Abstract:
In a hydraulic system that has first and second pumps, a pressure compensation circuit is provided to unload the output of the second pump when its operation is not required. The unloading reduces the power demanded of the engine that drives the pumps thus conserving energy. The first pump is directly connected to a supply line and a back flow check valve couples the second pump to the supply line. When pressure in the supply line is significantly greater than the load pressure of the actuators powered by the hydraulic fluid a bypass compensator valve opens to provide a path between the outlet of the second pump and the tank line. This action unloads the second pump and reduces its demand for engine power.

Description:
CROSS-REFERENCE TO RELATED APPLICATIONS  
       [0001]     Not Applicable  
       STATEMENT REGARDING FEDERALLY SPONSORED RESEARCH OR DEVELOPMENT  
       [0002]     Not Applicable  
       BACKGROUND OF THE INVENTION  
       [0003]     1. Field of the Invention  
         [0004]     The present invention relates to hydraulic systems which have multiple pumps connected to a common supply line, and particularly to mechanisms for unloading fluid supplied by one of the pumps when the output of that pump is not required.  
         [0005]     2. Description of the Related Art  
         [0006]     Numerous types of machines have members which are moved by a hydraulic system. Specifically, a member is driven by an actuator, such as a hydraulic cylinder and piston arrangement, that receives pressurized fluid via a proportional control valve. The control valve is opened varying degrees to proportionally control the rate the fluid flows to or from the associated actuator, thereby moving the machine member at different speeds, as desired by the user.  
         [0007]     It is common practice on a tractor loader/backhoe and similar machines to have two fixed displacement hydraulic pumps driven on a common shaft by the machine&#39;s engine. In many cases, one of the pumps has its working pressure reduced under certain circumstances in order that the hydraulic horsepower does not exceed the horsepower available from the engine or transmission system. This action, known as “pump unloading”, can be set to occur upon a given event, usually at a certain level of pressure in the main system. Hence this device can be controlled by a relief valve, which directs the pump output flow back to the reservoir at lower pressure.  
         [0008]     Hydraulic systems often control the unloading by producing a load-sense signal, which indicates the greatest load applied to the different hydraulic services on the machine. In a system with fixed displacement pumps, the load-sense signal operates a variable relief valve or bypass compensator that opens flow path from the pumps to system reservoir. This single compensator valve maintains the combined pump pressure a fixed amount above the load-sense pressure as determined by a spring force acting on that valve. This pressure difference is often referred to as the “margin.” If the flow required at the service ports of the control valves is greater than or equal to the combined capacity of the pumps, then the unloading path from the pump to tank is closed off. At this point the margin decays below the level set by the spring and is dependent upon the size of the opening which is presented to the downstream pump.  
         [0009]     The present inventor has recognized that for optimal engine power savings, it is desirable to provide independent unloading for each pump in a dual pump system and have one pump be subordinate to the other.  
       SUMMARY OF THE INVENTION  
       [0010]     A pressure compensation circuit is provided for a hydraulic system that controls flow of fluid to at least one hydraulic service connected to a supply line and a return line. The supply line is fed fluid from a primary pump and a secondary pump that is coupled to the supply line by a backflow prevention check valve. A load-sense circuit senses the load pressure at each hydraulic service.  
         [0011]     The pressure compensation circuit comprises a first bypass compensator valve that selectively provides a path between the supply line and the return line when pressure in the supply line is greater than pressure in the load-sense circuit by at least a first amount. A second bypass compensator valve selectively provides a path between an outlet of the secondary pump and the return line when pressure in the supply line is greater than pressure in the load-sense circuit by at least a second amount. The second amount is less than the first amount so that the second bypass compensator valve opens under lower pressure in the supply line than the first bypass compensator valve.  
         [0012]     One type of hydraulic system has a load-sense circuit that produces a pressure on a load-sense line corresponding to a greatest load among all of the hydraulic services. For this system, the pressure compensation-circuit includes a first orifice coupling the load-sense line to a first node and a second orifice coupling the load-sense line to a second node. A first bypass compensator valve selectively provides a path between the supply line and the return line in response to pressure in the supply line being the first amount greater than pressure at the first node. A second bypass compensator valve selectively provides a path between the second outlet of the secondary pump and the return line in response to pressure in the supply line being the second amount greater than pressure at the second node. The second bypass compensator valve opens before the first bypass compensator valve.  
         [0013]     Another type of hydraulic system has a load-sense circuit in which pressure in a first load-sense line indicates the load at one hydraulic service and pressure in a second load-sense line indicates the load pressure at another hydraulic service. For this system, the pressure compensation circuit includes a first orifice coupling the first load-sense line to a first node and a second orifice coupling the second load-sense line to a second node. A third orifice is connected between the first and second load-sense lines. A first bypass compensator valve selectively provides a path between the supply line and the return line in response to pressure in the supply line being greater than pressure at the first node. A second bypass compensator valve selectively provides a path between the outlet of the secondary pump and the return line in response to pressure in the supply line being greater than pressure at the second node. 
     
    
     BRIEF DESCRIPTION OF THE DRAWINGS  
       [0014]      FIG. 1  is a schematic diagram of a hydraulic system according to the present invention;  
         [0015]      FIG. 2  is cross sectional view through an assembly of pressure compensation components of the hydraulic system; and  
         [0016]      FIG. 3  is a schematic diagram of a second hydraulic system according to the present invention. 
     
    
     DETAILED DESCRIPTION OF THE INVENTION  
       [0017]     With initial reference to  FIG. 1 , a hydraulic system  10  has a primary pump with an output connected to a supply line  14  and a secondary pump  16  having an output coupled to the supply line by a backflow check valve  18 . Both pumps  12  and  16  are fixed displacement types being driven by the engine of an off-highway vehicle, for example. The supply line conveys pressurized hydraulic fluid to several services, or hydraulic functions,  20  and  22  on the machine. One service  22  has a first hydraulic cylinder  24  that moves a member on the machine and the other service  24  includes a second hydraulic cylinder  26  that drives a different machine member.  
         [0018]     The flow of hydraulic fluid to and from the first cylinder  24  is proportionally metered by a first directional control valve  28 . This three-position (or four-position) valve selectively connects the supply line  14  to one chamber of the first cylinder and connects the other cylinder chamber to a return line  30  the leads to the reservoir, or tank,  32  of the hydraulic system. Which one of the chambers of the first cylinder  24  receives the pressurized fluid determines the direction that the piston  34  in the cylinder moves and thus the direction of motion of the associated member of the machine. The flow of hydraulic fluid to and from the second hydraulic cylinder  26  is metered in a similar manner by a second directional control valve  36  that also is connected to the supply line  14  and the return line  30 .  
         [0019]     Each of the first and second directional control valves  28  and  36  has a load-sense port  29  and  37 , respectively, the pressure at which corresponds to the load pressure from the associated hydraulic cylinder  24  and  26 . These load-sense ports are connected to a conventional shuttle valve  38  which selectively applies the greater of those load pressures to a load-sense line  40  as a load sense pressure. In a more complex machine, the load-sense ports from other services are connected by cascaded shuttle valves to the load-sense line  40 .  
         [0020]     The load-sense line  40  is coupled to a pressure compensation circuit  41 . Specifically, a first orifice  42  couples the load-sense line  40  to a first node  44 . A spool or poppet type, first bypass compensator valve  48  controls a fluid path between the supply line  14  and the return line  30  in response to pressures at the first node  44  and the supply line. As will be described, the first bypass compensator valve  48  is biased closed by a spring  49  and opens when the supply line pressure is greater than the combined force of that spring and the pressure at the first node  44 . A load-sense pressure relief valve  46  connects the first node  44  to the return line  30  to relieve excessively high pressure from acting on the first bypass compensator valve  48 .  
         [0021]     The load-sense line  40  also is coupled by a second orifice  50  to a second node  52 . A spool or poppet type, second bypass compensator valve  54  controls a fluid path between the outlet  56  of the secondary pump  16  and the return line  30  in response to pressure at the second node  52  and the supply line pressure. As will be described, the second bypass compensator valve  54  is biased closed by a spring  55  and opens when the supply line pressure from both pumps is greater that the combined force of that spring and the pressure at the second node  52 . The biasing springs  49  and  55  of the bypass compensator valves are different, with the first bypass compensator valve  48  having a higher spring force than the second bypass compensator valve  54 . Thus the second bypass compensator valve  54  opens at a lower pressure differential than the first bypass compensator valve  48 . An unloader relief valve  58  connects the second node  52  to the return line  30  to relieve excessively high pressure from acting on the second bypass compensator valve  54 .  
         [0022]     The operation of the pressure compensation circuit  41  can be understood by first assuming that the pumps  12  and  16  are running and neither service  20  or  22  is active, so there is no load-sense pressure signal from the directional control valves  28  and  36 . As a result, the pump pressure in the supply line  14  acts on the first bypass compensator valve  48  against the force of the spring  49  thereby pushing the valve spool into an open position. The degree to which the first bypass compensator valve  48  opens is dependent upon a number of factors, including the characteristics of the valve&#39;s metering notches and the spring force. Hence the pump output flow into the supply line  14  passes to tank  32  at a pressure related to the spring  49  and metering notches of the first bypass compensator valve  48 . This pressure in ‘the supply ’ line  14  also is sensed by the spool in the second bypass compensator valve  54 , which pushes that valve&#39;s spool to open a relatively large path for the second pump output to pass to the tank  32 . Hence, the output flow from the secondary pump  16  passes to the tank at a lower pressure than the output flow from the primary pump  12 . The check valve  18  prevents fluid in the supply line  14  from flowing through the open second bypass compensator valve  54 .  
         [0023]     When one or both of the directional control valves  28  and  36  is operated, a load-sense pressure is generated in line  40  and acts on the spring ends of the spools in both bypass compensator valves  48  and  54 . In response, the first bypass compensator valve  48  closes down in order for the primary pump  12  to generate an output pressure equal to the load-sense pressure plus the effect of the compensator spring  49 . In this case, the first bypass compensator valve  48  fixes the margin, provided that flow to the active service is less than the capacity of the primary pump  12 . The fluid flow passing to the tank  32  via the first bypass compensator valve  48  is equal to the flow from the pumps minus the flow passing to the service(s)  20  and  22 . The pressure at the output of the primary pump  12  is sensed at the non-spring end of the second bypass compensator valve  54  as before, but in order for its spool to be in equilibrium with a lighter spring force than for the first bypass compensator valve  48 , the spool of the second bypass compensator valve  54  moves to a position determined by the margin and its spring  55 . Hence, the second bypass compensator valve  54  is again in a position where the output flow from the secondary pump  16  passes to the tank  32  through a relatively large valve orifice. Thus the output of the secondary pump  16  is maintained at a relatively low pressure. Under these circumstances, the power required from the tractor engine is lower than would normally be required if both pumps  12  and  16  were connected in a more conventional manner.  
         [0024]     As the flow required by the services  20  and  22  increases towards the maximum available from the primary pump  12 , the engine horsepower savings is reduced. The load-sense relief valve  46  sets a maximum load-sense pressure in the load-sense line  40 . This limit of load-sense pressure sets a corresponding limit on the system pressure and the first bypass compensator valve  48  behaves as a relief valve for the primary pump  12 .  
         [0025]     As the size of the metering orifice in one or both of the first and second directional control valves  28  and  36  increases, the flow to the services  20  and  22  increases proportionately. At the point where the required flow is equal to the capacity of the primary pump  12 , the first bypass compensator valve  46  is fully closed. Due to the nature of the spring rate (or slope) characteristic of the first bypass compensator valve  46 , the effective margin has reduced. The spring  55  of second bypass compensator valve  54  is arranged so that at a pre-determined point, such as when the first bypass compensator valve  48  closes, the second bypass compensator valve begins to raise appreciable pressure. As a result, at least a portion of the flow from the secondary pump  16  enters the supply line  14  via the check valve  18 . The pressure difference between the pump delivery and the load-sense pressure is now being maintained by the second bypass compensator valve  54 .  
         [0026]     If the load-sense pressure in line  40  reaches a level dictated by the unloader relief valve  58 , then any increase in the load-sense pressure from the directional control valves  28  and  36  no longer is met with a corresponding increase in pump pressure. The unloader relief valve  58  has a pressure-flow characteristic with a steep slope. Therefore, any increase in load-sense pressure above the level set at the unloader relief valve  58  results in a disproportionately lower increase in pressure at the spring end of the second bypass compensator valve  54 . This effect is related to the slope of the relief valve characteristic, and the size of the orifice between the load-sense line and that relief valve. Hence, the second bypass compensator valve  54  is pushed towards the open position and the secondary pump  16  is gradually unloaded to a low pressure. This function can also be achieved manually by activating a solenoid operated relief valve  59  to relieve the load-sense pressure acting on the second bypass compensator.  
         [0027]     Referring to  FIG. 1 , it is possible to use the first bypass compensator valve  48  to “time” the application of the load-sense pressure signal to the second bypass compensator valve  54 . In this case, the second bypass compensator valve  54  sees only full supply line pressure, which acts on the non-spring end of its spool, and thus remain fully open until a delayed application of the load-sense pressure to the spring end. Hence up to the point of load-sense pressure application to the second bypass compensator valve  54 , the second pump  16  experiences virtually no pressure at its output  56  and exerts minimal load on the tractor engine. The power savings are more pronounced even where the fluid flow to the services  20  and  22  approaches the limit of the capacity of the first pump  12 .  
         [0028]     A further embodiment of the pressure compensation circuit  60  is similar to that in  FIG. 1  and is shown in  FIG. 3 , with the common components being assigned the same reference numerals. The load-sense line  62  from the first directional control valve  28  is connected directly to the pressure compensation circuit  60  and then via the first orifice  42 . The load-sense line  64  from the second directional control valve  36  of the second service  22  is connected directly to a third node  65  that is between the second orifice  50  and the second node  52 . A third orifice  68  is provided between the second and third nodes  52  and  65 . A check valve  66  is connected in parallel with the orifice  68 , allowing free flow from node  65  to node to node  62 .  
         [0029]     By applying the load-sense pressure from the first directional control valve  28  to the second bypass compensator valve  54  via the third orifice  68  in additional to the second orifice  50 , it is possible to modify the characteristic of the unloading function of the second bypass compensator valve  54 . For example, that unloading function operates at a lower load-sense pressure from the first directional control valve  28  as compared&#39;to the load-sense pressure from the second directional control valve  36 , where a greater service load may occur.  
         [0030]     This is achieved because the maximum load-sense pressure at node  52  whilst activating control valve  22  is set by the flow passing across the orifice  50  but the maximum load-sense pressure at node  52  whilst activating control valve  20  is set by the flow passing across the orifices  50  and  68  in series. In the latter case less flow passes across the relief valve  58  and because this relief valve has a steep pressure rise characteristic it&#39;s effective setting is lower. Hence the second bypass compensator valve  54  unloads the second pump  16  at the lower level when control valve  20  is in use compared with control valve  22 . Hence the power requirement when using control valve  22  is less than when using control valve  20 .  
         [0031]     The foregoing description was primarily directed to a preferred embodiment of the invention. Although some attention was given to various alternatives within the scope of the invention, it is anticipated that one skilled in the art will likely realize additional alternatives that are now apparent from disclosure of embodiments of the invention. Accordingly, the scope of the invention should be determined from the following claims and not limited by the above disclosure.