Patent Publication Number: US-2002007631-A1

Title: Pressure compensating valve, unloading pressure control valve and hydraulically operated device

Description:
FIELD OF THE INVENTION  
       [0001] The present invention relates to a pressure compensating valve, an unloading pressure control valve, and a hydraulically operated device.  
       DESCRIPTION OF THE RELATED ART  
       [0002]FIG. 25 depicts a hydraulically operated device described in Japanese Unexamined-Patent Application 1-247805.  
       [0003] In this hydraulically operated device, a variable delivery pump A is connected to a low pressure hydraulic cylinder D via a pressure compensating valve B and a directional control valve (operating valve) C. The pump A is also connected to a high pressure cylinder D′ via a pressure compensating valve B′ and a directional control valve C′.  
       [0004] An actuator E for changing the displacement volume and a flow regulating valve F for controlling the actuator E are attached to the hydraulic pump A.  
       [0005] The higher load pressure among the load pressures that are produced during the operation of the cylinders D and DT is sensed by a shuttle valve G as the maximum load pressure P LS , and this maximum load pressure P LS  is output as the pilot pressure to the flow regulating valve F.  
       [0006] The flow regulating valve F controls the actuator E so that the discharge pressure P P  of the pump A is always greater than the maximum load pressure P LS  The cylinders D and D′ are jointly operated by the simultaneous operation of directional control valves C and C′ in the hydraulically operated device. At this time, the pressure compensating valve B controls the amount of oil supplied to the cylinder D so that the difference between the input pressure and the output pressure of the directional control valve C is constant, and the pressure compensating valve B′ similarly controls the amount of oil supplied to the cylinder D′ so that the difference between the input pressure and the output pressure of the directional control valve C′ is constant.  
       [0007] The hydraulically operated device equipped with the pressure compensating valves B and B′ can prevent the disadvantage of pressured oil accumulating and being supplied to the cylinder with the lighter load among the operating valve cylinders D and D′.  
       [0008] According to the aforementioned Japanese Unexamined Patent Application 1-247805, however, the discharge pressure of pump A decreases upon the supply of large amounts of pressured oil to the hydraulic cylinder D with the lower pressure during periods of considerable control input to the directional control valves C and C′. In such cases, the pressure difference before and after the pressure compensating valve B fails to reach the compensated pressure difference, and the pressure compensating valve B thus fails to achieve pressure compensation. That is, the pressure compensating valve B remains open.  
       [0009] While the pressure compensating valve B fails to achieve pressure compensation, the amount of pressured oil supplied to the hydraulic cylinder D with the lower pressure is uncontrolled, so no pressured oil is supplied to the hydraulic cylinder D′ with the higher pressure, and the hydraulic cylinder D′ with the higher pressure is thus not operated. The operator must then operate the directional control valve C in the slightly open direction to control the flow rate to the hydraulic cylinder D with the lower pressure.  
       [0010] To prevent such a situation from developing, the aforementioned hydraulically operated device is provided with a pressure difference sensing device H for sensing the pressure difference P P -P LS  between the pressure P P  of the pressured oil discharged from the hydraulic pump A and the maximum load pressure P LS , a control force set device I for setting the control force fc based on the pressure difference P P -P LS  and the relationship depicted in FIG. 26, and an electromagnetic valve J that is operated by means of the output signals from the control force setting device I.  
       [0011] The control force fc is given by the following equation.  
         fc=f−a ( P   P   −P   LS  )  
       [0012] Where  
       [0013] f: the pressing force of springs b and b −  in pressure compensating valves B and B′ 
       [0014] α: constant  
       [0015] The electromagnetic valve J allows pressured oil corresponding to the control force fc to act on the pressure receiving components of the pressure compensating valves B and B′ when the pressure difference P P -P LS  is at or below the specific pressure difference Pm shown in FIG. 26.  
       [0016] This allows the control force fc against the pressing force f of the aforementioned springs b and b −  to be exerted on the springs in the pressure compensating valves B and B′. The force fc increases the discharge pressure of the pump A by increasing the flow resistance of the pressure compensating valves B and B′, allowing pressured oil to be supplied to the hydraulic cylinder D′ with the higher pressure too.  
       [0017] When the cylinders D and D′ are cylinders that operate an operating device in construction machinery (such as a hydraulic shovel boom, arm, or bucket), the pressure compensation characteristics of the pressure compensating valves B and B′ are preferably modified in some cases to improve the operating characteristics, depending on the operating configuration of the aforementioned operating device.  
       [0018] A technique that is capable of changing the throttle levels for each pressure compensating valve and that is capable of suitably changing the pressure difference before and after the directional control valves C and C′ has been disclosed in the aforementioned patent publication. That is, in this technique, electromagnetic valves J as described above are provided for the pressure compensating valves B and B′, and the control force fc for the pressure compensating valves B and B′ are individually adjusted by these electromagnetic valves J. Accordingly, the throttle levels of the pressure compensating valves B and B′ are individually changed; that is, the pressure differences before and after the direction control valves C and C′ are different from each other.  
       [0019] A state in which the required flow rate is distributed completely irrespective of load is also referred to in particular as a fully compensated state.  
       [0020] The conventional devices described above suffer from the following drawbacks, however.  
       [0021] In some cases, pressure compensation is not possible when the mechanism for producing control force fc to modify the pressure compensation characteristics malfunctions. Furthermore, the electromagnetic valves J are operated by computations after the pressure difference has been sensed by a pressure difference detector  21 H, resulting in poor response.  
       [0022] In view of the foregoing, a first object of the present invention is to provide a pressure compensating valve that allows the pressure compensation characteristics to be arbitrarily modified, that has good response, and that is highly reliable.  
       [0023]FIG. 27 depicts a hydraulically operated device described in Japanese Unexamined Patent Application 4-250226. When the operating device A in this hydraulically operated device is operated, a flow regulating valve (operating valve) B is operated, by means of the pilot pressure produced by the operating device A, to an extent corresponding to the extent to which the operating device A has been operated, and the discharged pressured oil from a hydraulic pump D is consequently supplied to a hydraulic cylinder (hydraulic actuator) C.  
       [0024] A pressure compensating valve E for keeping the pressure difference before and after the flow regulating valve B at a constant level is located between the hydraulic pump D and the flow regulating valve (operating valve) B.  
       [0025] An operating device A′, flow regulating valve (operating valve) B′, hydraulic motor (hydraulic actuator) C′, and pressure compensating valve E′ each correspond to the operating device A, flow regulating valve (operating valve) B, hydraulic cylinder C, and pressure compensating valve E.  
       [0026] An unloading pressure control valve F is connected in parallel to the hydraulic pump D. The higher pressure between the load pressure acting on the hydraulic cylinder C and the load pressure acting on the hydraulic motor C′ is sensed as the maximum load pressure by a shuttle valve G, and this maximum load pressure is allowed to act on the unloading pressure control valve F.  
       [0027] The unloading pressure control valve F is provided to return the discharged oil from the hydraulic pump D to the tank. The amount of the aforementioned discharged oil returned by the unloading pressure control valve F is set by the difference between the maximum load pressure and the discharge pressure of the hydraulic pump D, and by control signals output from a control unit J.  
       [0028] A computer H connected to the control unit J computes the difference ΔP LS  between the discharge pressure of the hydraulic pump D and the load pressure of the hydraulic cylinder C or hydraulic motor C′ based on the functional relation shown in FIG. 28 and the output of sensors a 1 , a 2  and a 1 ′, a 2 ′ for sensing the control input of the operating devices A and A′.  
       [0029] The function shown in FIG. 28 defines a relation in which the pressure difference ΔP LS  increases proportionally until the control input St of the operating device A reaches a set value, and the pressure difference ΔP LS  stays at a value ΔP LS 1 when the control input St is at or beyond the set value.  
       [0030] When the control input St is 20%, for example, the pressure difference ΔP LS  is computed by the computer H, so a control signal corresponding to a pressure difference ΔP LS2  is output from the control unit J, and the unloading start pressure of the unloading pressure control valve F is set to pressure difference ΔP LS2 . As a result, the amount of pressured oil supplied through the pressure compensating valve E′ and flow regulating valve B′ to the hydraulic motor C′ is the amount defined by the pressure difference ΔP LS2 .  
       [0031]FIG. 29 shows the relation between the amount of oil Q supplied to the hydraulic motor C′ and the pressure difference AP before and after the flow regulating valve B′ when the control input St is 20%.  
       [0032] As shown in FIG. 29, the pressure compensating valve E′ supplies pressured oil in a constant oil amount q 2  to the hydraulic motor C′ so that the pressure difference AP of the flow regulating valve B′ is kept at a constant pressure difference ΔPc+ΔP LS  (ΔP LS  is the pressure loss of the pressure compensating valve E′). However, while the pressure difference ΔP has not yet reached the constant pressure difference ΔPc+ΔP LS  (compensated pressure difference), the pressured oil is supplied to the hydraulic motor C′ in the oil amount q 1  defined by the unloading start pressure ΔP LS2  of the unloading pressure control valve F.  
       [0033] Thus, according to this hydraulically operated device, when the control input of the operating device A is set to about 20% for moderate acceleration of the hydraulic motor C′, the amount of oil supplied to the hydraulic motor C′ is limited to the amount of oil q 1  defined by the unloading start pressure ΔP LS2 , and the hydraulic motor C′ is thus moderately accelerated.  
       [0034] Furthermore, in the case of the load sensing circuit of a variable delivery pump, when the unloading start pressure of the unloading pressure control valve F is pre-modified, the amount of pressured oil discharged from the hydraulic pump D is increased in advance. The response of the hydraulic cylinder C when operated by the operating unit A is thus better.  
       [0035] The unloading start pressure of the unloading pressure control valve F is variable. However, the unloading start pressure is set through the computer H and the control unit J. It is accordingly always set after the output from the sensors a 1 , a 2  and a 1 ′, a 2 ′ of the operating devices A and A′, and a resulting problem is the poor response in terms of the hydraulic cylinder C or the hydraulic motor C′. More specifically, when the fluctuations in the load pressure of the hydraulic cylinder C or hydraulic motor C′ are estimated, the unloading start pressure is hopefully pre-modified rapidly irrespective of the control input of the operating devices A and A′. For the reasons described above, however, the unloading start pressure is difficult to modify in advance.  
       [0036] In view of the foregoing, a second object of the present invention is to provide an unloading pressure control valve allowing the unloading start pressure to be preset so as to improve the response in terms of a hydraulic actuator.  
       [0037] A pump discharge pressure control means for controlling the displacement volume of a hydraulic pump (discharge volume per revolution) is provided in a hydraulically operated device in which the pressured oil discharged from a variable delivery pump is supplied to a hydraulic actuator such as a hydraulic cylinder by the operation of an operating valve. This pump discharge pressure control means is designed so as to control the displacement volume of a hydraulic pump based on the discharge pressure of a hydraulic pump and the load pressure acting on a hydraulic actuator, so that the aforementioned discharge pressure is greater by a specific pressure than the aforementioned load pressure.  
       [0038] According to the hydraulically operated device equipped with the pump discharge pressure control means, when the load pressure is increased during the operation of the operating valve, the displacement volume of the hydraulic pump immediately increases to a magnitude corresponding to the load pressure. The actuator is also connected via a pressure compensating valve. A flow rate corresponding to the control input of the operating valve can thus be supplied, irrespective of the magnitude of the load pressure, to the actuator.  
       [0039] To be supplied at flow rate corresponding to the control input is, in other words, a matter of the action of pressure corresponding to the load. The control input of the operating valve at this time and certain actuator conditions sometimes result in rapid start up with shocks.  
       [0040] When the aforementioned hydraulic actuator is a hydraulic motor or cylinder driving an operating unit in construction machinery (such as the revolving superstructure, boom, arm, or bucket in the case of a hydraulic shovel, for example), the rapid start up of the aforementioned hydraulic actuator results in lower operating performance, depending on the operating configuration.  
       [0041] Hydraulically operated devices such as the following have been proposed in patent publications.  
       [0042] That is, in the hydraulically operated device proposed in Japanese Unexamined Patent Application 9-222101, for example, a bleed valve is connected to the discharge channel of the aforementioned hydraulic pump, and part of the pressured oil discharged by the hydraulic pump is bled through the bleed valve to the tank.  
       [0043] According to the hydraulically operated device described in this patent publication, the rapid start up of the hydraulic actuator is controlled, resulting in better operating performance.  
       [0044] However, the bleed valve used in the hydraulically operated device of the aforementioned patent publication bleeds off part of the pressured oil discharged from the hydraulic pump to the tank. In other words, a large amount of the pressured oil that is supposed to be supplied to the hydraulic actuator ends up being returned to the tank when bled off. This results in significant energy loss.  
       [0045] Other resulting problems are the need for large-scale machines because of the large amounts of pressured oil that are bled off, poor sensitivity, and difficulties in achieving high-precision control.  
       [0046] In view of the foregoing, a third object of the present invention is to provide a hydraulically operated device that allows energy loss to be minimized to control rapid start up of hydraulic actuators, and that also allows machinery to be made more compact and high-precision control to be achieved.  
       [0047] Another object of the present invention is to simultaneously achieve the first and second objects.  
       [0048] Still another object of the present invention is to simultaneously achieve the first and third objects.  
       [0049] Yet another object of the present invention is to simultaneously achieve the second and third objects.  
       [0050] And finally another object of the present invention is to simultaneously achieve the first, second, and third objects.  
       SUMMARY OF THE INVENTION  
       [0051] To achieve the first object, the first of the present inventions is a pressure compensating valve through which passes pressured oil that is fed from a hydraulic pump  1  to a hydraulic actuator  5 , characterized by comprising a main valve  20  that is operated in such a way as to increase the area of the opening between an inlet port  24  and an outlet port  25  by means of pressure acting on a first pressure receiving component  21 , that is also operated in such a way as to reduce the area of the opening by means of pressure acting on a second pressure receiving component  22  and pressure acting on a third pressure receiving component  23 , and that is designed to allow the pressure Pa of the pressured oil flowing to the inlet port  24  to act on the first pressure receiving component  21  and the pressure Pb of the load  5  driven by the pressured oil flowing from the outlet port  25  to act on the second pressure receiving component  22 ; and control pressure producing means  7 B for allowing control pressure Pe resulting from a reduction in the pressure Pa of the inlet port  24  to act on the third pressure receiving component  23 .  
       [0052] The first invention allows the desired pressure compensation characteristics to be obtained by changing the control pressure Pe because the pressure compensation characteristics are changed according to the magnitude of the control pressure Pe.  
       [0053] Because the control pressure Pe resulting from a reduction in the pressure of the inlet port  24  is allowed to act on the third pressure receiving component  23  of the main valve  20 , fluctuations in the control pressure Pe also correspond to fluctuations in the pressure of the inlet port  24 . The pressure compensation characteristics are thus unaffected by the pressure fluctuation of the inlet port  24  of the main valve  20 .  
       [0054] To achieve the second object described above, the second invention is an unloading pressure control valve for introducing discharged pressured oil from a hydraulic pump  1  to a tank according to the pressure difference between the discharge pressure P P  of the hydraulic pump  1  and the load pressure P LS  of a hydraulic actuator  5 , characterized by comprising a main valve  20  that is constructed in such a way as to operate in the communicating direction by means of the discharge pressure P P  of the hydraulic pump  1  acting on a first pressure receiving component  123 , to operate in the cut-off direction upon load pressure P LS  to a second pressure receiving component  124 , and to change the balance of the operating force in each of the directions by means of control pressure Pg acting on a third pressure receiving component  125 ; and control pressure producing means  101  for producing the control pressure Pg.  
       [0055] The second invention allows the unloading start pressure to be set by means of the control pressure Pg acting on the third pressure receiving component  125 . The control pressure Pg is produced by means of the control pressure producing means. Accordingly, the unloading start pressure can be preset by the control pressure producing means, and the amount of pressured oil discharged from the hydraulic pump  1  can be increased in advance to improve the response in terms of the hydraulic actuator  5 .  
       [0056] To achieve the third object described above, the third invention is a hydraulically operated device comprising a plurality of hydraulic actuators  5  to which pressured oil discharged from a variable delivery pump  1  is supplied via pressure compensating valves  7  and directional control valves  4 ; means for outputting pressure P LS  to a load pressure sensing passage  9  according to the maximum load pressure among the load pressures acting on the actuators; and pump discharge pressure control means for controlling the discharge pressure of the hydraulic pump  1  based on the pressure P LS ; wherein the hydraulically operated device is characterized in that a variable bleed valve  11  is located in the load pressure sensing passage  9 .  
       [0057] The third invention allows the amount discharged from the hydraulic pump  1  to be controlled by bleeding off the pressured oil in the load pressure sensing passage  9 . The amount flowing in the load pressure sensing channel  9  is generally quite low. The pump pressure is controlled according to the pressure of the load pressure sensing passage  9 , whereas the pressure of the load pressure sensing passage  9  is the pressure corresponding to the load pressure of the actuator and thus reacts exactly to the fluctuations in the load pressure of the actuator. It also reacts promptly to fluctuations in the load pressure. Energy loss can thus be minimized, and machines can be made more compact. The amount discharged from the hydraulic pump  1  can be controlled with greater precision.  
       [0058] To achieve the first and second objects described above, the fourth of the inventions is a hydraulically operated device comprising a pressure compensating valve through which passes pressured oil that is fed from a hydraulic pump  1  to a hydraulic actuator  5 ; and an unloading pressure control valve for introducing discharged pressured oil from the hydraulic pump  1  to a tank according to the pressure difference between the discharge pressure P P  of the hydraulic pump  1  and the load pressure P LS  of the hydraulic actuator  5 ; wherein the hydraulically operated device is characterized by comprising a pressure compensating valve  7  itself comprising a pressure compensated main valve  20  that is operated in such a way as to increase the area of the opening between an inlet port  24  and an outlet port  25  by means of pressure acting on a first pressure receiving component  21  for a pressure compensating valve, that is also operated in such a way as to reduce the area of the opening by means of pressure acting on a second pressure receiving component  22  for a pressure compensating valve and pressure acting on a third pressure receiving component  23  for a pressure compensating valve, and that is designed to allow the pressure Pa of the pressured oil flowing to the inlet port  24  to act on the first pressure receiving component  21  for a pressure compensating valve and the pressure Pb of the load  5  driven by the pressured oil flowing from the outlet port  25  to act on the second pressure receiving component  22  for a pressure compensating valve, and control pressure producing means  7 B for allowing control pressure Pe resulting from a reduction in the pressure Pa of the inlet port  24  to act on the third pressure receiving component  23  for a pressure compensating valve; and an unloading pressure control valve  10  itself comprising a main valve  100  for an unloading pressure control valve, that is constructed in such a way as to operate in the communicating direction by means of the discharge pressure P P  of the hydraulic pump  1  acting on a first pressure receiving component  123  for an unloading pressure control valve, to operate in the cut-off direction upon load pressure P LS  to a second pressure receiving component  124  for an unloading pressure control valve, and to change the balance of the operating force in each of the directions by means of control pressure Pg acting on a third pressure receiving component  125  for an unloading pressure control valve, and control pressure producing means  101  for producing the control pressure Pg.  
       [0059] According to the fourth invention, the pressure compensation characteristics are changed according to the magnitude of the control pressure Pe, allowing the desired pressure compensation characteristics to be obtained by changing the control pressure Pe.  
       [0060] Because the control pressure Pe resulting from a reduction in the pressure of the inlet port  24  is allowed to act on the third pressure receiving component  23  for a pressure compensating valve in the main valve  20  for a pressure compensating valve, the control pressure Pe also fluctuates according to the fluctuations in the pressure of the inlet port  24 . The pressure compensation characteristics are thus unaffected by the pressure fluctuations in the inlet port  24  of the main valve  20 .  
       [0061] Furthermore, the unloading start pressure can be set by means of the control pressure Pg acting on the third pressure receiving component  125  for an unloading pressure control valve. The control pressure Pg is produced by the control pressure producing means. The unloading start pressure can thus be preset by the control pressure producing means, and the amount of pressured oil discharged from the hydraulic pump  1  can be increased in advance to improve the response in terms of the hydraulic actuator  5 .  
       [0062] To achieve the first and third objects described above, the fifth of the inventions is a hydraulically operated device comprising a plurality of hydraulic actuators  5  to which pressured oil discharged from a variable delivery pump  1  is supplied via pressure compensating valves and directional control valves  4 ; means  8  for outputting pressure P LS  to a load pressure sensing passage  9  according to the maximum load pressure among the load pressures acting on the actuators  5 ; and pump discharge pressure control means  12  for controlling the discharge pressure of the hydraulic pump  1  based on the pressure P LS ; wherein the hydraulically operated device is characterized by comprising a pressure compensating valve  7  itself comprising a main valve  20  that is operated in such a way as to increase the area of the opening between an inlet port  24  and an outlet port  25  by means of pressure acting on a first pressure receiving component  21 , that is also operated in such a way as to reduce the area of the opening by means of pressure acting on a second pressure receiving component  22  and pressure acting on a third pressure receiving component  23 , and that is designed to allow the pressure Pa of the pressured oil flowing to the inlet port  24  to act on the first pressure receiving component  21  and the pressure Pb of the load  5  driven by the pressured oil flowing from the outlet port  25  to act on the second pressure receiving component  22 , and control pressure producing means  7 B for allowing control pressure Pe resulting from a reduction in the pressure Pa of the inlet port  24  to act on the third pressure receiving component  23 ; and a variable bleed valve  11  is located in the load pressure sensing passage  9 .  
       [0063] According to the fifth invention, the pressure compensation characteristics are changed according to the magnitude of the control pressure Pe, allowing the desired pressure compensation characteristics to be obtained by changing the control pressure Pe.  
       [0064] Because the control pressure Pe resulting from a reduction in the pressure of the inlet port  24  is allowed to act on the third pressure receiving component  23  of the main valve  20 , the control pressure Pe also fluctuates according to the fluctuations in the pressure of the inlet port  24 . The pressure compensation characteristics are thus unaffected by the pressure fluctuations in the inlet port  24  of the main valve  20 .  
       [0065] Furthermore, the amount discharged from the hydraulic pump  1  can be controlled by bleeding off the pressured oil in the load pressure sensing passage  9 . The amount flowing in the load pressure sensing channel  9  is generally quite low. The pump pressure is controlled according to the pressure of the load pressure sensing passage  9 , whereas the pressure of the load pressure sensing passage  9  is the pressure corresponding to the load pressure of the actuator and thus reacts exactly to the fluctuations in the load pressure of the actuator. It also reacts promptly to fluctuations in the load pressure. Energy loss can thus be minimized, and machines can be made more compact. The amount discharged from the hydraulic pump  1  can be controlled with greater precision.  
       [0066] To achieve the second and third objects described above, the sixth of the present inventions is a hydraulically operated device comprising a plurality of hydraulic actuators  5  to which pressured oil discharged from a variable delivery pump  1  is supplied via pressure compensating valves  7  and directional control valves  4 ; means  8  for outputting pressure P LS  to a load pressure sensing passage  9  according to the maximum load pressure among the load pressures acting on the actuators  5 ; pump discharge pressure control means  12  for controlling the discharge pressure of the hydraulic pump  1  based on the pressure P LS ; and an unloading pressure control valve for introducing discharged pressured oil from the hydraulic pump  1  to a tank according to the pressure difference between the discharge pressure P P  of the variable delivery pump  1  and the load pressure P LS  of the hydraulic actuators  5 ; wherein the hydraulically operated device is characterized by comprising an unloading pressure control valve  10  itself comprising a main valve  100  that is constructed in such a way as to operate in the communicating direction by means of the discharge pressure P P  of the hydraulic pump  1  acting on a first pressure receiving component  123 , to operate in the cut-off direction upon load pressure P LS  to a second pressure receiving component  124 , and to change the balance of the operating force in each of the directions by means of control pressure Pg acting on a third pressure receiving component  125 , and control pressure producing means  101  for producing the control pressure Pg; and a variable bleed valve  11  is located in the load pressure sensing passage  9 .  
       [0067] According to the sixth invention, the unloading start pressure can be set by means of the control pressure Pg acting on the third pressure receiving component  25 . The control pressure Pg is produced by means of the control pressure producing means. Accordingly, the unloading start pressure can be preset by the control pressure producing means, and the amount of pressured oil discharged from the hydraulic pump  1  can be increased in advance to improve the response in terms of the hydraulic actuator  5 .  
       [0068] The amount discharged from the hydraulic pump  1  can be controlled by bleeding off the pressured oil in the load pressure sensing passage  9 . The amount flowing in the load pressure sensing channel  9  is generally quite low. The pump pressure is controlled according to the pressure of the load pressure sensing passage  9 , whereas the pressure of the load pressure sensing passage  9  is the pressure corresponding to the load pressure of the actuator and thus reacts exactly to the fluctuations in the load pressure of the actuator. It also reacts promptly to fluctuations in the load pressure. Energy loss can thus be minimized, and machines can be made more compact. The amount discharged from the hydraulic pump  1  can be controlled with greater precision.  
       [0069] To achieve the first, second, and third objects described above, the seventh of the present inventions is a hydraulically operated device comprising a plurality of hydraulic actuators  5  to which pressured oil discharged from a variable delivery pump  1  is supplied via pressure compensating valves and directional control valves  4 ; means  8  for outputting pressure P LS  to a load pressure sensing passage  9  according to the maximum load pressure among the load pressures acting on the actuators  5 ; pump discharge pressure control means  12  for controlling the discharge pressure of the variable delivery pump  1  based on the pressure P LS ; and an unloading pressure control valve for introducing discharged pressured oil from the variable delivery pump  1  to a tank according to the pressure difference between the discharge pressure P P  of the variable delivery pump  1  and the load pressure P LS  of the hydraulic actuators  5 ; wherein the hydraulically operated device is characterized by comprising a pressure compensating valve  7  itself comprising a pressure compensated main valve  20  for a pressure compensating valve, that is operated in such a way as to increase the area of the opening between an inlet port  24  and an outlet port  25  by means of pressure acting on a first pressure receiving component  21  for a pressure compensating valve, that is also operated in such a way as to reduce the area of the opening by means of pressure acting on a second pressure receiving component  22  for a pressure compensating valve and pressure acting on a third pressure receiving component  23  for a pressure compensating valve, and that is designed to allow the pressure Pa of the pressured oil flowing to the inlet port  24  to act on the first pressure receiving component  21  for a pressure compensating valve and the pressure Pb of the load  5  driven by the pressured oil flowing from the outlet port  25  to act on the second pressure receiving component  22  for a pressure compensating valve, and control pressure producing means  7 B for allowing control pressure Pe resulting from a reduction in the pressure Pa of the inlet port  24  to act on the third pressure receiving component  23  for a pressure compensating valve; and an unloading pressure control valve  10  itself comprising a main valve  100  for an unloading pressure control valve, that is constructed in such a way as to operate in the communicating direction by means of the discharge pressure P P  of the hydraulic pump  1  acting on a first pressure receiving component  123  for an unloading pressure control valve, to operate in the cut-off direction upon load pressure P LS  to a second pressure receiving component  124  for an unloading pressure control valve, and to change the balance of the operating force in each of the directions by means of control pressure Pg acting on a third pressure receiving component  125  for an unloading pressure control valve, and control pressure producing means  101  for producing the control pressure Pg; and a variable bleed valve  11  is located in the load pressure sensing passage  9 .  
       [0070] According to the seventh invention, the pressure compensation characteristics are changed according to the magnitude of the control pressure Pe, allowing the desired pressure compensation characteristics to be obtained by changing the control pressure Pe.  
       [0071] Because the control pressure Pe resulting from a reduction in the pressure of the inlet port  24  is allowed to act on the third pressure receiving component  23  for a pressure compensating valve in the main valve  20  for a pressure compensating valve, the control pressure Pe also fluctuates according to fluctuations in the pressure of the inlet port  24 . The pressure compensation characteristics are thus unaffected by the fluctuations in the inlet port  24  of the main valve  20  for a pressure compensating valve.  
       [0072] Furthermore, the unloading start pressure can be set by means of the control pressure Pg acting on the third pressure receiving component  125  for an unloading pressure control valve. The control pressure Pg is produced by means of the control pressure producing means. Accordingly, the unloading start pressure can be preset by the control pressure producing means, and the amount of pressured oil discharged from the hydraulic pump  1  can be increased in advance to improve the response in terms of the hydraulic actuators  5 .  
       [0073] The amount discharged from the hydraulic pump  1  can be controlled by bleeding off the pressured oil in the load pressure sensing passage  9 . The amount flowing in the load pressure sensing channel  9  is generally quite low. The pump pressure is controlled according to the pressure of the load pressure sensing passage  9 , whereas the pressure of the load pressure sensing passage  9  is the pressure corresponding to the load pressure of the actuators and thus reacts exactly to the fluctuations in the load pressure of the actuators. It also reacts promptly to fluctuations in the load pressure. Energy loss can thus be minimized, and machines can be made more compact. The amount discharged from the hydraulic pump  1  can be controlled with greater precision. 
     
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS  
     [0074]FIG. 1 is a circuit diagram of the oil pressure in a hydraulically operated device relating to the present invention;  
     [0075]FIG. 2 is a circuit diagram of oil pressure, depicting the structure of a pressure compensating valve relating to the present invention;  
     [0076]FIG. 3 is a longitudinal cross section depicting the attachment of a pressure compensating valve and an operating valve relating to the present invention;  
     [0077]FIG. 4 is a longitudinal cross section, depicting the structure of a pressure compensating valve relating to the present invention;  
     [0078]FIG. 5 is a longitudinal cross section, depicting the structure of another pressure compensating valve relating to the present invention;  
     [0079]FIG. 6 is a longitudinal cross section, depicting the structure of another pressure compensating valve relating to the present invention;  
     [0080]FIG. 7 is a longitudinal cross section, depicting the structure of another pressure compensating valve relating to the present invention;  
     [0081]FIG. 8 is a longitudinal cross section, depicting the structure of another pressure compensating valve relating to the present invention;  
     [0082]FIG. 9 is a longitudinal cross section, depicting the structure of another pressure compensating valve relating to the present invention;  
     [0083]FIG. 10 is a longitudinal cross section, depicting the structure of another pressure compensating valve relating to the present invention;  
     [0084]FIG. 11 is a longitudinal cross section, depicting the structure of another pressure compensating valve relating to the present invention;  
     [0085]FIG. 12 is a longitudinal cross section, depicting the structure of another pressure compensating valve relating to the present invention;  
     [0086]FIG. 13 is a longitudinal cross section, depicting the structure of another pressure compensating valve relating to the present invention;  
     [0087]FIG. 14 is a longitudinal cross section, depicting the structure of another pressure compensating valve relating to the present invention;  
     [0088]FIG. 15 is a longitudinal cross section, depicting the structure of another pressure compensating valve relating to the present invention;  
     [0089]FIG. 16 is a circuit diagram of oil pressure, depicting the structure of an unloading pressure control valve relating to the present invention;  
     [0090]FIG. 17 is a cross section depicting a specific structure for an unloading pressure control valve relating to the present invention;  
     [0091]FIG. 18 is a cross section depicting another embodiment of an unloading pressure control valve relating to the present invention;  
     [0092]FIG. 19 is a circuit diagram of oil pressure in another hydraulic system involving the application of an unloading pressure control valve relating to the present invention;  
     [0093]FIG. 20 is a circuit diagram of oil pressure, depicting an enlargement of the structure of a variable bleed valve used in the hydraulically operated device of FIG. 1;  
     [0094]FIG. 21 depicts an embodiment with a variable bleed valve attached to the hydraulically operated device in FIG. 1;  
     [0095]FIG. 22 is a cross section of line A-A in FIG. 21;  
     [0096]FIG. 23 is a graph depicting an example of the relation between input and output when a mode set memory means has been set and stored;  
     [0097]FIG. 24 is a graph depicting another example of the relation between input and output when a mode set memory means has been set and stored;  
     [0098]FIG. 25 is a circuit diagram of oil pressure, depicting the structure of a conventional hydraulic device equipped with a pressure compensating valve;  
     [0099]FIG. 26 is a graph depicting the relation between pressure difference and control force;  
     [0100]FIG. 27 is a circuit diagram of oil pressure in a conventional hydraulically operated device in which an unloading pressure control valve is used;  
     [0101]FIG. 28 is a graph depicting the relation between the pressure difference and the control input of an operating device; and  
     [0102]FIG. 29 is a graph depicting the relation between the pressure difference before and after a flow regulating valve and the amount of oil Q supplied to a hydraulic motor. 
    
    
     DESCRIPTION OF THE EMBODIMENTS  
     [0103] Embodiments of the present invention are described in detail below with reference to the attached drawings.  
     [0104]FIG. 1 depicts an embodiment of a hydraulically operated device relating to the present invention. The hydraulically operated device can be used for a hydraulic shovel, for example.  
     [0105] The hydraulically operated device comprises a variable delivery pump  1 , auxiliary hydraulic pump  2 , a plurality of closed center operating valves (directional control valves)  4  to which the oil discharged from the hydraulic pump  1  is supplied through an oil passage  3 , and a plurality of hydraulic cylinders  5  corresponding to each operating valve  4 .  
     [0106] The head oil chambers of the hydraulic cylinders  5  are connected by means of oil passages  6   a  and pressure compensating valves  7  to the operating valves  4 , and the bottom oil chambers are connected by means of a pressure compensating valve not shown in the figure in an oil passage  6   b  to the operating valves  4 . A pressure compensating valve is in fact interposed in the oil passage  6   b , but thus pressure compensating valve has been left out in FIG. 1 to avoid complicating the drawing.  
     [0107] The load pressure P 1  of the hydraulic cylinders  5  connected thereto act on each of the oil passages  6   a . The maximum load pressure among the load pressures P 1  acting on these oil passages  6   a  are sensed as the maximum load pressure P LS  by a shuttle valve  8 , and the sensed maximum load pressure P LS  is allowed by means of an oil passage (load pressure sensing passage)  9  to act on the hydraulic pump  1 , pressure compensating valves  7 , unloading pressure control valve  10 , and variable bleed valve  11 . A fixed throttle  13  is interposed between the tank and the oil passage  9  into which the pressured oil with the maximum load pressure P LS  is introduced.  
     [0108] A pump discharge pressure control means  12  is attached to the hydraulic pump  1 . The pump discharge pressure control means  12  introduces the discharge pressure P P  of the hydraulic pump  1  and the maximum load pressure P LS ′ and controls the displacement volume of the hydraulic pump  1  so that the discharge pressure P P  is always slightly higher than the maximum load pressure P LS .  
     [0109] The structure of the pressure compensating valve  7  relating to the present invention is described below with reference to FIG. 2. The pressure compensating valve  7  is composed of a compensator  7 A, a control pressure producing component  7 B, and a pilot pressure supply component  7 C.  
     [0110] The compensator  7 A has a main valve  20 . The main valve  20  comprises a first pressure receiving component  21 , a second pressure receiving component  22 , and a third pressure receiving component  23 . The pressure Pa acting on the first pressure receiving component  21  acts in such a way as to increase the area of the opening between the inlet port  24  and outlet port  25 . The pressure Pb acting on the second pressure receiving component  22  and the pressure Pc acting on the third pressure receiving component  23  act in such a way as to reduce the area of the opening along with the elastic force of a spring  26 .  
     [0111] The inlet port  24  is connected to the outlet port of the operating valve  4  depicted in FIG. 1. The pressure Pa of the inlet port  24  acts on the first pressure receiving component  21  via an oil passage  27 . The outlet port  25  is connected to the oil passage  6   a  through a load check valve  28 .  
     [0112] A shuttle valve  29  senses the load pressure P 1  acting on the oil passage  6   a  and the greater pressure Pb among the maximum load pressure P LS , and allows the pressure Pb to act on the second pressure receiving component  22  of the main valve  20 .  
     [0113] The control pressure producing component  7 B has a variable throttle valve  30 . The-variable throttle valve  30  is operated in such a way as to reduce the area of the opening between an inlet port  32  and an outlet port  33  by means of the elastic force of a spring  31 . It is also operated in such a way as to increase the area of the opening by means of the elastic force of a spring  35  and the pilot pressure P 2  acting on a pressure receiving component  34 .  
     [0114] In ordinary cases, the spring  35  is used only in the initial fine tuning of the variable throttle valve  30 , and is not indispensable. The tank port pressure is allowed to act constantly on the spring  31  to ensure that the variable throttle valve  30  is operated more rapidly.  
     [0115] The inlet port  32  of the variable throttle valve  30  is connected to the inlet port  24  of the main valve  20  by way of an oil passage  37  equipped with a throttle  36 . The pressure Pe of the inlet port  32  of the variable throttle valve  30  acts on the third pressure receiving component  23  of the main valve  20 .  
     [0116] The outlet port  33  of the variable throttle valve  30  is connected to the oil passage  6   a  by way of an oil passage  40  equipped with a check valve  39 . The pressure Pe acting on the third pressure receiving component  23  of the main valve  20  is thus determined by the pressure Pa of the inlet port  24  of the main valve  20 , the load pressure P 1 , and the throttle levels of the throttle  36  and the variable throttle valve  30 . Pe=Pa when the variable throttle valve  30  is closed.  
     [0117] The pilot pressure Pd is given as the output pressure of an electromagnetic proportional pressure control valve  50  located in the pilot pressure supply component  7 C. The electromagnetic proportional pressure control valve  50  introduces the pressured oil discharged from the pilot hydraulic pump  2  depicted in FIG. 1 to an inlet port  52 . The pressure Pa of this pressured oil is lowered to the pilot pressure Pd by means of the electricity applied to a solenoid  53 . The pilot pressure Pd displays a magnitude proportional to the amount of electricity to the solenoid  53 .  
     [0118] When zero electricity is supplied to the solenoid  53 , the outlet port  55  communicates with the tank port  56  by means of the elastic force of a spring  54 , as shown in the figure. The pressure Pd acting on the pressure receiving component  34  of the variable throttle valve  30  is thus zero. With this, the inlet port  32  and outlet port  33  of the variable throttle valve  30  are blocked off from each other by the elastic force of the spring  31 , both sides of which are acted upon by the tank port pressure.  
     [0119] The discharge pressure of the pilot hydraulic pump  2  is held constant by constant pressure means not shown in the figure.  
     [0120] The specific structures of the operating valve  4  and pressure compensating valve  7  are described below.  
     [0121] As noted above, pressure compensating valves  7  are interposed not only in oil passages  6   a  but also in oil passages  6   b  in the hydraulically operated device in FIG. 1.  
     [0122]FIG. 3 depicts an example of the structure of an operating valve  4  by which pressured oil is selectively supplied to the two pressure compensating valves  7  described above.  
     [0123] The operating valve  4  has a structure in which a body  60  is provided with a spool  61 , pairs of left and right outlet ports  62 , pairs of left and right pump ports  63 , pairs of left and right actuator ports  64 , and pairs of left and right of tank ports  65 .  
     [0124] The spool  61  blocks all of the ports  62  through  65  in the center valve state depicted in the figure. When the spool  61  moves left from the center valve state, the outlet ports  62  on one side communicate with the pump ports  63 , and the actuator ports  64  on the other side communicate with the tank ports  65 . When the spool  61  moves right from the center valve state, the outlet ports  65  on the other side communicate with the pump ports  63 , and the actuator ports  64  on the first side communicate with the tank ports  65 .  
     [0125] The main valve  20  located in the compensator  7 A of the pressure compensating valve  7  has a valve component  66 A interposed between the actuator port  64  and the outlet port  62  of the operating valve  4 , and a pressing component  67  connected to the valve component  66 A.  
     [0126]FIG. 4 depicts an enlargement of the pressure compensating valve  7 . As shown in FIG. 4, the valve component  66  comprises a hollow component  68  open at the left end, a hole  69  open in the outer peripheral surface through the hollow component  68 , and a seat surface  71  that presses into contact with a seat  70  formed in the body  60 . Pressure-receiving surfaces  66   a  and  66   b  of the valve component  66  form the first pressure receiving component  21  of the main valve  20  depicted in FIG. 2, and the hole  69  of the valve component  66  forms the outlet port  25  of the main valve  20 . The entire valve component  66  functions as the load check valve  39  depicted in FIG. 2.  
     [0127] The pressing component  67  is positioned on an extension of the central axis of the valve component  66 , and comprises a piston  73  that slides to the left and right in a sleeve  72  fixed to the body  60 , a sliding element  74  that slides to the left and right in the piston  73 , and a spring  26  (see FIG. 2) interposed between the sleeve  72  and the sliding element  74 .  
     [0128] An annular space  75  into which the pressured oil with the maximum  1  load pressure P LS  (see FIG. 2) is introduced is formed between the body  60  and the sleeve  72 . The pressured oil with the maximum load pressure P LS  introduced into this annular space  75  flows into a stepped hole  80  in the sliding element  74  through a fine hole  76  located in the sleeve  72 , an annular groove  77 , a hole  78  located in the piston  73 , and an inlet port  79  located in the sliding element  74 , and acts on the right side of a bore  81  located in the stepped hole  80 .  
     [0129] Meanwhile, the pressured oil in the actuator port  64  of the operating valve  4  depicted in FIG. 3, that is, the pressured oil with the pressure load P 1  flowing through the oil passage  6   a , flows through an inlet port  82  located in the left end of the piston  73  and into the stepped hole  80  of the sliding element  74 , and acts on the left side of the bore  81 .  
     [0130] When the relation between the pressures P LS  and P 1  is such that P LS  is greater than P 1 , the bore  81  rotates to the left position of the outlet port  84 , and when P LS  is less than P 1 , the bore  81  rotates to the right position of the outlet port  84 .  
     [0131] As shown in FIGS. 1 and 2, P LS &lt;P 1  is a state of transition, where the pressure P LS  increases so that P LS =P 1 .  
     [0132] The stepped hole  80  communicates with a pressure chamber  83  through the outlet port  84  and a convex groove  85  located in the outer peripheral surface thereof. Thus, when P LS &gt;P 1 , the pressured oil with the maximum load pressure P LS  is introduced into the pressure chamber  83 , and when P LS &lt;P 1 , the pressured oil with load pressure P 1  is introduced into the pressure chamber  83 .  
     [0133] As described above, the stepped hole  80  and bore  81  have the function of sensing the higher oil pressure between the oil pressure P LS  and P 1 , and of guiding it into the pressure chamber  83 . The shuttle valve  29  depicted in FIG. 2 is composed of the stepped hole  80  and the bore  81 . The pressure of the pressured oil introduced into the pressure chamber  83  is pressure Pb depicted in FIG. 2.  
     [0134] The pressure chamber  83  is a space enclosed by the inner surface of the sleeve  72 , the right end surface of the piston  73 , and the outer peripheral surface of the sliding element  74 , where the right end surface of the piston  73  functions as the second pressure receiving component  22  depicted in FIG. 2.  
     [0135] The control pressure producing component  7 B is described below. The control pressure producing component  7 B is located to the side of the compensator  7 A, and is equipped with the variable throttle valve  30  depicted in FIG. 2.  
     [0136] A spool  88  for changing the flow resistance (throttle level) between the inlet port  32  and outlet port  33  depicted in FIG. 2 is located in the vertical direction in the body  87  of the variable throttle valve  30 .  
     [0137] The spool  88  is such that downwardly directed force (the direction in which the flow resistance increases) is provided by the spring  31 , and upwardly directed force (the direction in which the flow resistance decreases) is given by the spring  35  in a pressure chamber  90  formed between the spool and an adjusting screw  89 .  
     [0138] The bottom end surface of the spool  88  facing the pressure chamber  90  forms the pressure receiving component  34  depicted in FIG. 2.  
     [0139] The inner surface of a concave component  91  located in the left surface of the body  87  forms a pressure chamber  92  along with the right end surface of the sliding element  74  and the right end surface of the sleeve  72  of the compensator  7 A. The right end surface of the sliding element  74  facing the pressure chamber  92  forms the second pressure receiving component  23  of the main valve  20  depicted in FIG. 2.  
     [0140] The inlet port  32  of the variable throttle valve  30  communicates through the oil passage  37  equipped with the throttle  36  to the outlet port  62  of the operating valve  4 , that is, to the inlet port  24  of the main valve  20  depicted in FIG. 2, and also communicates-through an oil passage  38  to the pressure chamber  92 . The outlet port  33  communicates through the oil passage  40  equipped with the check valve  39  (see FIG. 2) to the actuator port  64  of the operating valve  4 .  
     [0141] The pilot pressure producing component  7 C is located in the top of the body  87  of the control pressure producing component  7 B. The electromagnetic proportional pressure control valve  50  forming the pilot pressure producing component  7 C comprises a spool  94  arranged in the vertical direction in the body  93 , and a solenoid  53  that presses the spool  94  down against the spring  54 .  
     [0142] In this electromagnetic proportional pressure control valve  50 , the spool  94  is driven down by the thrust of the solenoid  53 , allowing the flow resistance to be reduced between the inlet port  52  and the outlet port  55 .  
     [0143] The outlet port  55  communicates through an oil passage  95  to the pressure chamber  90  of the variable throttle valve  30 . The spool  94  also is positioned on the axis of the spool  88  of the control pressure producing component  7 B.  
     [0144] The operation of the pressure compensating valve  7  having the aforementioned structure is described below with reference to FIG. 4.  
     [0145] The pressured oil with the pressure Pa flowing out of the outlet port  62  of the operating valve  4  presses the valve component  66  to the right by acting on the surfaces  66   a  and  66   b  of the valve component  66  forming the first pressure receiving component  21  of the main valve  20  depicted in FIG. 2.  
     [0146] Meanwhile, the pressured oil with the load pressure Pb (pressure P 1  or P LS ) flowing into the pressure chamber  83  presses the valve component  66  to the left by acting on the right end surface of the piston  73  (second pressure receiving component  22  depicted in FIG. 2), and the pressured oil with the control pressure Pe flowing into the pressure chamber  92  presses the valve component  66  to the left by acting on the right end surface of the sliding element  74  (third pressure receiving component  23  depicted in FIG. 2). The spring  26  also presses the valve component  66  to the left by means of the sliding element  74 .  
     [0147] The pressure balance in the main valve  20  can thus be expressed as in the following Eq. (1).  
       Pa×A   0   =Pe×A   1   +Pb ( A   0   −A   1 )+ F   0   (1)  
     [0148] A 0 &gt;A 1   
     [0149] A 0 : sum of the area of surfaces  66   a  and  66   b  of valve component  66   
     [0150] A 1 : area of right end surface of sliding element  74   
     [0151] A 0 −A 1 : area of right end surface of piston  73   
     [0152] F 0 : elastic force of spring  26   
     [0153] The pressure Pe in Eq. (1) is the control pressure that changes the pressure compensation characteristics of the pressure compensating valve  7 . The control pressure Pe results in pressure Pa when the variable throttle valve  30  of control pressure producing component  7 B depicted in FIG. 2 is closed. The relation in Eq. (2) below is obtained by substituting Pa into Pe in Eq. (1).  
       Pa−Pb=F   0 /( A   0   −A 1)  (2)  
     [0154] As can be seen from this relation, the pressure compensating valve  7  is operated in such a way that the pressure difference Pa−Pb is constant when Pe=Pa. In other words, pressure compensation is achieved.  
     [0155] Thus, operating both operating valves  4  depicted in FIG. 1 to bring about the joint operation of the cylinders  5  avoids the drawback of pressured oil becoming concentrated and supplied to only the cylinder  5  with the lighter load.  
     [0156] When the variable throttle valve  30  of the control pressure producing component  7 B is not closed, the pressured oil passing through the fixed throttle  36  flows through the variable throttle valve  30  and check valve  39  to the cylinder  5  end. Thus the control pressure Pe obtained by dividing the pressure difference between the pressures Pa and P 1  by the throttling ratio between the throttle  36  and variable throttle valve  30 , in other words, the control pressure Pe resulting from the reduction of the pressure Pa, acts on the third pressure receiving component  23  of the main valve  20  in the compensator  7 A.  
     [0157] In this state, the leftward moving force of the sliding element  74  depicted in FIG. 4 is lower than when Pe=Pa.  
     [0158] Reducing the leftward moving force of the sliding element  74  is equal to lowering the elastic force Fe of the spring  26  in the Eq. (2). That is because, when the control force Pe is lower than Pa, the pressure difference Pa−Pb is set lower than when Pe=Pa (change in the pressure compensation characteristics). Here, the function of keeping the pressure difference Pa−Pb constant is still maintained, despite the change in the pressure compensation characteristics.  
     [0159] In the case of two or more cylinders with different loads, more pressured oil flows to the one with the lower load under conditions where the control input of the operating valves  4  is constant.  
     [0160] The control pressure Pe drops as the amount of electricity to the solenoid  53  of the electromagnetic proportional pressure control valve  50  increases. Accordingly, when an operating unit in construction machinery, for example (such as the boom, arm, or bucket in hydraulic shovels), is driven by cylinders  5 , pressure compensation characteristics suitable for the operating configuration of such an operating unit can be set by controlling the amount of electricity to the solenoid  53 .  
     [0161] The pressure compensation characteristics of pressure compensating valves  7  for a plurality of cylinders, as shown in FIG. 1, can also be altered, of course. The pressure compensation characteristics of the pressure compensating valves  7  for the series of cylinders  5  depicted in FIG. 3 can each be varied so as to alter the operating speeds during extension and retraction of the cylinders  5 .  
     [0162] In the pressure compensating valves  7 , the pilot pressure Pd no longer acts on the variable throttle valve  39  of the control pressure producing component  7 B in the event of wire breakage in the solenoid  53  of the electromagnetic proportional pressure control valve  50  or in the event of malfunctions of the pilot pump  2  depicted in FIG. 1, for example. In other words, the variable throttle  39  is no longer capable of throttling operations.  
     [0163] Despite such accidents, however, there is no loss of the pressure compensation characteristics of the pressure compensating valves  7 . Only a fully compensated state results.  
     [0164] That is, when the variable throttle  39  is closed, the magnitude of the control pressure Pe is changed, making it impossible to change the pressure compensation characteristics. However, since the control pressure Pe is set to Pe=Pa, it is still possible to maintain pressure compensation operations keeping the pressure difference Pa Pb shown in the Eq. (2) at a constant level.  
     [0165]FIG. 5 depicts a second example of the structure of a pressure compensating valve  7 . This pressure compensating valve  7  differs from the pressure compensating valve  7  in FIG. 4 in that the spring  54  of the electromagnetic proportional pressure control valve  50  is brought into contact with the top end of the spool  88  of the variable throttle valve  30  of the control pressure producing component  7 B.  
     [0166] In this pressure compensating valve  7 , when the spool  88  of the variable throttle valve  30  is operated based on the pilot pressure Pd supplied from the electromagnetic proportional pressure control valve  50 , the operating force is mechanically fed back to the spool  94  of the electromagnetic proportional pressure control valve  50  through the spring  54 .  
     [0167] The operating characteristics (response) of the spool  88  of the variable throttle valve  30  are improved, allowing high-precision pressure compensation to be achieved.  
     [0168]FIG. 6 depicts a third example of the structure of the pressure compensating valve  7 . This pressure compensating valve  7  is such that the variable throttle valve  30  of the control pressure producing component  7 B and the electromagnetic proportional pressure control valve  50  have a shared body  218 , with the solenoid  53  of the electromagnetic proportional pressure control valve  50  located on the exterior of the body  218 . This allows the structure to be made more compact and the number of parts to be reduced.  
     [0169] Meanwhile, the control pressure producing component  7 B in this pressure compensating valve  7  forms a flange  88   a  having a tapered peripheral surface on the spool  88  of the variable throttle valve  30 , and the flange  88   a  is interposed between the inlet port  32  and outlet port  33  of the variable throttle valve  30 .  
     [0170] When pressured oil with the pressure P 1  flows through the oil passage  40  into the outlet port  33  of the variable throttle valve  30  by means of this structure, the top surface of the flange  88   a  is placed under pressure by the pressured oil.  
     [0171] The spool  88  is thus moved down, and the tapered peripheral surface of the flange  88   a  presses against the seat surface of the body  218 , so that the inlet port  32  and outlet port  33  are blocked off from each other.  
     [0172] In this way, the spool  88  functions as a check valve to prevent the pressured oil with the pressure P 1  from flowing toward the inlet port  32 . The body  218  of this pressure compensating valve  7  thus does not require the check valve  39  depicted in FIGS. 4 and 5, making the body  218  easier to fabricate.  
     [0173]FIG. 7 depicts a fourth example of the structure of the pressure compensating valve  7 . This pressure compensating valve  7  has a structure in which a joint  102  is attached to an attachment block  219  secured to the top surface of the body  87  of the control pressure producing component  7 B, and the pressure chamber  90  of the variable throttle valve  30  in the control pressure producing component  7 B communicates through the joint  102  and piping  95  to the outlet port  55  of the electromagnetic proportional pressure control valve  50 .  
     [0174] In this pressure compensating valve  7 , the pilot pressure Pd output from the electromagnetic proportional pressure control valve  50  or the pilot pressure output from a manual pilot valve can be allowed to act on the variable throttle valve  30  of the control pressure producing component  7 B by way of the joint  102 . This pressure compensating valve  7  is thus suitable for use in cases where the electromagnetic proportional pressure control valve  50  or pilot valve must be located at a distance from the control pressure producing component  7 B because of restricted space or the like.  
     [0175] The variable throttle valve  30  of the control pressure producing component  7 B in this pressure compensating valve  7  has a structure similar to that of the variable throttle valve  30  of the pressure compensating valve  7  depicted in FIG. 4.  
     [0176]FIG. 8 depicts a fifth example of the structure of the pressure compensating valve  7 . This pressure compensating valve  7  has a structure in which a joint  104  is attached to the exterior of the body  103  of the control pressure producing component  7 B, and the pressure chamber  90  located in the variable throttle valve  30  of the control pressure producing component  7 B communicates through the joint  104  and piping  95  to the electromagnetic proportional pressure control valve  50  or a manual pilot valve not shown in the figure.  
     [0177] The electromagnetic proportional pressure control valve  50  or pilot valve of this pressure compensating valve  7  can be located apart from the control pressure producing component  7 B. Since the joint  104  is located in the body  103  of the control pressure producing component  7 B in this pressure compensating valve  7 , the machine can be made more compact and the number of parts can be reduced.  
     [0178] The variable throttle valve  30  of the control pressure producing component  7 B has a structure similar to that of the variable throttle valve  30  in the pressure compensating valve  7  depicted in FIG. 6. The body  103  of the control pressure producing component  7 B in this pressure compensating valve  7  thus requires no check valve in a manner similar to that in the pressure compensating valve  7  depicted in FIG. 6.  
     [0179]FIG. 9 depicts a sixth example of the structure of the pressure compensating valve  7 . This pressure compensating valve  7  is composed of only the compensator  7 A and the control pressure producing component  7 B. The compensator  7 A has a structure similar to that of the compensator  7 A depicted in FIG. 4.  
     [0180] The control pressure producing component  7 B is equipped with a variable throttle valve  30  having a structure allowing the magnitude of the throttling to be manually altered. This variable throttle valve  30  has a vertical hole  106  in the body  105 , and a poppet type spool  107  is inserted into this vertical hole  106 . The top and bottom of the vertical hole  106  can be rendered communicable and are blocked by the vertical movement of the spool  107 .  
     [0181] The top of a vertical hole  106  communicates through the oil passage  40  to the actuator port  64  of the operating valve  4 . The bottom of the vertical hole  106  communicates through the oil passage  37  equipped with the throttle  36  to the outlet port  62  of the operating valve  4 , and also communicates through the oil passage  38  to the pressure chamber  92 .  
     [0182] An adjusting screw  108  is threaded into the top of the vertical hole  106 , and a spring  109  with weak elastic force is interposed between the adjusting screw  108  and the spool  107 .  
     [0183] In the variable throttle valve  30  constructed in this manner, the pressured oil with the pressure Pa discharged from the outlet port  62  of the operating valve  4  flows through the oil passage  37  into the bottom of the vertical hole  106 .  
     [0184] With this, the spool  107  is pushed up, and part of the pressured oil with the pressure Pa flows into the oil passage  40  while constricted by the spool  107 . The pressure Pe of the pressure chamber  92  is set according to the amount of pressured oil flowing into the oil passage  40 , that is, according to the throttle level of the spool  107 .  
     [0185] The upward moving stroke of the spool  107  defining the throttle level of the spool  107  can be adjusted by manually rotating the adjusting screw  108 . The pressure compensating valve  7  can thus alter the pressure Pe, that is, can alter the pressure compensation characteristics, when the screw  108  is rotated.  
     [0186] Since the spool  107  is a poppet valve type, when pressured oil flows from the cylinder  5  into the oil passage  40 , the spool  107  is pushed down, blocking off the top and bottom of the vertical hole  106  from each other. In other words, the spool  107  functions as a check valve.  
     [0187] Thus, with this pressure compensating valve  7 , there is no need to provide the body  105  with the check valve  39  depicted in FIG. 4, making the body  105  easier to fabricate.  
     [0188]FIG. 10 depicts a seventh example of the structure of the pressure compensating valve  7 . This pressure compensating valve  7  differs from the pressure compensating valve  7  depicted in FIG. 4 in terms of the structure of the compensator  7 A.  
     [0189] That is, the main valve  20  of the compensator  7 A depicted in FIG. 10 has a spool S comprising the unification of the valve component  66  and pushing component  67  depicted in FIG. 4.  
     [0190] In this pressure compensating valve  7 , the pressured oil with the maximum load pressure P LS  flowing into the annular space  75  flows through a hole  112  located in the sleeve  72  directly into the pressure chamber  83 , so the pressure Pb of the pressure chamber  83  results in the maximum load pressure P LS .  
     [0191] The spool S forms a communication hole  113  along the central axis, thereby allowing the outlet port  62  of the operating valve  4  and the pressure chamber  92  to communicate with each other. As a result, the pressured oil with the pressure Pa flowing from the outlet port  62  of the operating valve  4  flows through the communicating hole  113  into the pressure chamber  92 . In other words, the communicating hole  113  functions as the oil passage  37  in FIG. 4.  
     [0192] A fixed throttle  113   a  corresponding to the fixed throttle  36  depicted in FIG. 4 is formed at the end on the pressure chamber  92  side of the communication hole  113 .  
     [0193] In the pressure compensating valve  7  having the aforementioned structure, there is no need to provide the body  60  of the compensator  7 A with the oil passage  37  depicted in FIG. 4, nor is there-any need to provide the body  87  of the control pressure producing component  7 B with the throttle  36  depicted in FIG. 4. The bodies  60  and  87  are thus easier to fabricate.  
     [0194] The variable throttle valve  30  of the control pressure producing component  7 B has a structure similar to that of the variable throttle valve  30  depicted in FIG. 4.  
     [0195]FIG. 11 depicts an eighth example of the structure of the pressure compensating valve  7 . The structure of the compensator  7 A in this pressure compensating valve  7  is similar to that of the pressure compensating valve  7  depicted in FIG. 10, and the structures of the control pressure producing component  7 B and pilot pressure producing component  7 C are similar to those of the pressure compensating valve  7  depicted in FIG. 6.  
     [0196] Thus, in this pressure compensating valve  7 , the same effects in making the body  60  and the body  218  easier to fabricate can be obtained as in the pressure compensating valve  7  depicted in FIG. 10, and the same effects in making a more compact machine, reducing the number of parts, and making it easier to fabricate the body  100  can be obtained as in the pressure compensating valve  7  depicted in FIG. 6.  
     [0197]FIG. 12 depicts a ninth example of the structure of the pressure compensating valve  7 . The structure of the compensator  7 A in this pressure compensating valve  7  is the same as that of the pressure compensating valve  7  depicted in FIG. 10, while the structure of the control pressure producing component  7 B and the location for attaching the joint  102  are the same as that of the pressure compensating valve  7  depicted in FIG. 7.  
     [0198] In this pressure compensating valve  7 , the same effects in making the bodies  60  and  87  easier to fabricate can be obtained as in the pressure compensating valve  7  depicted in FIG. 10, and the same effects in locating the electromagnetic proportional pressure control valve  50  apart from the control pressure producing component  7 B can be obtained as in the pressure compensating valve  7  depicted in FIG. 7.  
     [0199]FIG. 13 depicts a tenth example of the structure of the pressure compensating valve  7 . The structure of the compensator  7 A of this pressure compensating valve  7  is similar to that of the pressure compensating valve  7  depicted in FIG. 10, and the structure of the control pressure producing component  7 B and the position for attaching the joint  140  are the same as in the pressure compensating valve  7  depicted in FIG. 8.  
     [0200] In this pressure compensating valve  7 , the same effects in making it easier to fabricate the bodies  60  and  218  can be obtained as in the pressure compensating valve  7  depicted in FIG. 10, and the same effects in locating the electromagnetic proportional pressure control valve  50  apart from the control pressure producing component  7 B can be obtained as in the pressure compensating valve  7  depicted in FIG. 8.  
     [0201] The variable throttle valve  30  of the control pressure producing component  7 B has a structure similar to that of the variable throttle valve  30  in the pressure compensating valve  7  depicted in FIG. 6. The same effects in dispensing with the need to provide the body  103  of the control pressure producing component  7 B with a check valve can be obtained as in the pressure compensating valve  7  depicted in FIG. 6.  
     [0202]FIG. 14 depicts an eleventh example of the structure of the pressure compensating valve  7 . The structure of the compensator  7 A of this pressure compensating valve  7  is similar to that of the pressure compensating valve  7  depicted in FIG. 10, and the structure of the control pressure producing component  7 B is similar to that of the pressure compensating valve  7  depicted in FIG. 9.  
     [0203] In this pressure compensating valve  7 , the same effects in making the bodies  60  and  105  easier to fabricate are obtained as in the pressure compensating valve  7  depicted in FIG. 10. The same effects in being able to manually adjust the throttle level and making the body  105  easier to fabricate can be obtained as in the pressure compensating valve  7  depicted in FIG. 9.  
     [0204]FIG. 15 depicts a twelfth example of the structure of the pressure compensating valve  7 . The structure of the compensator  7 A of this pressure compensating valve  7  differs from that of the pressure compensating valve  7 A depicted in FIG. 4.  
     [0205] The spool S of the main valve  20  of the compensator  7 A depicted in FIG. 15 is equipped with a piston  116  featuring the unification of the valve component  66  and the piston  73  depicted in FIG. 4, and a sliding element  117  located in the piston  116 .  
     [0206] The piston  116  and the sliding element  117  are located along the central axis through the communication holes  118  and  119 , respectively. One end of the communication hole  119  in the sliding element  117  communicates through a check valve  120  to the communication hole  118  of the piston  116 , and the other end communicates through a throttle  119   a  corresponding to the throttle  36  depicted in FIG. 2 to the pressure chamber  92 .  
     [0207] In the pressure compensating valve  7  with the aforementioned structure, the pressured oil with the pressure Po supplied from the outlet port  62  flows into the pressure chamber  92  through the communication hole  118 , a check valve  120 , a slit  121  formed around the check valve  120 , a port  122  passing through the peripheral wall of the sliding element  117 , the communicating hole  119 , and the throttle  119   a . In other words, the communication holes  118  and  119  function as the oil passage  37  depicted in FIG. 2.  
     [0208] Accordingly, there is no need to provide the body  60  of the compensator  7 A with the oil passage  37  depicted in FIG. 4, and there is no need to provide the body  87  of the control pressure producing component  7 B with the throttle  36  depicted in FIG. 4. It is thus easier to fabricate the bodies  60  and  87 .  
     [0209] Meanwhile, when the pressure P 1  of the pressured oil in the actuator port  64  becomes greater than the pressure Po of the oil pressure in the outlet port  62 , the check valve  120  closes. The pressured oil in the actuator port  64  is thus prevented by the check valve  120  from flowing into the outlet port  62 .  
     [0210] The check valve  120  thus has the same function as the check valve  39  depicted in FIG. 2. Accordingly, in this pressure compensating valve  7 , there is no need to provide the body  87  of the control pressure producing component  7 B with the check valve  39  depicted in FIG. 4, which makes the body  87  easier to fabricate.  
     [0211] In the pressure compensating valves  7  described above, the oil passage  40  connected to the outlet port  33  of the variable throttle valve  30  was connected to the actuator port  64  (oil passage  6   a ) of the operating valve  4  depicted in FIG. 3, but this oil passage  40  may also be connected to the tank port  65 .  
     [0212] The structure of the unloading pressure control valve  10  relating to the present invention is described below with reference to FIG. 16.  
     [0213]FIG. 16 is a circuit diagram of oil pressure, depicting the structure of the unloading pressure control valve  10 . The unloading pressure control valve  10  is used to return the oil discharged from a hydraulic pump  1  directly to a tank to keep the hydraulic pump  1  in an unloaded state in a hydraulic system comprising, for example, a variable delivery pump  1 , an auxiliary hydraulic pump (pilot hydraulic pump)  2 , an operating valve  4  to which the oil discharged from the hydraulic pump  1  is supplied through an oil passage  3 , and a hydraulic cylinder (hydraulic actuator)  5  located opposite the operating valve  4 .  
     [0214] The unloading pressure control valve  10  comprises a main valve  100  and an electromagnetic proportional pressure control valve  101 .  
     [0215] The main valve  100  has a first pressure receiving component  123 , a second pressure receiving component  124 , a third pressure receiving component  125 , and a fourth pressure receiving component  126 . The main valve  100  sets the throttle level (unloading start pressure) between a first inlet port  127  and outlet port  128  by means of the elastic force of a spring  130  and the pressure acting on the first pressure receiving component  123 , second pressure receiving component  124 , third pressure receiving component  125 , and fourth pressure receiving component  126 .  
     [0216] The first pressure receiving component  123  is connected to the variable delivery pump  1  along with the first inlet port  127 , and receives the discharge pressure P P  of the hydraulic pump  1 . The second pressure receiving component  124  receives the maximum load pressure P LS  by way of a throttle  129 . The third pressure receiving component  125  receives the control pressure Pg described below. The fourth pressure receiving component  126  is connected to the tank. The main valve  100  determines the unloading set pressure by means of the elastic force of the spring  130  and the pressure area of the second pressure receiving component  124  and third pressure receiving component. The main valve  100  does not require the spring  130 . In other words, the unloading start pressure can be set by just the difference between the pressure area of the second pressure receiving component  124  and the third pressure receiving-component.  
     [0217] The control pressure Pg is given from the electromagnetic proportional pressure control valve  101 . That is, the electromagnetic proportional pressure control valve  101  introduces the pressured oil discharged from an auxiliary hydraulic pump  2  through the inlet port  132 , and the oil pressure resulting from a reduction in the pressure Pc of this pressured oil is output as the control pressure Pg. The control pressure Pg changes proportionally to the amount of electricity sent to the solenoid  133 .  
     [0218] When zero electricity is supplied to the solenoid  133 , the outlet port  135  communicates with the tank port  136  by means of the elastic force of a spring  134 , as shown in the figure. The control pressure Pg acting on the third pressure receiving component  125  of the main valve  100  is thus zero.  
     [0219] The specific structure of the unloading pressure control valve  10  is described below with reference to FIG. 17.  
     [0220] A sliding element  145  is slidably inserted into the left side of the valve body  140  of the main valve  100 , and the left end of a sleeve  148  is fitted to the right side of the valve body  140 .  
     [0221] The sliding element  145  has a U-shaped cross section, and is brought into contact on the left end surface with an adjusting screw  147  threaded into the left end of the valve body  140 . The adjusting screw  147  is locked by a lock nut  148 . The interior of the sliding element  145  communicates through a hole  145   a  to the tank.  
     [0222] A spool  150  has a first small diameter component  151  forming a left half, a large diameter component  152  forming a central component, and a second small diameter component  153  forming a right half. The left tip of the first small diameter component  151  of the spool  150  is slidably inserted into the sliding element  145 . The large diameter component  152  is slidably inserted into a large diameter hole  154  in a sleeve  146 . The second small diameter component  153  is slidably inserted into a small diameter hole  155  in the sleeve  146 .  
     [0223] The right end surface  150   a  of the spool  150  forms the first pressure receiving component  123  depicted in FIG. 16. The left end surface  150   b  of the spool  150  forms the fourth pressure receiving component  126 .  
     [0224] The spool  150  is designed so that the cross sectional area of the second small diameter component  153  is a size equal to that obtained by subtracting the cross sectional area of the first small diameter component  151  from the cross sectional area of the large diameter component  152 .  
     [0225] The right end of the sleeve  146  is positioned in the valve body  180  of the operating valve  4 . The sleeve  146  forms the first inlet port  127  depicted in FIG. 16 by opening the right end. The inlet port  127  communicates with the pump port  181  of the operating valve  4 .  
     [0226] Meanwhile, the sleeve  146  forms the outlet port  128  depicted in FIG. 16 at a position located slightly to the left of the right end opening. The outlet port  128  communicates with the tank port  182  of the operating valve  4 .  
     [0227] The sleeve  146  further comprises a load pressure introduction port  157  and a control pressure introduction port  158 . The load pressure introduction port  157  introduces pressured oil with the maximum load pressure P LS  The control pressure introduction port  158  introduces control pressure Pg through the electromagnetic proportional pressure control valve  101 .  
     [0228] The load pressure introduction port  157  communicates through an annular space  159 , an oil hole  160 , and a fine hole  161  to a spring chamber  162 . The annular space  159  is formed between the inner peripheral surface of the sleeve  146  and the outer peripheral surface of the second small diameter component  153  of the spool  150 . The oil hole  60  is formed along the central axis of the spool  150 . The fine hole  161  passes diametrically through the spool  150 , forming the throttle  129  depicted in FIG. 16.  
     [0229] Meanwhile, the control pressure introduction port  158  communicates with a space  163  formed between the large diameter component  152  of the spool  150  and the sleeve  146 . The right end surface  152   a  of the spool large diameter component  152  located in the space  163  forms the third pressure receiving component  125  depicted in FIG. 16.  
     [0230] The spring  130  depicted in FIG. 16 is located in the spring chamber  162 . The spring  130  is interposed between a spring receiver  162   a  inserted into the first small diameter component  151  of the spool  150  and the right end surface of the sliding element  145 , and pushes the spool  150  to the right.  
     [0231] While the spring receiver  162   a  is in contact with the left end of the sleeve  146  in the state depicted in the figure, the first inlet port  127  and outlet port  128  are blocked off from each other by the right end of the spool  150 . The left end surface  152   b  of the spool large diameter component  152  facing the spring chamber  162  forms the elastic force creating component of the spring  130  as well as the second pressure receiving component  124  depicted in FIG. 16.  
     [0232] The electromagnetic proportional pressure control valve  101  of the unloading pressure control valve  10  is described below.  
     [0233] The electromagnetic proportional pressure control valve  101  is disposed over the valve body  140  of the main valve  100 . A spool  167  for allowing the inlet port  132  and outlet port  135  depicted in FIG. 16 to communicate with each other and to be blocked off from each other is located in the valve body  166  of the electromagnetic proportional pressure control valve  101 . The top of the valve body  166  has a solenoid  133  that pushes the spool  167  down against the spring  134 .  
     [0234] The inlet port  132  is connected to the auxiliary hydraulic pump  2 . The outlet port  135  communicates through an oil passage  168  to the control pressure introduction port  158 .  
     [0235] The operation of the unloading pressure control valve  10  having the aforementioned structure is described below.  
     [0236] When the discharge pressure P P  of the hydraulic pump  1  acts on the right end surface  150   a  of the spool  150  which is the first pressure receiving component  123 , the spool  150  is pushed to the left (the direction passing through the first inlet port  127  and outlet port  128 ).  
     [0237] Meanwhile, the control pressure Pg supplied from the electromagnetic proportional pressure control valve  101  acts on the right end surface  152   a  of the large diameter component of the spool  150  serving as the third pressure receiving component  125 , by way of the oil passage  168  and the control pressure introduction port  158 , so that the spool  150  is pushed to the left.  
     [0238] The spring  130  located in the spring chamber  162  pushes the spool  150  to the right. The load pressure P LS  is introduced through the load pressure introduction port  157 , annular space  159 , oil hole  160 , and fine hole  161  (throttle  129 ) into the spring chamber  162 . The load pressure P LS  thus acts on the left end surface  152   b  of the large diameter component of the spool  150  which is the second pressure receiving component  124 , and the spool  150  is pushed to the right.  
     [0239] The balance of force determining the position of the spool  150  in the unloading pressure control valve  10  is represented by the following Eq. (3).  
       P   P ×A 1   =P   LS ×( A   2   −A   1 )+ F   0   −Pg ×( A   2 −A1)  (3)  
     [0240] Where  
     [0241] A 1 : area of right end surface  150   a  of spool  150   
     [0242] A 2 : area of large diameter component  152  of spool  150   
     [0243] A 3 : area of left end surface  150   b  of spool  150   
     [0244] F 0 : elastic force of spring  130   
     [0245] As noted above, the relation between area A 1 , A 2 , and A 3  is A 1 =(A 2 −A 3 ). Eq. (3) thus results in Eq. (4) below.  
     ( P 2 −P   LS )× A   1   =F   0   −Pg ×( A   2   −A   1 )  (4)  
     [0246] It is evident from the Eq. (4) that a constant pressure difference P P −P LS  is obtained irrespective of fluctuations in the load pressure P LS  when the control pressure Pg is constant.  
     [0247] The pressure difference P P −P LS  determines the unloading start pressure. The unloading pressure control valve  10  thus allows the unloading start pressure to be arbitrarily set by controlling the amount of electricity to the solenoid  133  of the electromagnetic proportional pressure control valve  101  to change the control pressure Pg.  
     [0248] The main valve  100  of the unloading pressure control valve  10  is interposed between the hydraulic pump  1  and the tank. Thus, when the pressure difference P P −P LS  reaches the unloading start pressure, the oil discharged from the hydraulic pump  1  is returned to the tank during continuous operation.  
     [0249] When the operating valves  4  are operated in the center valve position, the pressure difference P P −P LS  increases to the unloading start pressure. With this, the oil discharged from the hydraulic pump  1  is returned through the unloading pressure control valve  10  to the tank, so the hydraulic pump  1  is in an unloaded state.  
     [0250] The electromagnetic proportional pressure control valve  101  of the unloading pressure control valve  10  produces pilot control pressure Pg resulting from the reduction of the discharge oil pressure Pc of the auxiliary hydraulic pump  2 . Meanwhile, in the main valve  100 , the operating start pressure (unloading start pressure) changes according to the control pressure Pg given by the electromagnetic proportional pressure control valve  101 .  
     [0251] Thus, according to the unloading pressure control valve  10 , control signals to the solenoid  133  of the electromagnetic proportional pressure control valve  101  can be changed to set the unloading start pressure to the desired magnitude.  
     [0252]FIG. 18 depicts another embodiment of the unloading pressure control valve relating to the present invention.  
     [0253] This unloading pressure control valve  10  comprises an attachment block  185 , a piping joint  187 , and an oil pressure pilot valve  188 . The attachment block  185  is fixed to the upper surface of the valve body  140 . The piping joint  187  is screwed into a threaded hole  186  located in the attachment block  185 , and is thus secured. The oil pressure pilot valve  188  is manually operated.  
     [0254] The threaded hole  186  passes through the control pressure introduction port  158 . The inlet port  188   b  of the oil pressure pilot valve  188  is connected to the auxiliary hydraulic pump  2 . The outlet port  188   a  is connected to the piping joint  187 .  
     [0255] In this unloading pressure control valve  10 , the pressured oil with the control pressure Pg supplied from the oil pressure pilot valve  188  acts on the right end surface  152   a  (third pressure receiving component  125 ) of the spool large diameter component  152  by way of the control pressure introduction port  158 .  
     [0256] This unloading pressure control valve  10  allows the unloading start pressure to be arbitrarily set according to the control pressure Pg. The oil pressure pilot valve  188  which is the means for producing the control pressure Pg can also be disposed apart from the main valve  100 . It can thus be freely disposed, enabling manual remote control of the unloading start pressure, and the like.  
     [0257] In the unloading pressure control valves  10  depicted in FIGS. 17 and 18, the control pressure Pg acted as the force moving the spool  150  to the left (the direction passing through the first inlet port  127  and outlet port  128  of the main valve  100 ).  
     [0258] In contrast to the above, it is also possible to allow the control pressure Pg to act as the force moving the spool  150  to the right. In this case, the pressing force of the spring  130  acts in the direction opposite that described above (the direction in which the spool is pushed to the left).  
     [0259] When the control pressure Pg is allowed to act in the opposite direction as described above, the unloading start pressure increases as the control pressure Pg increases.  
     [0260]FIG. 19 depicts a hydraulic system featuring the use of two hydraulic pumps  1 A and  1 B.  
     [0261] In this hydraulic system, the hydraulic pumps  1 A and  1 B are connected to corresponding operating valves  4 A and  4 B by means of a switching valve  191  in a converged flow component  190 . A switching valve  192  switches between the communication and blockage of pressured oil, with a maximum load pressure P LS-A  sensed by one shuttle valve  8 A, and pressured oil with a maximum load pressure P LS-B  sensed by another shuttle valve  8 B.  
     [0262] The switching valves  191  and  192  of the converged flow component  190  are always simultaneously switched over by means of the pilot pressure Ph.  
     [0263] In the state depicted in the figure, the switching valves  191  and  192  of the converged flow component  190  in this case allow the oil discharged by the hydraulic pumps  1 A and  1 B to converge, and also allow the pressured oil with the load pressures P LS-A  and P LS-B  to converge.  
     [0264] Meanwhile, when the switching valves  191  and  192  of the converged flow component  190  are switched over by the pilot pressure Ph, the oil discharged by the hydraulic pumps  1 A and  1 B and that with the load pressures P LS-A  and P LS-B  are separated from each other, resulting in the independent operation of the unloading pressure control valves  10 A and  10 B.  
     [0265] The maximum load pressure P LS-A  sensed by the shuttle valve  8 A is the highest among the plurality of hydraulic cylinders  5  driven by the hydraulic pump  1 A. The maximum load pressure P LS-B  sensed by the shuttle valve  8 B is the highest among the plurality of hydraulic cylinders  5  driven by the hydraulic pump  1 B.  
     [0266] The load pressure P LS-A  is supplied to the unloading pressure control valve  10  and the volume control component (pump discharge pressure control means)  12  of the hydraulic pump  1 A. The load pressure P LS-A  is also supplied through a check valve  193 A to the load pressure bleed valve  11 .  
     [0267] The load pressure P LS-B  is supplied to the unloading pressure control valve  10  and the volume control component  12  of the hydraulic pump  1 B. The load pressure P LS-B  is also supplied through a check valve  193 B to the load pressure bleed valve  11 .  
     [0268] As described above, when the two pump circuits are separated, the switching valves  191  and  192  of the converged flow component  190  are switched to a blocking state.  
     [0269] However, even though the switching valves  191  and  192  are in a blocked state, minute amounts of oil leakage always occur. For example, when one operating valve  4 A is in the center valve state, and the other operating valve  4 B is in the operating state, the maximum load pressure P LS-A  sensed by the shuttle valve  8 A should be zero as long the switching valve  192  is operating in an ideal manner. In fact, however, the oil leakage from the switching valve  192  results in an increase in the maximum load pressure P LS-A .  
     [0270] In this case, when the maximum load pressure P LS-A  increases, the discharge pressure P P  of the hydraulic pump  1 A also increases, resulting in the maximum load pressure P LS-A +pump set pressure.  
     [0271] The pressured oil with the maximum load pressure P LS-A  is allowed to communicate with the tank during the operation of the one unloading pressure control valve  10 A. That is, the pressured oil with the maximum load pressure P LS-A  is introduced from in front of the throttle  129  through the branched piping into the unloading pressure control valve  10 A, and this pressured oil is also output through a throttle  169  from the unloading pressure control valve  10 A so as to be returned to the tank. This allows the maximum load pressure P LS  confined in the piping leading from the shuttle valve  8 A to the main valve  20  to escape to the tank, and prevents the discharge pressure P P  of the hydraulic pump  1 A from increasing. The discharge pressure P P  of the hydraulic pump  1 B can similarly be prevented from increasing during the operation of the other unloading pressure control valve  10 B.  
     [0272] The structure of the variable bleed valve  11  relating to the present invention is described below with reference to FIG. 20.  
     [0273] The variable bleed valve  11  comprises a variable throttle valve  110  and an electromagnetic proportional pressure control valve  111 , as shown in the enlargement in FIG. 20.  
     [0274] The variable throttle valve  110  is operated so as to increase the area of the opening-between an inlet port  196  and an outlet port  197  by means of the elastic force of a spring  95  and the pilot pressure Pg acting on a pressure receiving component  194 , and is operated so as to reduce the area of the opening by means of the elastic force of a spring  198 .  
     [0275] The electromagnetic proportional pressure control valve  111  introduces pressured oil with a standard pressure Pc discharged from the auxiliary hydraulic pump  2  into the inlet port  199 , and the pressure Pc of the pressured oil is reduced to the pilot pressure Pg. The pressured oil with the pilot pressure Pg is allowed to act on the pressure receiving component  194  of the variable throttle valve  110  by way of the outlet port  200 . The pilot pressure Pg changes proportionally to the amount of electricity to the solenoid  201 .  
     [0276] The variable bleed valve  11  is connected to a controller  300 . The controller  300  gives a corresponding control signal to the solenoid  201  of the electromagnetic proportional pressure control valve  111  based on operation commands such as a command to open the operating valve  4  by the operation of an operating lever (not shown in figure).  
     [0277]FIG. 21 depicts the variable bleed valve  11  while mounted. It may also be seen from FIG. 21 that the variable bleed valve  11  is provided as a valve block along with a plurality of operating valves  4  and  4 . That is, the variable bleed valve  11  is attached by means of a support block  202  to the operating valve  4  located on the outermost side of the plurality of operating valves  4  joined in parallel. The symbol  4   a  indicates the spool of the operating valve  4 .  
     [0278]FIG. 22 is a cross section of line A-A in FIG. 21. It may be seen from FIG. 22 that the variable throttle valve  110  is such that the spool  206  is inserted into the spool hole  205  of the valve body  204 . The spool hole  205  is formed in the vertical direction.  
     [0279] The spool  206  is interposed between the inlet port  196  and outlet port  197  of the variable throttle valve  110 . The spool  206  is such that downward force (the direction in which the area of the opening between the ports  196  and  197  is reduced) is urged by the spring  215 . Meanwhile, the upward force (the direction in which the area of the opening between the ports  196  and  197  is increased) is urged by the spring  195  in the pressure chamber  209  formed between the spool and an adjustment screw  62 .  
     [0280] The bottom end surface of the spool  205  facing the pressure chamber  209  forms the pressure receiving component  194  depicted in FIG. 20. The elastic force of the spring  195  can be fine tuned by the operation of the adjustment screw  217 .  
     [0281] The inlet port  196  communicates with the load pressure introduction hole  203  through a load pressure introduction oil passage  210  leading from the valve body  204  to the support block  202 . The outlet port  197  communicates with the tank through a tank oil hole  211  that opens into the attachment surface  204  of the valve body  204 .  
     [0282] The electromagnetic proportional pressure control valve  111  is disposed on the upper surface of the aforementioned valve body  204 . The electromagnetic proportional pressure control valve  111  comprises a spool  214  and the solenoid  201 . The spool  214  is vertically disposed in the valve body  213 . The spool  214  is disposed coaxially relative to the spool  206  of the variable throttle valve  110 . The solenoid  201  pushes the spool  214  down against the spring  215  according to the amount of electricity.  
     [0283] The spool  214  is constantly pushed upward by the elasticity of the spring  215 . In this state, the inlet port  199  and outlet port  200  of the electromagnetic proportional pressure control valve  111  are blocked off from each other. The standard pressure Pc discharged by the auxiliary hydraulic pump  2  acts on the inlet port  199 .  
     [0284] The outlet port  200  communicates with the pressure chamber  209  by way of an oil passage  216  located in the valve body  213  and an oil passage  212  located in the valve body  204  of the variable throttle valve  110 . The spring  215  is in contact with the upper tip of the spool  206  of the variable throttle valve  110 . Here, the spring chamber  207  of the valve body  204  communicates with the tank by way of a tank oil hole  208  that opens into the attachment surface  204   a  of the valve body  204 .  
     [0285] The operation of the variable bleed valve  11  is described below.  
     [0286] Pressured oil present in the oil passage  9  gradually flows through the fixed throttle  112  into the tank when the operating valves  4  are operated in the center valve position. The maximum load pressure P LS  acting on the discharge pressure control means  12  thus gradually decreases. When the maximum load pressure P LS  decreases to zero, the displacement volume of the hydraulic pump  1  is reduced to the minimum preset volume by the pump discharge pressure control means  12 .  
     [0287] When the operating valve  4  is operated from this state to supply pressured oil to the hydraulic cylinder  5 , the maximum load pressure P LS  increases. The pressured oil of the maximum load pressure P LS  is introduced into the inlet port  196  of the variable throttle valve  110  through the load pressure introduction hole  203  connected to the oil passage  9  and the load pressure introduction oil passage  210 . When the inlet port  196  and outlet port  197  of the variable throttle valve  110  thus communicate with each other, part of the pressured oil with the maximum load pressure P LS  is bled off into the tank through the outlet port  197 .  
     [0288] The amount of the aforementioned pressured oil bled off at this time increases as the area of the opening between the inlet port  196  and outlet port  197  increases. The greater the amount that is bled off, the lower the rate of increase of the maximum load pressure P LS .  
     [0289] When the solenoid  201  of the electromagnetic proportional pressure control valve  111  is in a noncommunicating state, the spool  214  remains pushed up by the spring  215 . The inlet port  199  and outlet port  200  of the electromagnetic proportional pressure control valve  111  are thus blocked off from each other.  
     [0290] In this state, the pilot pressure Pg given from the outlet port  200  of the electromagnetic proportional pressure control valve  111  to the pressure chamber  209  of the variable throttle valve  110  is zero. The spool  206  of the variable throttle valve  110  is thus pushed down by the spring  198 .  
     [0291] When the spool  206  is pushed down, the inlet port  196  and outlet port  197  of the variable throttle valve  110  are blocked off from each other by the spool  206 . The area of the opening between the ports  196  and  197  is thus reduced to the minimum, and the amount of pressured oil that is bled off by the variable throttle valve  110  is zero.  
     [0292] When electricity is applied to the solenoid  201  of the electromagnetic proportional pressure control valve  111 , the spool  214  of the electromagnetic proportional pressure control valve  111  is pressed down by the thrust of the solenoid  201 , allowing the inlet port  199  and outlet port  200  to communicate with each other.  
     [0293] With this, the pilot pressure Pg resulting from a reduction in the discharge pressure Pc of the auxiliary hydraulic pump  2  acts on the pressure chamber  209  of the variable throttle  20 . The spool  206  of the variable throttle valve  110  thus moves up against the spring  198 .  
     [0294] When the spool  206  moves up, the inlet port  196  and outlet port  197  of the variable throttle valve  110  communicate with each other. As a result, the pressured oil with the maximum load pressure P LS  introduced into the inlet port  196  is bled off through the outlet port  197  into the tank.  
     [0295] The greater the pilot pressure Pg at this time, in other words, the greater the amount of electricity to the solenoid  201  of the electromagnetic proportional pressure control valve  111 , the greater the amount of pressure bled off by the variable throttle valve  110 .  
     [0296] As is evident from the description above, the variable bleed valve  11  allows the amount of pressured oil with the maximum load pressure P LS  that is bled off to be arbitrarily adjusted by controlling the amount of electricity to the electromagnetic proportional pressure control valve  111 . In other words, the rate of increase in the maximum load pressure P LS  in the oil passage  9  can be arbitrarily adjusted by controlling the aforementioned amount of electricity.  
     [0297] Adjusting the amount of pressured oil that is bled off by the variable throttle valve  110  to zero results in a higher rate of increase in the maximum load pressure P LS  acting on the pump discharge pressure control means  12 , so the pump discharge pressure control means  12  rapidly increases the displacement volume (discharge oil amount) of the hydraulic pump  1 .  
     [0298] As a result, the hydraulic cylinder  5  starts rapidly at the same time the operating valve  4  is operated.  
     [0299] In contrast, when the variable throttle valve  110  is in bleed off operating mode, the rate of increase in the maximum load pressure P LS  acting on the pump discharge pressure control means  12  is lower than when the aforementioned bleed off amount is zero. In this case, the pump discharge pressure control means  12  moderately increases the displacement volume of the hydraulic pump  1 , so the start up speed of the hydraulic cylinder  5  decreases.  
     [0300] Accordingly, the variable bleed valve  11  allows the start up response of the hydraulic cylinder  5  to be adjusted by controlling the amount of electricity to the solenoid  201  of the electromagnetic proportional pressure control valve  111 .  
     [0301] The amount discharged by the hydraulic pump  1  is controlled to bleed off the pressured oil in the oil passage  9  for sensing the maximum load pressure P LS  serving as the pilot pressure. The amount flowing in the load pressure sensing channel  9  is generally quite low. The pump pressure is controlled according to the pressure of the load pressure sensing passage  9 , whereas the pressure of the load pressure sensing passage  9  is the pressure corresponding to the load pressure of the actuator and thus reacts exactly to the fluctuations in the load pressure of the actuator. It also reacts promptly to fluctuations in the load pressure. Energy loss can thus be minimized, and machines can be made more compact. The amount discharged from the hydraulic pump  1  can be controlled with greater precision.  
     [0302] In the aforementioned hydraulically operated device, only bleed off operations actually stop in the event of accidents such as malfunctions of the electromagnetic proportional pressure control valve  111  which lead to interruption of the pilot pressure Pg. In other words, the operation of the hydraulic cylinder  5  by the hydraulic pump  1  is unaffected even when accidents such as those described above occur. The reliability of the hydraulically operated device can thus be improved.  
     [0303] Moreover, the amount of pressured oil that is bled off can be arbitrarily adjusted by means of control signals output by a controller  300  described below, making such control easier to manage. Since pressured oil should be supplied by the application of electricity to the electromagnetic proportional pressure control valve  111  only when bleed off is needed, not only can pressured oil energy loss be further minimized, but electrical energy can also be economized.  
     [0304] Here, the aforementioned hydraulically operated device is equipped with a controller  300  connected to the variable bleed valve  11  as described above, and this controller  300  comprises, as shown in FIG. 20, a mode setting memory component  310 , a mode select setting component  320 , and a control signal output component  330 .  
     [0305] The mode setting memory component  310  sets and stores a plurality of input-output relations according to the operating configuration of the hydraulic cylinder  5 . As shown in FIG. 23, for example, three different modes comprising an ordinary mode which is the ordinary operating state, a heavy operating mode requiring considerable force, and a more precise operating mode requiring highly precise manipulations are set and stored in terms of the input-output relations between the open command to the operating valve  4  and the control signals to the solenoid  201  of the electromagnetic proportional pressure control valve  111 , that is, the area of the opening of the variable throttle valve  110 . Although these three input-output relations have the same degree of variation relative to each other, the area of the opening of the variable throttle valve  110  in terms of open commands to the same operating valve  4  is preset and stored so as to increase in ascending order from heavy operating mode, to ordinary mode, to precision operating mode.  
     [0306] The mode select setting component  320  selects and sets one of the three input-output relations set and stored in the mode setting memory component  310 . This mode select setting component  320  selects and sets a corresponding input-output relation according to the operation of a mode select switch not shown in the figure and located in the driver seat of a hydraulic shove, for example.  
     [0307] The control signal output component  330  converts the open command for the operating valve  4  based on the input-output relation selected by the mode select setting component  320 , and the converted control signal is given to the solenoid  201  of the electromagnetic proportional pressure control valve  111 .  
     [0308] Thus, in the aforementioned hydraulically operated device, a control signal output from the controller  300  in response to an open command for the operating valve  4  can be modified according to the operating configuration of the hydraulic cylinder  5 . In other words, when heavy operating mode is selected and set by the mode select setting component  320 , the area of the opening of the variable throttle valve  110  for open commands to the operating valve  4  can be further reduced. Thus, in this heavy operating mode, more pressured oil can be supplied to the hydraulic cylinder  5  and the hydraulic cylinder can be rapidly operated, even though the control input of the operating lever (not shown in figure) is the same.  
     [0309] Meanwhile, when precision operating mode is selected and set, the area of the opening in the variable throttle valve  110  can be further increased for open commands to the operating valve  4 . Thus, in precision operating mode, less pressured oil can be supplied to the hydraulic cylinder  5  and the hydraulic cylinder can be moderately operated, even though the control input of the operating lever (not shown in figure) is the same.  
     [0310] The hydraulic cylinder  5  is provided to drive the operating unit of the hydraulic shovel (such as a boom, arm, or bucket). A hydraulically operated device equipped with the variable bleed valve  11  can thus provide operating speeds and operating sensitivity for an operating unit that are suitable for the operating configuration of the aforementioned hydraulic shovel.  
     [0311] The plurality of input-output relations set and stored in the mode setting memory component  310  are not limited to those depicted in FIG. 5.  
     [0312]FIG. 24 is a graph depicting another example of input-output relations set and stored by the mode setting memory component  310 . The input-output relations depicted in FIG. 24 are designed so that the rate of change increases in the order from the heavy operating mode, to ordinary mode, to precision operating mode. The use of this mode setting selection means results in a different proportion of change in the speed by which the pressured oil in the load pressure sensing passage  9  is bled off into the tank, allowing the operating speeds and operating sensitivity of the hydraulic cylinder  5  to be set with even greater precision according to the operating configuration.  
     [0313] A combination of the input-output relations depicted in FIGS. 23 and 24 can provide input-output relations such as that indicated by the broken lines in FIG. 23 for the heavy operating mode and precision operating mode in relation to ordinary mode. In this case, the input-output relations can be set even more precisely than those depicted in FIG. 24.  
     [0314] The variable throttle valve  110  is constructed in such a way as to increase the area of the opening between the inlet port  196  and outlet port  197  by means of the action of the pilot pressure Pg, but conversely it can also be constructed in such a way as to reduce the area of the aforementioned opening by means of the action of the pilot pressure Pg.  
     [0315] The variable throttle valve  110  is also constructed in such a way that the spool  206  is pressed in the cut-off direction (downward in FIG. 22) by the spring  198  and the spool  206  is pressed in the communicating direction (upward in FIG. 22) by the pressure in the pressure chamber  209 , but it can also be constructed in such a way that the elastic force of the spring  198  and the pressure of the pressure chamber  209  act in directions opposite those described above.  
     [0316] The variable bleed valve  11  is such that the spool  206  of the variable throttle valve  110  and the spool  214  of the electromagnetic proportional pressure control valve  111  are located coaxially, making it possible to achieve more compact shapes with a shorter lateral length. That is, when the lay out of the variable bleed valve  11 , for example, is like that depicted in FIG. 21, a more compact embodiment can be devised because the electromagnetic proportional pressure control valve  111  can be mounted further inside than the spring case  4   b  of the operating valve  4 , that is, inside the surface defined by the spring case  4   b  when a valve block is used in a generally right-angled parallelepiped form.  
     [0317] The variable bleed valve  11  is such that the spring  215  of the electromagnetic proportional pressure control valve  111  is in contact with the upper end of the spool  206  of the variable throttle valve  110 . According to this structure, the operating force of the spool  206  is mechanically fed back to the spool  214  of the electromagnetic proportional pressure control valve  111  through the spring  215  when the spool  206  of the variable throttle valve  110  is operated. The operating characteristics (response) of the spool  206  of the variable throttle valve  110  can thus be improved, allowing more precise bleed off operations to be managed.  
     [0318] The variable bleed valve  11  is also designed to allow the elastic force of the spring  195  of the variable throttle valve  110  to be fine tuned by rotating the adjustment screw  217 . When a plurality of variable bleed valves  11  are manufactured, the machining precision of the various parts and the elastic force of the spring  198  used in the individual variable bleed valves  11  are not uniform. Despite the uneven elastic force of the spring  198 , however, it is possible to compensate for the uneven elastic force of the spring  198  by adjusting the elastic force of the spring  195  by rotating the adjustment screw  217 .