MULTIPLE PUMP SYSTEM WITH HORSEPOWER LIMITING CONTROL

Abstract: Flow-pressure compensated valves are employed in the servo-systems for variable displacement pumps to maintain pump discharge pressure above a minimum pressure level and also above the load pressure in a fluid actuator. The utilization of multiple pumps in a particular fluid circuit, each functioning to charge a separate fluid actuator, dictates the need for closely controlling the displacements of the pumps simultaneously to obtain maximum performance efficiency from the prime mover for the pumps. It is further desirable to provide a control system which minimizes the total pump envelope size. The improved fluid circuit (10) of this invention includes a summing valve (22) for creating a control signal (P<uC>u) in response to the collective pump pressure signals (P<uP>u) in a plurality of fluid actuators (13) and a control arrangement (36) for controlling modulation of flow-pressure compensated valves of the above type in response to variations in the control signal (P<uC>u).

Description: Description ?:ultlple Pump System with Horsepower Limiting Control Technical Field This invention relates generally to a fluid circuit having multiple pumps and means for automatically controlling the displacement of the pumps and more particularly to summing means for creating a control signal in response to the collective load pressures in fluid actuators, and control means responsive to the control signal for closely modulating a load pressure signal used to automatically vary pump displacement. Background Art It is well known in the art to employ a flow- pressure compensated or "load-plus" valve to maintain the discharge pressure of a variable displacement pump above a minimum pressure level and also above a load pressure generated in a fluid actuator. This type of valve is fully disclosed in U.S. Patent No. ,116,587, issued on September 26, 197δ to Kenneth P. Liesener, and assigned to the assignee of this application. The valve functions to sense load pressure and to automatically actuate a swash-plate of the pump in response to the load pressure to maintain the pump at its desired displacement. ' when multiple pumps are employed in a fluid circuit, such as when a separate actuator is connected to each pump for performing work on a construction vehicle, there is a need for closely controlling the displacements of the pumps simultaneously to obtain maximum performance efficiency from the prime mover OMPI wvI iPOu- for the pumps. In particular, it is desirable to provide a control system for modulating the load pressure signals communicated to the "load-plus" valves employed in the servo-systems for the pumps during all phases of operation of the pumps. In providing such a control system, it is further desirable to provide a relatively non-complex control system which minimizes total pump envelope sizes. The present invention is directed to overcoming one or more of the problems as set forth above. Disclosure of Invention In one aspect of the present invention, fluid circuit has a plurality of actuators and a variable displacement pump connected to each of the actuators, first biasing means for urging a control member of each pump towards a first displacement position, second biasing means for urging the control meinber towards a second displacement position in response to a load pressure signal received from a respective one of the fluid actuators, and modulating means for modulating the load pressure signal in the second biasing means to vary the displacement of the pump in response to the magnitude of the load pressure signal and the position of the control member. The improved fluid circuit further comprises summing means for creating a control signal in response to the collective load pressure signals prevalent in the fluid actuators and control means for controlling the modulating means in response to variations in the control signal. The improved fluid circuit will thus provide maximum performance efficiency from the prime mover O PI | for the pumps, such as an internal combustion engine, by continuously communicating the control signal to the modulating means or servo-systems for the pumps. The control circuit is torque-limiting and ensures that the actuators will be supplied with the required fluid pressures within a range wherein both pumps are operating under fully loaded conditions and wherein only one of the pumps is so operating. Furthermore, the above summing means and control means aid in minimizing total pump envelope sizes which provides obvious benefits in 'both monetary costs, servicing, and operation. Brief Description of the Drawings Other objects and advantages of this invention will become apparent from the following description and accompanying drawings wherein: Figure 1 schematically illustrates a fluid circuit having a pair of variable displacement pumps, each associated with a fluid actuator and incorporating a summing and control means embodiment of the present invention therein to control displacement of the pumps in response to the collective load pressures of the actuators; Figure 2 is a longitudinal sectional view through one of the pumps and the control system therefor; Figure 3 is an enlarged sectional view of a modulating or horsepower limiting valve employed in the control system; Figure 4 graphically illustrates curves A and A 1 , plotting pump flow versus load pressure, and a horsepower curve H Figure 5 is a view similar to Figure 2, but OMPI WIPO illustrates a modification of the control system; and Figure 6 is an enlarged sectional view of a modulating or horsepower limiting valve employed in the Figure 5 control system. Best Mode of Carrying Out the Invention Figure 1 illustrates a fluid circuit 10 comprising a pair of variable displacement pumps 11, each adapted to communicate pressurized fluid from a source 12 to a fluid actuator 13 under the control of a directional control valve 14. A prime mover 15, such as an internal combustion engine, is adapted to drive pumps 11, with each pump preferably taking the form of a hydraulic pump of the type illustrated in Figure 2. Each fluid actuator 13 may comprise a double-acting hydraulic cylinder, for example, adapted for use on a construction vehicle or the like in a conventional manner. Upon selective actuation of a respective directional control valve 14, head and rod ends of a connected actuator 13 may be alternately pressurized and exhausted in a conventional manner via lines 16 and 17 and lines 18 and 19. Upon presεurization of one of the ends of a selected actuator 13, a line 20 will communicate a pump pressure signal ? p to an actuating chamber 21 of a summing valve 22. As described more fully hereinafter, summing valve 22 provides a summing means for creating a control pressure signal ?„ in a line 23 in response to collective pump pressure signals P to control the actuation of servo-systems 24 employed for pumps 11. Control pressure signal ? c is generated by an engine-driven pump 25 which is connected to summing valve 22 by a line 26. As shown in Figure 1, when the averaged pump pressure signals P prevalent in actuating chambers 21 exceed a predetermined level, a spring-biased spool 27 of the summing valve will shift leftwardly to throttle and meter fluid pressure in a 5 controlled and modulated manner from line 26 to line 23 to create control pressure signal P in the latter line. The magnitude of control pressure signal P„ is closely controlled by restricted orifice 28 in a drain line 29 connected to fluid source or tank 12. ]_0 Figure 1 further illustrates a line 30 interconnected between each directional control valve 14 and a respective servo-system 24 for communicating load pressure signal P τ to the servo-system upon pressurization of the head or rod end of a respective l ~ cylinder 13. Referring to Figures 2 and 3, line 30 communicates load pressure signal P, through an orifice 31 mounted in a housing 32 of servo-system 24 and into a passage 33 defined in the housing. As described more fully hereinafter, passage 33 2o communicates load pressure signal P L to a flow- pressure compensated or "load-plus" valve 34 for maintaining pump discharge pressure at a specified level above the required load pressure. As further described more fully hereinafter, load pressure signal P L is modulated by a modulating means or horsepower limiting valve 35 under the control of a control means 36 which controls the modulating means in response to variations in control pressure signal P_ in line 23. Still referring to Figures 2 and 3, pump 11 0 comprises a barrel 37 which is adapted to be driven by an output shaft 38 of engine 15, a plurality of reciprocal pistons 39, and a control member or swash plate 40 having the pistons connected thereto in a conventional manner. The displacement of pump 11 is determined by the rotational orientation of swash plate 40 which has one side thereof interconnected between first and second biasing means 41 and 42, respectively. In the position shown in Figure 2, swash plate 40 will effect maximum pump displacement whereas horizontal orientation of the swash plate will effect zero or minimum displacement of the pump. First biasing means 41 is shown in the form of a compression coil spring 43 interconnected between housing 32 and swash plate 40 to bias the swash plate towards its illustrated maximum pump displacement position. Second biasing means 42 may be considered to include valve 34, which functions substantially identically to the corresponding valve disclosed in above-referenced U.S. Patent No. 4,116,587- Pump discharge pressure in a main discharge passage 44 communicates with a branch passage 45 which, in turn, communicates with an actuating chamber 46 via an annulus 47 and a passage 48 defined in housing 32. Actuating chamber 46 is defined in a tubular member 49, secured within housing 32, and the force generated by fluid pressure in the chamber will tend to urge swash plate 40 counterclockwise in Figure 2 tovrards its maximum displacement position through a piston 50 and a rod 51 interconnected between the piston and swash plate. Piston 50, rod 51, and a second rod 52 may be considered to comprise a follow- up linkage 53 for purposes hereinafter explained. It should be noted that since the force generated in actuating chamber 46 to urge swash plate counterclockwise in Figure 2 and the force imposed on the follow-up linkage by a hereinafter described spring means 79 are additive to the force of spring _C H_ 43, that they may be considered to collectively comprise first biasing means 41. As shown in Figure 2, a second branch passage 54 communicates pump discharge passage 44 to an annulus 55 surrounding valve 34. Valve 34 includes a spool 56 having lands 57, 58, and 59 formed thereon to define annuluses 60 and 61 about the spool. Spool 56 is slidably mounted in a bore 62 defined in a tubular member 63 secured within housing 32 with the bore being blocked at its lower end by a threaded plug 64. An actuating chamber 65 is thus defined between reciprocal spool 56 and plug 64 and the spool has an extension 66 secured thereon to limit the spool's downward movement, as shown in Figure 2. An actuating chamber 67 is also defined between plug 64 and a piston 68 which is attached to a rod 69 which is further attached to swash plate 40. In the illustrated position of spool 56, pressurized fluid communicated to branch passage 54 is communicated to actuating chamber 65 via annulus 55, a port 70, and a restricted orifice 71 formed in the port. Such communication will shift spool 56 upwardly in Figure 2 under certain operating conditions and against the opposed biasing force of a compression coil spring 72 and the force generated by load pressure signal P τ Li' in an actuating chamber 73 of valve 34. Land 58 of spool 56 will thus uncover a passage 74 to communicate pump discharge pressure in branch passage 54 to actuating chamber 67 via annulus 55, a port 75, annulus 61, and passage 74. Pressurization of chamber 67 will function to rotate swash plate 40 clockwise in Figure 2, against the opposed biasing forces of spring 43, the force of spring means 79, and the fluid pressure prevalent in chamber 46, to destroke the pump by moving the swash plate towards its minimum displacement position of operation. It should be noted that a drain passage 76 is formed in member 63 for venting chamber 67 upon downward movement of spool 56 from Its position shown in Figure 2. The function of "load-plus" valve 34 is more fully described in above-referenced U.S. Patent No. 4,116,587. Referring to Figures 2 and 3, modulating means or horsepower limiting valve 35 includes a first spool 77 slidably mounted in a second spool 78,- the latter spool being slidably mounted in member 49• Spools 77 and 78 are urged downwardly by two-stage biasing means 79, including a pair of compression coil springs 80 and 81. The lower end of spring 80 seats on a retainer 82, which overlies member 49 and spools 77 and 78, whereas the lower end of spring 81 is spaced from the retainer for purposes hereinafter explained. As discussed more fully hereinafter, springs oO and 81 provide that modulating means 35 will cause the horsepower consumption curve to closely match that of engine horsepower curve H, shown in Figure 4. Load pressure signal P L In line 30 communicates with passage 33 via an annulus 83 which further communicates such fluid pressure, through orifice 31, to an annulus or chamber 84 defined about spool 77, and between a pair of lands 85 and 86 of the spool, via ports 87 and 88. Upward movement of spool 77 relative to spool 78 will place one or more modulating slots 89, formed on spool 77, in communication with a drain passage 90 for modulating load pressure signal P L in passage 33 and in actuating chamber 73 of "load-plus" valve 34 by venting fluid pressure from chamber 84. When spool 77 is in its illustrated closed position, it engages an upper end of a rod 91, secured on a plug 92 threaded into a lower end of spool 78. Rod 52 of the follow-up linkage engages the underside of plug 92, whereby upon 10 clockwise rotation of swash plate 40, the follow-up linkage will move spools 77 and 78 upwardly simultaneously against the opposed force of biasing means 79. As suggested above, control means 36 l ~ functions to control modulating means 35 and, in particular, the reciprocal position of spool 77 within spool 78, to modulate load pressure signal P L in response to variations in control pressure signal P-, in line 23. Control pressure signal P,, communicates 2o with an actuating or control chamber 93, defined in spool 78, via an annulus 94, a port 95, an annulus 96, and a port 97. It can thus be seen in Figure 3 that when the control pressure signal ?- exceeds a predetermined level in chamber 93 that spool 77 will 5 move upwardly to modulate the load pressure in chamber 73 of "load-plus" valve 34 through modulating slots 89. As shown in Figure 2, a valve 9δ functions as a pressure relief valve to prevent the level of fluid pressure in chamber 73 from exceeding a predetermined 0 maximum. Figures 5 and 6 Illustrate a modified servo- system 24a which functions substantially similar to above-described servo-system 24. Identical numerals ' H A . OMPI depict corresponding constructions and arrangements with numerals depicting modified constructions and arrangements in Figures 5 and 6 being accompanied by an As shown in Figure 5, spring means 79 and pressurization of an actuating chamber 67a will tend to urge swash plate 40 of pump 11 towards its maximum displacement position against the opposed biasing force of a first biasing means 41a comprising a compression coil spring 43a, connected to the opposite side of the swash plate. A second biasing means 42a may be considered to comprise expansible actuating chamber 67a and biasing means 79. In this respect, servo-system 24a is different than servo-system 24 in that the pressurization of corresponding actuating chamber 67 of the latter system will tend to move swash plate 40 thereof towards its minimum displacement position and against the opposing biasing force of spring 43 (Figure 2). Pressurization of chamber 67a in Figure 5 is effected by communicating pump discharge pressure from outlet 44 and through passages 54a, 54a', and 74a, in response to pressurization of an actuating chamber 73a of "load-plus" valve 34a. Chamber 73a is connected to line 30 to receive load pressure signal P τ therein across orifice 31a, whereby increased fluid pressures in the chamber will move a rod 72a' downwardly to compress springs 72a against the upper end of a spool 56a of valve 34a to connect lines 54a and 54a' . When the differential fluid pressure in an actuating chamber 65a of valve 34a exceeds a predetermined level, greater than the combined forces of springs 72a and 72b and the fluid pressure in chamber 73a, spool : * MPI 56a will move upwardly to block communication between passages 54a and 54a ? and to vent chamber 67a via passages 74a and 54a', and vent passages 76a' and 76a. Load pressure signal P L is modulated in chamber 73a via passages 33a which connect the chamber with an annulus or chamber 84a defined around a spool 77a of a modulating means or horsepower limiting valve 35a, as shown in Figure 6. Line 33a communicates load pressure signal P τ Li to chamber 84a via an annulus 83a and a port 87a formed through a second spool 78a. A plurality of modulating slots 89a are formed on spool 77a to modulate load pressure signal P τ in chamber 84 by opening the slots to drain passage 76a in response to variations in control pressure signal P„ in line 23. Control pressure signal P.. is communicated to an actuating chamber 93a disposed within spool 78a and a -a lower end of spool 77a, via ports 95a and 97a. When the fluid pressure in chamber 93a exceeds a predetermined level, spool 77a will .move upwardly relative to spool 7δa to at least partially open modulating slots 89a, against the opposed biasing force of above-described biasing means 79- A rod 52a is interconnected between piston 68 ar.d spool 78a to maintain follow-up under the force of biasing means 79 and to urge the spool in the opposite direction under the force of spring 43a via swashplate 40 (Figure 5). Indus-rial Applicability Fluid circuit 1C finds particular application to hydraulic circuits for construction vehicles and the like wherein close and efficient control of fluid actuators or cylinders 13 is required. In this respect, the fluid circuit utilizes pressure compensation in conjunction with a displacement follower which through control pressure signal P r will change the null point pressure along a constant horsepower envelope. This invention provides for Instant and correct sensing and response to system energy consumption or demand over a wide temperature range. This invention also provides for a minimum total pump envelope size, as exemplified in Figure 5 which suggests that "load-plus" valve 34a may be conveniently located at any desired position in the pump envelope, such as horizontally along the top of. the head of pump 11 whereat it is convenient for connection to line 30 and passages 33a, 54a, and 54a' . Referring to Figures 1-4, "load-plus" valve 34 will function as a conventional pressure- compensated flow control valve operating in a normal manner throughout the working range of its associated pump 11 to provide a load-sensitive control of pump discharge pressure in line 19, . relative to load pressure signalP L by continuously providing a margin between these pressures, as described In above- referenced U.S. Patent No. 4,116,587. Summing valve 22 is arranged to receive pump pressure signals P via lines 20 to determine the actual load or horsepower drain on engine 15 to, in turn, create and modulate control pressure signal P p in line 23 for controlling pump displacement. In particular, when the summed pump pressure signals P_ are equal to or are less than a predetermined pressure level, spool 27 will remain in its closed position illustrated in Figure 1 to prevent communication of pump 25 with line 23. Thus, control chamber 93 will remain vented via drain line 29 to prevent upward shifting of spool 77 against -*ΌREA OMPI V.'IPO spring means 79. Thus, as long as the pumps are operating in their normal range of working pressures, fluid circuit 10 will remain under full control of "load-plus" valves 34, associated with pumps 11. During this period of operation in the normal working pressure range, both modulating means 35 will remain inactivated and will not affect the operation of the pumps. Under operating conditions in which pumps 11 are consuming all of the available horsepower from engine 15, the summed pump pressure signals P in lines 20 will exceed a predetermined level to shift spool 27 leftwardly to communicate pump 25 with line 23- Throttled and modulated control pressure signal ? thus communicates with chamber 93 to shift spool 77 upwardly against the opposed modulating force of spring means 79 to at least partially open modulating slots 8 to vent a controlled amount of load pressure signal P L to drain passage 90. The resultant reduction in fluid pressure in chamber 73 of valve 3 will thus permit upward shifting of spool 56 to pressurize chamber 67 in the manner described above to rotate swash plate 40 towards its minimum displacement position. This motion of the swash plate will feed- back to modulating means 35 via rods 51 and 52 to move sleeve-like spool 78 upwardly relative to spool 77 to a reset position closing-off modulating slots 89. As the reduction In pump displacement reduces the horsepower consumption from the engine, spool 27 of summing valve 22 will maintain a position therein respective of the system pressure to thus modulate control pressure signal P_, in line 23. The pumps will continue to operate at their restaged * ^ 5 ?*E OMPI .. viro displacement settings until such time as the summed pump pressure signals P exceed a level whereby the horsepower consumption exceeds that available from the engine. Upon attaining this condition of operation, control pressure signal P c will be Increased ir. control chamber 93 and modulating means 35 will again function in the manner described above to further reduce pump displacement and thus closely control the total horsepower consumption from engine 15. Conversely, reduction in summed pump pressure signals P will permit the displacements of pumps 11 to P increase by permitting the swash plates thereof to move back towards their maximum displacement positions, illustrated in Figure 2. It should be further noted in Figure 2 that the dual-spring arrangement comprising springs 80 and 81 provides that modulating 1"* means 3 will cause the horsepower consumption curve to closely match that of the engine horsepower curve shown at "H" in Figure 4. In particular, when spool 77 moves upwardly against spring 80 to open modulating slots 89, pump flew or displacement will drop as depicted at point A, on curve A. It should be further noted that when spring retainer 82 moves sufficiently upwardly In Figure 2 to engage second spring 81, that a restaging effect will occur at point A~. Curve A depicts a pressure trace upon opening of both directional control valves 14 to actuate cylinders 13 simultaneously, whereas curve A' depicts the opening of only one valve to operate only one of the cylinders. As discussed above, modified servo-system 24a of Figures 5 and 6 functions- substantially identically to servo-system 24. As discussed above, _OMPI ' - Figures 5 and 6 teach tha ' t flow-pressure compensated valve 34a may be located at any convenient position in the pump package, such as along the top of the head of pump 11. Other aspects, objects, and advantages of this invention can be obtained from a study of the drawings, the disclosure, and the appended claims.

Claims: Clai s 1. In a fluid circuit (10) having a plurality of actuators (13), a variable displacement pump (11) connected to each of said actuators (13), and including a control member (40) movable between first and second displacement positions, first biasing means (41) for urging said control member (40) towards its first displacement position, second biasing means (42) for urging said control member (40) towards its second displacement position in opposition to said first biasing means (41) and in response to a load pressure signal communicated thereto from a respective one of said fluid actuators (13), and modulating means (35) for modulating said load pressure signal in said second biasing means (42) to vary the displacement of said pump (11) in response to the magnitude of said • load pressure signal and the position of said control member (40), the improvement comprising: summing means (22) for creating a control signal in response to the collective pump pressure signals in said fluid actuators (13) and control means (36) for controlling said modulating means (35) in response to variations in said control signal. 2. The fluid circuit (10) of claim 1 wherein said summing means (22) includes a summing valve having spool means (27) for movement between a closed position when said collective pump pressure signals fall below a predetermined level and a throttling position modulating said control signal to said modulating means (35) when said collective pump pressure signals exceed said level. 3. The fluid circuit (10) of claim 2 wherein said modulating means (35) includes a housing (32), a first spool (77), a second spool (78) reciprocally mounted in said housing (32) and having said first spool (77) reciprocally mounted therein, and means (89) for modulating said load pressure signal in response to relative reciprocation between said first (77) and second (78) spools. 4. The fluid circuit (10) of claim 3 wherein said means for modulating (89) is formed on said first spool (77). 5- The fluid circuit (10) of claim 3 wherein said control means (36) includes a chamber (93) defined in said second spool (78) and exposed to a lower end of said first spool (77), and means (95) for communicating said control signal to said chamber (93).

Inventors: RUSEFF W, SCHIMPF J
Applicants: CATERPILLAR TRACTOR CO, RUSEFF W, SCHIMPF J
Application year: 1980
Main classification: F16H39
IPC codes: F04B1, F04B49, F16H61