive actuator-extension quadrant, a third load-resistive actuator-retraction quadrant and a fourth load-assistive actuator-retraction quadrant; wherein the at least one charging-control valve comprises first and second pilot-operated charging-control valves (POCV.sub.A & POCV.sub.B) respectively installed in the first and second charging lines, with a pilot of the first pilot-operated charging-control valve connected to the second main fluid line and a pilot of the second pilot-operated charging-control valve connected to the first main fluid line, and the hydraulic circuit further comprises a first and second pilot-operated vibration damping valves (CBV.sub.A, CBV.sub.B) respectively installed in the first and second main lines between the first and second pilot-operated charging-control valves and the differential hydraulic actuator, and configured to throttle fluid during low loading conditions of the differential hydraulic actuator, and to freely pass fluid in an unthrottled manner during higher loading conditions of the differential hydraulic actuator.

5. The hydraulic circuit of claim 4 wherein said pilot-operated vibration damping valves (CBV.sub.A, CBV.sub.B) comprise pilot-operated counterbalance valves (CBV.sub.A CBV.sub.A), with a pilot of the first pilot-operated counterbalance valve connected to the second main fluid line and a pilot of the second pilot-operated counterbalance valve connected to the first main fluid line.

6. A pump-controlled hydraulic circuit for operating a differential hydraulic actuator, said circuit comprising: a reversible hydraulic pump; a first main fluid line connecting a first side of the reversible hydraulic pump to an extension side of the differential hydraulic actuator; a second main fluid line connecting a second side of the reversible hydraulic pump to a retraction side of the differential hydraulic actuator; a hydraulic charging system for supplying/releasing charging fluid to and from the first and second main fluid lines to compensate for differential flow on opposing sides of the differential hydraulic actuator, said charging system having two different outlets respectively providing higher and lower pressure supplies of charging fluid; first and second charging lines respectively connecting the charging system to the first and second main fluid lines, and each being connected to a different one of said two different outlets of the charging system; and at least one charging-control valve (32, or p.sub.crA & p.sub.crB), operably installed in the first and/or second charging lines and operable to switch between at least a first charging fluid supply/release state enabling flow through the first charging line between the first main fluid line and the charging circuit, and a second charging fluid supply/release state enabling flow through the second charging line between the second main fluid line and the charging circuit, thereby enabling supply and release of the charging fluid to and from the first and second main fluid lines, whereby the reversible hydraulic pump cooperates with the differential hydraulic cylinder via the main charging lines, the charging lines and the charging system to operate to provide a four quadrant mode operation including a first load-resistive actuator-extension quadrant, a second load-assistive actuator-extension quadrant, a third load-resistive actuator-retraction quadrant and a fourth load-assistive actuator-retraction quadrant; wherein the at least one charging-control valve comprises a charging-control valve (32) having first and second piston areas for driving of said charging-control valve in opposing directions using fluid from opposing ones of said main fluid lines and resisted by first and second springs, and wherein said first and second piston areas differ from one another in size, and/or said first and second springs have different spring constants.

7. A pump-controlled hydraulic circuit for operating a differential hydraulic actuator, said circuit comprising: a reversible hydraulic pump; a first main fluid line connecting a first side of the reversible hydraulic pump to an extension side of the differential hydraulic actuator; a second main fluid line connecting a second side of the reversible hydraulic pump to a retraction side of the differential hydraulic actuator; a hydraulic charging system for supplying/releasing charging fluid to and from the first and second main fluid lines to compensate for differential flow on opposing sides of the differential hydraulic actuator, said charging system having two different outlets respectively providing higher and lower pressure supplies of charging fluid; first and second charging lines respectively connecting the charging system to the first and second main fluid lines, and each being connected to a different one of said two different outlets of the charging system; and at least one charging-control valve (32, or p.sub.crA & p.sub.crB), operably installed in the first and/or second charging lines and operable to switch between at least a first charging fluid supply/release state enabling flow through the first charging line between the first main fluid line and the charging circuit, and a second charging fluid supply/release state enabling flow through the second charging line between the second main fluid line and the charging circuit, thereby enabling supply and release of the charging fluid to and from the first and second main fluid lines, whereby the reversible hydraulic pump cooperates with the differential hydraulic cylinder via the main charging lines, the charging lines and the charging system to operate to provide a four quadrant mode operation including a first load-resistive actuator-extension quadrant, a second load-assistive actuator-extension quadrant, a third load-resistive actuator-retraction quadrant and a fourth load-assistive actuator-retraction quadrant; wherein the at least one charging-control valve comprises a shuttle valve (32) having a center position presenting closure or throttling points between the first and second charging lines and two differently pressured outlets of the charging system, a first shifted position opening the first charging line to the charging system and closing the second charging line from the charging system to define the first charging fluid supply/release state, and a second shifted position opening the second charging line to the charging system and closing the first charging line from the charging system to define the second charging fluid supply/release state.

8. A pump-controlled hydraulic circuit for operating a differential hydraulic actuator, said circuit comprising: a reversible hydraulic pump; a first main fluid line connecting a first side of the reversible hydraulic pump to an extension side of the differential hydraulic actuator; a second main fluid line connecting a second side of the reversible hydraulic pump to a retraction side of the differential hydraulic actuator; a hydraulic charging system for supplying/releasing charging fluid to and from the first and second main fluid lines to compensate for differential flow on opposing sides of the differential hydraulic actuator, said charging system having two different outlets respectively providing higher and lower pressure supplies of charging fluid; first and second charging lines respectively connecting the charging system to the first and second main fluid lines, and each being connected to a different one of said two different outlets of the charging system; and at least one charging-control valve (32, or p.sub.crA & p.sub.crB), operably installed in the first and/or second charging lines and operable to switch between at least a first charging fluid supply/release state enabling flow through the first charging line between the first main fluid line and the charging circuit, and a second charging fluid supply/release state enabling flow through the second charging line between the second main fluid line and the charging circuit, thereby enabling supply and release of the charging fluid to and from the first and second main fluid lines, whereby the reversible hydraulic pump cooperates with the differential hydraulic cylinder via the main charging lines, the charging lines and the charging system to operate to provide a four quadrant mode operation including a first load-resistive actuator-extension quadrant, a second load-assistive actuator-extension quadrant, a third load-resistive actuator-retraction quadrant and a fourth load-assistive actuator-retraction quadrant; wherein the at least one charging-control valve comprises a shuttle valve (32) having a center position closing or throttling both the first and second charging lines, a first shifted position opening the first charging line to the charging system and closing the second charging line from the charging system to define the first charging fluid supply/release state, a second shifted position opening the second charging line to the charging system and closing the first charging line from the charging system to define the second charging fluid supply/release state, first and second piston areas arranged to shift the valve into the first and second shifted positions respectively when acted upon by sufficient fluid pressure, and first and second springs respectively resisting movement into the first and second shifted positions, wherein the piston areas differ from one another in size and/or the springs differ from one another in stiffness.

9. The hydraulic circuit of claim 8 wherein the shuttle valve (32) closes the first and second charging lines in the center position.

10. The hydraulic circuit of claim 8 wherein the shuttle valve (32) throttles the first and second charging lines in the center position.

11. A pump-controlled hydraulic circuit for operating a differential hydraulic actuator, said circuit comprising: a reversible hydraulic pump; a first main fluid line connecting a first side of the reversible hydraulic pump to an extension side of the differential hydraulic actuator; a second main fluid line connecting a second side of the reversible hydraulic pump to a retraction side of the differential hydraulic actuator; a hydraulic charging system for supplying/releasing charging fluid to and from the first and second main fluid lines to compensate for differential flow on opposing sides of the differential hydraulic actuator, said charging system having two different outlets respectively providing higher and lower pressure supplies of charging fluid; first and second charging lines respectively connecting the charging system to the first and second main fluid lines, and each being connected to a different one of said two different outlets of the charging system; at least one charging-control valve (32, or p.sub.crA & p.sub.crB), operably installed in the first and/or second charging lines and operable to switch between at least a first charging fluid supply/release state enabling flow through the first charging line between the first main fluid line and the charging circuit, and a second charging fluid supply/release state enabling flow through the second charging line between the second main fluid line and the charging circuit, thereby enabling supply and release of the charging fluid to and from the first and second main fluid lines, whereby the reversible hydraulic pump cooperates with the differential hydraulic cylinder via the main charging lines, the charging lines and the charging system to operate to provide a four quadrant mode operation including a first load-resistive actuator-extension quadrant, a second load-assistive actuator-extension quadrant, a third load-resistive actuator-retraction quadrant and a fourth load-assistive actuator-retraction quadrant; and further comprising one or more pilot-operated vibration-damping valves (32, or CBV.sub.A & CBV.sub.B), wherein the at least one charging-control valve comprises first and second charging control valves (p.sub.crA & p.sub.crB), and the one or more pilot-operated vibration-damping valves are installed in one or both of the main lines at one or more locations between the first and second charging control valves and the differential hydraulic actuator, and are configured to throttle fluid during low loading conditions of the differential hydraulic actuator, and to freely pass fluid in an unthrottled manner during higher loading conditions of the differential hydraulic actuator.

12. The hydraulic circuit of claim 11 wherein the one or more vibration-damping valves comprise one or more variable flow area valves (32, CBV.sub.A, CBV.sub.B) each having a variable and controllable flow area, and arranged to maintain a smaller flow area during the low loading conditions before enlarging the flow area for the higher loading conditions.

13. The hydraulic circuit of claim 11 wherein the one or more vibration-damping valves comprise first and second pilot-operated counterbalance valves (CBV.sub.A CBV.sub.A) respectively installed in the first and second main fluid lines, with a pilot of the first pilot-operated counterbalance valve connected to the second main fluid line and a pilot of the second pilot-operated counterbalance valve connected to the first main fluid line.

Description

BRIEF DESCRIPTION OF THE DRAWINGS

(1) One embodiment of the invention will now be described in conjunction with the accompanying drawings in which:

(2) FIG. 1 schematically illustrates a prior art hydraulic circuit for pump-based control of a differential linear hydraulic actuator using piloted-operated check valves in a cross-pump line fed by a singular charging pressure.

(3) FIG. 2 shows a prior art outline of critical zones during pump mode of operation switching between the second and first quadrants and the fourth and third quadrants which, for simplicity, will be designated to be in the first and fourth quadrants of a four-quadrant operational area of a pump-controlled differential linear hydraulic actuator of FIG. 1.

(4) FIG. 3 shows more elaborate features of the critical zones for the FIG. 1 circuit taking into account the effect of transmission line losses, Coulomb and viscous frictions and cracking pressures of the POCVs.

(5) FIG. 4 schematically illustrates a first embodiment hydraulic circuit of the present invention for pump-based control of a differential linear hydraulic actuator using pair of piloted-operated check valves (potentially having different cracking pressures) in charging lines fed by two different charging pressures to shift the critical zones to lower loading ranges.

(6) FIG. 5 schematically illustrates a second embodiment hydraulic circuit using a singular biased shuttle valve operated by a singular charging pressure to instead perform the critical zone shifting effected by the different charged POVCs of the first embodiment.

(7) FIG. 6 schematically illustrates a third embodiment hydraulic circuit using a singular 4-way 3 position shuttle valve actuated in opposing directions by two different pilot pressures to both shift the critical zones and provide a leakage control action within the shifted critical zones.

(8) FIG. 6A schematically illustrates a variant of the FIG. 8 circuit in which the 4-way 3-position shuttle valve has a closed center position rather than an open center position allowing some intentional leakage flow through the valve.

(9) FIG. 7 schematically illustrates a fourth embodiment hydraulic circuit using the two differently charged pilot-operated check valves of the first embodiment for zone-shifting functionality together with a single dual-piloted selective-throttling valve on one of the main fluid lines to throttle flow therethrough only at the low loading values of the shifted critical zones.

(10) FIG. 8 schematically illustrates fifth embodiment hydraulic circuit in which the single dual-piloted selective-throttling valve from the fourth embodiment is replaced by two counterbalancing valves respectively installed in the two main fluid lines to perform the selective throttling at the low loading values, and a single-charging pressure is used for simplification.

(11) FIG. 8A schematically illustrates a variant of the FIG. 8 circuit modified to include the differently charged pilot-operated check valves of the first and fourth embodiments for shifting of the critical loading zones.

(12) FIG. 9 schematically illustrates a sixth embodiment hydraulic circuit in which both the pilot-operated check valves and counterbalancing valves of the fifth embodiment variant of FIG. 8A are replaced with pilot-operated selective-throttling valves installed in the charging lines to both shift the critical oscillatory zone in the load-assistive fourth quadrant retraction of the actuator, and throttle the differential flow during this critical zone.

(13) FIG. 10 schematically illustrates a seventh embodiment hydraulic circuit in which the pilot-operated selective-throttling valves of the sixth embodiment are replaced with sequence valves.

(14) FIG. 11 schematically illustrates an eighth embodiment hydraulic circuit in which one of the sequence valves of the seventh embodiment is replaced with a pilot-operated check valve.

(15) FIG. 12 shows a test rig used for experimentation testing of the second, fifth, seventh and eighth embodiments of FIGS. 5, 8, 10 and 11, including (1) JD-48 backhoe attachment, (2) main pump unit, (3) charge pump unit, (PS) pressure sensors, and (DS) displacement sensor.

(16) FIG. 13 shows experimental identification of critical zones (shown by hashed lines) given the prior art circuit of FIG. 1 utilizing POCVs.

(17) FIG. 14 shows typical performance results of the prior art shown in FIG. 1 circuit with POCVs only in extension and retraction at 2.54 kN external load (marked by distinguished points in FIG. 13), and more specifically shows the (a) control signal applied to pump swash plate system; (b) actuator velocity.

(18) FIG. 15 shows performance of the FIG. 8 circuit at retraction and extension of 2.54 kN external load, and more specifically shows the: (a) control signal: and (b) actuator velocity.

(19) FIG. 16 shows the control signal applied for experimental evaluation of the FIG. 8 circuit compared to performance of FIG. 1 circuit.

(20) FIG. 17 shows the actuator velocity performance of the FIG. 1 circuit utilizing only POCVs at 4 quadrants of operation and 0.4 kN external load.

(21) FIG. 18 shows the actuator velocity performance of the FIG. 8 circuit at 4 quadrants of operation and 0.4 kN external load.

(22) FIG. 19 shows energy delivered/received by main pump in the FIG. 1 circuit that utilizes only POCVs (dotted line) and the FIG. 8 circuit (solid line).

(23) FIG. 20 schematically illustrates a 4-way 3-position shuttle valve employed in the third embodiment of FIG. 6.

(24) FIG. 21 schematically illustrates a dual-piloted selective-throttling valve employed in the fourth embodiment of FIG. 7.

(25) FIGS. 4A, 5A, 6B, 7A, 8B, 8C, 9A, 10A and 11A show the flow of hydraulic fluid through the circuits of FIGS. 4, 5, 6, 7, 8, 8A, 9, 10 and 11, respectively, in each of the four quadrants of operation, with the first to fourth quadrant operations shown sequentially counter-clockwise from the top right corner of the figure.

(26) In the drawings like characters of reference indicate corresponding parts in the different figures.

DETAILED DESCRIPTION

(27) FIG. 4 illustrates a first embodiment hydraulic circuit of the present invention that, like the prior art circuit of FIG. 1, features the same layout of a reversible hydraulic pump 10, a single-rod differential linear actuator 12, and first and second main fluid lines L.sub.A, L.sub.B respectively connecting the first and second sides of the reversible pump 10 to the extension and retraction sides 12a, 12b of the actuator, and likewise includes first and second pilot-operated check valves POCV.sub.A, POCV.sub.B respectively installed on first and second charging lines 22, 24 that connect the first and second main fluid lines L.sub.A, L.sub.B to a charging system 14 with a unidirectional pump 16. Once again, the POCVs are operated by way of cross pilot lines 26, 28 each connecting the pilot port of the respective POCV to the opposing main fluid line, whereby the differential flow to and from the cylinder in all four quadrants is accommodated in the same manner described for the prior art in the preceding background. The first and second pilot-operated check valves POCV.sub.A, POCV.sub.B thus serve as the two charging-control valves of this embodiment.

(28) However, the circuit differs from that of FIG. 1 in that the two charging lines 22, 24 are independent from one another and fed by two different outputs of the charging system 14. The second charging line 24 and POCV.sub.B installed thereon are fed directly by the unidirectional charging pump 16, like in the circuit of FIG. 1, but the first charging line 22 and POCV.sub.A installed thereon are instead fed indirectly by the unidirectional charging pump 16 via a pressure reducing valve 30 that reduces the pressure of the charging fluid pumped by the charging pump 16. The feeding of POCV.sub.A by a lower charging pressure than POCV.sub.B causes the critical operation zones of FIG. 3 to shift toward the origin of the actuator-velocity/load-force plot along the x-axis, thus lowering the load force range spanned by each critical zone. Since the oscillation in the hydraulic circuit occurs at lower loading values due to this shifting of the critical oscillatory zone in the fourth quadrant, the effective degree of vibration experienced by the operator of the excavator or other machine is less pronounced, thus improving the overall operability of same.

(29) FIG. 5 shows a second embodiment which likewise performs shifting of the critical zones to lower ranges on the load force axis of the four quadrant operational plot, but instead of using two different respective charging pressures to uniquely characterize the two different actuating inputs respectively acting on the two POVCs, the circuit instead employs a singular 3-way 3-position double-piloted shuttle valve 32 as a singular charging-control valve of this embodiment that relies on a conventional single-pressure charging system 14 and is driven by two unique pilot inputs 32a, 32b from the two main lines L.sub.A and L.sub.B. The purpose of the charge system's unilateral low pressure pump, low pressure relief valve and tank/reservoir is feeding or releasing flow from each of the main lines as the operation requirements. In quadrants 1 and 2 the charge pump 16 of the charging system feeds the line L.sub.B and L.sub.A to balance the flow to the main pump and actuator respectively. In quadrants 3 and 4, the relief valve in the charging system allows the release of the extra flow from lines L.sub.A and L.sub.B, respectively. Rather than differing in terms of their charge pressure source, these uniquely characterized pilot inputs 32a, 32b instead differ from one another in terms of the piston surface area and/or spring constant used at each input. The shuttle valve is connected between the singular output of the single-pressure charging system 14 and each of the two charging lines 22, 24, and is biased into a center position by a pair of springs 34a, 34b. In this default center position, the valve 32 closes both of the charging lines 22, 24 from the singular outlet of the charging system, thus defining a normally-closed condition of the valve 32. The first pilot input 32a is fed from the first charging line 22 by a first pilot path 36a, where the fluid pressure from the first charging line 22 acts on the piston area A.sub.PA of the first pilot input 32a to drive movement of the shuttle valve in one direction. The second pilot input 32b is fed from the second charging line 24 by a second pilot path 36b, where the fluid pressure from the second charging line 24 acts on the piston area A.sub.PB of the second pilot input 32b to drive movement of the shuttle valve in the opposing direction. First spring 34a, has a first spring constant k.sub.SA that opposes actuation of the shuttle valve in the first direction by the pilot pressure at first input 32a, while second spring 34b has a different second spring constant k.sub.SB that opposes actuation of the shuttle valve in the second direction by the pilot pressure at second input 32b. The ratio between the two charge pressures and the ratio between the two spring stiffnesses are related to the ratio of the two piston areas.

(30) In a first shifted position of the valve resulting from actuation of the valve 32 via first pilot input 32a against the resistance of first spring 34a, the valve connects the second charging line 24 to the charging system 14, while closing off the first charging line 22 therefrom. In the second shifted position of the valve resulting from actuation of the valve 32 via second pilot input 32b against the resistance of second spring 34b, the valve 32 connects the first charging line 22 to the charging system 14, while closing off the second charging line 24 therefrom. So like the POCVs in the first embodiment circuit of FIG. 4, the shuttle valve 32 connects the charging system to the first main fluid line L.sub.A via the first charging line 22 in the second and third quadrants of operation, and connects the charging system to the second main fluid line L.sub.B via the second charging line 24 in the first and fourth quadrants of operation, thereby accommodating the differential flow into and out of the actuator in all operational modes. However, by characterising the two actuation inputs of the shuttle valve 32 from one another by either piston area, resistive spring constant, or both, the singular charging pressure can accomplish the critical zone shifting function performed by the differently charged POCVs of the first embodiment. To accomplish this result, first input 32a is characterized by a larger piston area than second input 32b and/or by lesser spring stiffness at spring 34a than at spring 34b.

(31) If the valve 32 instead had two identical pilot areas and springs of equal stiffness, undesirable switching back and forth between the two shifted positions of the valve (i.e. critical zone conditions) would occur around the area where the two pilot pressures from lines 22 and 24 are close to each other. At this condition, there would be a bias force exerted on the actuator due to the area difference between the two faces of the actuator piston 18. By using the differently characterized inputs, the shuttle valve of the inventive circuit accomplishes bias-balancing pressures because shifting the pressure balance at valve where switching occurs shifts the bias-force at the actuator (and consequently the load) to null value.

(32) Shifting the critical zones causes the proper matching between the main pump null position (zero control volt.fwdarw.zero swash angle.fwdarw.-zero flow) and the actuator null position (zero actuation force.fwdarw.zero velocity), thereby avoiding the bias force created in the prior art by the single charge pressure and the identical valve(s) resulting in undesirable and uncontrollable motion, especially if there is no resistive load, which can create dangerous conditions in various applications, including applications other than excavation machine actuator control.

(33) FIG. 6 shows a third embodiment hydraulic circuit again using a singular shuttle valve 32 having two pilot inputs 32a, 32b for driving the valve in opposing directions out of a default center position against the resistance of respective springs 34a, 34b, and using different piston areas and/or resistive spring constants for the two inputs. Like in FIG. 5, the first and second pilot inputs 32a, 32b are respectively fed by first and second pilot paths 36a, 36b coming off the first and second charging lines 22, 24. However, instead of using the conventional single-pressure charging system 14 of FIG. 5, the circuit instead uses the dual-pressure charging system 14 of FIG. 4, with a lower charging pressure provided from the pressure reducing valve 30 than directly from the charge pump 16. Accordingly, the shuttle valve 32 in this embodiment is a 4-way 3-position shuttle valve. In the default center position, the valve 32 provides a throttled connection of first charging line 22 to the lower pressure side of the dual-pressure charging system 14, and a throttled connection of second charging line 24 to the higher pressure side of the dual-pressure charging system 14. In the first shifted position caused by sufficient pressurization of pilot input 32a against the resistance of spring 34a, second charging line 24 is connected to the higher pressure side of the dual-pressure charging system 14 for free-flowing unthrottled connection therebetween, while first charging line 22 is closed off from the charging system. In the second shifted position caused by sufficient pressurization of pilot input 32b against the resistance of spring 34b, first charging line 22 is connected to the lower pressure side of the dual-pressure charging system 14 for free-flowing unthrottled connection therebetween, while second charging line 24 is closed off from the charging system.

(34) The initially centered position of shuttle valve 32 thus allows some intentional leakage of fluid between the main lines L.sub.A, L.sub.B to the charging system 14 at lower loading conditions, until enough pilot pressure builds up to drive the shuttle valve into one of its two shifted free-flowing unthrottled conditions. Like in the first two embodiments, the use of different charging pressures and the use of different piston areas and/or spring constants cause the critical loading zones to shift to lower loading conditions of the operational map, during which dampening of the oscillations in the oscillatory critical zone is performed by the intentional leakage to the charging system through the throttled center position ports of the valve. The amplitude of the oscillations are thus dampened, thereby reducing the vibrational effect on the overall machine to improve the performance quality thereof. In the meantime, differential flow to and from the actuator is accommodated over the full operational area by opening up of second charging line 24 between the charging system and the second main fluid line in quadrants 1 and 4, and by opening up of first charging line 22 between the first main fluid line and the charging system in quadrants 2 and 3. In brief, the circuit acts to reduce the critical load value corresponding to the undesirable regions, thereby shifting the undesirable/critical performance region/zones in the oscillatory zone 6 towards the central origin of the load-force/actuator-velocity plot along the load-force axis to a lower range of loading values within which the undesirable performance may be induced, and applies leakage to dampen vibration at this shifted critical region. This reduces the leakage needed to stabilize the system and saves energy compared to the prior art. The shuttle valve 32 in this embodiment thus singularly serves as both a charging-control valve and vibration-damping valve of the hydraulic circuit. This embodiment is believed to possess improved performance compared to the first two embodiments, but has a more complex design.

(35) FIG. 20 schematically illustrates the shuttle valve 32 of the FIG. 6 circuit. In the illustrated example, the valve is a spool valve in which an internal spool member 100 is linearly displaceable back and forth on a longitudinal axis of an outer housing 102 in which four flow connection ports 104a, 104b, 105a, 105b open radially into the housing. First and second connection ports 104a, 104b respectively connect to charging lines 22, 24, while third and fourth connection ports 105a, 105b respectively connect to the lower and higher pressure sides of the charging system. The third and fourth charging system ports are closer to one another and closer to the center of the valve than the first and second charging line ports. The displaceable spool member features a flow-blocking central land 106, two neighbouring flow-enabling valleys 107 on opposing sides thereof, and two flow-obstructing outer lands 108a, 108b at opposing ends of the spool. A respective chamber is defined between each end of the displaceable spool member and a respective closed end of the housing, and each chamber is fed by a respective channel in the housing wall that connects the chamber to a respective one of flow connection ports 104a, 104b. Each chamber and the respective outer landed end of the spool thus collectively define a respective one of the pilot inputs 32a, 32b, at which the respective end of the spool defines the piston area of this pilot input, while the respective channel of each chamber defines the respective pilot path 36a, 36b for fluid-based operation of the pilot input.

(36) Springs 34a, 34b each reside between one end of the displaceable spool member and a respective end of the housing to bias the spool into the centered position, where the central land 106 of the spool resides between the first and second charging line connection ports 104a, 104b and between the third and fourth charging system connection ports 105a, 105b. In the centered spool position, the first and second flow-obstructing outer lands 108a, 108b respectively block off the substantial majority of the charging line connection ports 104a, 104b, but leave a small fraction of each charging line connection port open at the side thereof nearest the other charging line connection port. In the centered spool position, the third charging system connection port 105a is left open at the first flow-enabling spool valley 107a, and the fourth charging system connection port 105b is likewise left open at the second flow-enabling spool valley 107b. This way, in the normal centered position of the valve spool, some intentional fluid leakage can occur between the first charging line connection port 104a and the third charging system connection port 105a, and also between the second charging line connection port 104b and the fourth charging system connection port 105b.

(37) Under application of sufficient pressure against the first landed end of the spool at the first pilot input 32a, the spool shifts in first direction along the longitudinal axis of the housing, moving the first outer land 108a into a position fully sealed with an intact area of the housing's internal periphery at a location situated axially between the first charging line connection port 104a and the third charging system connection port 105a, thereby fully closing off these two ports from one another. At the same time, the second outer land 108b is pushed toward the nearest end of the housing in order to further open the second charging line connection port 104b. This travel is short enough that the central land 106 remains between the third and fourth charging system ports 105a, 105b and thus does not close off the fourth charging system connection port 105b from the fully opened second charging line connection port 104b. Accordingly, the second charging line connection port 104b and the fourth charging system connection port 105b are open to one another in this first shifted position to enable flow between the second charging line and the higher pressure side of the dual-pressure charging system, while the first charging line and the lower pressure side of the dual-pressure charging system are closed off from one another by the first outer land 108a of the spool. With sufficient pilot pressure at the second input 32b, shifting in the reverse direction likewise uses the second outer land 108b to close the second charging line connection port 104b and the fourth charging system connection port 105b from one another while further opening the first charging line connection 104a to enable flow between the first charging line and the lower pressure side of the dual-pressure charging system.

(38) FIG. 6A shows a variant of the FIG. 6 circuit in which the 4-way 3-position shuttle valve is not open in its default center position to allow throttled leakage therethrough, and instead is fully closed in the center position.

(39) FIG. 7 illustrates a fourth embodiment hydraulic circuit of the present invention, which like the first embodiment circuit of FIG. 4 features first and second pilot-operated check valves POCV.sub.A, POCV.sub.B respectively installed on first and second charging lines 22, 24 that connect the first and second main fluid lines L.sub.A, L.sub.B to lower and higher pressure sides of the dual-pressure charging system 14, and are operated by way of cross pilot lines 26, 28 each connecting the pilot port of the POCV to the opposing main fluid line. The fourth embodiment thus features the same critical zone-shifting functionality as the first embodiment to reduce oscillatory behaviour in the actuator of the machine by reducing the load range over which critical loading oscillation occurs in the fourth quadrant of operation.

(40) The fourth embodiment circuit differs from the first embodiment in the addition of a selective-throttling valve 32, and differs from the second and third embodiments in both the type of valve employed for this dampening function and its position within the circuit. Particularly, the illustrated valve 32 is a 2-way valve installed in the first main fluid line L.sub.A near the connection thereof to the extension side 12a of the actuator 12. Like the correspondingly numbered valves 32, 32 of the preceding embodiments, the purpose of this vibration dampening valve 32 is to reduce oscillations under critical loading conditions. This valve 32 may alternatively be installed in the second main fluid line L.sub.B, but locating the valve 32 in the first main line L.sub.A is preferred, since experimental results have showed that oscillatory motions are more noticeable during actuator retraction of assistive load (quadrant 4), where the load is acting to pressurize the fluid in the capped extension side of the actuator. The pilot-operated actuation inputs at 32a, 32b at opposing ends of the valve 32 are activated via pilot paths 36a, 36b from the two pilot lines 26, 28 of the POCVs, whereby fluid pressure from first main fluid line L.sub.A drives the valve in one direction out of a normally centered position, while fluid pressure from second main fluid line L.sub.B drives the valve in an opposing direction out of the normally centered position. Once again, motion of the valve 32 in each direction out of center is resisted by a respective spring 34a, 34b, whereby the springs cooperate to normally center the valve. Spring 34a resists pressure-based operated of piloted input 32a, while spring 34b resists pressure-based actuation of piloted input 32b.

(41) The valve has a variable flow area controlled as a function of the piloting pressure differential, for example using a spool-sleeve throttling configuration and balance springs to achieve the flow-area profile shown in the inset of FIG. 7, where it can be seen that at its centered position (zero-displacement), the open flow area of the valve is zero. In each direction from the centered position, the flow-area gradually increases at a first rate denoted by the gradual slope shown rising slowly away from the origin of the graphical represented flow-area profile in the FIG. 7 inset, until the flow-area's rate of increase rises dramatically at a predetermined point of displacement, as shown by the transition to a notably steeper slope in the graphically represented profile. Within the displacement range between the predetermined displacement points in the positive and negative directions from center, the low flow-through area of the valve performs a throttling action on the fluid passing therethrough. Beyond these points the flow-through area of the valve increases quickly to a free-flow state allowing the fluid to pass freely therethrough with no throttling action thereon. The pre-set displacement points at which the valve transitions from its throttling condition to its free-flowing state are set for a given circuit according to the pilot pressures at which the load value F.sub.L has moved beyond the critical range, whereby throttling of the fluid in the hydraulic circuit is only performed in the critical zones to dampen the vibration/oscillation experienced therein, while the free-flowing state of the valve avoids unnecessary throttling in all other regions, which represent the majority of the overall operating area of the circuit. The energy inefficiencies of throttling are therefore only exploited where needed, while efficient unthrottled operation of the circuit is retained elsewhere.

(42) In other words, the main idea behind the FIG. 7 circuit is to utilize flow throttling to control the actuator motion, exclusively, in the regions where responses are not satisfactory. In other regions, motion is controlled in a throttle-less manner. Throttling of hydraulic fluid creates pressure drop across the valve orifices maintaining increased pressure in cylinder chambers compared to pump ports which contribute towards a stiffer actuator [24, 28]. The circuit of FIG. 7 possesses a comparable energy efficiency and energy regeneration ability to the prior art circuit with POCVs (FIG. 1) at high loading conditions, and the stability of the prior art circuits with throttling valves (not shown) at low loading conditions. Furthermore, the present design does not require additional electronic control, which is desirable in industrial settings. Instead, the valve 32 is pilot-operated through the same pilot lines that actuate the POCVs in order to dampen the undesirable responses in the regions of interest. The valve also throttles the flow in the transmission line when the two pilot pressures are close to each other, but allow free flow in and out of the actuator when the two pilot pressures are not close to each other and throttling is unnecessary. This embodiment thus uses the two POCVs as its charging-control valves, and its shuttle valve 32 as a singular pilot-operated vibration damping valve.

(43) FIG. 21 schematically illustrates the dual-piloted selective-throttling valve 32 employed in the fourth embodiment of FIG. 7. In the illustrated example, the valve is a spool valve in which an internal spool member 200 is linearly displaceable back and forth on a longitudinal axis of an outer sleeve-shaped housing 202 in which two flow connection ports 204a, 204b open radially into the housing in alignment with one another at diametrically opposing points of the housing near an axial center thereof. Pilot ports 205a, 205b open into the housing at longitudinally opposing ends thereof and feed into respective chambers defined between the ends of the displaceable spool member and the respective ends of the housing. Each chamber, the respective pilot port, and the respective end of the spool thus define a respective one of the pilot inputs 32a, 32b, at which the respective end of the spool defines the piston area of this pilot input. Springs 34a, 34b each reside between one end of the displaceable spool member and a respective end of the housing to bias the spool into the centered position, where a central land 206 of the displaceable spool member forms a flow-blocking portion of the spool closing off the two flow connection ports 204a, 204b to define the normally closed state of the valve. The flow-blocking central land 206 is neighboured by two flow-enabling valleys 207 on opposing sides thereof to define two flow-enabling portions of the spool.

(44) With continued reference to FIG. 21, each flow connection port has a non-uniform cross section having a narrow portion of smaller cross-sectional area intersecting the exterior of the housing and a wider portion of larger cross-sectional area intersecting the interior of the housing. The wider portion of this stepped-width port structure spans a shorter axial length of the connection port (i.e. radial thickness of the housing walls) than the smaller diameter portion of the connection port. The central land 206 of the displaceable spool member 200 is wide enough to fully span the wider portion of each connection port at the interior of the housing wall, thus fully closing off the two flow connection ports from one another.

(45) When the pilot pressure in one of the pilot inputs 32a, 32b of the FIG. 21 valve is high enough to overcome the bias of the respective spring 34a, 34b at the opposing end of the valve, the shifting of the spool 200 toward the opposing end of the housing 202 starts to open up the two flow connection ports 204a, 204b by moving the flow-blocking central land 206 out of alignment between the flow connection ports and shifting the neigbouring flow-enabling valley 207 into place between the flow connection ports. During this initial movement, flow through each connection port 204a, 204b is restricted to a path moving around the central land of the spool via a small axial flow path travelling axially of the housing and delimited between the outer periphery 206a of the central land and the shoulder or step 208 created at the transition between the two differently-sized portions of the port, and a small radial flow path opening into the respective flow-enabling valley 207 that is moving into place between the widened inner ends of the connection ports 204a, 204b. As the pilot pressure increases and more of the flow-enabling valley 207 moves into the space between the connection ports 204a, 204b, the radial flow path increases in size while the axial flow path remains constant, until the flow enabling-valley 207 reaches the space between the narrowed outer ends of the connection ports 204a, 204b.

(46) At this point, the fluid is no longer limited to a flow path around the central land 206 via the constricted axial-flow path, as direct radial flow straight through the narrower outer portion of each port is now also allowed. As the flow-enabling valley 207 of the spool moves into full alignment between the connection ports, the overall available flow area thus now increases at a greater rate, as more and more area of the narrower outer portions of the flow connections points are opened by movement of the flow-blocking land fully out from between the connection ports. In the fully shifted position of the spool, the respective flow-enabling valley 207 spans the full width of the widened inner ends of the connection ports, thus maximizing the available flow area to enable unthrottled free flow through the valve. Outer flow-blocking lands 214a, 214b at the opposing ends of the spool seal off the flow-enabling valleys 207 and the connection ports 204a, 204b from the pilot inputs 32a, 32b at the ends of the housing. Accordingly, the flow through the valve is only throttled during initial displacement of the spool at low loading conditions of the hydraulic circuit, until central flow-obstructing land if the displaceable spool 206 clears the respective shoulder 208 of each stepped-width connection ports.

(47) FIG. 8 shows a fifth embodiment circuit which employs the same selective-throttling operation principle as the fourth embodiment, but uses readily available off-the-shelf parts in place of the unique valve 32 to provide similar selective-throttling effect. In the FIG. 8 implementation, first and second counterbalance valves CBV.sub.A, CBV.sub.B are instead installed in the first and second main fluid lines L.sub.A, L.sub.B, respectively, near the connections to the extension and retraction sides of the actuator 12 to serve as the embodiment's two pilot-operated vibration-damping valves, while two POCVs serve as the embodiment's two charging-control valves. Generally, CBVs are throttling valves typically used for safety requirements through the whole working range actuator operation. They have been used in some pump-controlled applications [6, 24, 15, 29], but with no ability to regenerate energy [24]. Here, the CBVs are utilized to only restrict flow at low loading conditions to enhance the performance while allowing free flow at high loading conditions to allow energy regeneration. CBV.sub.A, is operable by pressure at a respective pilot input port 32a fed by a cross pilot line 38a connected to the second main fluid line L.sub.B, while CBV.sub.B is operable by pressure at a respective pilot input port 32b fed by a cross pilot line 38b connected to the first main fluid line L.sub.A. In addition to the cross pilot line from the opposing main fluid line, the pilot input of each CBV is also fed by a respective pilot path from the same main fluid line on which the valve is installed, from a point situated on the actuator-side of the valve. This is shown in the figure by pilot path 36a of CBV.sub.A and pilot path 36b of CBV.sub.B.

(48) Each CBV is normally closed, and is only opened on the presence of the sufficient pilot pressure from either or both of its pilot sources 36a, 32a/36b, 32b. In its initial stages of opening, each CBV is only partially opened, and has a reduced flow area relative to the respective main fluid line, thus throttling the fluid passing through it. However, as the respective pilot pressure increases due to the rising pressure at the other main fluid line, the CBV opens further, exposing an unrestricted flow area allowing free, unthrottled flow therethrough. So like the pilot-controlled spool and sleeve valve 32 of FIG. 7, the CBV only throttles at low loading values, thus limiting throttling primarily, if not entirely, to the critical zones shifted down to such lower loading ranges in the operational performance map. In brief, this embodiment employs a singular charge pressure source and two POCVs and two counterbalance valves (CBVs) for limited throttling. Compared to the prior art, this design reduces the throttling margin and saves energy, while providing more flexibility, including use of separate settings for each CBV to deal with the two different regions of undesirable performance.

(49) FIG. 8 shows the circuit during load-resisting extension of the actuator in a pumping-mode of the reversible pump 10 (Quadrant 1, FIG. 3), where the check-valve equipped bypass 40a of CBV.sub.A allows pumped fluid from the reversible pump 10 to freely flow in an unthrottled manner to the extension side of the actuator, while the check-valve equipped bypass 40b of CBV.sub.B prevents the fluid exiting the retraction side of the actuator from bypassing CBV.sub.B, which due to the pilot pressure provided from first main fluid line L.sub.A through cross pilot line 38b is opened initially into a throttling position, and eventually into a free-flowing state as the pilot pressure increases. During load-assisting extension of the actuator in a motoring-mode of the reversible pump (Quadrant 2, FIG. 3), where the check-valve equipped bypass 40a of CBV.sub.A allows output fluid from the motoring reversible pump 10 to again flow freely in an unthrottled manner to the extension side of the actuator, while the check-valve equipped bypass 40b of CBVs prevents the fluid exiting the retraction side of the actuator from bypassing CBV.sub.B, which due to the pilot pressure in pilot path 36b is opened initially into a throttling position, and eventually into a free-flowing state as the pilot pressure increases.

(50) During load-resisting retraction of the actuator in a pumping-mode of the reversible pump (Quadrant 3, FIG. 3), the check-valve equipped bypass 40b of CBVs allows pumped fluid from the reversible pump 10 to flow freely in an unthrottled manner to the retraction side of the actuator, while the check-valve equipped bypass 40a of CBV.sub.A prevents the fluid exiting the extension side of the actuator from bypassing CBV.sub.A, which due to the pilot pressure in cross pilot line 38a is opened initially into a throttling position, and eventually into a free-flowing state as the pilot pressure increases. Finally, during load-assisting retraction of the actuator in a motoring-mode of the reversible pump (Quadrant 4, FIG. 3), the check-valve equipped bypass 40b of CBVs allows output fluid from the motoring reversible pump 10 to flow freely in an unthrottled manner to the retraction side of the actuator, while the check-valve equipped bypass 40a of CBV.sub.A prevents the fluid exiting the extension side of the actuator from bypassing CBV.sub.A, which due to the pilot pressure in the pilot path 36a is opened initially into a throttling position, and eventually into a free-flowing state as the pilot pressure increases.

(51) In addition to the described throttling at low loading conditions in each quadrant by one of the two CBVs, FIG. 8 employs the same use of two POCVs fed by a singular charge pressure to accommodate the differential flow across the actuator, as described above in relation to FIG. 1, unlike the FIG. 7 circuit which uses two different charge pressures for the respective POCVs to shift the critical loading zones to lower loading ranges. The charging system in FIG. 8 is denoted solely by accumulator 20, with the remainder of the charging system, including the charge pump 16, omitted for illustrative simplicity. The two CBVs are thus set such that the throttling occurs up to the upper limit of the unshifted critical zone, beyond which the CBV fully opens to a non-throttling condition.

(52) FIG. 8A shows a variant of the FIG. 8 circuit, which employs the same use of two CBVs to perform select throttling only below the upper loading limits of the critical loading zones, but includes the FIG. 7 arrangement of two different charging pressures respectively applied to the two POCVs. This way, the shifting of the critical load value and surrounding critical loading zone to a lower range of load values means that the upper limit of the critical loading zone at which the CBV switches from throttled to unthrottled operation is lower, whereby throttling is performed over a lesser overall fraction of the total operating area, thus improving the efficiency of the circuit.

(53) FIG. 9 shows a sixth embodiment circuit, which employs both concepts of centering the critical zones and throttling the flow only in the shifted critical zones. This embodiment replaces each POCV of the first embodiment with a respective 2-way single-pilot select-throttling valve 42a, 42b that serves both as a charging-control and vibration damping valve. Like the 2-way dual-pilot select-throttling valve of FIG. 7, each single-pilot throttling valve 42a, 42b has a controllable variable flow area that increases at a first rate during initial displacement, before increasing more rapidly under further displacement. However, displacement out of the normal default position is only possible in one direction. The first throttling valve 42a has a single pilot input 32a at one end thereof, actuation of which is resisted by a respective spring 34a at the opposing end thereof. The second throttling valve 42b likewise has a single pilot input 32b at one end thereof, actuation of which is resisted by a respective spring 34b at the opposing end thereof. The pilot input 32a of the first throttling valve 42a is fed by a cross-pilot line 26 from the second main fluid line L.sub.B, while the pilot input 32b of the second throttling valve 42b is fed by a cross-pilot line 28 from the first main fluid line L.sub.A. The first throttling valve 42a is connected between the first charging line 22 and the lower pressure side of the dual-pressure charging system 14, while the second throttling valve 42b is connected between the second charging line 24 and the higher pressure side of the dual-pressure charging system 14. Each selective-throttling valve 42a, 42b is a normally closed valve that closes off the charging system from the respective charging line in the default valve position, but then initially throttles the fluid passing therethrough during the initial portion of its displacement due to the low flow-area opened therein, and then allows unthrottled flow during later stages of displacement due to the larger flow-area opened up therein. As with the other selective-throttling embodiments, each valve is set so that the free-flow state is achieved once the critical zone has been cleared, whereby throttling only occurs at low loading conditions below the upper limit of the critical zone, which is shifted toward center due to the use of two different charging pressures for the two valves 42a, 42b. This embodiment is more efficient than the fourth embodiment, as it only restricts the differential flow (i.e. the flow passing through the charging lines), which is only around 25% of the main flow. Consequently, this reduces the energy losses due to throttling, and reduces the number of components and complexity of the circuit required to accomplish both critical zone shifting and vibration damping within the shifted critical zone.

(54) FIG. 10 shows a seventh embodiment that like the sixth embodiment accomplishes both critical zone shifting functionality and selective-throttling functionality within the shifted critical zones using only a single set of off-the-shelf valves, which in this case are sequence valves 44a, 44b that serve as both charge-control valves and vibration-damping valves. The first sequence valve 44a is operated by a first cross pilot line 26 connected to the second main fluid line L.sub.B, while the second sequence valve 44b is operated by a second cross pilot line 28 connected to the first main fluid line L.sub.A. The resulting effect is similar that of the sixth embodiment, wherein the normally closed sequence valve normally closes off the respective charging line from the charging system, and throttles the fluid only during an initial part of its opening stroke before fully opening its through-path to enable free unthrottled flow between the charging system and the respective charging line. Once again, only the differential flow in the charging lines is throttled, not the main flow in the main lines L.sub.A, L.sub.B.

(55) Finally, FIG. 11 shows an eight embodiment employing a singular pilot-operated check valve POCV.sub.A installed between the first charging line 22 and the lower pressure side of the of the dual-pressure charging system 14 to serve as one of the embodiments two charging control valves, and a singular sequence valve 44b between the second charging line 24 and the higher pressure side of the dual-pressure charging system 14 to serve as both the other charging-control valve and the vibration-damping valve. The POCV and the sequence valve 44b are respectively operated by cross pilot lines 26, 28, whereby the circuit once again provides both critical zone shifting and selective-throttling functionality.

(56) Each of the forgoing embodiment uses valves that are exclusively pilot-operated (requiring no electronic monitoring and control components) not only to perform the acceptable switching necessary to accommodate differential flow to and from a single rod actuator (i.e. switching between a first circuit-charging state enabling flow through the first circuit-charging line between the first main fluid line and the charging circuit, and a second circuit-charging state enabling flow through the second circuit-charging line between the second main fluid line and the charging circuit), but also to use one or more varying characteristics (applied charge source, piston area, spring constant) between the two respective valve-actuating inputs such that the critical load value and associated range at which problematic operation would otherwise occur is shifted toward the center of the four quadrant operational map along the load force axis thereof. Select embodiments additionally or alternatively employ one or more valves in the main lines or charging lines that are again exclusively pilot-operated (requiring no electronic monitoring and control components) to provide selective throttling only below the upper limits of the critical loading zones, while allowing more efficient throttle-less flow in the larger operational areas outside the critical loading zones. In each case, four-quadrant operation is fully retained whereby motoring of the pump in two quadrants can be used for regeneration purposes for optimal efficiency.

(57) FIG. 12 shows a test rig constructed for this study and its schematic drawing. The test rig was a John Deere backhoe attachment (JD-48) equipped with a variable displacement pump unit, a charge pressure unit and instrumentations. It was designed to facilitate the implementation of different hydraulic actuation circuits.

(58) In testing the fifth embodiment circuit of FIG. 8, different loading conditions were applied to the stick actuator and responses were obtained at different velocities in each of the four quadrants. Experimental results showed good performance when pump runs only in single mode of operation away from the switching regions shown in FIG. 3.

(59) FIG. 13 shows the results categorized based on quality of performance and plotted on the F.sub.L-v.sub.a plane. Each vertical set of points in the figure represents different actuator velocities for one load value. Areas hatched with dashed lines are regions where the pump switches mode of operation during actuator extension and retraction. Operation in these regions using the prior art exhibits deteriorated performance. FIG. 14 shows prior art circuit performance covering two regions. The experiment was done for a load of 2.54 kN during extension (v.sub.a=5 cm/s) and retraction (v.sub.a=9 cm/s). As it is seen the second portion illustrates the circuit performance at oscillatory retraction. These experimental results validate the discussion presented in the earlier background.

(60) A first experiment using the FIG. 8 circuit was designed to demonstrate performance improvements at low loading conditions. A second set of tests was performed to show the circuit performance and energy consumption during operation spanning all four quadrants. FIG. 15 shows the performance in a typical retractionextension of actuator with constant load (similar to test shown in FIG. 14). Actuator velocity and pressure graphs show that the circuit response is non-oscillatory.

(61) In the second set of experiments, the load of 0.4 kN was applied to the full setup shown in FIG. 12. The experiments were repeated for both the inventive FIG. 8 circuit and the prior art FIG. 1 circuit that utilizes the POCVs. The wave square control signal input (FIG. 16) was applied to the pump to move the stick link carrying the external load of 0.4 kN.

(62) Results for both circuits are shown in FIGS. 17 to 19. It is clear that the prior art FIG. 1 circuit with the POCVs exhibits oscillation during switching from assistive to resistive loading modes in actuator retraction. The oscillatory response is shown clearly in velocity plot. Results also show that performance of the proposed circuit is smooth without any significant oscillation during switching modes.

(63) The inventive FIG. 8 circuit, however, consumes more energy than the prior art FIG. 1 circuit with only POCVs as shown in FIG. 19. The delivered hydraulic energy from the pump to the circuit is calculated as the multiplication of pressure differential across the pump by the flow rate, W.sub.pmh=(p.sub.ap.sub.b)Q. Q was calculated by multiplying the actuator measured velocity and the piston effective area. Results showed that both circuits consume energy when load is resistive and recuperate energy when load is assistive. For this experiment, the average delivered hydraulic energy from the pump to the circuit was 17.1 W for the prior art FIG. 1 circuit that utilizes only POCVs and was 36 W for the inventive FIG. 8 circuit. The average received (recuperated) hydraulic energy from the circuit to the pump are 7.2 W and 2.9 W for the prior art FIG. 1 circuit that utilizes only POCVs and the inventive FIG. 8 circuit, respectively. The extra energy consumed by the inventive FIG. 8 circuit was used to overcome the hydraulic resistance generated by the CBVs to stabilize the system. Note that, the extra needed energy reduces as the load increases.

(64) Comparison was also made of the energy consumed by the inventive FIG. 8 circuit to a valve-controlled circuit. Considering a valve-controlled hydraulic system is equipped with a pressure compensated pump, the pump energy consumption equals to the nominal pump pressure multiplied by the flow rate. Knowing that the maximum pressure value in the experiment shown in FIGS. 16, 17 and 18 is 8 MPa, the pump nominal pressure was set in the valve-controlled circuit at 8 MPa. The average consumed hydraulic energy by the pump in a valve-controlled circuit performing the same task as in FIG. 19 is 1081.8 W. Thus the inventive FIG. 8 circuit consumed only 8.9% of energy needed by a comparable valve-controlled circuit to deliver the same amount of hydraulic energy to the actuator, and at the same time produces a performance better than at least the prior art of FIG. 1.

(65) Since various modifications can be made in the invention as herein above described, and many apparently widely different embodiments of same made within the scope of the claims without departure from such scope, it is intended that all matter contained in the accompanying specification shall be interpreted as illustrative only and not in a limiting sense.

REFERENCES

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