Hydraulic drive system for construction machine

It is an object of the present invention to accurately detect the absorption torque of the other of two hydraulic pumps by a purely hydraulic structure and feed the absorption torque to one of the two hydraulic pumps, thereby to accurately perform a total torque control, effectively utilize a rated output torque of a prime mover, and enhance mountability. To achieve the object, there are provided: a torque feedback circuit 31 to which the delivery pressure of a first hydraulic pump 1a and a load sensing drive pressure are introduced, which modifies the delivery pressure of a second hydraulic pump 1b to provide a characteristic simulating the absorption torque of the second hydraulic pump 1b, and which outputs the modified pressure; and torque feedback pistons 32a, 32b to which the output pressure of the torque feedback circuit 31 is introduced, and which control the capacity of the first hydraulic pump 1a to decrease the capacity of the first hydraulic pump 1a and decrease a maximum torque T1max as the output pressure becomes higher. The torque feedback circuit 31 includes pressure dividing restrictor parts 34a, 34b, pressure dividing valves 35a, 35b, and relief valves 37a, 37b.

TECHNICAL FIELD

The present invention relates to a hydraulic drive system for a construction machine such as hydraulic excavator. Particularly, the invention relates to a hydraulic drive system for a construction machine that includes at least two variable displacement hydraulic pumps, one of which has a pump control unit (regulator) performing at least a torque control and the other of which has a pump control unit (regulator) performing a load sensing control and a torque control.

BACKGROUND ART

As a hydraulic drive system for a construction machine such as hydraulic excavator, one having a regulator that controls the capacity (flow rate) of a hydraulic pump in such a manner that the delivery pressure of the hydraulic pump becomes higher than a maximum load pressure of a plurality of actuators by a target differential pressure is widely used, and this is called load sensing control. Patent Document 1 describes a two-pump load sensing system in a hydraulic drive system for a construction machine provided with a regulator for performing such a load sensing control, in which two hydraulic pumps are provided, and the respective two hydraulic pumps perform the load sensing control.

Besides, in a regulator of a hydraulic drive system for a construction machine, normally, a torque control is conducted such that the absorption torque of a hydraulic pump does not exceed a rated output torque of a prime mover, by decreasing the capacity of the hydraulic pump as the delivery pressure of the hydraulic pump rises, thereby to prevent stoppage of the prime mover (engine stall) due to an overtorque. In the case where the hydraulic drive system is provided with two hydraulic pumps, the regulator of one hydraulic pump performs a torque control (total torque control) by using not only its own delivery pressure but also a parameter concerning the absorption torque of the other hydraulic pump, thereby to attain both prevention of stoppage of the prime mover and effective utilization of a rated output torque of the prime mover.

For instance, in Patent Document 2, a total torque control is carried out by introducing the delivery pressure of one of the two hydraulic pumps to the regulator of the other hydraulic pump through a pressure reduction valve. A set pressure of the pressure reduction valve is fixed, and this set pressure is set at a value simulating a maximum torque in the torque control of the regulator of the other hydraulic pump. This ensures that in an operation of driving only the actuators concerning the one hydraulic pump, the one hydraulic pump can effectively use substantially the whole of the rated output torque of the prime mover, and, in a combined operation of simultaneously driving the actuators concerning the other hydraulic pump, the absorption torque of the whole of the pumps does not exceed the rated output torque of the prime mover, so that stoppage of the prime mover can be prevented from occurring.

In Patent Document 3, in order to carry out a total torque control for two variable displacement hydraulic pumps, the tilting angle of the other hydraulic pump is detected as an output pressure of a pressure reduction valve, and the output pressure is introduced to the regulator of the one hydraulic pump. In Patent Document 4, control accuracy of a total torque control is enhanced by detecting the tilting angle of the other hydraulic pump by replacing the tilting angle with the arm length of an oscillating arm.

PRIOR ART DOCUMENTS

Patent Documents

SUMMARY OF THE INVENTION

Problem to be Solved by the Invention

By applying the technology of the total torque control described in Patent Document 2 to the two-pump load sensing system described in Patent Document 1, it is possible to perform a total torque control also in the two-pump load sensing system described in Patent Document 1. In the total torque control of Patent Document 2, however, the set pressure of the pressure reduction valve is set at a fixed value simulating the maximum torque for the torque control of the other hydraulic pump, as aforementioned. Therefore, in a combined operation of simultaneously driving the actuators concerning the two hydraulic pumps, when the other hydraulic pump is in such an operating state that the other hydraulic pump is limited by the torque control and operates at the maximum torque for the torque control, it is possible to contrive effective utilization of a rated output torque of the prime mover. However, when the other hydraulic pump is in such an operating state that the other hydraulic pump is not limited by the torque control and performs a capacity control by the load sensing control, there occurs the following problem: notwithstanding the absorption torque of the other hydraulic pump being smaller than the maximum torque for the torque control, the output pressure of the pressure reduction valve simulating the maximum torque is introduced to the one regulator of the hydraulic pump, and a control such as to decrease the absorption torque of the one hydraulic pump more than necessary would be performed. Consequently, it has been impossible to accurately perform the total torque control.

In Patent Document 3, it is attempted to enhance the accuracy of the total torque control, by detecting the tilting angle of the other hydraulic pump as the output pressure of the pressure reduction valve and introducing the output pressure to the regulator of the one hydraulic pump. However, there occurs a problem. In general, the torque of a pump is determined as the product of delivery pressure and capacity, specifically, (delivery pressure×pump capacity)/2π. On the other hand, in Patent Document 3, the delivery pressure of the one hydraulic pump is introduced to one of two pilot chambers of a stepped piston, whereas the output pressure of the pressure reduction valve (the delivery amount proportional pressure for the other hydraulic pump) is introduced to the other pilot chamber of the stepped piston, and the capacity of the one hydraulic pump is controlled using the sum of the delivery pressure and the delivery amount proportional pressure as a parameter of the output torque. Consequently, there would be generated a considerable error between the parameter and the torque being actually used.

In Patent Document 4, the control accuracy of the total torque control is enhanced by detecting the tilting angle of the other hydraulic pump by replacing the tilting angle with the arm length of an oscillating arm. However, the regulator in Patent Document 4 has a very complicated structure in which the oscillating arm and a piston provided in a regulator piston structure are slid relative to each other while transmitting a force. To provide a sufficiently durable structure, therefore, it is necessary to cause parts such as the oscillating arm and the regulator piston to be rigid, which makes it difficult to miniaturize the regulator. Particularly, in the small-type hydraulic excavator such as so-called rear small swing type having a small rear end radius, there have been the cases where the space for accommodating the hydraulic pump is so small that it is difficult to mount the hydraulic pump.

It is an object of the present invention to provide a hydraulic drive system for a construction machine that is provided with two variable displacement hydraulic pumps, one having a pump control unit to perform at least a torque control and the other performing a load sensing control and a torque control, in which the absorption torque of the other hydraulic pump is accurately detected by a purely hydraulic structure and fed back to the one hydraulic pump side, whereby it is possible to accurately carry out the total torque control, effectively utilize a rated output torque of a prime mover, and enhance mountability.

Means for Solving the Problem

(1) To achieve the above object, the present invention provides a hydraulic drive system for a construction machine, including: a prime mover; a variable displacement first hydraulic pump driven by the prime mover; a variable displacement second hydraulic pump driven by the prime mover; a plurality of actuators driven by hydraulic fluids delivered by the first and second hydraulic pumps; a plurality of flow control valves that control flow rates of hydraulic fluids supplied from the first and second hydraulic pumps to the plurality of actuators; a plurality of pressure compensating valves that control differential pressures across the plurality of flow control valves; a first pump control unit that controls a delivery flow rate of the first hydraulic pump; and a second pump control unit that controls a delivery flow rate of the second hydraulic pump, the first pump control unit including a first torque control section that, when at least one of delivery pressure and capacity of the first hydraulic pump increases and absorption torque of the first hydraulic pump increases, controls the capacity of the first hydraulic pump such that the absorption torque of the first hydraulic pump does not exceed a first maximum torque, the second pump control unit including a second torque control section that, when at least one of delivery pressure and capacity of the second hydraulic pump increases and absorption torque of the second hydraulic pump increases, controls the capacity of the second hydraulic pump such that the absorption torque of the second hydraulic pump does not exceed a second maximum torque, and a load sensing control section that, when the absorption torque of the second hydraulic pump is lower than the second maximum torque, controls the capacity of the second hydraulic pump such that the delivery pressure of the second hydraulic pump becomes higher by a target differential pressure than a maximum load pressure of the actuators driven by a hydraulic fluid delivered by the second hydraulic pump, wherein the first torque control section includes a first torque control actuator that receives the delivery pressure of the first hydraulic pump and that, when the delivery pressure rises, controls the capacity of the first hydraulic pump to decrease the capacity of the second hydraulic pump and decrease the absorption torque thereof, and first biasing means that sets the first maximum torque, the second torque control section includes a second torque actuator that receives the delivery pressure of the second hydraulic pump and, when the delivery pressure rises, controls the capacity of the second hydraulic pump to decrease the capacity of the second hydraulic pump and decrease the absorption torque thereof, and second biasing means that sets the second maximum torque, the load sensing control section includes a control valve that varies a load sensing drive pressure such that the load sensing drive pressure is lowered as a differential pressure between the delivery pressure of the second hydraulic pump and the maximum load pressure becomes smaller than the target differential pressure, and a load sensing control actuator that controls the capacity of the second hydraulic pump to increase the capacity of the second hydraulic pump and increase the delivery flow rate as the load sensing drive pressure becomes lower, the first pump control unit further includes a torque feedback circuit that receives the delivery pressure of the second hydraulic pump and the load sensing drive pressure and modifies the delivery pressure of the second hydraulic pump based on the delivery pressure of the second hydraulic pump and the load sensing drive pressure to provide a characteristic simulating the absorption torque of the second hydraulic pump both in the cases of when the second hydraulic pump is limited by control of the second torque control section and operates at the second maximum torque and when the second hydraulic pump is not limited by control of the second torque control section and the load sensing control section controls the capacity of the second hydraulic pump, and then outputs the modified delivery pressure as a torque control pressure, and a third torque control actuator that receives the torque control pressure and controls the capacity of the first hydraulic pump to decrease the capacity of the first hydraulic pump and decrease the first maximum torque as the torque control pressure becomes higher, the torque feedback circuit includes a fixed restrictor that receives the delivery pressure of the second hydraulic pump, a variable restrictor valve located on a downstream side of the fixed restrictor and connected to a tank in the downstream side thereof, and a pressure limiting valve connected to a hydraulic line between the fixed restrictor and the variable restrictor valve to control the pressure in the hydraulic line such that the pressure does not increase beyond a pressure that initiates the control of the second torque control section, the variable restrictor valve is configured such that the variable restrictor valve is fully closed when the load sensing drive pressure is at a lowest pressure and that the opening area of the variable restrictor valve increases as the load sensing drive pressure rises, and the torque feedback circuit generates the torque control pressure based on the pressure in the hydraulic line between the fixed restrictor and the variable restrictor valve, the torque control pressure being introduced to the third torque control actuator.

In the present invention configured as above, when the second hydraulic pump is not limited by control of the second torque control section and the load sensing control section controls the capacity of the second hydraulic pump (when the delivery pressure of the second hydraulic pump is lower than a pressure that initiates the control of the second torque control section), the pressure in the hydraulic line between the fixed restrictor and the variable restrictor valve increases as the delivery pressure of the second hydraulic pump increases, and decreases as the load sensing drive pressure rises. This variation in the pressure is approximate to variation in the absorption torque of the second hydraulic pump that increases as the delivery pressure of the second hydraulic pump increases and that decreases as the load sensing drive pressure rises (the capacity of the second hydraulic pump decreases), in the case when the second hydraulic pump is not limited by the control of the second torque control section and the load sensing control controls the capacity of the second hydraulic pump. In addition, the torque control pressure is generated based on the pressure in the hydraulic line between the fixed restrictor and the variable restrictor valve, and variation in the torque control pressure is also approximate to variation in the absorption torque of the second hydraulic pump. As a result, the absorption torque of the second hydraulic pump can be accurately detected by a purely hydraulic structure, and the torque feedback circuit can modify the delivery pressure of the second hydraulic pump to provide a characteristic simulating the absorption torque of the second hydraulic pump and can output the modified pressure as a torque control pressure.

Besides, the torque control pressure is introduced to the third torque control actuator and the absorption torque of the second hydraulic pump is fed back to the side of the first hydraulic pump (the one hydraulic pump), whereby the first maximum torque set in the first torque control section of the first hydraulic pump can be decreased by the amount of the absorption torque of the second hydraulic pump, both in the cases of when the second hydraulic pump is limited by control of the second torque control section and operates at the second maximum torque and when the second hydraulic pump is not limited by the control of the second torque control section and the load sensing control section controls the capacity of the second hydraulic pump; accordingly, the total torque control can be carried out accurately and a rated output torque of the prime mover can be utilized effectively. In addition, since the absorption torque of the second hydraulic pump is detected on a purely hydraulic structure basis, the first pump control unit can be miniaturized, and mountability is enhanced.

(2) In the above paragraph (1), preferably, the torque feedback circuit further includes a pressure reduction valve that receives the delivery pressure of the second hydraulic pump as a primary pressure, the pressure in the hydraulic line between the fixed restrictor and the variable restrictor valve is introduced to the pressure reduction valve as a target control pressure for providing a set pressure of the pressure reduction valve, and the pressure reduction valve outputs the delivery pressure of the secondary hydraulic pump as a secondary pressure without reduction when the delivery pressure of the second hydraulic pump is lower than the set pressure, and reduces the delivery pressure of the second hydraulic pump to the set pressure and outputs the thus lowered pressure when the delivery pressure of the second hydraulic pump is higher than the set pressure, the output pressure of the pressure reduction valve being introduced to the third torque control actuator as the torque control pressure.

By thus generating the torque control pressure from the delivery pressure of the second hydraulic pump by the pressure reduction valve, it is possible to secure a flow rate at the time of driving the third torque control actuator by the torque control pressure and to improve the responsiveness at the time of driving the third torque control actuator.

In addition, since the pressure in the hydraulic line between the fixed restrictor and the variable restrictor valve is not directly used as the torque control pressure, the setting of the fixed restrictor and the variable restrictor valve for obtaining a required target control pressure and the setting of the responsiveness of the third torque control actuator can be performed independently, and thus the setting of the torque feedback circuit for exhibiting a required performance can be performed easily and accurately.

Further, since fluctuations in the delivery pressure of the second hydraulic pump are blocked by the pressure reduction valve and therefore do not influence the third torque control actuator when the delivery pressure of the second hydraulic pump is higher than the set pressure of the pressure reduction valve, the stability of the system is secured.

(3) In the above paragraph (1) or (2), preferably, the pressure limiting valve is a relief valve.

Effect of the Invention

According to the present invention, the absorption torque of the second hydraulic pump can be accurately detected by a purely hydraulic structure (torque feedback circuit). Besides, by feeding the absorption torque back to the side of the first hydraulic pump (the one hydraulic pump), it is possible to accurately perform the total torque control and to effectively utilize a rated output torque of the prime mover. In addition, since the absorption torque of the second hydraulic pump is detected on a purely hydraulic basis in this structure, the first pump control unit can be miniaturized, and mountability is enhanced. As a result, it is possible to provide a construction machine that is good in energy efficiency, low in fuel consumption, and is practical.

MODES FOR CARRYING OUT THE INVENTION

Embodiments of the present invention will be described below, referring to the drawings.

FIGS. 1A, 1B and 2are diagrams showing a hydraulic drive system for a hydraulic excavator (construction machine) according to a first embodiment of the present invention.FIG. 1Ais a hydraulic circuit diagram showing the whole of the hydraulic drive system, andFIG. 2is a block diagram showing the whole of the hydraulic drive system.FIG. 1Bis a hydraulic circuit diagram showing the details of a torque feedback circuit shown inFIGS. 1A and 2.

InFIGS. 1A and 2, the hydraulic drive system according to this embodiment includes: a variable displacement first hydraulic pump1ahaving two delivery ports, namely, first and second delivery ports P1and P2; a variable displacement second hydraulic pump1bhaving two delivery ports, namely, third and fourth delivery ports P3and P4; a prime mover2that is connected to the first and second hydraulic pumps1aand1band drives the first and second hydraulic pumps1aand1b; a plurality of actuators3ato3hdriven by hydraulic fluid delivered from the first and second delivery ports P1and P2of the first and second hydraulic pumps1aand hydraulic fluid delivered from the third and fourth delivery ports P3and P4of the second hydraulic pump1b; and a control valve4that is disposed between the first to fourth delivery ports P1to P4of the first and second hydraulic pumps1aand1band the plurality of actuators3ato3hand controls flows of the hydraulic fluid supplied from the first to fourth delivery ports P1to P4of the first and second hydraulic pumps1aand1bto the plurality of actuators3ato3h.

The capacity of the first hydraulic pump1aand the capacity of the second hydraulic pump1bare the same. The capacity of the first hydraulic pump1aand the capacity of the second hydraulic pump1bmay be different.

The first hydraulic pump1ahas a first pump control unit (regulator)5aprovided in common to the first and second delivery ports P1and P2. Similarly, the second hydraulic pump1bhas a second pump control unit (regulator)5bprovided in common to the third and fourth delivery ports P3and P4.

In addition, the first hydraulic pump1ais a split flow type hydraulic pump provided with a single capacity control element (swash plate), and the first pump control unit5adrives the single capacity control element to control the capacity (tilting angle of the swash plate) of the first hydraulic pump1a, thereby controlling delivery flow rates of the first and second delivery ports P1and P2. Similarly, the second hydraulic pump1bis a split flow type hydraulic pump provided with a single capacity control element (swash plate), and the second pump control unit5bdrives the single capacity control element to control the capacity (tilting angle of the swash plate) of the second hydraulic pump1b, thereby controlling delivery flow rates of the third and fourth delivery ports P3and P4.

Each of the first and second hydraulic pumps1aand1bmay be a combination of two variable displacement hydraulic pumps each having a single delivery port. In that case, the two capacity control elements (swash plates) of the two hydraulic pumps of the first hydraulic pump1amay be driven by the first pump control unit5a, and the two capacity control elements (swash plates) of the two hydraulic pumps of the second hydraulic pump1bmay be driven by the second pump control unit5b.

The prime mover2is, for example, a diesel engine. As publicly known, a diesel engine has, for example, an electronic governor, which controls fuel injection amount, whereby revolution speed and torque are controlled. The engine resolution speed is set by operation means such as an engine control dial. The prime mover2may be an electric motor.

The control valve4includes: a plurality of closed center type flow control valves6ato6m; pressure compensating valves7ato7mthat are connected to the upstream side of the flow control valves6ato6mand control differential pressures across meter-in restrictor parts of the flow control valves6ato6m; a first shuttle valve group8athat is connected to load pressure ports of the flow control valves6ato6cand detects a maximum load pressure of the actuators3a,3band3e; a second shuttle valve group8bthat is connected to load pressure ports of the flow control valves6dto6fand detects a maximum load pressure of the actuators3a,3cand3d; a third shuttle valve group8cthat is connected to load pressure ports of the flow control valves6gto6iand detects a maximum load pressure of the actuators3e,3fand3h; a fourth shuttle valve group8dthat is connected to load pressure ports of the flow control valves6jand6mand detects a maximum load pressure of a spare actuator when the spare actuator is connected to the actuators3d,3gand3hand the flow control valve6m; first and second unloading valves10aand10bthat are connected respectively to the delivery ports P1and P2of the first hydraulic pump1a, and that are put into an open state when the delivery pressures of the delivery ports P1and P2become higher than pressures obtained by adding set pressures (unloading pressures) of springs9aand9bto the maximum load pressure detected by the first and second shuttle valve groups8aand8b, so that the hydraulic fluid from the delivery ports P1and P2is returned into a tank, thereby limiting a rise in the delivery pressures; third and fourth unloading valves10cand10dthat are connected respectively to the delivery ports P3and P4of the second hydraulic pump1b, and that are put into an open state when the delivery pressures of the delivery ports P3and P4become higher than pressures obtained by adding set pressures (unloading pressures) of springs9cand9dto the maximum load pressure detected by the third and fourth shuttle valve groups8cand8d, so that the hydraulic fluid from the delivery ports P3and P4is returned into a tank, thereby limiting a rise in the delivery pressures; a first communication control valve15adisposed between respective delivery hydraulic lines of the first and second delivery ports P1and P2of the first hydraulic pump1aand between respective output hydraulic lines of the first and second shuttle valve groups8aand8b; and a second communication control valve15bdisposed between respective delivery hydraulic lines of the third and fourth delivery ports P3and P4of the second hydraulic pump1band between respective output hydraulic lines of the third and fourth shuttle valve groups8cand8d. The set pressures of the springs9ato9dof the first to fourth unloading valves10ato10dare set to be equal to or slightly higher than a target differential pressure in a load sensing control described later.

Besides, though not shown in the drawings, the control valve4includes first and second main relief valves that are connected respectively to the delivery ports P1and P2of the first hydraulic pump1aand function as safety valves, and third and fourth main relief valves that are connected respectively to the delivery ports P3and P4of the second hydraulic pump1band function as safety valves.

The pressure compensating valves6ato6fare configured such that differential pressures between the delivery pressures of the delivery ports P1and P2of the first hydraulic pump1aand the maximum load pressure detected by the first and second shuttle valve groups8aand8bare set as target compensation pressures. The pressure compensating valves7gto7mare configured such that differential pressures between the delivery pressures of the delivery ports P3and P4of the second hydraulic pump1band the maximum load pressure detected by the third and fourth shuttle valve groups8cand8dare set as target compensation pressures. Specifically, the pressure compensating valves7ato7cperform such a control that the delivery pressure of the first delivery port P1is introduced to an opening direction operation side, the maximum load pressure of the actuators3ato3edetected by the first and second shuttle valve groups8aand8bis introduced to a closing direction operation side, and differential pressures across the meter-in restrictor parts of the flow control valves6ato6cbecome equal to the differential pressure between the delivery pressure and the maximum load pressure. The pressure compensating valves7dto7fperform such a control that the delivery pressure of the second delivery port P2is introduced to an opening direction operation side, the maximum load pressure of the actuators3ato3edetected by the first and second shuttle valve groups8aand8bis introduced to a closing direction operation side, and differential pressures across the meter-in restrictor arts of the flow control valves6dto6fbecome equal to the differential pressure between the delivery pressure and the maximum load pressure. The pressure compensating valves7gto7iperform such a control that the delivery pressure of the third delivery port P3is introduced to an opening direction operation side, the maximum load pressure of the actuators3dto3hdetected by the third and fourth shuttle valve groups8cand8dis introduced to a closing direction operation side, and differential pressures across the meter-in restrictor parts of the flow control valves6gto6ibecome equal to the differential pressure between the delivery pressure and the maximum load pressure. The pressure compensating valves7jto7mperform such a control that the delivery pressure of the fourth delivery port P4is introduced to an opening direction operation side, the maximum load pressure of the actuators3dto3hdetected by the third and fourth shuttle valve groups8cand8dis introduced to a closing direction operation side, and differential pressures across the meter-in restrictor parts of the flow control valves6jto6mbecome equal to the differential pressure between the delivery pressure and the maximum load pressure. This structure ensures that at the time of a combined operation of simultaneously driving the plurality of actuators respectively in the first hydraulic pump1aand the second hydraulic pump1b, a distribution of flow rates according to the opening area ratios of the flow control valves can be performed irrespectively of the magnitude of the load pressures of the actuators. In addition, even in a saturation state in which the delivery flow rates of the first to fourth delivery ports P1to P4are deficient, it is possible to reduce the differential pressures across the meter-in restrictor parts of the flow control valves according to the degree of saturation, and thereby to secure good properties for the combined operation.

The plurality of actuators3ato3dare, for example, an arm cylinder, a bucket cylinder, a swing cylinder, and a left travelling motor, respectively, of a hydraulic excavator. The plurality of actuators3eto3hare, for example, a right travelling motor, a swing cylinder, a blade cylinder, and a boom cylinder, respectively.

Here, the arm cylinder3ais connected to the first and second delivery ports P1and P2through the flow control valves6aand6eand the pressure compensating valves7aand7esuch that both the hydraulic fluids delivered from the first and second delivery ports P1and P2of the first hydraulic pump1aare supplied in a joining manner. The boom cylinder3his connected to the third and fourth delivery ports P3and P4through the flow control valves6hand6land the pressure compensating valves7hand7lsuch that both the hydraulic fluids delivered from the third and fourth delivery ports P3and P4of the second hydraulic pump1bare supplied in a joining manner.

The travelling-left travelling motor3dis connected to the second and fourth delivery ports P2and P4through the flow control valves6fand6jand the pressure compensating valves7fand7jsuch that the hydraulic fluid delivered from the second delivery port P2as one delivery port of the first and second delivery ports P1and P2of the first hydraulic pump1aand the hydraulic fluid delivered from the fourth delivery port P4as one of the third and fourth delivery ports P3and P4of the second hydraulic pump1bare supplied in a joining manner. The travelling-right travelling motor3eis connected to the first and third delivery ports P1and P3through the flow control valves6cand6gand the pressure compensating valves7cand7gsuch that the hydraulic fluid delivered from the first delivery port P1as the other delivery port of the first and second delivery ports P1and P2of the first hydraulic pump1aand the hydraulic fluid delivered from the third delivery port P3as the other delivery port of the third and fourth delivery ports P3and P4of the second hydraulic pump1bare supplied in a joining manner.

Besides, the bucket cylinder3bis connected to the first delivery port P1of the first hydraulic pump1athrough the flow control valve6band the pressure compensating valve7bso that the hydraulic fluid delivered from the first delivery port P1is supplied to the bucket cylinder3b. The swing motor3cis connected to the second delivery port P2of the first hydraulic pump1athrough the flow control valve6dand the pressure compensating valve7dso that the hydraulic fluid delivered from the second delivery port P2is supplied to the swing motor3c.

The swing cylinder3fis connected to the third delivery port P3of the second hydraulic pump1bthrough the flow control valve6iand the pressure compensating valve7iso that the hydraulic fluid delivered from the third delivery port P3is supplied to the swing cylinder3f. The blade cylinder3gis connected to the fourth delivery port P4of the second hydraulic pump1bthrough the flow control valve6kand the pressure compensating valve7kso that the hydraulic fluid delivered from the fourth delivery port P4is supplied to the blade cylinder3g.

The flow control valve6mand the pressure compensating valve7mare for use as spare (accessory); for example, in the case where the bucket308is replaced by a crusher, an opening/closing cylinder of the crusher is connected to the fourth delivery port P4through the flow control valve6mand the pressure compensating valve7m.

The first communication control valve15ais in an interruption position of the upper side in the drawing at the time other than the combined operation of simultaneously driving the travelling motors3dand3eand at least one of the other actuators (the boom cylinder3c, the bucket cylinder3b, and the swing motor3c) concerning the first hydraulic pump1a(hereinafter referred to as the time other than the travelling combined operation), and is changed over to a communication position of the lower side in the drawing at the time of the combined operation of simultaneously driving the travelling motors3dand3eand at least one of the other actuators (hereinafter referred to as the time of the travelling combined operation).

The second communication control valve15bis in an interruption position of the upper side in the drawing at the time other than the combined operation of simultaneously driving the travelling motors3dand3eand at least one of the other actuators (the swing cylinder3f, the blade cylinder3g, and the boom cylinder3h) concerning the second hydraulic pump1b(hereinafter referred to as the time other than the travelling combined operation), and is changed over to a communication position of the lower side in the drawing at the time of the combined operation of simultaneously driving the travelling motors3dand3eand at least one of the other actuators (hereinafter referred to as the time of the travelling combined operation).

When the first communication control valve15ais in the interruption position of the upper side in the drawing, it interrupts the communication between respective delivery hydraulic lines of the first and second delivery ports P1and P2of the first hydraulic pump1a, and, when changed over to the communication position of the lower side in the drawing, the first communication control valve15acauses the respective delivery hydraulic lines of the first and second delivery ports P1and P2of the first hydraulic pump1ato communicate with each other.

Similarly, when the second communication control valve15bin the interruption position of the upper side in the drawing, it interrupts the communication between respective delivery hydraulic lines of the third and fourth delivery ports P3and P4of the second hydraulic pump1b, and, when changed over to the communication position of the lower side in the drawing, the second communication control valve15bcauses the respective delivery hydraulic lines of the third and fourth delivery ports P3and P4of the second hydraulic pump1bto communicate with each other.

In addition, the first communication control valve15aincorporates a shuttle valve therein. When in the interruption position of the upper side in the drawing, the first communication control valve15ainterrupts the communication between an output hydraulic line of the first shuttle valve group8aand an output hydraulic line of the second shuttle valve group8b, and causes the respective output hydraulic lines of the first and second shuttle valve groups8aand8bto communicate with the downstream side. When changed over to the communication position of the lower side in the drawing, the first communication control valve15acauses the respective output hydraulic lines of the first and second shuttle valve groups8aand8bto communicate with each other through the shuttle valve, thereby to introduce a maximum load pressure on the high-pressure side to the downstream side.

Similarly, the second communication control valve15bincorporates a shuttle valve therein. When in the interruption position of the upper side in the drawing, the second communication control valve15binterrupts the communication between an output hydraulic line of the third shuttle valve group8cand an output hydraulic line of the fourth shuttle valve group8d, and causes the respective output hydraulic lines of the third and fourth shuttle valve groups8cand8dto communicate with the downstream side. When changed over to the communication position of the lower side in the drawing, the second communication control valve15bcauses the respective output hydraulic lines of the third and fourth shuttle valve groups8cand8dto communicate with each other through the shuttle valve, thereby to introduce a maximum load pressure on the high-pressure side to the downstream side.

When the first communication control valve15ais in the interruption position of the upper side in the drawing, in the side of the first delivery port P1of the first hydraulic pump1a, the maximum load pressure of the actuators3a,3band3edetected by the first shuttle valve group8ais introduced to the first unloading valve10aand the pressure compensating valves7ato7c, so that based on the maximum load pressure, the first unloading valve10alimits a rise in the delivery pressure of the first delivery port P1, and the pressure compensating valves7ato7ccontrol the differential pressures across the meter-in restrictor parts of the flow control valves6ato6c. In the side of the second delivery port P2of the second hydraulic pump1a, the maximum load pressure of the actuators3a,3cand3ddetected by the second shuttle valve group8bis introduced to the second unloading valve10band the pressure compensating valves7dto7f, so that based on the maximum load pressure, the second unloading valve10blimits a rise in the delivery pressure of the second delivery port P2, and the pressure compensating valves7dto7fcontrol the differential pressures across the meter-in restrictor parts of the flow control valves6dto6f.

When the first communication control valve15ais changed over to the communication position of the lower side in the drawing, in the side of the first delivery port P1of the first hydraulic pump1a, the maximum load pressure of the actuators3ato3edetected by the first and second shuttle valve groups8aand8bis introduced to the first unloading valve10aand the pressure compensating valves7ato7c, so that based on the maximum load pressure, the first unloading valve10alimits a rise in the delivery pressure of the first delivery port P1, and the pressure compensating valves7ato7ccontrol the differential pressures across the meter-in restrictor parts of the flow control valves6ato6c. Similarly, in the side of the second delivery port P2of the second hydraulic pump1a, the maximum load pressure of the actuators3ato3edetected by the first and second shuttle valve groups8aand8bis introduced to the second unloading valve10band the pressure compensating valves7dto7f, so that based on the maximum load pressure, the second unloading valve10blimits a rise in the delivery pressure of the second delivery port P2, and the pressure compensating valves7dto7fcontrol the differential pressures across the meter-in restrictor parts of the flow control valves6dto6f.

When the second communication control valve15bis in the interruption position of the upper side in the drawing, in the side of the third delivery port P3of the second hydraulic pump1b, the maximum load pressure of the actuators3e,3fand3hdetected by the third shuttle valve group8cis introduced to the third unloading valve10cand the pressure compensating valves7gto7i, so that based on the maximum load pressure, the third unloading valve10climits a rise in the delivery pressure of the third delivery port P3, and the pressure compensating valves7gto7icontrol the differential pressures across the meter-in restrictor parts of the flow control valves6gto6i. In the side of the fourth delivery port P4of the second hydraulic pump1b, the maximum load pressure of the actuators3d,3gand3hdetected by the fourth shuttle valve group8dis introduced to the fourth unloading vale10dand the pressure compensating valves7jto7m, so that based on the maximum load pressure, the fourth unloading valve10dlimits a rise in the delivery pressure of the fourth delivery port P4, and the pressure compensating valves7jto7mcontrol the differential pressures across the meter-in restrictor parts of the flow control valves6jto6m.

When the second communication control valve15bis changed over to the communication position of the lower side in the drawing, in the side of the third delivery port P3of the second hydraulic pump1b, the maximum load pressure of the actuators3dto3hdetected by the third and fourth shuttle valve groups8cand8dis introduced to the third unloading valve10cand the pressure compensating valves7gto7i, so that based on the maximum load pressure, the third unloading valve10climits a rise in the delivery pressure of the third delivery port P3, and the pressure compensating valves7gto7icontrol the differential pressures across the meter-in restrictor parts of the flow control valves6gto6i. Similarly, in the side of the fourth delivery port P4of the second hydraulic pump1b, the maximum load pressure of the actuators3dto3hdetected by the third and fourth shuttle valve groups8cand8dis introduced to the fourth unloading valve10dand the pressure compensating valves7jto7m, so that based on the maximum load pressure, the fourth unloading valve10dlimits a rise in the delivery pressure of the fourth delivery port P4, and the pressure compensating valves7jto7mcontrol the differential pressures across the meter-in restrictor parts of the flow control valves6jto6m.

The first pump control unit5aincludes: a first load sensing control section12afor controlling the tilting angle of the swash plate (capacity) of the first hydraulic pump1ain such a manner that the delivery pressures of the first and second delivery ports P1and P2of the hydraulic pump1abecome higher by a predetermined pressure than the maximum load pressure of the actuators3ato3edriven by the hydraulic fluids delivered from the first and second delivery ports P1and P2in the plurality of actuators3ato3h; and a first torque control section13afor limiting and controlling the tilting angle of the swash plate (capacity) of the first hydraulic pump1ain such a manner that the absorption torque of the first hydraulic pump1adoes not exceed a predetermined value.

The second pump control unit5bincludes: a second load sensing control section12bfor controlling the tilting angle of the swash plate (capacity) of the second hydraulic pump1bin such a manner that the delivery pressures of the third and fourth delivery ports P3and P4of the second hydraulic pump1bbecome higher by a predetermined angle than the maximum load pressure of the actuators3dto3hdriven by the hydraulic fluids delivered from the third and fourth delivery ports P3and P4in the plurality of actuators3ato3h; and a second torque control section13bfor limiting and controlling the tilting angle of the swash plate (capacity) of the second hydraulic pump1bin such a manner that the absorption torque of the second hydraulic pump1bdoes not exceed a predetermined value.

The first load sensing control section12aincludes: load sensing control valves16aand16bfor generating load sensing drive pressures (hereinafter referred to as LS drive pressures); a low pressure selection valve21afor selecting and outputting the lower pressure side of the LS drive pressures generated by the load sensing control valves16aand16b; and a load sensing control piston (load sensing control actuator)17ato which the LS drive pressure selected and outputted by the low pressure selection valve21ais introduced and which varies the tilting angle of the swash plate of the first hydraulic pump1aaccording to the LS drive pressure.

The second load sensing control section12bincludes: load sensing control valves16cand16dfor generating load sensing drive pressures (hereinafter referred to as LS drive pressures); a low pressure selection valve21bfor selecting and outputting a lower pressure side of the LS drive pressures generated by the load sensing control valves16cand16d; and a load sensing control piston (load sensing control actuator)17bto which the LS drive pressure selected and outputted by the low pressure selection valve21bis introduced and which varies the tilting angle of the swash plate of the second hydraulic pump1baccording to the LS drive pressure.

In the first load sensing control section12a, a control valve16aincludes: a spring16a1for setting a target differential pressure for a load sensing control; a pressure receiving part16a2which is located opposite to the spring16a1and to which the delivery pressure of the first delivery port P1is introduced; and a pressure receiving part16a3located on the same side as the spring16a1. When the first communication control valve15ais in the interruption position of the upper side in the drawing, the maximum load pressure of the actuators3a,3band3edetected by the first shuttle valve group8ais introduced to the pressure receiving part16a3of the control valve16a. When the first communication control valve15ais changed over to the communication position of the lower side in the drawing, the maximum load pressure of the actuators3ato3edetected by the first and second shuttle valve groups8aand8bis introduced to the pressure receiving part16a3of the control valve16a. The control valve16ais displaced according to the balance among the delivery pressure of the first delivery port P1introduced to the pressure receiving part16a2, the maximum load pressure of the actuators3a,3band3eor the actuators3ato3eintroduced to the pressure receiving part16a3, and a biasing force of the spring16a1, thereby to vary the LS drive pressure.

In other words, when the delivery pressure of the first delivery port P1introduced to the pressure receiving part16a2becomes higher than a pressure obtained by adding the target differential pressure (predetermined pressure) set by the spring16a1to the maximum load pressure introduced to the pressure receiving part16a2, the control valve16ais moved leftward in the drawing to cause its secondary port to communicate with a hydraulic fluid source (the first delivery port P1), thereby raising the LS drive pressure. When the delivery pressure on the high pressure side of the first delivery port P1introduced to the pressure receiving part16a2becomes lower than a pressure obtained by adding the target differential pressure (predetermined pressure) set by the spring16a1to the maximum load pressure introduced to the pressure receiving part16a2, the control valve16ais moved rightward in the drawing to cause the secondary port to communicate with the tank, thereby lowering the LS drive pressure. The hydraulic fluid source that the secondary port communicates with when the control valve16ais moved leftward in the drawing may be a pilot hydraulic fluid source that is formed in a delivery hydraulic line of a pilot pump and generates a fixed pilot pressure.

The control valve16bincludes: a spring16b1for setting a target differential pressure for a load sensing control; a pressure receiving part16b2which is located opposite to the spring16b1and to which the delivery pressure of the second delivery port P2is introduced; and a pressure receiving part16b3located on the same side as the spring16b1. When the first communication control valve15ais situated in the interruption position of the upper side in the drawing, the maximum load pressure of the actuators3a,3cand3ddetected by the second shuttle valve group8bis introduced to the pressure receiving part16b3of the control valve16b. When the first communication control valve15ais changed over to the communication position of the lower side in the drawing, the maximum load pressure of the actuators3ato3edetected by the first and second shuttle valve groups8aand8bis introduced to the pressure receiving part16a3of the control valve16b. The control valve16bis displaced according to the balance among the delivery pressure of the second delivery port P2introduced to the pressure receiving part16b2, the maximum load pressure of the actuators3a,3cand3dor the actuators3ato3eintroduced to the pressure receiving part16b3, and the biasing force of the spring16b1, thereby varying the LS drive pressure, like the control valve16a.

The low pressure selection valve21aselects the lower pressure side of the LS drive pressures generated by the load sensing control valves16aand16b, and outputs the selected LS drive pressure to the load sensing control piston17a. Based on the LS drive pressure, the load sensing control piston17avaries the tilting angle of the swash plate of the first hydraulic pump1a, and thereby varies the delivery flow rates of the first and second delivery ports P1and P2.

In the second load sensing control section12b, the control valve16cincludes: a spring16c1for setting a target differential pressure for a load sensing control; a pressure receiving part16c2which is located opposite to the spring16c1and to which the delivery pressure of the third delivery port P3is introduced; and a pressure receiving part16c3located on the same side as the spring16c1. When the second communication control valve15bis located in the interruption position of the upper side in the drawing, the maximum load pressure of the actuators3e,3fand3hdetected by the third shuttle valve group8cis introduced to the pressure receiving part16c3of the control valve16c. When the second communication control valve15bis changed over to the communication position of the lower side in the drawing, the maximum load pressure of the actuators3dto3hdetected by the third and fourth shuttle valve groups8cand8dis introduced to the pressure receiving part16c3of the control valve16c. The control valve16cis displaced according to the balance among the delivery pressure of the third delivery port P3introduced to the pressure receiving part16c2, the maximum load pressure of the actuators3e,3fand3hor the actuators3dto3hintroduced to the pressure receiving part16c3, and a biasing force of the spring16c1, thereby varying the LS drive pressure, like the control valve16a.

The control valve16dincludes: a spring16d1for setting a target differential pressure for a load sensing control; a pressure receiving part16d2which is located opposite to the spring16d1and to which the delivery pressure of the fourth delivery port P4is introduced; and a pressure receiving part16dlocated on the same side as the spring16d1. When the second communication control valve15bis located in the interruption position of the upper side in the drawing, the maximum load pressure of the actuators3d,3gand3hdetected by the fourth shuttle valve group8dis introduced to the pressure receiving part16d3of the control valve16d. When the second communication control valve15bis changed over to the communication position of the lower side in the drawing, the maximum load pressure of the actuators3dto3hdetected by the third and fourth shuttle valve groups8cand8dis introduced to the pressure receiving part16d3of the control valve16d. The control valve16dis displaced according to the balance among the delivery pressure of the fourth delivery port P4introduced to the pressure receiving part16d2, the maximum load pressure of the actuators3d,3gand3hor the actuators3dto3hintroduced to the pressure receiving part16d3, and a biasing force of the spring16d1, thereby varying the LS drive pressure, like the control valve16a.

The low pressure selection valve21bselects the lower pressure side of the LS drive pressures generated by the load sensing control valves16cand16d, and outputs the selected LS drive pressure to the load sensing control piston17b. Based on the LS drive pressure, the load sensing control piston17bvaries the tilting angle of the swash plate of the second hydraulic pump1b, and thereby varies the delivery flow rates of the third and fourth delivery ports P3and P4.

FIG. 3is a diagram showing the relation between LS drive pressures and tilting angles of swash plates of the first and second hydraulic pumps1aand1bwhen the load sensing control pistons17aand17boperate. In the diagram, the LS drive pressures acting on the load sensing control pistons17aand17bare denoted by Px1and px2, and the tilting angles of the swash plates of the first and second hydraulic pumps1aand1bare denoted by q1and q2.

As shown inFIG. 3, when the LS drive pressure Px1rises, the load sensing control piston17areduces the tilting angle q1of the swash plate of the first hydraulic pump1a, thereby decreasing the delivery flow rates of the first and second delivery ports P1and P2. When the LS drive pressure Px1is lowered, the load sensing control piston17aenlarges the tilting angle q1of the swash plate of the first hydraulic pump1a, thereby increasing the delivery flow rates of the first and second delivery ports P1and P2. With such arrangement, the first load sensing control section12acontrols the tilting angle of the swash plate (capacity) of the first hydraulic pump1ain such a manner that the delivery pressure on the high pressure side of the first and second delivery ports P1and P2of the first hydraulic pump1abecomes higher by a predetermined pressure than the maximum load pressure of the actuators3ato3edriven by the hydraulic fluids delivered from the first and second delivery ports P1and P2. In the diagram, K is the rate of change of the tilting angle q1of the swash plate of the first hydraulic pump1ain relation to the LS drive pressure Px1, and is a value determined by the relation between constants of springs S3and S4described later and the tilting angle q2(capacity) of the second hydraulic pump1b.

Like the load sensing control piston17a, the load sensing control piston17bvaries the tilting angle q2of the swash plate of the second hydraulic pump1bin accordance with variation in the LS drive pressure Px2, thereby to control the tilting angle of the swash plate (capacity) of the second hydraulic pump1bin such a manner that the delivery pressure on the high pressure side of the third and fourth delivery ports P3and P4of the second hydraulic pump1bbecomes higher by a predetermined pressure than the maximum load pressure of the actuators3dto3hdriven by the hydraulic fluids delivered from the third and fourth delivery ports P3and P4.

In the first and second load sensing control sections12and12b, the target differential pressures for the load sensing control that are set by the springs16a1and16b1and the springs16c1and16d1are each, for example, about 2 MPa.

Besides, in the first pump control unit5a, the first torque control section13aincludes: a first torque control piston (first torque control actuator)18ato which the delivery pressure of the first delivery port P1is introduced; a second torque control piston (first torque control actuator)19ato which the delivery pressure of the second delivery port P2is introduced; and springs S1and S2(inFIG. 1, only one spring is illustrated for simplification) as biasing means for setting a maximum torque T1max (first maximum torque).

The second torque control section13bincludes: a third torque control piston (second torque control actuator)18bto which the delivery pressure of the third delivery port P3is introduced; a fourth torque control piston (second torque control actuator)19bto which the delivery pressure of the fourth delivery port P4is introduced; and springs S3and S4(inFIG. 1, only one spring is illustrated for simplification) as biasing means for setting a maximum torque T2max (second maximum torque).

In addition, the first torque control section13aincludes: a torque feedback circuit30to which the delivery pressures of the third and fourth delivery ports P3and P4of the second hydraulic pump1band the LS drive pressure acting on the load sensing control piston17bof the second load sensing control section12bare introduced, which modifies the delivery pressures of the third and fourth delivery ports P3and P4of the second hydraulic pump1bbased on the delivery pressures of the third and fourth delivery ports P3and P4and the LS drive pressure to provide a characteristic simulating the absorption torque of the second hydraulic pump1bboth in the cases of when the second hydraulic pump1bis limited by control of the second torque control section13band operates at the maximum torque T2max (second maximum torque) and when the second hydraulic pump1bis not limited by the control of the second torque control section13band the second load sensing control section12bcontrols the capacity of the second hydraulic pump1b(when lower than a starting pressure Pb of an absorption torque constant control of the second hydraulic pump1bdescribed later), and which outputs the modified pressures; a first torque reduction control piston (third torque control actuator)31ato which an output pressure of the torque feedback circuit30obtained by modification of the delivery pressure of the third delivery port P3of the second hydraulic pump1bis introduced, and which, as the output pressure rises, decreases the tilting angle of swash plate (capacity) of the first hydraulic pump1aand decreases the maximum torque T1max set by the springs S1and S2; and a second torque reduction control piston (third torque control actuator)31bto which an output pressure of the torque feedback circuit30obtained by modification of the delivery pressure of the fourth delivery port P4of the second hydraulic pump1bis introduced, and which, as the output pressure rises, decreases the tilting angle of swash plate (capacity) of the first hydraulic pump1aand decreases the maximum torque T1max set by the springs S1and S2.

FIG. 4Ais a torque control diagram for the first torque control section13a, andFIG. 4Bis a torque control diagram for the second torque control section13b. In these torque control diagrams, the axis of ordinates represents the tilting angle (capacity) q1, q2, and these diagrams are turned to be horsepower control diagrams when the axis of ordinates is replaced by delivery flow rate Q1, Q2or delivery flow rate Q3, Q4. Besides, the axis of abscissas represents pump delivery pressure; specifically, the axis of abscissas represents average delivery pressure (P1p+P2p/2) of the first and second delivery ports P1and P2inFIG. 4A, and represents average delivery pressure (P3p+P4p/2) of the third and fourth delivery ports P3and P4inFIG. 4B.

InFIG. 4A, when the hydraulic oil delivered by the second hydraulic pump1bis not supplied to the actuators3dto3h, the torque feedback circuit30and the first and second torque reduction control pistons31aand31bdo not function, and the maximum torque T1max is set in the first torque control section13aby the springs S1and S2. TP1a and TP1b are characteristic curves of the springs S1and S2for setting the maximum torque T1max.

In this condition, when the hydraulic fluid delivered by the first hydraulic pump1ais supplied to one of the actuators3ato3econcerning the first hydraulic pump1aand the average delivery pressure of the first and second delivery ports P1and P2rises, the first torque control section13adoes not operate during when the average delivery pressure is not more than a pressure (torque control start pressure) Pa at a starting end of the characteristic curve TP1a. In this case, the tilting angle of swash plate (capacity) q1of the first hydraulic pump1ais not limited by the control of the first torque control section13a, and can be increased to the maximum tilting angle q1max possessed by the first hydraulic pump1aaccording to an operation amount of a control lever device (demanded flow rate), under the control of the first load sensing control section12a.

When the average delivery pressure of the first and second delivery ports P1and P2exceeds Pa in a condition where the swash plate of the first hydraulic pump1ais at the maximum tilting angle q1max, the first torque control section13aoperates to perform an absorption torque constant control (or horsepower constant control) so as to decrease the maximum tilting angle (maximum capacity) of the first hydraulic pump1aalong the characteristic curves TP1a and TP1b as the average delivery pressure rises. In this case, the first load sensing control section12acannot increase the tilting angle of the first hydraulic pump1ain excess of a tilting angle determined by the characteristic curves TP1a and TP1b.

As shown in the diagram, the characteristic curves TP1a and TP1b are set to be approximate to an absorption torque constant curve (hyperbola) TP1by the two springs S1and S2. With such setting, the first torque control section13aperforms the absorption torque constant control (or horsepower constant control) such that the absorption torque of the first hydraulic pump1adoes not exceed the maximum torque T1max when the average delivery pressure of the first hydraulic pump1arises. The maximum torque T1max is set to be slightly lower than a rated output torque TER of an engine2.

InFIG. 4B, a maximum torque T2max is set in the second torque control section13bby the springs S3and S4, irrespectively of the operating conditions of the first hydraulic pump1a. TP2a and TP2b are characteristic curves of the springs S3and S4for setting the maximum torque T1max.

When the hydraulic fluid delivered by the second hydraulic pump1bis supplied to one of the actuators3dto3hconcerning the second hydraulic pump1band the average delivery pressure of the third and fourth delivery ports P3and P4rises, the second torque control section13bdoes not operate while the average delivery pressure is not more than a pressure (torque control start pressure) Pb at a starting end of the characteristic curve TP2a. In this case, the tilting angle of swash plate (capacity) q2of the second hydraulic pump1bis not limited by control of the second torque control section13b, and the tilting angle can be increased to a maximum tilting angle q2max possessed by the second hydraulic pump1baccording to an operation amount of the control lever device (demanded flow rate), under control of the second load sensing control section12b.

When the average delivery pressure of the third and fourth delivery ports P3and P4exceeds Pb in a condition where the swash plate of the second hydraulic pump1bis at the maximum tilting angle q2max, the second torque control section13boperates to perform an absorption torque constant control so as to decrease the maximum tilting angle (maximum capacity) of the second hydraulic pump1balong the characteristic curves TP2a and TP2b as the average delivery pressure rises. In this case, the second load sensing control section12bcannot increase the tilting angle of the second hydraulic pump1bin excess of a tilting angle determined by the characteristic curves TP2a and TP2b.

As shown in the diagram, the characteristic curves TP2a and TP2b are set to be approximate to an absorption torque constant curve (hyperbola) TP2by the two springs S3and S4. With such setting, the second torque control section13bperforms an absorption torque constant control (or horsepower constant control) such that the absorption torque of the second hydraulic pump1bdoes not exceed the maximum torque T2max when the average delivery pressure of the second hydraulic pump1brises. The maximum torque T2max is lower than the maximum torque T1max set in the first torque control section13a, and is set to be about ½ times the rated output torque TER of the engine2.

In addition, when the hydraulic fluid delivered by the second hydraulic pump1bis supplied to one of the actuators3dto3hconcerning the second hydraulic pump1band the one of the actuators3dto3his driven by the hydraulic fluid delivered by the second hydraulic pump1b, the torque feedback circuit30modifies the delivery pressures of the third and fourth delivery ports P3and P4of the second hydraulic pump1bso as to attain a characteristic simulating the absorption torque of the second hydraulic pump1b, and outputs the modified delivery pressures. In addition, the first and second torque reduction control pistons31aand31bdecrease the maximum torque T1max set in the first torque control section13aas the output pressure of the torque feedback circuit30rises.

InFIG. 4A, the two arrows R1and R2represent the effects of the first and second torque reduction control pistons31aand31bto decrease the maximum torque T1max. When the delivery pressures of the third and fourth delivery ports P3and P4of the second hydraulic pump1brise and when the absorption torque of the second hydraulic pump1bin that instance is T2which is lower than the maximum torque T2max and the absorption torque simulated by the torque feedback circuit30is T2s (≈T2max), the torque feedback pistons32aand32bdecrease the maximum torque T1max to T1max−T2s, as indicated by the arrow R1inFIG. 4A. In addition, when the absorption torque of the second hydraulic pump1bis the maximum torque T2max and the absorption torque simulated by the torque feedback circuit30is T2maxs (≈T2max), the torque feedback pistons32aand32bdecrease the maximum torque T1max to T1max−T2maxs, as indicated by the arrow R2inFIG. 4A.

Here, the maximum torque T1max set in the first torque control section13ais lower than the rated output torque TER of the engine2, as aforementioned. In addition, when the hydraulic fluid delivered by the second hydraulic pump1bis not supplied to the actuators3dto3hand the hydraulic fluid delivered by the first hydraulic pump1ais supplied to one of the actuators3ato3eto drive the one of the actuators3ato3e, the first torque control section13aperforms an absorption torque constant control (or horsepower constant control) such that the absorption torque of the first hydraulic pump1adoes not exceed the maximum torque T1max, whereby the absorption torque of the first hydraulic pump1ais controlled not to exceed the rated output torque TER of the engine2. With such arrangement, stoppage of the engine2(engine stall) can be prevented, while making the most of the rated output torque TER of the engine2.

In addition, when the hydraulic fluid delivered by the second hydraulic pump1bis supplied to one of the actuators3dto3hand the one of the actuators3dto3his driven by the hydraulic fluid delivered by the second hydraulic pump1b, the torque feedback pistons32aand32bdecrease the maximum torque T1max to T1max−T2s or T1max−T2maxs, as indicated by the arrow X inFIG. 4A, as aforementioned. With such arrangement, also in a combined operation of simultaneously driving one of the actuators3ato3econcerning the first hydraulic pump1aand one of the actuators3dto3hconcerning the second hydraulic pump1b, a total torque control is conducted such that the total absorption torque of the first hydraulic pump1aand the second hydraulic pump1bdoes not exceed the rated output torque TER of the engine2. In this case, also, stoppage of the engine2(engine stall) can be prevented, while making the most of the rated output torque TER of the engine2.

FIG. 1Bis a diagram showing the details of the torque feedback circuit30.

The torque feedback circuit30includes: a first torque feedback circuit section30athat modifies the delivery pressure of the third delivery port P3of the second hydraulic pump1bso as to attain a characteristic simulating the absorption torque of the second hydraulic pump1b, and outputs the modified delivery pressure; and a second torque feedback circuit section30bthat modifies the delivery pressure of the fourth delivery port P4of the second hydraulic pump1bso as to attain a characteristic simulating the absorption torque of the second hydraulic pump1b, and outputs the modified delivery pressure.

The first torque feedback circuit section30aincludes: a first torque pressure reduction valve32ato which the delivery pressure of the third delivery port P3is introduced; and a first pressure dividing circuit33athat generates a target control pressure for setting a set pressure of the first torque pressure reduction valve32a. When the delivery pressure of the third delivery port P3is lower than the set pressure, the first torque pressure reduction valve32aoutputs the delivery pressure of the third delivery port P3as a secondary pressure without reduction, whereas when the delivery pressure of the third delivery port P3is higher than the set pressure, the first torque pressure reduction valve32areduces the delivery pressure of the third delivery port P3to the set pressure (target control pressure) and outputs the thus reduced pressure. The output pressure (secondary pressure) is introduced to the first torque reduction control piston31aas a torque control pressure.

The first pressure dividing circuit33aincludes: a first pressure dividing restrictor part34ato which the delivery pressure of the third delivery port P3is introduced; a first pressure dividing valve35alocated on a downstream side of the first pressure dividing restrictor part34a; and a first relief valve (pressure limiting valve)37athat is connected to a first hydraulic line36abetween the first pressure dividing restrictor part34aand the first pressure dividing valve35aand causes the pressure in the first hydraulic line36anot to increase beyond a set pressure (relief pressure). The first pressure dividing restrictor part34ais a fixed restrictor, and has a fixed opening area. The first pressure dividing valve35ais a variable restrictor valve to which an LS drive pressure Px2acting on the load sensing control piston17bof the second load sensing control section12bis introduced and which varies the opening area according to the LS drive pressure Px2. When the LS drive pressure Px2is a tank pressure, the opening area of the first pressure dividing valve35ais zero (fully closed). As the LS drive pressure Px2rises, the opening area of the first pressure dividing valve35aincreases. When the LS drive pressure Px2rises to be equal to or higher than a predetermined pressure, the opening area of the first pressure dividing valve35abecomes maximum (fully opened). The target control pressure generated in the first hydraulic line36abetween the first pressure dividing restrictor34aand the first pressure dividing valve35aaccording to the variation in the opening area of the first pressure dividing valve35avaries continuously from the set pressure of the first relief valve37ato the tank pressure (zero). According to the variation in the target control pressure, a torque control pressure generated by the first torque pressure reduction valve32ais also varied continuously. The set pressure of the first relief valve37ais set to be equal to a torque control start pressure Pb (FIG. 4B) of the second torque control section13b, in conformity with Pb.

The second torque feedback circuit section30balso is configured similarly to the first torque feedback circuit section30a. Specifically, the second torque feedback circuit section30bincludes: a second torque pressure reduction valve32bto which the delivery pressure of the fourth delivery port P4is introduced as a primary pressure; and a second pressure dividing circuit33bthat generates a target control pressure for providing a set pressure of the second torque pressure reduction valve32b. When the delivery pressure of the fourth delivery port P4is lower than the set pressure, the second torque pressure reduction valve32boutputs the delivery pressure of the fourth delivery port P4as a secondary pressure without reduction. When the delivery pressure of the fourth delivery port P4is higher than the set pressure, the second torque pressure reduction valve32breduces the delivery pressure of the fourth delivery port P4to the set pressure (target control pressure), and outputs the reduced pressure. The output pressure (secondary pressure) is introduced to the second torque reduction control piston31bas a torque control pressure.

The second pressure dividing circuit33bincludes: a second pressure dividing restrictor part34bto which the delivery pressure of the fourth delivery port P4is introduced; a second pressure dividing valve35blocated on a downstream side of the second pressure dividing restrictor part34b; and a second relief valve (pressure limiting valve)37bthat is connected to a second hydraulic line36bbetween the second pressure dividing restrictor part34band the second pressure dividing valve35band causes the pressure in the second hydraulic line36bnot to increase beyond a set pressure (relief pressure). The second pressure dividing restrictor part34bis a fixed restrictor, and has a fixed opening area. The second pressure dividing valve35bis a variable restrictor valve to which the LS drive pressure Px2acting on the load sensing control piston17bof the second load sensing control section12bis introduced, and which varies the opening area according to the LS drive pressure Px2. When the LS drive pressure Px2is the tank pressure, the opening area of the first pressure dividing valve35ais zero (fully closed). As the LS drive pressure Px2rises, the opening area of the first pressure dividing valve35aincreases. When the LS drive pressure Px2rises to be equal to or higher than a predetermined pressure, the opening area of the first pressure dividing valve35abecomes maximum (fully opened). A target control pressure generated in the second hydraulic line36bbetween the second pressure dividing restrictor34band the second pressure dividing valve35baccording to the variation in the opening area of the second pressure dividing valve35bvaries continuously from the set pressure of the second relief valve37bto the tank pressure (zero). According to the variation in the target control pressure, a torque control pressure generated by the second torque pressure reduction valve32bis also varied continuously. The set pressure of the second relief valve37bis set to be equal to a torque control start pressure Pb (FIG. 4B) of the second torque control section13b, in conformity with Pb.

FIG. 5Ais a diagram showing the relation between the LS drive pressure Px2and the opening area of the first and second pressure dividing valves35aand35b;FIG. 5Bis a diagram showing the relation between the opening area of the first and second pressure dividing valves35aand35band a target control pressure;FIG. 5Cis a diagram showing the relation between the delivery pressure of the third and fourth delivery ports and the target control pressure when the LS drive pressure Px2varies; andFIG. 5Dis a diagram showing the relation between the delivery pressure of the third and fourth delivery ports and a torque control pressure when the LS drive pressure Px2varies. In the diagrams, AP3and AP4are opening areas of the first and second pressure dividing valves35aand35b; P3tref and P4tref are the target control pressures generated in the first and second hydraulic lines36aand36b; P3p and P4p are delivery pressures of the third and fourth delivery ports; and P3t and P4t are the torque control pressures generated by the first and second torque pressure reduction valves32aand32b.

As shown inFIG. 5A, when the LS drive pressure Px2acting on the load sensing control piston17bof the second load sensing control section12bis the tank pressure, the opening areas AP3and AP4of the first and second pressure dividing valves35aand35bare zero (fully closed). As the LS drive pressure Px2rises, the opening areas AP3and AP4of the first and second pressure dividing valves35aand35bincrease. When the LS drive pressure Px2rises to be equal to or higher than a predetermined pressure Px2a, the opening areas of the first and second pressure dividing valves35aand35bbecome maximum (fully opened).

As shown inFIG. 5B, when the opening areas AP3and AP4of the first and second pressure dividing valves35aand35bare zero (fully closed), the pressures in the first and second hydraulic lines36aand36bare equal to the delivery pressures P3p and P4p of the third and fourth delivery ports. It is to be noted, however, that the pressures in the first and second hydraulic lines36aand36bcannot become equal to or higher than the set pressures of the first and second relief valves37aand37b. As the opening areas AP3and AP4of the first and second pressure dividing valves35aand35bincrease from the zero (fully closed), the target control pressures P3tref and P4tref are lowered. When the opening areas AP3and AP4of the first and second pressure dividing valves35aand35bbecome maximum APmax (fully opened), the target control pressures P3tref and P4tref become the tank pressure (zero).

As shown inFIG. 5C, when the LS drive pressure is the tank pressure (zero), the opening areas AP3and AP4of the first and second pressure dividing valves35aand35bare zero (fully closed), and the target control pressures P3tref and P4tref are equal to the delivery pressures of the third and fourth delivery ports. As a result, when the delivery pressures of the third and fourth delivery ports rise, the target control pressures P3tref and P4tref also rise while remaining equal to the delivery pressures of the third and fourth delivery ports. The gradients of straight lines representing the rates of rise in the target control pressures P3tref and P4tref in this instance are 1. When the delivery pressures of the third and fourth delivery ports reach the set pressures of the first and second relief valves37aand37b, the target control pressures P3tref and P4tref become constant at the set pressures of the first and second relief valves37aand37b.

When the LS drive pressure rises from the tank pressure, the opening areas AP3and AP4of the first and second pressure dividing valves35aand35bincrease accordingly. As the delivery pressures of the third and fourth delivery ports rise, the target control pressures P3tref and P4tref rise at smaller rates (with smaller gradients of straight lines) as compared to the case where the opening areas AP3and AP4of the first and second pressure dividing valves35aand35bare zero (fully closed). As the LS drive pressure rises, the rates of rise (gradients of straight lines) in the target control pressures P3tref and P4tref are reduced, and the target control pressures P3tref and P4tref obtained at the same delivery pressures of the third and fourth delivery ports are lowered. When the delivery pressures of the third and fourth delivery ports reach the torque control start pressure Pb which is the set pressure of the first and second relief valves37aand37b, the target control pressures P3tref and P4tref become constant at the set pressure (Pb) of the first and second relief valves37aand37b.

When the LS drive pressure rises to a predetermined pressure Px2, the opening areas AP3and AP4of the first and second pressure dividing valves35aand35bbecome a max APmax (fully opened), and the target control pressures P3tref and P4tref become the tank pressure (zero).

As a result of that the target control pressures P3tref and P4tref thus vary when the delivery pressures of the third and fourth delivery ports rise, the torque control pressures P3t and P4t also vary like the target control pressures P3tref and P4tref, as illustrated inFIG. 5D. Specifically, when the LS drive pressure is the tank pressure (zero), the torque control pressures P3t and P4t are equal to the delivery pressures of the third and fourth delivery ports. As the LS drive pressure rises, the rates of rise (gradients of straight lines) in the torque control pressures P3t and P4t are reduced, and the torque control pressures P3t and P4t obtained at the same delivery pressures of the third and fourth delivery ports are lowered. When the delivery pressures of the third and fourth delivery ports reach the torque control start pressure Pb which is a set pressure of the first and second relief valves37aand37b, the torque control pressures P3t and P4t become constant at the set pressure (Pb) of the first and second relief valves37aand37b. When the LS drive pressure reaches a predetermined pressure Px2, the torque control pressures P3t and P4t become the tank pressure (zero).

It will be explained below that the torque control pressures P3t and P4t generated by the torque feedback circuit sections30aand30bare characteristics simulating the absorption torque of the second hydraulic pump1bas aforementioned.

In the second pump control unit5bshown inFIGS. 1A and 1B, assuming that the actual absorption torques of the third and fourth delivery ports P3and P4of the second hydraulic pump1bare τ3and τ4, the absorption torques τ3and τ4are calculated according to the following equations.
τ3=(P3p×q2)/2π  (1)
τ4=(P4p×q2)/2π  (2)

As aforementioned, P3p and P4p are the delivery pressures of the third and fourth delivery ports P3and P4, and q2is the tilting angle of the second hydraulic pump1b.

In addition, in the case when limitation by the absorption torque constant control (or horsepower constant control) of the second torque control section13bis not received, the tilting angle of the second hydraulic pump1is controlled by the second load sensing control section12b. In this instance, the swash plate of the second hydraulic pump1breceives the LS drive pressure Px2and springs S3and S4, and the tilting angle q2is expressed by the following equation.
q2=q2max−K×Px2  (3)

Here, K is a constant determined by the relation between the constants of the springs S3and S4and the tilting angle q2(capacity) of the second hydraulic pump1b, and is a value corresponding to the gradient K shown inFIG. 3.

On the other hand, in order to cause the torque control pressures P3t and P4t to be characteristics simulating the absorption torque of the second hydraulic pump1b, it is necessary that biasing forces generated at the first and second torque reduction control pistons31aand31bby application of the torque control pressures P3t and P4t should be values proportional to the absorption torques τ3and τ4of the third and fourth delivery ports P3and P4, and for ensuring this, the following relations must be established.
τ3=C(A×P3t)  (4)
τ4=C(A×P4t)  (5)

Here, A is a pressure-receiving area of the first and second torque reduction control pistons31aand31b, and C is a proportionality factor.

From the equations (1) to (5) above, the torque control pressures P3t and P4t are expressed by the following equations.
τ3=(P3p×(q2max−K×Px2))/2π=C(A×P3t)
τ4=(P4p×(q2max−K×Px2))/2π=C(A×P4t)

Deformation of these gives the following equations.
P3t=((P3p×(q2max−K×Px2))/2π)/C×A
P4t=((P4p×(q2max−K×Px2))/2π)/C×A

Setting the values of A and C such that D×q2max is 1 gives the following equations.
P3t=P3p×(1−(K×Px2/D))  (6)
P4t=P4p×(1−(K×Px2/D))  (7)

FIG. 6is a diagram showing relations among the delivery pressures P3p and P4p of the third and fourth delivery ports, the torque control pressures P3t and P4t, and the LS drive pressure Px2expressed by the equations (6) and (7).

As shown inFIG. 6, when the LS drive pressure Px2is the tank pressure (zero) in the equations (6) and (7), the torque control pressures P3t and P4t are the same as the delivery pressures P3p and P4p of the third and fourth delivery ports. Besides, as the LS drive pressure Px2rises, the value of (1−(K×Px2/D)) which is the gradients of straight lines representing the rates of rise in the torque control pressures P3t and P4t is reduced, and the torque control pressures P3t and P4t obtained at the same delivery pressures P3p and P4p of the third and fourth delivery ports are lowered. When the delivery pressures of the third and fourth delivery ports rise to the torque control start pressure Pb, the absorption torque constant control (or horsepower constant control) of the second torque control section13bis started, and the absorption torque of the second hydraulic pump1bbecomes constant. Therefore, it is sufficient to set the torque control pressures P3t and P4t to be also constant at the torque control start pressure Pb.

As seen fromFIGS. 5D and 6, the rates of increase (gradients of straight lines) of the torque control pressures P3t and P4t when the delivery pressures P3p and P4p of the third and fourth delivery ports rise as shown inFIG. 5Dvary in such a manner as to be reduced as the LS drive pressure Px3rises, like the rates of increase (gradients of straight lines) of the torque control pressures P3t and P4t when the delivery pressures P3p and P4p of the third and fourth delivery ports rise as shown inFIG. 6. When the torque control pressures P3t and P4t reach the torque control start pressure Pb which is a set pressure of the first and second relief valves37aand37b, the rates of increase (gradients of straight lines) become at the set pressure (Pb).

In this way, the torque control pressures P3t and P4t generated by the torque feedback circuit sections30aand30bare characteristics simulating the absorption torque of the second hydraulic pump1b. The torque feedback circuit sections30aand30bhave the function of modification, and outputting, the delivery pressure of a main pump202in such a manner as to provide characteristics simulating the absorption torque of the main pump202both in the cases of when the second hydraulic pump1bis limited by control of the second torque control section13band operates at a maximum torque T2max (second maximum torque) and when the second hydraulic pump1bis not limited by the second torque control section13band the second load sensing control section12bcontrols the capacity of the second hydraulic pump1b(when lower than the start pressure Pb of the absorption torque constant control).

FIG. 7shows an external appearance of a hydraulic excavator.

InFIG. 7, the hydraulic excavator includes an upper swing structure300, a lower track structure301, and a front work device302. The upper swing structure300is swingably mounted on the lower track structure301, and the front work device302is connected to a front end portion of the upper swing structure300through a swing post303in such a manner as to rotate upward and downward and leftward and rightward. The lower track structure301includes left and right crawlers310and311, and is provided on the front side of a track frame304with an earth removing blade305which is movable up and down. The upper swing structure300includes a cabin (operating room)300a, in which are provided control lever devices309aand309b(only one of them is shown) for the front work device and for swing, and control lever/pedal devices309cand309d(only one of them is shown) for travelling. The front work device302is configured by connecting a boom306, an arm307, and a bucket308by using pins.

The upper swing structure300is driven to swing relative to the lower track structure301by a swing motor3c. The front work device302is rotated horizontally by turning a swing post303by a swing cylinder3f(seeFIG. 1A). The left and right crawlers310and311of the lower track structure301are driven by left and right travelling motors3dand3e. The blade305is driven up and down by a blade cylinder3g. In addition, the boom306, the arm307, and the bucket308are vertically rotated by extension/contraction of a boom cylinder3h, an arm cylinder3a, and a bucket cylinder3b.

Operation of this embodiment will be described below.

<<Single Drive of Actuator on First Hydraulic Pump1aSide>>

When an arm operation is conducted by singly driving one of actuators connected to the first hydraulic pump1aside, for example, the arm cylinder3a, an arm control lever is operated, whereon the flow control valves6aand6eare changed over, and hydraulic fluids delivered from the first and second delivery ports P1and P2are supplied to the arm cylinder3ain a joining manner. Besides, in this instance, the delivery flow rates of the first and second delivery ports P1and P2are controlled by the load sensing control of the first load sensing control section12aand the absorption torque constant control of the first torque control section13a, as aforementioned.

When a bucket operation or a swing operation is conducted by singly driving the bucket cylinder3bor the swing motor3c, a relevant control lever is operated, whereon the flow control valve6bor the flow control valve6dis changed over, and the hydraulic fluid delivered from the delivery port P1or P2on one side is supplied to the bucket cylinder3bor the swing motor3c. Besides, in this instance, the delivery flow rates of the first and second delivery ports P1and P2are controlled by the load sensing control of the first load sensing control section12aand the absorption torque constant control of the first torque control section13a. The hydraulic fluid delivered from the delivery port P2or P1on the side of not supplying the hydraulic fluid to the bucket cylinder3bor the swing motor3cis returned to the tank by way of the unloading valve10bor10a.

<<Single Drive of Actuator on Second Hydraulic Pump1bSide>>

When a boom operation is conducted by singly driving one of the actuators connected to the second hydraulic pump1bside, for example, the boom cylinder3h, a boom control lever is operated, whereon the flow control valves6hand6lare changed over, and hydraulic fluids delivered from the third and fourth delivery ports P3and P4are supplied to the boom cylinder3hin a joining manner. Besides, in this instance, the delivery flow rates of the third and fourth delivery ports P3and P4are controlled by the load sensing control of the second load sensing control section12band the absorption torque constant control of the second torque control section13b, as aforementioned.

When a swing operation or a blade operation is performed by singly driving the swing cylinder3for the blade cylinder3g, a relevant control lever is operated, whereon the flow control valve6ior the flow control valve6kis changed over, and the hydraulic fluid delivered from the delivery port P3or P4on one side is supplied to the swing cylinder3for the blade cylinder3g. Besides, in this instance also, the delivery flow rates of the third and fourth delivery ports P3and P4are controlled by the load sensing control of the second load sensing control section12band the absorption torque constant control of the second torque control section13b. The hydraulic fluid delivered from the delivery port P4or P3on the side of not supplying the hydraulic fluid to the swing cylinder3for the blade cylinder3gis returned to the tank by way of the unloading valve10dor10c.

<Simultaneous Drive of Actuator on First Hydraulic Pump1aSide and Actuator on Second Hydraulic Pump1bSide>

<<Simultaneous Drive of Arm Cylinder and Boom Cylinder>>

When a combined operation of the arm307and the boom306is conducted by simultaneously driving the arm cylinder3aand the boom cylinder3h, the arm control lever and the boom control lever are operated, whereon the flow control valves6aand6eand the flow control valves6hand6lare changed over, the hydraulic fluids delivered from the first and second delivery ports P1and P2are supplied to the arm cylinder3ain a joining manner, and the hydraulic fluids delivered from the third and fourth delivery ports P3and P4are supplied to the boom cylinder3hin a joining manner. Besides, on the first hydraulic pump1aside and the second hydraulic pump1bside, the delivery flow rates of the first and second delivery ports P1and P2and the delivery flow rates of the third and fourth delivery ports P3and P4are controlled by the load sensing control of the first and second load sensing control sections12aand12band the absorption torque constant control of the first and second torque control sections13aand13b, as aforementioned. In addition, in the absorption torque constant control of the first torque control section13a, the total torque control shown inFIG. 4Ais conducted.

<<Simultaneous Drive of Swing Arm and Boom Cylinder>>

When a combined operation of the upper swing structure300(swing) and the boom306by simultaneously driving the swing motor3cand the boom cylinder3h, a swing control lever and the boom control lever are operated, whereon the flow control valve6dand the flow control valves6hand6lare changed over, whereon the hydraulic fluid delivered from the second delivery port P2is supplied to the swing motor3c, and the hydraulic fluids delivered from the third and fourth delivery ports P3and P4are supplied to the boom cylinder3hin a joining manner. Besides, on the first hydraulic pump1aside and the second hydraulic pump1bside, the delivery flow rates of the first and second delivery ports P1and P2and the delivery flow rates of the third and fourth delivery ports P3and P4are controlled by the load sensing control of the first and second lead sensing control sections12aand12band the absorption torque constant control of the first and second torque control sections13aand13b, as aforementioned. In addition, in the absorption torque constant control of the first torque control section13a, the total torque control shown inFIG. 4Ais performed. The hydraulic fluid delivered from the first delivery port P1on the side where the flow control valves6ato6care closed is returned to the tank by way of the unloading valve10a.

<<Simultaneously Drive of Other Combination of Actuator on First Hydraulic Pump1aSide and Actuator on Second Hydraulic Pump1bSide>>

In a combined operation other than the above-mentioned in which at least one of the actuators (arm cylinder3a, bucket cylinder3b, and swing motor3c) connected only to the first and second delivery ports P1and P2of the first hydraulic pump1aand at least one of the actuators (swing cylinder3f, blade cylinder3g, and boom cylinder3h) connected only to the third and fourth delivery ports P3and P4of the second hydraulic pump1bare simultaneously driven, also, the delivery flow rates of the first and second delivery ports P1and P2and the delivery flow rates of the third and fourth delivery ports P3and P4are controlled by the load sensing control and the absorption torque constant control, similarly to the above. Besides, in the absorption torque constant control of the first torque control section13a, the total torque control shown inFIG. 4Ais conducted. The hydraulic fluid delivered from the delivery port on the side where the flow control valve is closed is returned to the tank by way of the unloading valve.

<Simultaneous Drive of Two Actuators on First Hydraulic Pump1aSide>

In a combined operation in which at least one of the actuators (arm cylinder3a, bucket cylinder3b, and travelling-right travelling motor3e) connected to the first delivery port P1of the first hydraulic pump1aand at least one of the actuators (arm cylinder3a, swing motor3c, and travelling-left travelling motor3d) connected to the second delivery port P2of the first hydraulic pump1bare simultaneously driven, the delivery flow rates of the first and second delivery ports P1and P2are controlled by the load sensing control of the first load sensing control section12aand the absorption torque constant control of the first torque control section13a, like in the case of the arm operation in which the arm cylinder3ais singly driven. In addition, a surplus flow rate of the hydraulic fluid delivered from the delivery port on the side where the demanded flow rate is low or the hydraulic fluid delivered from the delivery port on the side where the flow control valve is closed is returned to the tank by way of the unloading valve. In this instance, a load pressure (maximum load pressure) of the actuators on the first delivery port P1side that is detected by the first shuttle valve group208ais introduced to the pressure compensating valves7ato7cand the first unloading valve210a, whereas a load pressure (maximum load pressure) of the actuators on the second delivery port P2side that is detected by the second shuttle valve group208bis introduced to the pressure compensating valves7dto7fand the second unloading valve210b, and controls by the pressure compensating valves and the unloading valve are performed separately on the first delivery port P1side and on the second delivery port P2side. This ensures that when the surplus flow rate of the delivery port on the low load pressure side is returned to the tank, the pressure of the delivery port is limited in rise based on the low load pressure by the unloading valve on the relevant delivery port side, and, accordingly, the pressure loss at the unloading valve at the time of returning of the surplus flow rate to the tank is reduced, and an operation with little energy loss can be achieved.

<Simultaneous Drive of Two Actuators on Second Hydraulic Pump1bSide>

In a combined operation in which two actuators on the second hydraulic pump1bside are simultaneously driven, also, the delivery flow rates of the third and fourth delivery ports P3and P4are controlled by the load sensing control of the second load sensing control section12band the second torque control section13b, like in the aforementioned case of the combined operation in which two actuators on the first hydraulic pump1aare simultaneously driven. In addition, a surplus flow rate of hydraulic fluid delivered from the delivery port on the side where the demanded flow rate is low or the hydraulic fluid delivered from the delivery port on the side where the flow control valve is closed is returned to the tank by way of the unloading valve, and, accordingly, the pressure loss at the unloading valve in this instance is reduced, and an operation with little energy loss can be achieved.

When a travelling operation is conducted by driving the travelling-left travelling motor3dand the travelling-right travelling motor3e, left and right travelling control levers or pedals are operated, whereon the flow control valves6fand6jand the flow control valves6cand6gare changed over, whereby the hydraulic fluid delivered from the second delivery port P2of the first hydraulic pump1aand the hydraulic fluid delivered from the fourth delivery port P4of the second hydraulic pump1bare supplied to the travelling-left travelling motor3din a joining manner, whereas the hydraulic fluid delivered from the first delivery port P1of the first hydraulic pump1aand the hydraulic fluid delivered from the third delivery port P3of the second hydraulic pump1bare supplied to the travelling-right travelling motor3ein a joining manner. Therefore, even if the tilting angle of the swash plate of the first hydraulic pump1aand the tilting angle of the swash plate of the second hydraulic pump1bare different and a difference in delivery flow rate is generated between the first and second delivery ports P1and P2and the third and fourth delivery ports P3and P4, the supply flow rate to the travelling-left travelling motor3dand the supply flow rate to the travelling-right travelling motor3eare the same, and, accordingly, the vehicle body can travel straight without meandering.

Specifically, assuming that the delivery flow rate of the first delivery port P1is Q1, the delivery flow rate of the second delivery port P2is Q2, the delivery flow rate of the third delivery port P3is Q3, and the delivery flow rate of the fourth delivery port P4is Q4, then the supply flow rate to the travelling-left travelling motor3dand the supply flow rate to the travelling-right travelling motor3eare as follows.

Here, the relations of Q1=Q2(because of the same swash plate) and Q3=Q4(because of the same swash plate) are established. Therefore, even if Q1=Q2≠Q3=Q4, the relation of
Q2+Q4=Q1+Q3
is established, and, therefore, the supply flow rate to the travelling-left travelling motor3dand the supply flow rate to the travelling-right travelling motor3eare the same.

In this way, even if a difference in delivery flow rate is generated between the first and second delivery ports P1and P2and the third and fourth delivery ports P3and P4, the supply flow rate to the travelling-left travelling motor3dand the supply flow rate to the travelling-right travelling motor3eare the same, and, accordingly, the vehicle body can travel straight without meandering.

A case of performing a travelling combined operation in which the travelling motors3dand3eand at least one of other actuators, for example, the arm cylinder3aare simultaneously driven will be described.

When the left and right travelling control levers or pedals and the arm control lever are operated with an intention to perform a travelling combined operation, the flow control valves6fand6j, the flow control valves6cand6gand the flow control valves6aand6eare changed over, and, simultaneously, the first communication control valve215ais changed over to the communication position of the lower side in the drawing. With such arrangement, the hydraulic fluids delivered from the first and second delivery ports P1and P2are supplied from the first hydraulic pump1aside in a joining manner and the hydraulic fluid delivered from the fourth delivery port P4is supplied from the secondary hydraulic pump1bside, to the travelling-left travelling motor3d, whereas the hydraulic fluids delivered from the first and second delivery ports P1and P2are supplied from the first hydraulic pump1aside in a joining manner and the hydraulic fluid delivered from the third delivery port P3is supplied from the second hydraulic pump1bside, to the travelling-right travelling motor3e. The arm cylinder3ais supplied with the remainder of the hydraulic fluids supplied to the travelling motors3dand3efrom the first and second delivery ports P1and P2.

In this instance, besides, on the first hydraulic pump1aside, the first communication control valve215ais changed over to the communication position of the lower side in the drawing. Therefore, the maximum load pressure of the actuators3ato3ethat is detected by the first and second shuttle valve groups208aand208bis introduced to the load sensing control valves216aand216b, the pressure compensating valves7ato7cand7dto7fand the first unloading valves210aand210b, whereby the load sensing control and the controls of the pressure compensating valves and the unloading valves are performed. On the other hand, on the second hydraulic pump1bside, the second communication control valve215bis held in the interruption position of the upper side in the drawing. Therefore, the maximum load pressures are detected separately on the third delivery port P3side and on the fourth delivery port P4side, and the respective maximum load pressures are introduced to the load sensing control valves216cand216d, the pressure compensating valves7gto7iand7jto7mand the third and fourth unloading valves210cand210d, whereby the load sensing control and the controls of the pressure compensating valves and the unloading valves are performed.

Here, a case where straight travelling is conducted by a travelling combined operation will be described.

When the left and right travelling control levers or pedals are operated by the same amount with the intention to perform straight travelling by a travelling combined operation, the flow control valves are changed over such that the stroke amount (opening area) of the flow control valves6fand6jand the stroke amount (opening area−demanded flow rate) of the flow control valves6cand6gwill be the same. In addition, as aforementioned, the hydraulic fluid delivered from the second delivery port P2of the first hydraulic pump1aand the hydraulic fluid delivered from the fourth delivery port P4of the second hydraulic pump1bare supplied to the travelling-left travelling motor3din a joining manner; the hydraulic fluids delivered from the first and second delivery ports P1and P2are supplied from the first hydraulic pump1aside in a joining manner and the hydraulic fluid delivered from the fourth delivery port P4is supplied from the second hydraulic pump1bside, to the travelling-left travelling motor3d; the hydraulic fluids delivered from the first and second delivery ports P1and P2are supplied from the first hydraulic pump1aside in a joining manner and the hydraulic fluid delivered from the third delivery port P3is supplied from the second hydraulic pump1bside, to the travelling-right travelling motor3e. This ensures that in the travelling combined operation, also, the supply flow rate to the travelling-left travelling motor3dand the supply flow rate to the travelling-right travelling motor3eare the same, and, therefore, the vehicle body can travel straight without meandering.

Specifically, assuming that the delivery flow rate of the first delivery port P1is Q1, the delivery flow rate of the second delivery port P2is Q2, the delivery flow rate of the third delivery port P3is Q3, and the delivery flow rate of the fourth delivery port P4is Q4, and that the flow rate of the hydraulic fluid supplied to the travelling-left travelling motor3dis Qd, the flow rate of the hydraulic fluid supplied to the travelling-right travelling motor3eis Qe, and the flow rate of the hydraulic fluid supplied to the boom cylinder3awhich is an actuator other than the travelling motors is Qa, the flow rates Qd and Qe of the hydraulic fluids supplied to the left and right travelling motors3dand3eare as follows.

First, each of the left and right travelling motor3dand3eis supplied with hydraulic fluid from the first hydraulic pump1aside in an amount of ½ of Q1+Q2−Qa, the amount obtained by subtracting the flow rate Qa of the hydraulic fluid supplied to the boom cylinder3afrom the total flow rate Q1+Q2of the hydraulic fluids delivered from the first and second deliver ports P1and P2. The amount supplied is ½ of Q1+Q2−Qa because the stroke amount (opening area) of the flow control valve6fand the stroke amount (opening area−demanded flow rate) of the flow control valve6care the same. In addition, each of the left and right travelling motors3dand3eis supplied with hydraulic fluid from the second hydraulic pump1bside in an amount of ½ of the total flow rate Q3+Q4of the hydraulic fluids delivered from the first and second delivery ports P1and P2. In this case, also, the amount supplies is ½ of Q3+Q4because the stroke amount (opening area) of the flow control valve6jand the stroke amount (opening area−demanded flow rate) of the flow control valve6gare the same. Accordingly, the flow rates Qd and Qe of the hydraulic fluids supplied to the left and right travelling motors3dand3eare expressed as follows.
Travelling-right supply flow rateQd=(Q1+Q2−Qa)/2+(Q3+Q4)/2
Travelling-left supply flow rateQe=(Q1+Q2−Qa)/2+(Q3+Q4)/2

In other words, Qd=Qe, and according, the vehicle body can travel straight without meandering.

The above-mentioned example of the travelling combined operation corresponds to the case where the travelling motors3dand3eand the arm cylinder3aare simultaneously driven. As other example of the travelling combined operation, there is a travelling combined operation in which an actuator (bucket cylinder3b, swing motor3c) driven by the hydraulic fluid delivered only from the first delivery port P1or the second delivery port P2of the first hydraulic pump1aor an actuator (swing cylinder3f, blade cylinder3g) driven by the hydraulic fluid delivered only from the third delivery port P3or the fourth delivery port P4of the second hydraulic pump1bis driven simultaneously with the travelling motors. In this embodiment, in the case of performing such a travelling combined operation, also, the vehicle body can travel straight without meandering.

Note that in this embodiment, the first to fourth shuttle valve groups208ato208d, the first and second communication control valves15aand15b, the load sensing control valves216ato216dand the low pressure selection valves221aand221bare provided, and communication is established and interrupted with respect to both the delivery ports and the output hydraulic line of the maximum load pressure by the first and second communication control valves15aand15b. However, a structure in which communication is established and interrupted with respect to the delivery ports by the first and second communication control valves15aand15bmay be adopted, and the other circuit structure may be the same as in the first embodiment. In this case, also, the first and second communication control valves15aand15bare changed over to the communication positions at the time of the travelling combined operation, whereby an effect to secure the straight travelling properties can be obtained.

The effects obtained by this embodiment will be described below.

FIG. 8is a diagram showing, as a comparative example, a hydraulic system in the case where the total torque control technology described in Patent Document 2 is incorporated into the two-pump load sensing system provided with the first and second hydraulic pumps1aand1bshown inFIG. 1. In the diagram, members equivalent to the elements shown inFIG. 1are denoted by the same reference symbols as used above.

The hydraulic system of the comparative example shown inFIG. 8includes pressure reduction valves41aand41bin place of the torque feedback circuit30(the first torque feedback circuit section30aand the second torque feedback circuit section30b). The pressure reduction valves41aand41breduce the delivery pressures of the third and fourth delivery ports of the second hydraulic pump1bin such a manner that the secondary pressures (torque control pressures) does not exceed a set pressure, and outputs the thus reduced pressures. The set pressure of the pressure reduction valves41aand41bis set to be a value (the start pressure Pb of the absorption torque constant control shown inFIG. 4B) corresponding to the maximum torque T2max set by the springs S3and S4in the torque control section of the second hydraulic pump1b.

FIG. 9is a diagram showing the total torque control in the comparative example shown inFIG. 8. In the comparative example illustrated inFIG. 8, when the delivery pressures of the third and fourth delivery ports of the second hydraulic pump are equal to or higher than the start pressure of the absorption torque constant control, it is assumed that the second hydraulic pump1bis under the absorption torque constant control. In this case, the pressure reduction valves41aand41breduce the delivery pressures of the third and fourth delivery ports of the second hydraulic pump to a pressure corresponding to the maximum torque T2max, and introduce the thus reduced pressure to the torque reduction control pistons31aand31bof the first hydraulic pump1a. On the first hydraulic pump1aside, the maximum torque is reduced from T1max by an amount of T2max. In this way, the total torque control is carried out.

However, even when the delivery pressures of the third and fourth delivery ports of the second hydraulic pump are equal to or higher than the start pressure of the absorption torque constant control, there is a case where the second hydraulic pump1bis not under the absorption torque constant control, and the second hydraulic pump1bis controlled to a tilting angle smaller than the tilting that is limited under the absorption torque constant control by the load sensing control. In this case, the absorption torque of the second hydraulic pump1bestimated with the pressure corresponding to the maximum torque T2max would be a value greater than the actual absorption torque of the second hydraulic pump1b.

As a result, in the first hydraulic pump1awhere a pressure corresponding to the maximum torque T2max is introduced and the total torque control is conducted with the maximum torque of T1max−T2max, such a control as to reduce the maximum torque more than necessary would be performed, and, accordingly, the output torque of the prime mover cannot be used effectively.

FIG. 10is a diagram showing a total torque control in this embodiment.

In this embodiment, the torque feedback circuit30modifies the delivery pressures of the third and fourth delivery ports P3and P4of the second hydraulic pump1bin such a manner as to provide characteristics simulating the absorption torque of the second hydraulic pump1bboth in the cases of when the second hydraulic pump1bis limited by control of the second torque control section13band operates at the maximum torque T2max (second maximum torque) and when the second hydraulic pump1bis not limited by the control of the second torque control section13band the second load sensing control section12bcontrols the capacity of the second hydraulic pump1b(when lower than the start pressure Pb of the absorption torque constant control of the second hydraulic pump1b), and outputs the thus modified pressures. The first and second torque reduction control pistons31aand31breduce the maximum torque T1max set in the first torque control section13a, as the output pressure of the torque feedback circuit30becomes higher.

For example, as aforementioned, when the delivery pressures of the third and fourth delivery ports P3and P4of the second hydraulic pump1brise, the absorption torque of the second hydraulic pump1bin that instance is T2which is lower than the maximum torque T2max, and the absorption torque simulated by the torque feedback circuit30is T2s (≈T2), the torque feedback pistons32aand32breduce the maximum torque T1max to T1max−T2s, as shown inFIG. 10, and the total torque control is conducted with the maximum torque T1max−T2s. As a result, the maximum torque is not reduced more than necessary, and stoppage of the engine2(engine stall) can be prevented, while making the most of the rated output torque TER of the engine2.

As above-mentioned, according to this embodiment, the absorption torque of the second hydraulic pump1bcan be accurately detected by a purely hydraulic structure (torque feedback circuit30). In addition, by feeding back the absorption torque to the first hydraulic pump1aside, it is possible to accurately perform the total torque control and to effectively utilize the rated output torque TER of the prime mover2. Besides, owing to the structure in which the absorption torque of the second hydraulic pump1bis detected on a purely hydraulic basis, the first pump control unit5acan be miniaturized, and the mountability of the hydraulic pump inclusive of the pump control unit is enhanced. Consequently, it is possible to provide a construction machine that is good in energy efficiency, is low in fuel cost, and is practical.

In addition, as shown inFIGS. 5C and 5D, the target control pressures formed in the first and second hydraulic lines36aand36bbetween the first and second pressure dividing restrictor parts (fixed restrictors)34aand34band the first and second pressure dividing valves (variable restrictor valves)35aand35band the torque control pressures outputted by the first and second pressure reduction valves32aand32bare pressures of the same values, and the pressures formed in the first and second hydraulic lines36aand36bcan also be used directly as torque control pressures.

In the case where the pressures formed in the first and second hydraulic lines36aand36bare used directly as the torque control pressures, however, at the time of driving the third torque control actuators32aand32bwith the torque control pressures, the first and second pressure dividing restrictor parts (fixed restrictors)34aand34bconstitute resistances to make it difficult to supply sufficient quantities of hydraulic fluid to the third torque control actuators32aand32b, so that the responsiveness of the third torque control actuators32aand32bmay be worsened.

Besides, in the case where hydraulic fluid is supplied from the first and second hydraulic lines36aand36bto the third torque control actuators32aand32b, pressure variations are liable to occur due to variations in the quantities of hydraulic fluid in the first and second hydraulic lines36aand36b, making it difficult for the pressures formed in the first and second hydraulic lines36aand36bto be accurately set to attain pressure variations as shown inFIG. 5C. Further, when the delivery pressure of the second hydraulic pump1bfluctuates, the fluctuations in the delivery pressure may be transmitted directly to the third torque control actuators32aand32b, whereby stability of the system may be damaged.

In this embodiment, the pressures in the first and second hydraulic lines36aand36bbetween the first and second pressure dividing restrictor parts (fixed restrictors)34aand34band the first and second pressure dividing valves (variable restrictor valves)35aand35bare introduced to the first and second pressure reduction valves32aand32bas target control pressures, thereby providing the set pressures for the first and second pressure reduction valves32aand32b, and the torque control pressure is generated from the delivery pressure of the second hydraulic pump1bby the first and second pressure reduction valves32aand32b. Therefore, it is possible to secure the flow rates at the time of driving the third torque control actuators32aand32bwith the torque control pressure, and to obtain good responsiveness at the time of driving the third torque control actuators32aand32b.

In addition, since the pressures in the first and second hydraulic lines36aand36bbetween the first and second pressure dividing restrictor parts (fixed restrictors)34aand34band the first and twenty-second pressure dividing valves (variable restrictor valves)35aand35bare not used directly as the torque control pressures, the setting of the first and second pressure dividing restrictor parts (fixed restrictors)34aand34band the first and twenty-second pressure dividing valves (variable restrictor valves)35aand35bfor obtaining the required target control pressures and the setting of the responsiveness of the third torque control actuators32aand32bcan be performed independently, so that the setting of the torque feedback circuit30for exhibiting required performance can be performed easily and accurately.

Further, when the delivery pressure of the second hydraulic pump1bis higher than the set pressures of the first and second pressure reduction valves32aand32b, fluctuations in the delivery pressure of the second hydraulic pump1bis blocked by the first and second pressure reduction valves32aand32b, and therefore do not influence the third torque control actuators32aand32b. Accordingly, the stability of the system is secured.

While the case where the first and second hydraulic pumps are split flow type hydraulic pumps having the first and second delivery ports P1and P2and the third and fourth delivery ports P3and P4, respectively, has been described in the embodiment above, both or one of the first and second hydraulic pumps may be a single flow type hydraulic pump having a single delivery port. In the case where the first and second hydraulic pumps are single flow type hydraulic pumps, it is sufficient that the torque feedback circuit30has one circuit section and one torque reduction control piston to which the torque control pressure is introduced. Besides, the axis of abscissas inFIGS. 4A and 4Bthen represents the pressure of the single delivery port (the delivery pressure of the hydraulic pump).

In addition, since in the torque feedback circuit30the target control pressures formed in the first and second hydraulic lines36aand36bbetween the first and second pressure dividing restrictor parts (fixed restrictors)34aand34band the first and second pressure dividing valves (variable restrictor valves)35aand35band the torque control pressures outputted by the first and second pressure reduction valves32aand32bare pressures of the same values as aforementioned, a structure may be adopted in which the pressures formed in the first and second hydraulic lines36aand36bare introduced directly to the torque reduction control actuators31aand31bas torque control pressures.

Besides, while in the embodiment above the first and second relief valves37aand37bhave been provided in the torque feedback circuit30in such a manner that the pressures in the first and second hydraulic lines36aand36bbetween the first and second pressure dividing restrictor parts (fixed restrictors)34aand34band the first and second pressure dividing valves (variable restrictor valves)35aand35bdo not increase beyond the set pressure (torque start pressure Pb), pressure reduction valves may be used in place of the relief valves. In this case, by providing the set pressure of the pressure reduction valves at the torque start pressure Pb and using the output pressures of the pressure reduction valves as the target control pressures P35ref and P4tref, the same or similar function to the above can be obtained.

In addition, while the first pump control unit5ahas had the first load sensing control section12aand the first torque control section18a, the first load sensing control section12ain the first pump control unit5ais not indispensable, and other control system, such as the so-called positive control or negative control system may also be used so long as the system can control the capacity of the first hydraulic pump according to the operation amount of the control lever (flow control valve's opening area−demanded flow rate).

Further, the load sensing system in the embodiment above is an example, and the load sensing system may be modified variously. For instance, while the differential pressure reduction valve outputting the pump delivery pressure and the maximum load pressure as absolute pressures has been provided and its output pressure has been introduced to the pressure compensating valve to set the target compensating pressure and introduced to the LS control valve to set the target differential pressure for the load sensing control in the embodiment above, the pump delivery pressure and the maximum load pressure may be introduced to the pressure control valve and the LS control valve by way of different hydraulic lines.

DESCRIPTION OF REFERENCE CHARACTERS

1a: First hydraulic pump1b: Second hydraulic pump2: Prime mover (diesel engine)3a-3h: Actuators3a: Arm cylinder3d: Left travelling motor3e: Right travelling motor3h: Boom cylinder4: Control valve5a: First pump control unit5b: Second pump control unit6a-6m: Flow control valves7a-7m: Pressure compensating valves8a: First shuttle valve group8b: Second shuttle valve group8c: Third shuttle valve group8d: Fourth shuttle valve group9a-9d: Springs10a-10d: Unloading valves12a: First load sensing control section12b: Second load sensing control section13a: First torque control section13b: Second torque control section15a: First communication control valve15b: Second communication control valve16a-16d: Load sensing control valves17a,17b: Load sensing control pistons (load sensing control actuators)18a: First torque control piston (first torque control actuator)19a: Second torque control piston (first torque control actuator)18b: Third torque control piston (second torque control actuator)19b: Fourth torque control piston (second torque control actuator)21a,21b: Low pressure selection valves30: Torque feedback circuit30a: First torque feedback circuit section30b: Second torque feedback circuit section31a: First torque reduction control piston (third torque control actuator)31b: Second torque reduction control piston (third torque control actuator)32a: First torque pressure reduction valve32b: Second torque pressure reduction valve33a: First pressure dividing circuit33b: Second pressure dividing circuit34a: First pressure dividing restrictor part34b: Second pressure dividing restrictor part35a: First pressure dividing valve35b: First pressure dividing valve36a: First hydraulic line36b: Second hydraulic line37a: First relief valve (pressure limiting valve)37b: Second relief valve (pressure limiting valve)P1, P2: First and second delivery portsP3, P4: Third and Fourth delivery portsS1, S2: SpringsS3, S4: Springs