阅读说明：本技术 工程机械的液压驱动装置 (Hydraulic drive device for construction machine ) 是由 高桥究 前原太平 石井刚史 于 2018-03-28 设计创作，主要内容包括：在与各驱动器取得关联的方向切换阀的前后差压非常小的情况下,也能够稳定地进行液压泵的流量控制和多个方向切换阀的分流控制,即使在从复合动作向单独动作转移时等要求流量发生了急速变动的情况下,也防止向各驱动器供给的压力油的流量急剧变换,实现优异的复合操作性,降低方向切换阀的入口节流损失,实现高能效。为此,在多个方向切换阀(6a、6b、6c)的下游侧分别配置以使它们的入口节流开口的下游侧的压力等于最高负载压的方式控制的多个压力补偿阀(7a、7b、7c),根据各操作杆的输入量计算各方向切换阀(6a、6b、6c)的要求流量,而且根据各方向切换阀(6a、6b、6c)的要求流量和入口节流开口面积计算预定的方向切换阀的入口节流压损,并使用该值控制卸载阀(15)的设定压。(Even when the differential pressure between the front and rear sides of the directional control valve associated with each actuator is very small, the flow rate of the hydraulic pump and the flow dividing control of the plurality of directional control valves can be stably performed, and even when the required flow rate rapidly changes such as when the composite operation is shifted to the single operation, the flow rate of the pressure oil supplied to each actuator is prevented from rapidly changing, so that excellent composite operability is realized, the meter-in loss of the directional control valve is reduced, and high energy efficiency is realized. A plurality of pressure compensating valves (7a, 7b, 7c) controlled so that the pressure on the downstream side of the meter-in openings of the directional switching valves (6a, 6b, 6c) is equal to the maximum load pressure are arranged on the downstream side of the directional switching valves (6a, 6b, 6c), the required flow rates of the directional switching valves (6a, 6b, 6c) are calculated from the input amounts of the respective control levers, the meter-in pressure loss of a predetermined directional switching valve is calculated from the required flow rate and the meter-in opening area of the directional switching valves (6a, 6b, 6c), and the set pressure of the unloader valve (15) is controlled using the calculated value.)

1. A hydraulic drive device for a construction machine, comprising:

a hydraulic pump of a variable displacement type;

a plurality of actuators driven by the pressure oil discharged from the hydraulic pump;

a control valve device for distributing and supplying the pressure oil discharged from the hydraulic pump to the plurality of actuators;

a plurality of operation lever devices indicating a driving direction and a speed of each of the plurality of drivers;

a pump control device for controlling a discharge flow rate of the hydraulic pump so as to discharge a flow rate corresponding to an input amount of an operation lever of the plurality of operation lever devices;

an unloading valve configured to discharge the pressurized oil from the pressurized oil supply path to a tank when a pressure in the pressurized oil supply path of the hydraulic pump exceeds a set pressure, the set pressure being a value obtained by adding at least a target differential pressure to a maximum load pressure of the plurality of actuators; and

a controller for controlling the above-mentioned control valve device,

the hydraulic drive apparatus for a construction machine described above is characterized in that,

the control valve device includes:

a plurality of direction switching valves that are switched by the plurality of lever devices, are associated with the plurality of actuators, and adjust the driving direction and speed of each actuator; and

a plurality of pressure compensating valves which are respectively arranged at the downstream side of the plurality of directional control valves and control the pressure at the downstream side of the meter-in openings of the plurality of directional control valves to be equal to the maximum load pressure,

the controller calculates a required flow rate of each of the plurality of actuators and an opening area of an inlet throttle of each of the plurality of directional control valves based on an input amount of an operation lever of the plurality of operation lever devices, calculates a pressure loss of the inlet throttle of a specific directional control valve among the plurality of directional control valves based on the opening area of the inlet throttle and the required flow rate, and controls a set pressure of the unloader valve using the pressure loss as the target differential pressure output.

2. The hydraulic drive apparatus of a construction machine according to claim 1,

the controller selects a maximum value of the pressure loss of the meter-in ports of the plurality of directional control valves as the pressure loss of the meter-in ports of the specific directional control valve, and controls the set pressure of the unloader valve by outputting the pressure loss as the target differential pressure.

3. The hydraulic drive apparatus of a construction machine according to claim 1,

further comprises a maximum load voltage detection device for detecting the maximum load voltages of the plurality of drivers,

the controller calculates an meter-in pressure loss of a directional switching valve corresponding to an actuator of the highest load pressure detected by the highest load pressure detection device among the plurality of directional switching valves as a meter-in pressure loss of the specific directional switching valve, and controls the set pressure of the unloader valve by outputting the pressure loss as the target differential pressure.

4. The hydraulic drive apparatus of a construction machine according to claim 1,

further provided with: a maximum load voltage detection device for detecting the maximum load voltages of the plurality of drivers; and

a pressure sensor for detecting the discharge pressure of the hydraulic pump,

the controller calculates a command value for making the discharge pressure of the hydraulic pump detected by the pressure sensor equal to a pressure obtained by adding the target differential pressure to the maximum load pressure detected by the maximum load pressure detection device, and outputs the command value to the pump control device to control the discharge flow rate of the hydraulic pump.

5. The hydraulic drive apparatus of a construction machine according to claim 1,

the controller calculates a sum of the required flow rates of the plurality of actuators based on input amounts of the control levers of the plurality of control lever devices, calculates a command value for making a discharge flow rate of the hydraulic pump equal to the sum of the required flow rates, and outputs the command value to the pump control device to control the discharge flow rate of the hydraulic pump.

Technical Field

The present invention relates to a hydraulic drive device for a construction machine such as a hydraulic excavator that performs various operations, and more particularly to a hydraulic drive device for a construction machine that guides and drives pressure oil discharged from one or more hydraulic pumps to two or more actuators via two or more control valves.

Background

As a hydraulic drive device for a construction machine such as a hydraulic excavator, for example, as described in patent document 1, a load sensing control is widely used, in which the displacement of a hydraulic pump is controlled so that a differential pressure between the discharge pressure of a variable displacement hydraulic pump and the maximum load pressure of a plurality of actuators is maintained at a predetermined set value.

Patent document 2 describes a hydraulic drive device including: a hydraulic pump of a variable displacement type; a plurality of drivers; a plurality of meter-in orifices that control the flow rate of pressure oil supplied from the hydraulic pump to the plurality of actuators; a plurality of pressure compensating valves disposed downstream of the plurality of meter-in orifices; and a controller that controls a discharge flow rate of the hydraulic pump in accordance with a lever input of the operation lever device, adjusts the plurality of meter-in orifices in accordance with the lever input, and performs full-open control of the meter-in orifice associated with the actuator having the highest load pressure based on the lever input. In this hydraulic drive apparatus, the plurality of pressure compensation valves provided downstream of the plurality of meter-in orifices are controlled so that the pressure on the downstream side of the meter-in orifice is equal to the maximum load pressure, without using the differential pressure (LS differential pressure) between the pump pressure and the maximum load pressure.

Patent document 3 proposes a drive system including: a hydraulic pump of a variable displacement type; a plurality of drivers; a plurality of control valves which have a throttling function at intermediate positions and supply pressure oil discharged from the hydraulic pump to the plurality of actuators; an unloading valve provided in a pressure oil supply path of the hydraulic pump; a controller for controlling the discharge flow rate of the hydraulic pump according to the lever input of the operation lever device; and a pressure sensor for detecting the discharge pressure of the hydraulic pump and the load pressure of at least one actuator, wherein the controller controls the opening of the regulating valve having a throttling action at the intermediate position based on the differential pressure between the discharge pressure of the hydraulic pump and the load pressure of the actuator detected by the pressure sensor. In this drive system, the set pressure of the unload valve is set by a spring set in the same direction as the highest load pressure of each actuator that guides the unload valve in the closing direction, and the discharge pressure of the hydraulic pump is controlled so as not to exceed the sum of the highest load pressure and the spring force.

Disclosure of Invention

Problems to be solved by the invention

In the conventional load sensing control described in patent document 1, a differential pressure between the discharge pressure (pump pressure) of the hydraulic pump and the maximum load pressure, which is referred to as an LS differential pressure generated by a differential pressure between the front and rear of the meter-in opening of each main spool (flow control valve), is used for pump flow rate control and flow division control of each main spool by the pressure compensating valve.

In order to improve the energy efficiency of the hydraulic system, the LS differential pressure may be reduced by extremely increasing the meter-in final opening (the meter-in opening area of the full stroke of the main spool) of each main spool, but cannot be extremely reduced to 0 or the like in the current load sensing control. The reason is as follows.

The pressure compensating valve that performs the split control of each main spool controls the opening thereof in such a manner that the front-rear differential pressure of each main spool becomes the same as the LS differential pressure. When the final opening of the meter-in port of the main spool is increased and the LS differential pressure is set to 0 as described above, the openings of the pressure compensation valves are adjusted so that the front-rear differential pressure of each main spool becomes 0. However, in this case, since the target differential pressure for determining the opening of the pressure compensating valve itself becomes 0, the opening of the pressure compensating valve, that is, the position of the spool in the case of the spool type and the lift amount of the poppet in the case of the poppet type, are determined carelessly, and the pressure control of the pressure compensating valve becomes unstable, which causes a problem of occurrence of a bounce.

According to the configuration described in patent document 2, since the meter-in opening of the actuator having the highest load pressure is fully opened, the LS differential pressure, which is one of factors that hinder energy efficiency in the conventional load sensing control, can be discharged, and a hydraulic system with high energy efficiency can be realized.

Here, the pressure compensating valve has a system of controlling the differential pressure between the front and rear of the meter-in opening of each main spool to be equal to a fixed value predetermined by a spring or the like or the differential pressure between the pump pressure and the maximum load pressure (LS differential pressure), and a system of controlling the pressure on the downstream side of the meter-in opening of each main spool to be equal to the maximum load pressure of the plurality of actuators without using the LS differential pressure. The former is generally called a load sensing valve, and the pressure compensating valve described in patent document 1 belongs to this type. The latter is called a flow sharing valve, and the pressure compensating valve described in patent document 2 belongs to this type. In any case, the combination with the load sensing control of the hydraulic pump is collectively referred to as a load sensing system.

In patent document 2, since a flow rate sharing valve that does not use the LS differential pressure is used as the pressure compensating valve, there is no problem that the control of the pressure compensating valve becomes unstable when the LS differential pressure is set to 0 by the load sensing control using the load sensing valve as the pressure compensating valve as in patent document 1.

However, the conventional technique described in patent document 2 has the following problems.

That is, since the throttle orifice (meter-in opening) associated with the actuator having the highest load pressure is always fully opened, for example, when the operation of the actuator having the lower load pressure is suddenly stopped from a state in which the actuator having the highest load pressure and the actuator having the lower load pressure are simultaneously operated, a certain amount of time may be consumed to reduce the flow rate discharged due to the limit of the responsiveness of the flow rate control of the hydraulic pump.

In such a case, the throttle orifice of the highest load pressure actuator is controlled to be opened at the maximum, and the pressure oil discharged from the hydraulic pump flows into the highest load pressure actuator without being throttled by the opening of the throttle orifice, and therefore the speed of the highest load pressure actuator may suddenly increase.

When the operation lever of the maximum load pressure actuator is fully operated and the actuator is operated quickly and a large flow rate is supplied, the influence on the behavior of the working machine is small, but when the operation lever of the maximum load pressure actuator is half operated, the influence when the flow rate supplied to the actuator rapidly increases cannot be ignored as described above, and there is a case where an uncomfortable shock is given to the operator of the working machine.

According to the configuration described in patent document 3, the pressure oil from the hydraulic pump supplied in accordance with the input of each rod can be branched only by the plurality of adjustment valves without using the pressure compensating valve, and therefore the cost of the hydraulic system can be reduced.

In addition, in patent document 3, since the openings of the plurality of adjustment valves are determined by calculation in the electronic control device based on the target flow rates to the actuators set in accordance with the respective operation levers and the differential pressure between the pump pressure detected by the pressure sensor and the maximum load pressure, when the LS differential pressure is set to 0 as in the conventional load sensing control, there is no problem that the control of the pressure compensating valve is unstable.

However, the conventional technique described in patent document 3 has the following problems.

That is, as described above, the relief valve is provided in the pressure oil supply path from the hydraulic pump, but the set pressure thereof is set by the highest load pressure and the spring force.

On the other hand, since the openings (meter-in openings) of the plurality of adjustment valves are determined by the differential pressure between the pump pressure and the actuator load pressure and the target flow rate of each actuator set in accordance with each operation lever, the pump pressure may be increased by the pressure loss of the adjustment valve associated with the highest load pressure actuator with respect to the highest load pressure.

However, since the set pressure of the unload valve is set only by the maximum load pressure and the spring force as described above, for example, when the pressure loss of the adjustment valve associated with the maximum load pressure actuator is high as described above, the pump pressure exceeds the pressure set by the maximum load pressure and the spring force, and the unload valve may be in the open position to discharge the pressurized oil supplied from the hydraulic pump to the tank. The pressure oil discharged through the unloading valve is an excessive bleed loss, and therefore the energy efficiency of the hydraulic system is impaired.

On the other hand, in order not to increase the pressure loss of the adjustment valve associated with the highest load pressure actuator and cause excessive relief loss by exceeding the set pressure of the unloader valve as described above, the spring force of the unloader valve may be increased (the set pressure may be increased), but in this case, for example, when only one actuator is urgently stopped from a state in which two or more actuators are simultaneously operated, the unload valve cannot suppress a sudden increase in the pump pressure due to the timing when the flow rate reduction control of the hydraulic pump is not timely, and therefore, as in the case of using patent document 2, an uncomfortable shock may be caused to the operator.

The present invention has an object to provide a hydraulic drive apparatus for a construction machine, which includes a variable displacement hydraulic pump and supplies pressure oil discharged from the hydraulic pump to a plurality of actuators via a plurality of control valves to drive the plurality of actuators, and which is capable of (1) stably performing flow division control of the plurality of directional control valves even when a differential pressure between the front and rear sides of the directional control valves associated with the respective actuators is very small, (2) suppressing a loss of discharge of the pressure oil from an unloading valve to a tank to the minimum even when a rapid change of a required flow rate is caused, such as when a composite operation is shifted to a single operation, suppressing a decrease in energy efficiency, preventing a rapid change of an actuator speed caused by a rapid change of a flow rate of the pressure oil supplied to the actuators, and suppressing generation of an uncomfortable shock, excellent combined operability is achieved, and (3) the inlet throttling loss of the directional control valve is reduced, thereby achieving high energy efficiency.

Means for solving the problems

In order to achieve the above object, the present invention provides a hydraulic drive device for a construction machine, including: a hydraulic pump of a variable displacement type; a plurality of actuators driven by the pressure oil discharged from the hydraulic pump; a control valve device for distributing and supplying the pressure oil discharged from the hydraulic pump to the plurality of actuators; a plurality of operation lever devices indicating a driving direction and a speed of each of the plurality of drivers; a pump control device for controlling a discharge flow rate of the hydraulic pump so as to discharge a flow rate corresponding to an input amount of an operation lever of the plurality of operation lever devices; an unloading valve configured to discharge the pressurized oil from the pressurized oil supply path to a tank when a pressure in the pressurized oil supply path of the hydraulic pump exceeds a set pressure, the set pressure being a value obtained by adding at least a target differential pressure to a maximum load pressure of the plurality of actuators; and a controller that controls the control valve device, wherein the control valve device includes: a plurality of direction switching valves that are switched by the plurality of lever devices, are associated with the plurality of actuators, and adjust the driving direction and speed of each actuator; and a plurality of pressure compensating valves that are respectively disposed downstream of the plurality of directional control valves and that control the pressure downstream of the meter-in openings of the plurality of directional control valves so as to be equal to the maximum load pressure, wherein the controller calculates a required flow rate for each of the plurality of actuators and an opening area of the meter-in opening of each of the plurality of directional control valves based on input amounts of the control levers of the plurality of control lever devices, calculates a pressure loss of the meter-in opening of a specific directional control valve among the plurality of directional control valves based on the opening areas of the meter-in openings and the required flow rates, and controls the set pressure of the unloader valve using the pressure loss as the target differential pressure output.

As described above, in the present invention, since the flow dividing control of the plurality of directional control valves is performed using the plurality of pressure compensating valves (flow sharing valves) which are respectively disposed on the downstream sides of the plurality of directional control valves and are controlled so that the pressure on the downstream side of the meter-in openings of the plurality of directional control valves is equal to the maximum load pressure, the flow dividing control of the plurality of directional control valves can be stably performed even when the differential pressure (meter-in pressure loss) between the front and rear sides of the directional control valves associated with the actuators is very small.

In the present invention, the controller calculates an opening area of the meter-in of each of the plurality of directional control valves based on an input amount of the operation lever of the plurality of operation lever devices, calculates an opening area of the meter-in of each of the plurality of directional control valves based on the opening area of the meter-in and the required flow rate, calculates a pressure loss of the meter-in of a specific directional control valve among the plurality of directional control valves based on the opening area of the meter-in and the required flow rate of each of the plurality of actuators, and controls the set pressure of the unload valve by outputting the pressure loss as the target differential pressure.

Thus, the set pressure of the unloader valve is controlled to a value in which the maximum load pressure is added to at least the target differential pressure corresponding to the meter-in pressure loss, and therefore, when the meter-in opening of the directional switching valve is throttled by a half operation of the operating lever of the specific directional switching valve or the like, the set pressure of the unloader valve is finely controlled in accordance with the pressure loss of the meter-in opening of the directional switching valve. As a result, even when the required flow rate abruptly changes at the time of transition from the combined operation to the single operation, or the like, and the responsiveness of the pump flow rate control is insufficient and the pump pressure abruptly increases, the release loss of the pressure oil from the unload valve to the tank in excess can be minimized, the decrease in energy efficiency can be suppressed, the abrupt change in the actuator speed due to the abrupt change in the flow rate of the pressure oil to be supplied can be prevented, the occurrence of an uncomfortable shock can be suppressed, and excellent combined operability can be achieved.

Further, according to the present invention, even when the differential pressure between the front and rear sides of each directional control valve is very small, the flow-dividing control of the plurality of directional control valves can be stably performed, and the set pressure of the unload valve can be finely controlled according to the pressure loss of the meter-in opening of the directional control valve, so that the final opening of the meter-in of each directional control valve (the meter-in opening area of the full stroke of the main spool) can be extremely increased, thereby reducing the meter-in loss and achieving high energy efficiency.

Effects of the invention

According to the present invention, in a hydraulic drive device for a construction machine which includes a variable displacement hydraulic pump and which drives a plurality of actuators by supplying pressure oil discharged from the hydraulic pump to the plurality of actuators via a plurality of control valves, the hydraulic drive device is capable of:

(1) even when the differential pressure between the front and rear sides of the directional control valve associated with each actuator is very small, the flow-dividing control of the plurality of directional control valves can be stably performed;

(2) even when the flow rate is required to change rapidly when the hybrid operation is shifted to the single operation, the relief loss of the pressure oil discharged from the relief valve to the tank is minimized, the decrease in energy efficiency is suppressed, the rapid change in the speed of the actuator due to the rapid change in the flow rate of the pressure oil supplied to the actuator is prevented, the occurrence of an uncomfortable shock is suppressed, and excellent hybrid operability is achieved; and

(3) the inlet throttling loss of the direction switching valve is reduced, and high energy efficiency is achieved.

Drawings

Fig. 1 is a diagram showing a configuration of a hydraulic drive device for a construction machine according to a first embodiment of the present invention.

Fig. 2 is an enlarged view of a peripheral portion of an unloading valve of the hydraulic drive apparatus of the first embodiment.

Fig. 3 is an enlarged view of a main pump peripheral portion including a regulator of the hydraulic drive apparatus of the first embodiment.

Fig. 4 is a diagram showing an external appearance of a hydraulic excavator as a typical example of a construction machine on which the hydraulic drive device of the present invention is mounted.

Fig. 5 is a functional block diagram of a controller of the hydraulic drive apparatus of the first embodiment.

Fig. 6 is a functional block diagram of a main pump actual flow rate calculation unit of the controller.

Fig. 7 is a functional block diagram of a requested flow rate calculation unit of the controller.

Fig. 8 is a functional block diagram of the required flow rate correcting section of the controller.

Fig. 9 is a functional block diagram of an meter-in opening calculation unit of the controller.

Fig. 10 is a functional block diagram of a target differential pressure calculation unit of the controller.

Fig. 11 is a functional block diagram of a main pump target tilt angle calculation unit of the controller.

Fig. 12 is a diagram showing a configuration of a hydraulic drive device for a construction machine according to a second embodiment of the present invention.

Fig. 13 is a functional block diagram of a controller of the hydraulic drive apparatus of the second embodiment.

Fig. 14 is a functional block diagram of a highest load voltage driver determination section of the controller.

Fig. 15 is a functional block diagram of a directional control valve meter-in opening calculation unit of the highest load pressure actuator of the controller.

Fig. 16 is a functional block diagram of a post-correction required flow rate calculation unit of the highest load pressure driver of the controller.

Fig. 17 is a functional block diagram of a target differential pressure calculation unit of the controller.

Fig. 18 is a diagram showing a configuration of a hydraulic drive device for a construction machine according to a third embodiment of the present invention.

Fig. 19 is a functional block diagram of a controller of the hydraulic drive apparatus of the third embodiment.

Fig. 20 is a functional block diagram of a requested flow rate calculation unit of the controller.

Fig. 21 is a functional block diagram of a main pump target tilt angle calculation unit of the controller.

Detailed Description

Hereinafter, embodiments of the present invention will be described with reference to the drawings.

< first embodiment >

A hydraulic drive device for a construction machine according to a first embodiment of the present invention will be described with reference to fig. 1 to 15.

Structure ^ E

Fig. 1 is a diagram showing a configuration of a hydraulic drive device for a construction machine according to a first embodiment of the present invention.

In fig. 1, the hydraulic drive device of the present embodiment includes: a power machine 1; a main pump 2 as a variable displacement hydraulic pump driven by the power machine 1; a fixed capacity type pilot pump 30; boom cylinder 3a, arm cylinder 3b, swing motor 3c, bucket cylinder 3d (see fig. 4), swing cylinder 3e (see fig. 4), travel motors 3f and 3g (see fig. 4), and blade cylinder 3h (see fig. 4), which are actuators driven by pressure oil discharged from the main pump 2; a pressure oil supply path 5 for introducing pressure oil discharged from the main pump 2 into the plurality of actuators 3a, 3b, 3c, 3d, 3f, 3g, and 3 h; and a control valve block 4 connected downstream of the pressure oil supply line 5 and guiding pressure oil discharged from the main pump 2. Hereinafter, "drivers 3a, 3b, 3c, 3d, 3f, 3g, 3 h" will be referred to simply as "drivers 3a, 3b, 3c.

A plurality of directional control valves 6a, 6b, 6c, for controlling the plurality of actuators 3a, 3b, 3c, and a plurality of pressure compensation valves 7a, 7b, 7c, which are located downstream of the inlet throttle openings of the plurality of directional control valves 6a, 6b, 6c. The pressure compensation valves 7a, 7b, and 7c. are provided with springs that bias the spools of the pressure compensation valves 7a, 7b, and 7c. in the closing direction, and further guide the pressure on the downstream side of the inlet throttle openings of the plurality of direction switching valves 6a, 6b, and 6c. to the side that biases the spools of the pressure compensation valves 7a, 7b, and 7c. in the opening direction, and guide the maximum load pressure Plmax of the plurality of actuators 3a, 3b, and 3c. described later to the side that biases the spools of the pressure compensation valves 7a, 7b, and 7c.

The plurality of directional control valves 6a, 6b, 6c and the plurality of pressure compensating valves 7a, 7b, 7c form a control valve device that distributes and supplies the pressure oil discharged from the main pump 2 to the plurality of actuators 3a, 3b, 3c.

Further, in the control valve block 4, a relief valve 14 that discharges the pressure oil of the pressure oil supply path 5 to the tank when the pressure is equal to or higher than a predetermined set pressure, and an unloading valve 15 that discharges the pressure oil of the pressure oil supply path 5 to the tank when the pressure is equal to or higher than a certain set pressure are provided downstream of the pressure oil supply path 5.

Further, shuttle valves 9a, 9b, 9c, which are connected to load pressure detection ports of the plurality of directional control valves 6a, 6b, 6c, are arranged in the control valve block 4. The shuttle valves 9a, 9b, 9c. are used to detect the highest load pressure of the plurality of actuators 3a, 3b, 3c. Shuttle valves 9a, 9b, 9c are connected in race mode, respectively, the uppermost shuttle valve 9a detecting the highest load pressure.

Fig. 2 is an enlarged view of a peripheral portion of the unloader valve. The unloader valve 15 includes a pressure receiving portion 15a and a spring 15b that guide the highest load pressure of the plurality of actuators 3a, 3b, 3c. Further, an electromagnetic proportional pressure reducing valve 22 for generating a control pressure to the unload valve 15 is provided, and the unload valve 15 includes a pressure receiving portion 15c for guiding an output pressure (control pressure) of the electromagnetic proportional pressure reducing valve 22 in a direction to close the unload valve 15.

Further, the hydraulic drive device of the present embodiment includes: a regulator 11 associated with the main pump 2 and used to control the capacity thereof; and an electromagnetic proportional pressure reducing valve 21 for generating a command pressure to the regulator 11.

Fig. 3 is an enlarged view of the main pump peripheral portion including the regulator 11. The regulator 11 includes a differential piston 11b driven by a pressure receiving area difference, a horsepower control tilt control valve 11e, and a flow rate control tilt control valve 11i, wherein a large diameter side pressure receiving chamber 11c of the differential piston 11b is connected to an oil passage 31a (a pilot hydraulic pressure source) or the flow rate control tilt control valve 11i, which is a pressure oil supply passage of the pilot pump 30, via the horsepower control tilt control valve 11e, the small diameter side pressure receiving chamber 11a is always connected to the oil passage 31a, and the flow rate control tilt control valve 11i guides the pressure or the tank pressure of the oil passage 31a to the horsepower control tilt control valve 11 e.

The horsepower control tilt control valve 11e includes: a sleeve 11f that moves together with the differential piston 11 b; a spring 11d positioned on the side where the flow rate control tilt control valve 11i and the large diameter side pressure receiving chamber 11c of the differential piston 11b communicate with each other; and a pressure receiving chamber 11g for guiding the pressure of the pressure oil supply passage 5 of the main pump 2 through the oil passage 5a in a direction in which the oil passage 31a communicates with the small-diameter side and large-diameter side pressure receiving chambers 11a and 11c of the differential piston 11 b.

The flow rate control tilt control valve 11i includes: a sleeve 11j that moves together with the differential piston 11 b; a pressure receiving portion 11h that guides an output pressure (control pressure) of the electromagnetic proportional pressure reducing valve 21 in a direction in which pressure oil of the horsepower control tilt control valve 11e is discharged to the oil tank; and a spring 11k located on the side of guiding the pressure oil of the oil passage 31a to the horsepower control tilt control valve 11 e.

When the large-diameter side pressure receiving chamber 11c communicates with the oil passage 31a via the horsepower control tilt control valve 11e and the flow rate control tilt control valve 11i, the differential piston 11b moves leftward in the drawing in accordance with the pressure receiving area difference, and when the large-diameter side pressure receiving chamber 11c communicates with the oil tank via the horsepower control tilt control valve 11e and the flow rate control tilt control valve 11i, the differential piston 11b moves rightward in the drawing in accordance with the force received from the small-diameter side pressure receiving chamber 11 a. When the differential piston 11b moves leftward in the drawing, the discharge flow rate decreases as the pump capacity, which is the tilting angle of the variable capacity main pump 2, decreases, and when the differential piston 11b moves rightward in the drawing, the discharge flow rate increases as the tilting angle and the pump capacity of the main pump 2 increase and decrease.

A pilot relief valve 32 is connected to the pressure oil supply path (oil path 31a) of the pilot pump 30, and a constant pilot pressure (Pi0) is generated in the oil path 31a by the pilot relief valve 32.

A pilot valve for controlling a plurality of operation lever devices 60a, 60b, 60c., of a plurality of directional control valves 6a, 6b, 6c., is connected downstream of the pilot relief valve 32 via a switching valve 33, and by operating the switching valve 33 with a valve lock lever 34 provided in a driver seat 521 (see fig. 4) of a construction machine such as a hydraulic excavator, it is possible to switch whether to supply pilot pressure (Pi0) generated by the pilot relief valve 32 as pilot primary pressure to the pilot valves of the plurality of operation lever devices 60a, 60b, 60c., or to discharge pressure oil of the pilot valves to a tank.

The hydraulic drive device of the present embodiment further includes: a pressure sensor 40 for detecting the highest load pressure of the plurality of drivers 3a, 3b, 3 c.; pressure sensors 41a1 and 41a2 for detecting operation pressures a1 and a2 of the pilot valve of the operation lever device 60a of the boom cylinder 3 a; pressure sensors 41b1 and 41b2 for detecting respective operation pressures b1 and b2 of the pilot valve of the operation lever device 60b of the boom cylinder 3 b; a pressure sensor 41c for detecting the operating pressures c1, c2 of the pilot valve of the operating lever device 60c of the swing motor 3 c; a pressure sensor, not shown, for detecting an operating pressure of a pilot valve of an operating lever device of another actuator, not shown; a pressure sensor 42 for detecting the pressure in the pressure oil supply passage 5 of the main pump 2 (discharge pressure of the main pump 2); a tilt angle sensor 50 that detects a tilt angle of the main pump 2; a rotational speed sensor 51 for detecting the rotational speed of the power machine 1; and a controller 70.

The controller 70 is constituted by a microcomputer including a CPU, a storage unit, and the like, which are not shown, and peripheral circuits thereof, the storage unit being constituted by a ROM (read Only memory), a ram (random access memory), a flash memory, and the like, and the controller 70 operates according to a program stored in the ROM, for example.

The controller 70 inputs detection signals of the pressure sensor 40, the pressure sensors 41a1, 41a2, 41b1, 41b2, and 41c., the pressure sensor 42, the tilt angle sensor 50, and the rotation speed sensor 51, and outputs control signals to the electromagnetic proportional pressure reducing valves 21 and 22.

Fig. 4 shows an external appearance of a hydraulic excavator on which the hydraulic drive device is mounted.

The hydraulic excavator includes an upper revolving structure 502, a lower traveling structure 501, and a swing type front work machine 504, and the front work machine 504 includes a boom 511, an arm 512, and a bucket 513. The upper revolving structure 502 is revolvable with respect to the lower traveling structure 501 by the revolution of the revolving motor 3c. A swing post 503 is attached to a front portion of the upper revolving structure, and a front work implement 504 is attached to the swing post 503 so as to be vertically movable. The swing post 503 is rotatable in the horizontal direction with respect to the upper revolving body 502 by extension and contraction of the swing cylinder 3e, and the boom 511, the boom 512, and the bucket 513 of the front working machine 504 are rotatable in the vertical direction by extension and contraction of the boom cylinder 3a, the boom cylinder 3b, and the bucket cylinder 3 d. A blade 506 that moves up and down by extension and contraction of the blade cylinder 3h is attached to the center frame 505 of the lower traveling body 501. The lower traveling structure 501 travels by driving the left and right crawler belts by the rotation of the traveling motors 3f and 3 g.

The upper slewing body 502 is provided with an operator cab 50, and an operator's seat 521, boom cylinders 3a, boom cylinders 3b, bucket cylinders 3d, operation lever devices 60a, 60b, 60c, 60d for the slewing motor 3c, an operation lever device 60e for the swing cylinder 3e, an operation lever device 60h for the blade cylinder 3h, operation lever devices 60f, 60g for the traveling motors 3f, 3g, and a door lock lever 24 are provided in the operator's seat 508.

Fig. 5 shows a functional block diagram of the controller 70 of the hydraulic drive apparatus shown in fig. 1.

The output of the tilt angle sensor 50 indicating the tilt angle of the main pump 2 and the output of the rotation speed sensor 51 indicating the rotation speed of the power machine 1 are input to the main pump actual flow rate calculation unit 71, the output of the rotation speed sensor 51 and the outputs of the pressure sensors 41a1, 41b1, 41c indicating the lever operation amount (operation pressure) are input to the required flow rate calculation unit 72, and the outputs of the pressure sensors 41a1, 41b1, 41c are input to the meter-in opening calculation unit 74. Note that, in fig. 5 to 11 and the following description, elements not shown in fig. 1 may be omitted for simplicity.

The output Plmax of the pressure sensor 40 indicating the highest load pressure of the plurality of actuators 3a, 3b, and 3c is introduced into the adder 81, and the output Ps of the pressure sensor 42 indicating the discharge pressure (pump pressure) of the main pump 2 is introduced into the differentiator 82.

The required flow rates Qr1, Qr2, and Qr3, which are the outputs of the required flow rate computing unit 72, and the flow rate Qa', which is the output of the main pump actual flow rate computing unit 71, are introduced into the required flow rate correcting unit 73.

The outputs Qr1 ', Qr2 ', Qr3 ' of the requested flow rate correcting unit 73 and the outputs Am1, Am2, Am3 of the meter-in opening computing unit 74 are introduced into the target differential pressure computing unit 75.

The target differential pressure calculation unit 75 outputs a command pressure (command value) Pi _ ul to the electromagnetic proportional pressure reducing valve 22 for the relief valve, and outputs a target differential pressure Δ Psd to the adder 81.

The adder 81 calculates a target pump pressure Psd (═ Plmax + Δ Psd) at which the target differential pressure Δ Psd and the maximum load pressure Plmax are added, and outputs the target pump pressure Psd to the differentiator 82.

The differentiator 82 calculates a differential pressure Δ P (which is Psd-Ps) obtained by subtracting a pump pressure (actual pump pressure) Ps, which is an output of the pressure sensor 42, from the target pump pressure Psd, and inputs the differential pressure Δ P to the main pump target tilt angle calculation unit 83.

The main pump target tilt angle calculation unit 83 calculates a command pressure Pi _ fc from the input differential pressure Δ P (Psd-Ps), and outputs the command pressure Pi _ fc as a command value to the electromagnetic proportional pressure reducing valve 21.

The controller 70 calculates the required flow rate of each of the plurality of actuators 3a, 3b, 3c and the throttle opening area of each of the plurality of directional control valves 6a, 6b, 6c based on the input amounts of the operation levers of the plurality of operation lever devices 60a, 60b, 60c in the required flow rate calculation unit 72, the required flow rate correction unit 73, the throttle opening calculation unit 74, and the target differential pressure calculation unit 75, calculates the pressure loss of the throttle opening of a specific one of the plurality of directional control valves 6a, 6b, 6c based on the throttle opening area and the required flow rate, outputs the pressure loss as the target differential pressure Δ Psd, and controls the set pressure of the unload valve 15.

The controller 70 selects the maximum value of the meter-in pressure loss of the plurality of directional control valves 6a, 6b, and 6c as the meter-in pressure loss of the specific directional control valve in the target differential pressure calculation unit 75, and outputs the pressure loss as the target differential pressure Δ Psd to control the set pressure of the unload valve 15.

Then, the controller 70 calculates a command value Pi _ fc for equalizing the discharge pressure of the main pump 2 (hydraulic pump) detected by the pressure sensor 42 to a pressure obtained by adding the maximum load pressure detected by the maximum load pressure detection device (shuttle valves 9a, 9b, 9c) to the target differential pressure in the main pump target tilt angle calculation unit 83, and outputs the command value Pi _ fc to the regulator 11 (pump control device) to control the discharge flow rate of the main pump 2.

Fig. 6 is a functional block diagram of the main pump actual flow rate calculation unit 71.

In the main pump actual flow rate calculation unit 71, the multiplier 71a multiplies the tilt angle qm input from the tilt angle sensor 50 and the rotation speed Nm input from the rotation speed sensor 51 to calculate the actual flow rate Qa' discharged from the main pump 2.

Fig. 7 is a functional block diagram of the required flow rate calculation unit 72.

In the required flow rate calculation unit 72, the operating pressures Pi _ a1, Pi _ b1, and Pi _ c input from the pressure sensors 41a1, 41b1, and 41c are converted into the reference required flow rates Qr1, Qr2, and Qr3 by the tables 72a, 72b, and 72c, respectively, and are multiplied by the rotation speed Nm input from the rotation speed sensor 51 by the multipliers 72d, 72e, and 72f, respectively, to calculate the required flow rates Qr1, Qr2, and Qr3 of the plurality of drivers 3a, 3b, and 3c.

Fig. 8 is a functional block diagram of the required flow rate correction unit 73.

The required flow rate correction unit 73 inputs the required flow rates Qr1, Qr2, and Qr3, which are the outputs of the required flow rate calculation unit 72, to the multipliers 73c, 73d, and 73e and the adder 73a, calculates the total value Qra by the adder 73a, and inputs the total value Qra to the denominator side of the divider 73b via the limiter 73f that limits the minimum value and the maximum value. On the other hand, the flow rate Qa 'as the output of the main pump actual flow rate calculation unit 71 is input to the numerator side of the divider 73b, and the divider 73b outputs the value Qa'/Qra to the multipliers 73c, 73d, and 73 e. Multipliers 73c, 73d, and 73e multiply Qr1, Qr2, and Qr3 described above and Qa '/Qra described above, respectively, to calculate corrected required flow rates Qr 1', Qr2 ', and Qr 3'.

Fig. 9 is a functional block diagram of the meter-in opening calculator 74.

In the meter-in opening calculator 74, the operating pressures Pi _ a1, Pi _ b1, and Pi _ c input from the pressure sensors 41a1, 41b1, and 41c are converted into meter-in opening areas Am1, Am2, and Am3 of the directional control valves by tables 74a, 74b, and 74 c. The tables 74a, 74b, and 74c store the meter-in opening areas of the directional control valves 6a, 6b, and 6c in advance, and are set to output 0 when the operating pressure is 0 and output a large value as the operating pressure increases. The maximum value of the meter-in opening area is set extremely large so that the meter-in pressure loss (LS differential pressure), which is the pressure loss that can occur in the meter-in openings of the directional control valves 6a, 6b, and 6c, is extremely small.

Fig. 10 is a functional block diagram of the target differential pressure calculation unit 75.

The inputs Qr1 ', Qr2 ', and Qr3 ' from the required flow rate correction unit 73 are input to the calculators 75a, 75b, and 75c, respectively. Further, the inputs Am1, Am2, Am3 from the meter-in opening arithmetic unit 74 are input to the arithmetic units 75a, 75b, 75c via limiters 75f, 75g, 75h that limit the minimum value and the maximum value, respectively. The calculators 75a, 75b, and 75c calculate the inlet throttle pressure losses Δ Psd1, Δ Psd2, and Δ Psd3 of the direction switching valves 6a, 6b, and 6c using the inputs Qr1 ', Qr2 ', Qr3 ', Am1, Am2, and Am3, respectively, by the following equations. Here, C is a predetermined contraction coefficient, and ρ is the density of the hydraulic oil.

[ number 1 ]

These pressure losses Δ Psd1, Δ Psd2, and Δ Psd3 are input to a maximum value selector 75d via limiters 75i, 75j, and 75k that limit the minimum value and the maximum value, respectively, and in the maximum value selector 75d, the maximum value among the pressure losses Δ Psd1, Δ Psd2, and Δ Psd3 is output to an adder 81 as a target differential pressure Δ Psd (an adjusted pressure for variably controlling the set pressure of the unloading valve 15), and then the target differential pressure Δ Psd is converted into a command pressure Pi _ ul by a table 75e and output to the electromagnetic proportional pressure reducing valve 22 as a command value.

Fig. 11 is a functional block diagram of the main pump target tilt angle calculation unit 83.

In the main pump target tilting angle calculation unit 83, the table 83a is inputted with the differential pressure Δ P (Psd-Ps) calculated by the differentiator 82, and converted into the target displacement increase/decrease amount Δ q. Δ q is added to the target capacity q 'before the 1 control cycle output from the delay element 83c by the adder 83b, output to the limiter 83d as a new target capacity q, limited to a value between the minimum value and the maximum value, and introduced into the table 83e as the target capacity q' after the limitation. The target capacity q' is converted into the command pressure Pi _ fc of the pair-wise proportional solenoid pressure reducing valve 21 in the table 83e and output as a command value.

Work ^ E

The operation of the hydraulic drive device configured as described above will be described.

The pressure oil discharged from the fixed displacement pilot pump 30 is supplied to the pressure oil supply path 31a, and the pilot relief valve 32 generates the fixed pilot primary pressure Pi0 in the pressure oil supply path 31 a.

(a) All the operating levers are neutral

Since the operating levers of all the operating lever devices 60a, 60b, and 60c · are neutral, all the pilot valves are neutral, and since the operating pressures a1, a2, b1, b2, c1, and c2.· are tank pressures, all the directional control valves 6a, 6b, and 6c · are in neutral positions.

Since all the directional control valves 6a, 6b, and 6c are in the neutral position, the load pressure detection oil passage of each actuator is connected to the tank via the directional control valve associated with each actuator.

Therefore, the tank pressure is detected as the maximum load pressure Plmax via the shuttle valves 9a, 9b, and 9c as the maximum load pressure detecting means, and the maximum load pressure Plmax is introduced into the pressure receiving portion 15a of the unload valve 15 and the pressure sensor 40.

The boom raising operation pressure a1, the boom excavating operation pressure b1, and the turning operation pressure c are detected by the pressure sensors 41a1, 41b1, and 41c, respectively, and the outputs Pi _ a1, Pi _ b1, and Pi _ c of the pressure sensors are introduced into the requested flow rate calculation unit 72 and the meter-in opening calculation unit 74.

The tables 72a, 72b, and 72c of the required flow rate calculation unit 72 store reference required flow rates for each lever input for each of boom raising, arm excavating, and turning operations in advance, and are set such that 0 is output when the input is 0, and a large value is output as the input increases.

As described above, when all the operation levers are neutral, the operation pressures Pi _ a1, Pi _ b1, and Pi _ c are all equal to the tank pressure, and therefore the reference required flow rates qr1, qr2, and qr3 calculated in the tables 72a, 72b, and 72c are all 0. Since Qr1, Qr2, and Qr3 are all 0, the required flow rates Qr1, Qr2, and Qr3, which are the outputs of the multipliers 72d, 72e, and 72f, are all 0.

The tables 74a, 74b, and 74c of the meter-in opening calculation unit 74 store the meter-in opening areas of the directional control valves 6a, 6b, and 6c in advance, and are configured to output 0 when the input is 0 and output a large value as the input increases.

As described above, when all the operation levers are neutral, the operation pressures Pi _ a1, Pi _ b1, and Pi _ c are all equal to the tank pressure, and therefore the meter-in opening areas Am1, Am2, and Am3, which are the outputs of the tables 74a, 74b, and 74c, are all 0.

The required flow rates Qr1, Qr2, and Qr3 are input to the required flow rate correction unit 73.

The required flow rates Qr1, Qr2, and Qr3 input to the required flow rate correction unit 73 are introduced into the adder 73a and the multipliers 73c, 73d, and 73 e.

Although the adder 73a calculates Qra-Qr 1+ Qr2+ Qr3, Qra-0 +0+0 when all the levers are neutral as described above.

The restriction by the limiter 73f is made between the minimum value and the maximum value that the main pump 2 can deliver. Here, when the minimum value is Qmin and the maximum value is Qmax, the limiter 73f is limited to Qmin because Qra ═ 0 < Qmin when all the operation levers are neutral, and Qra ═ Qmin is introduced to the denominator side of the divider 73 b.

On the other hand, as will be described later, when all the control levers are neutral, the divider 73b outputs Qr '/Qra' ═ 1 to the multipliers 73c, 73d, and 73e because the main pump actual flow rate is kept at the minimum value Qmin.

As described above, when all the operation levers are neutral, since Qr1, Qr2, and Qr3 are all 0, the outputs Qr1 ', Qr2 ', and Qr3 ' of the multipliers 73c, 73d, and 73e are all 0 × 1 — 0.

The target differential pressure calculation unit 75 calculates the pressure loss generated in the meter-in openings of the directional control valves 6a, 6b, and 6c by the above-described expressions, based on the corrected required flow rates Qr1 ', Qr2 ', Qr3 ', and the meter-in opening areas Am1, Am2, and Am 3.

First, the meter-in opening areas Am1, Am2, Am3 are limited by the limiters 75f, 75g, 75h to predetermined minimum values Am1 ', Am2 ', Am3 ' larger than 0.

When all the operation levers are neutral, as described above, the meter-in opening areas Am1, Am2, Am3 and the corrected required flow rates Qr1 ', Qr2 ', Qr3 ' are all 0, but as described above, the meter-in opening areas Am1, Am2, Am3 are limited to values greater than 0, and therefore the pressure losses Δ Psd1, Δ Psd2, Δ Psd3 as the outputs of the calculators 75a, 75b, 75c are all 0. The pressure losses Δ Psd1, Δ Psd2, and Δ Psd3, which are outputs of the calculators 75a, 75b, 75c, are limited to values of 0 or more and a predetermined maximum value Δ Psd _ max or less by the limiters 75i, 75j, 75k, and the maximum value of the pressure losses Δ Psd1, Δ Psd2, and Δ Psd3 is output as the target differential pressure Δ Psd by the maximum value selector 75 d.

As described above, when all the operation levers are neutral, the target differential pressure Δ Psd is 0.

The target differential pressure Δ Psd is converted into a command pressure Pi _ ul by the table 75e, and is output as a command value to the electromagnetic proportional pressure reducing valve 22 for the relief valve.

When all the operation levers are neutral as described above, the maximum load pressure Plmax is equal to the tank pressure.

The set pressure of the unloader valve 15 is determined by the maximum load pressure Plmax introduced into the pressure receiving portion 15a, the spring 15b, and the output pressure (═ Δ Psd) of the proportional solenoid pressure reducing valve 22 introduced into the pressure receiving portion 15c, but since the maximum load pressure Plmax and the output pressure (═ Δ Psd) of the proportional solenoid pressure reducing valve 22 are both tank pressures, the set pressure of the unloader valve 15 is kept at a very small value determined by the spring 15 b.

Therefore, the pressure oil discharged from the variable displacement main pump 2 is discharged from the unload valve 15 to the tank, and the pressure in the pressure oil supply passage 5 is maintained at the low pressure.

On the other hand, the target differential pressure Δ Psd, which is the output of the target differential pressure calculation unit 75, is added to the maximum load pressure Plmax by the adder 81, but when all the operation levers are neutral as described above, Plmax and Δ Psd are tank pressures 0, and therefore the target pump pressure Psd, which is the output thereof, is also 0.

The target pump pressure Psd and the pump pressure Ps detected by the pressure sensor 42 are introduced to the positive side and the negative side of the differential 82, respectively, and the difference Δ P therebetween is set to Psd-Ps, which is input to the main pump target tilt angle calculation unit 83.

In the main pump target tilting angle calculation unit 83, the above-described Δ P (Psd-Ps) is converted into the target displacement increase amount Δ q by the table 83a using the table 83 a. As shown in fig. 11, table 83a is configured such that Δ q < 0 is set when Δ P < 0, Δ q is set to 0 when Δ P is set to 0, Δ q > 0 is set when Δ P > 0, and the limit is set to a predetermined value when Δ P is large or small to some extent or more.

The target displacement increase amount Δ q is added to the target displacement q 'before the control step 1 described later by the adder 83b to be q, and is limited to a value between the physical minimum and maximum values of the main pump 2 by the limiter 83d, and is output as the target displacement q'.

The target capacity q' is converted into the command pressure Pi _ fc of the pair of proportional solenoid pressure reducing valves 21 in the table 83e, and the proportional solenoid pressure reducing valves 21 are controlled.

As described above, when all the operation levers are neutral, Psd (the maximum load pressure Plmax + the target differential pressure Δ Psd) is equal to the tank pressure.

On the other hand, the pressure in the pressure oil supply passage 5, that is, the pump pressure Ps is maintained at a pressure greater than the tank pressure by an amount determined by the spring 15b by the unload valve 15 as described above.

Therefore, when all the levers are neutral, Δ P (═ Psd-Ps) < 0, and therefore Δ q < 0 is shown in table 83 a. The target capacity q 'before 1 step obtained by the delay element 83c is added as new q by the adder 83b, but the target capacity q' before 1 step is kept at the minimum value thereof because the limiter 83d limits the minimum and maximum tilting of the main pump 2.

(b) In the case of a boom raising operation

The boom raising operation pressure a1 is output from the pilot valve of the boom operation lever device 60 a. The boom raising operation pressure a1 is introduced into the direction switching valve 6a and the pressure sensor 41a1, and the direction switching valve 6a is switched rightward in the drawing.

Since the direction switching valve 6a is switched, the load pressure of the boom cylinder 3a is introduced into the unload valve 15 and the pressure sensor 40 as the maximum load pressure Plmax via the shuttle valve 9 a.

The pressure oil introduced from the pressure oil supply passage 5 to the directional control valve 6a is guided to the upstream side of the pressure compensating valve 7a via the meter-in opening thereof.

The pressure compensating valve 7a is controlled so that the pressure on the downstream side of the meter-in opening is equal to the maximum load pressure Plmax, but when the boom raising is operated alone, the maximum load pressure Plmax becomes equal to the load pressure of the boom cylinder 3a, and therefore the pressure compensating valve 7a does not throttle, and the opening thereof is kept fully open.

The pressure oil that has passed through the pressure compensating valve 7a is supplied to the bottom side of the boom cylinder 3a again via the direction switching valve 6 a. Since the pressure oil is supplied to the bottom side of the boom cylinder 3a, the boom cylinder is extended.

On the other hand, the boom raising operation pressure a1 is input to the required flow rate calculation unit 72 as the output Pi _ a1 of the pressure sensor 41a1 to calculate the required flow rate Qr 1.

The actual discharge flow rate of the variable displacement main pump 2 is calculated by the main pump actual flow rate calculation unit 71 based on the inputs from the tilt angle sensor 50 and the rotation speed sensor 51, but after the boom raising operation is performed with all the operation levers from the neutral state, the tilting of the variable displacement main pump 2 is kept to the minimum as described in the case (a) where all the operation levers are neutral, and therefore the actual main pump discharge flow rate Qa' also has the minimum value.

The required flow rate Qr1 is corrected to Qr1 'by the required flow rate correction portion 73 while being restricted to the main pump actual flow rate Qa'.

The boom raising operation pressure a1 is also introduced into the meter-in opening arithmetic unit 74 as the output Pi _ a1 of the pressure sensor 41a1, and is converted into the meter-in opening area Am1 by the table 74a and output.

The target differential pressure calculation unit 75 calculates the pressure loss occurring in the meter-in opening of each directional control valve based on the corrected required flow rates Qr1 ', Qr2 ', Qr3 ', and the meter-in opening areas Am1, Am2, and Am3 according to the above expressions.

When the boom raising operation is performed, the corrected required flow rate Qr 1' and the boom-raising meter-in opening area Am1 are input to the arithmetic unit 75a, and the meter-in pressure loss Δ Psd1 of the directional control valve 6a is calculated according to the following equation.

Number 2

Similarly, the meter-in pressure losses Δ Psd2 and Δ Psd3 of the directional control valves 6b and 6c are calculated, but since Δ Psd2 is equal to Δ Psd3 equal to 0 as in the case where all the levers are neutral, the pressure loss Δ Psd1 which is the maximum value is selected by the maximum value selector 75d, and Δ Psd is equal to Δ Psd1, and the command pressure Pi _ ul of the proportional solenoid pressure reducing valve 22 for the load relief valve is converted and output by using the table 75e, and the target differential pressure Δ Psd is output to the adder 81.

The output Δ Psd of the electromagnetic proportional pressure reducing valve for unload valve 22 is introduced into the pressure receiving portion 15c of the unload valve 15, and functions to increase the set pressure of the unload valve 15 by the amount of Δ Psd.

As described above, since the load pressure Pl1 of the boom cylinder 3a is introduced as Plmax into the pressure receiving portion 15a of the unload valve 15, the set pressure of the unload valve 15 is set to Plmax + Δ Psd + spring force, that is, Pl1 (the load pressure of the boom cylinder 3 a) + Δ Psd (the differential pressure generated at the inlet throttle opening of the directional control valve 6a for controlling the boom cylinder 3 a) + spring force, and the pressure oil supply path 5 blocks the oil path discharged to the tank.

On the other hand, the adder 81 adds the maximum load pressure Plmax and the target differential pressure Δ Psd to calculate the target pump pressure Psd equal to Plmax + Δ Psd, but when the boom raising operation is performed alone, as described above, Plmax is equal to Pl1, the adder calculates the target pump pressure Psd equal to Pl1 (the load pressure of the boom cylinder 3 a) + Δ Psd (the differential pressure generated at the inlet throttle opening of the directional control valve 6a for controlling the boom cylinder 3 a) and outputs the target pump pressure Psd to the differentiator 82.

The differentiator 82 calculates a difference between the target pump pressure Psd and the pressure (actual pump pressure Ps) of the pressure oil supply passage 5 detected by the pressure sensor 42 as Δ P (Psd-Ps), and outputs the difference to the main pump target tilt angle calculation unit 83.

In the main pump target tilt angle calculation unit 83, the differential pressure Δ P is converted into the increase/decrease amount Δ q of the target capacity by the table 83a, but when the boom raising operation is performed from a state in which all the levers are neutral, the actual pump pressure Ps is kept at a value smaller than the target pump pressure Psd at the beginning of the operation (described in the case in which all the levers are neutral in (a)), and therefore Δ P (which is Psd-Ps) is a positive value.

The characteristic of the table 83a is that when the differential pressure Δ P is a positive value, the target amount of increase in capacity Δ q is also positive, and therefore the target amount of increase in capacity Δ q is also positive.

The target capacity q 'before the control step 1 is added to the above-mentioned capacity increase/decrease amount Δ q by the adder 83b and the delay element 83c to calculate a new capacity q, and since the target capacity increase/decrease amount Δ q is positive as described above, the target capacity q' increases.

The target capacity q 'is converted into a command pressure Pi _ fc for the proportional solenoid pressure reducing valve 21 for the main pump tilt control by the table 83e, and the output (Pi _ fc) of the proportional solenoid pressure reducing valve 21 is introduced into the pressure receiving portion 11h of the flow control tilt control valve 11i in the regulator 11 of the main pump 2, and the tilt angle of the main pump 2 is controlled so as to be equal to the target capacity q'.

The target capacity q' and the discharge rate of the main pump 2 continue to increase until the actual pump pressure Ps becomes equal to the target pump pressure Psd, and finally, the actual pump pressure Ps and the target pump pressure Psd are maintained in the same state.

In this way, the main pump 2 performs load sensing control in which the target differential pressure is varied by increasing or decreasing the flow rate of a target pressure that is obtained by adding the maximum load pressure Plmax to the pressure loss Δ Psd that may occur at the meter-in opening of the directional control valve 6a associated with the boom cylinder 3 a.

(c) The situation of simultaneous lifting operation of the boom and digging and installing operation of the cantilever

The boom raising operation pressure a1 is output from the pilot valve of the boom operation lever device 60a, and the boom excavating operation pressure b1 is output from the pilot valve of the boom operation lever device 60 b.

The boom raising operation pressure a1 is introduced into the direction switching valve 6a and the pressure sensor 41a1, and the direction switching valve 6a is switched in the right direction in the drawing.

The arm excavating operation pressure b1 is introduced into the direction switching valve 6b and the pressure sensor 41b1, and the direction switching valve 6b is switched in the right direction in the drawing.

Since the directional control valves 6a and 6b are switched, the load pressure of the boom cylinder 3a is introduced into the shuttle valve 9a via the directional control valve 6a, and the load pressure of the boom cylinder 3b is introduced into the shuttle valve 9a via the directional control valve 6b and the shuttle valve 9 b.

The shuttle valve 9a selects the higher one of the load pressure of the boom cylinder 3a and the load pressure of the boom cylinder 3b as the highest load pressure Plmax. When the operation in the air is assumed, the load pressure of the boom cylinder 3a is usually larger than the load pressure of the arm cylinder 3b, and therefore, if the load pressure of the boom cylinder 3a is larger than the load pressure of the arm cylinder 3b, the maximum load pressure Plmax is equal to the load pressure of the boom cylinder 3 a.

The highest load pressure Plmax is introduced into the pressure receiving portion 15a of the unload valve 15 and the pressure sensor 40.

The pressure compensation valve 7a associated with the boom cylinder 3a is controlled so that the pressure on the downstream side of the meter-in opening of the directional control valve 6a associated with the boom cylinder 3a is equal to the maximum load pressure Plmax, but when the load pressure of the boom cylinder 3a is greater than the load pressure of the boom cylinder 3b as described above, the maximum load pressure Plmax is equal to the load pressure of the boom cylinder 3a, and therefore the pressure compensation valve 7a does not throttle and the opening thereof is kept fully open.

The pressure compensating valve 7b associated with the arm cylinder 3b is controlled so that the pressure on the downstream side of the meter-in opening of the directional control valve 6b associated with the arm cylinder 3b is equal to the maximum load pressure Plmax, that is, in this case, equal to the load pressure of the boom cylinder 3 a. Thus, the pressure on the downstream side of the meter-in opening of the directional control valve 6b is kept Plmax equal to the load pressure of the boom cylinder 3 a.

Since the differential pressures before and after the directional switching valves 6a and 6b, that is, the pump pressures (common) and the pressures downstream of the meter-in openings are kept equal, the directional switching valves 6a and 6b distribute the pressure oil in the pressure oil supply passage 5 according to the sizes of the meter-in openings of the boom cylinder 3a and the arm cylinder 3b, regardless of the sizes of the load pressures.

The pressure oil having passed through the pressure compensating valves 7a and 7b is supplied to the bottom side of the boom cylinder 3a and the bottom side of the arm cylinder 3b via the directional control valves 6a and 6b, respectively.

Since the pressure oil is supplied to the bottom side of the boom cylinder 3a and the bottom side of the boom cylinder 3b, the boom cylinder and the boom cylinder are extended.

On the other hand, the boom raising operation pressure a1 and the arm excavating operation pressure b1 are input to the required flow rate calculation unit 72 as the outputs Pi _ a1 and Pi _ b1 of the pressure sensors 41a1 and 41b1, respectively, and calculate the required flow rates Qr1 and Qr 2.

The main pump actual flow rate calculation unit 71 calculates the flow rate actually discharged by the variable displacement main pump 2 based on the inputs from the tilt angle sensor 50 and the rotation speed sensor 51, but after the boom raising and boom raising operations are performed with all the operation levers from the neutral state, the tilting of the variable displacement main pump 2 is kept to the minimum as described in the case (a) where all the operation levers are neutral, and therefore the main pump actual flow rate Qa' is also the minimum value.

The required flow rate correction unit 73 introduces the boom raising required flow rate Qr1 and the boom excavating required flow rate Qr2 into the adder 73a, and calculates Qra (Qr 1+ Qr2+ Qr3 ═ Qr1+ Qr 2).

The Qra calculated by the adder 73a is limited to a certain range of values by the limiter 73f, and the divider 73b divides the output of the main pump actual flow rate calculation unit 71 and the main pump actual flow rate Qa 'by Qa'/Qra, and supplies the outputs to the multipliers 73c, 73d, and 73 e.

That is, in the required flow rate correcting unit 73, the boom raising required flow rate Qr1 and the boom excavating required flow rate Qr2 are newly distributed in accordance with the specific gravity of Qr1 and Qr2 within the range of the flow rate Qa' actually discharged from the variable capacity main pump 2.

For example, when Qa 'is 30L/min, Qr1 is 20L/min, and Qr2 is 40L/min, Qra-Qr 1+ Qr2+ Qr 3-60L/min, and thus Qa'/Qra-1/2.

The corrected boom lifting required flow rate Qr1 ═ Qr1 × 1/2 ═ 20L/min × 1/2 ═ 10L/min, and the corrected boom digging required flow rate Qr2 ═ Qr2 × 1/2 ═ 40L/min × 1/2 ═ 20L/min.

Further, the boom raising operation pressure a1 and the arm excavating operation pressure b1 are introduced into the meter-in opening calculation unit 74 as the outputs Pi _ a1 and Pi _ b1 of the pressure sensors 41a1 and 41b1, and are converted into meter-in opening areas Am1 and Am2 by the tables 74a and 74b, and are output.

The target differential pressure calculation unit 75 calculates pressure losses Δ Psd1, Δ Psd2, and Δ Psd3 generated in the meter-in openings of the directional switching valves from the corrected required flow rates Qr1 ', Qr2 ', Qr3 ', and the meter-in opening areas Am1, Am2, and Am 3.

When the boom raising operation and the boom excavating operation are performed simultaneously, the corrected required flow rates Qr1 ', Qr 2' and the meter-in opening areas Am1 and Am2 are input to the calculators 75a and 75b, and Δ Psd1 and Δ Psd2 are calculated according to the following equations.

[ number 3 ]

Similarly, since the Δ Psd3 is calculated, and since the Δ Psd3 is equal to 0 as in the case where all the levers are neutral, the higher one of the Δ Psd1 and the Δ Psd2 is selected as Δ Psd by the maximum value selector 75d, and the command pressure Pi _ ul of the proportional solenoid pressure reducing valve 22 for the relief valve is converted by the table 75e and output as a command value, and at the same time, Δ Psd is output to the adder 81.

The output of the electromagnetic proportional pressure reducing valve 22 for the relief valve is introduced into the pressure receiving portion 15c of the relief valve 15, and functions to increase the set pressure of the relief valve 15 by Δ Psd.

As described above, when the load pressure of the boom cylinder 3a is greater than the load pressure of the arm cylinder 3b, the load pressure Pl1 of the boom cylinder 3a is introduced as Plmax into the pressure receiving portion 15a of the unload valve 15, and therefore the set pressure of the unload valve 15 is set to Plmax + Δ Psd + spring force, that is, Pl1 (the load pressure of the boom cylinder 3 a) + Δ Psd (the larger of the differential pressure generated at the inlet orifice of the directional control valve 6a associated with the boom cylinder 3a and the differential pressure generated at the inlet orifice of the directional control valve 6b associated with the arm cylinder 3 b) + spring force, and the pressure oil of the pressure oil supply passage 5 blocks the oil passage discharged to the tank.

On the other hand, the adder 81 adds the maximum load pressure Plmax and the above-described Δ Psd to calculate the target pump pressure Psd of Plmax + Δ Psd, but when the load pressure of the boom cylinder 3a is greater than the load pressure of the arm cylinder 3b, as described above, Plmax is equal to Pl1, and therefore the target pump pressure Psd is calculated to be Pl1 (the load pressure of the boom cylinder 3 a) + Δ Psd (the larger of the differential pressure generated at the inlet throttle opening of the directional control valve 6a associated with the boom cylinder 3a and the differential pressure generated at the inlet throttle opening of the directional control valve 6b associated with the arm cylinder 3 b), and the calculated target pump pressure Psd is output to the differentiator 82.

The differentiator 82 calculates a difference between the target pump pressure Psd and the pressure (actual pump pressure Ps) of the pressure oil supply passage 5 detected by the pressure sensor 42 as Δ P (Psd-Ps), and outputs the difference to the main pump target tilt angle calculation unit 83.

In the main pump target tilt angle calculation unit 83, the differential pressure Δ P is converted into the increase/decrease amount Δ q of the target capacity by the table 83a, but when the boom raising operation and the boom excavating operation are performed with all the levers in a neutral state, the actual pump pressure Ps is kept at a value smaller than the target pump pressure Psd at the beginning of the operation (described in the case where (a) all the levers are neutral), and therefore Δ P (Psd-Ps) is a positive value.

The characteristic of the table 83a is that when the differential pressure Δ P is a positive value, the target amount of increase in capacity Δ q is also positive, and therefore the target amount of increase in capacity Δ q is positive.

The target capacity q 'before the control step 1 is added to the above-mentioned capacity increase/decrease amount Δ q by the adder 83b and the delay element 83c to calculate a new capacity q, and since the target capacity increase/decrease amount Δ q is positive as described above, the target capacity q' increases.

The target capacity q 'is converted into a command pressure (command value) Pi _ fc with respect to the main pump proportional solenoid pressure reducing valve 21 for the main pump tilt control by the table 83e, and the output Pi _ fc of the main pump proportional solenoid pressure reducing valve 21 for the main pump tilt control is introduced into the pressure receiving portion 11h of the flow control tilt control valve 11i in the regulator 11 of the variable capacity main pump 2, and the tilt angle of the variable capacity main pump 2 is controlled so as to be equal to the target capacity q'.

The target capacity q' and the discharge rate of the variable capacity type main pump 2 continue to increase until the actual pump pressure Ps becomes equal to the target pump pressure Psd, and finally, the state where the actual pump pressure Ps becomes equal to the target pump pressure Psd is maintained.

In this way, the variable displacement main pump 2 compares the pressure loss that can occur at the meter-in opening of the directional control valve 6a associated with the boom cylinder 3a and the pressure loss that can occur at the meter-in opening of the directional control valve 6b associated with the arm cylinder 3b, calculates the larger one of the pressure losses as the target differential pressure Δ Psd, and increases the flow rate of the pressure obtained by adding the maximum load pressure Plmax and the target differential pressure Δ Psd as the target pressure, thereby performing load sensing control in which the target differential pressure is variable.

Effect E

According to the present embodiment, the following effects can be obtained.

1. In the present embodiment, the flow-dividing control of the plurality of directional control valves 6a, 6b, and 6c is performed using the plurality of pressure compensating valves (flow rate sharing valves) 7a, 7b, and 7c which are respectively disposed downstream of the plurality of directional control valves 6a, 6b, and 6c and are controlled so that the pressure on the downstream side of the meter-in openings of the plurality of directional control valves 6a, 6b, and 6c is equal to the maximum load pressure, and therefore, even when the differential pressures (meter-in pressure loss) before and after the directional control valves 6a, 6b, and 6c associated with the actuators 3a, 3b, and 3c are extremely small, the flow-dividing control of the plurality of directional control valves 6a, 6b, and 6c can be stably performed.

2. In the present embodiment, the controller 70 calculates the meter-in pressure loss in each of the directional control valves 6a, 6b, and 6c associated with the actuators 3a, 3b, and 3c, selects the maximum value of the meter-in pressure loss (calculates the meter-in pressure loss of the inlet of the specific directional control valve), and outputs the pressure loss of the maximum value as the target differential pressure Δ Psd to control the set pressure (Plmax + Δ Psd + spring force) of the unload valve 15. Thus, the set pressure of the unload valve 15 is controlled to a value obtained by adding the target differential pressure Δ Psd and the spring force to the maximum load pressure, and therefore, for example, in a directional control valve associated with an actuator other than the maximum load pressure actuator, even when the meter-in opening is throttled to an extremely small value, the set pressure of the unload valve 15 can be finely controlled in accordance with the pressure loss of the meter-in opening of the directional control valve. As a result, even when the required flow rate changes rapidly and the pump pressure rises rapidly due to insufficient responsiveness of the pump flow rate control when the directional control valve having the meter-in pressure loss at the maximum value shifts from the combined operation of the lever operation including the operation lever to the semi-independent operation, it is possible to minimize the leakage loss of the pressurized oil from the unloading valve 15 to the tank in excess, suppress a decrease in energy efficiency, prevent a rapid change in the actuator speed due to a rapid change in the flow rate of the pressurized oil supplied to each actuator, suppress occurrence of an uncomfortable shock, and realize excellent combined operability.

3. In addition, in the present embodiment, even when the differential pressure between the front and rear sides of each of the directional control valves 6a, 6b, and 6c is very small as described above, since the flow-dividing control of the plurality of directional control valves 6a, 6b, and 6c can be stably performed and the set pressure of the unload valve 15 can be finely controlled according to the pressure loss of the meter-in opening of the directional control valves 6a, 6b, and 6c, the final opening of the meter-in of each of the directional control valves 6a, 6b, and 6c (the meter-in opening area in the full stroke of the main spool) can be extremely increased, and thus the meter-in loss can be reduced and energy efficiency can be improved.

4. In the conventional load sensing control described in patent document 1, the discharge flow rate of the hydraulic pump is increased or decreased so that the LS differential pressure becomes equal to the predetermined target LS differential pressure, but when the final opening of the inlet throttle of the main spool is extremely large as described above, the LS differential pressure becomes substantially equal to 0, and therefore the hydraulic pump projects the maximum flow rate within the allowable range, and there is a problem that the flow rate control according to the input of each operation lever cannot be performed.

In the present embodiment, the controller 70 calculates a target differential pressure Δ Psd for adjusting the set pressure of the unload valve 15, and controls the discharge flow rate of the main pump 2 using the target differential pressure Δ Psd so that the discharge pressure of the main pump 2 detected by the pressure sensor 42 becomes equal to the pressure obtained by adding the maximum load pressure to the target differential pressure Δ Psd. Therefore, even if the final opening of the inlet throttle of each of the directional control valves 6a, 6b, and 6c is extremely large, the problem that the pump flow rate control cannot be performed as in the case where the LS differential pressure is set to 0 by the conventional load sensing control does not occur, and the discharge flow rate of the main pump 2 can be controlled in accordance with the input of the operation lever.

5. Further, since the main pump 2 performs load sensing control in consideration of the meter-in pressure loss and the main pump 2 sufficiently discharges the pressure oil required for each actuator in accordance with the input of each control lever, a highly energy-efficient hydraulic system can be realized as compared with a flow control in which a target flow rate is determined simply by the input of each control lever.

6. Further, compared to the conventional technique described in patent document 2, the number of pressure sensors for detecting the load pressure of each actuator and the electromagnetic proportional pressure reducing valve can be reduced, and the cost of electronic control can be reduced.

< second embodiment >

Hereinafter, a hydraulic drive system for a construction machine according to a second embodiment of the present invention will be described mainly focusing on differences from the first embodiment.

Structure ^ E

Fig. 12 is a diagram showing a configuration of a hydraulic drive device of a construction machine according to a second embodiment.

In fig. 12, the second embodiment is configured to cancel the pressure sensor 40 for detecting the highest load pressure, to provide the pressure sensors 40a, 40b, and 40c for detecting the load pressures of the plurality of actuators 3a, 3b, and 3c, and to provide the controller 90 instead of the controller 70, as compared with the first embodiment.

Fig. 13 is a functional block diagram of the controller 90 according to the present embodiment.

Fig. 13 differs from the ID first embodiment shown in fig. 5 in that a maximum value selector 76, a maximum load pressure driver determination unit 77, a direction switching valve meter-in opening calculation unit 78 of the maximum load pressure driver, a post-correction required flow rate calculation unit 79 of the maximum load pressure driver, and a target differential pressure calculation unit 80 are provided instead of the target differential pressure calculation unit 75. These functional block diagrams are explained below.

In fig. 13, the outputs of the pressure sensors 40a, 40b, and 40c indicating the load pressures of the respective drivers are introduced into a maximum selector 76 and a maximum load pressure driver determination unit 77.

The maximum load pressure Plmax, which is the output of the maximum selector 76, is introduced into the maximum load pressure driver determination unit 77 together with the outputs Pl1, Pl2, Pl3 of the pressure sensors 40a, 40b, 40c, and the determination unit 77 introduces an identifier i indicating the maximum load pressure driver into the direction switching valve inlet throttle opening calculation unit 78 of the maximum load pressure driver and the post-correction required flow rate calculation unit 79 of the maximum load pressure driver. The maximum load pressure Plmax is introduced into the adder 81.

The direction switching valve meter-in opening calculation unit 78 of the highest load pressure actuator inputs the identifier i and the meter-in opening areas Am1, Am2, Am3 as outputs of the meter-in opening calculation unit 74, and outputs the meter-in opening area Ami of the direction switching valve of the highest load pressure actuator.

The corrected required flow rate computing unit 79 of the highest load pressure driver inputs the identifier i and the corrected required flow rates Qr1 ', Qr 2', and Qr3 'as the output of the required flow rate correcting unit 73, and outputs the corrected required flow rate Qri' of the highest load pressure driver.

The meter-in opening area Ami of the directional control valve of the maximum load pressure actuator and the corrected required flow rate Qri' of the maximum load pressure actuator are introduced into the target differential pressure calculation unit 80, and the target differential pressure calculation unit 80 outputs the target differential pressure Δ Psd to the adder 81 and outputs the command pressure (command value) Pi _ ul to the electromagnetic proportional pressure reducing valve 22.

The controller 90 calculates the required flow rate of each of the plurality of actuators 3a, 3b, and 3c and the opening area of the meter-in of each of the plurality of directional switching valves 6a, 6b, and 6c based on the input amounts of the operation levers of the plurality of operation lever devices 60a, 60b, and 60c in the required flow rate calculation unit 72, the required flow rate correction unit 73, the meter-in opening calculation unit 74, the maximum value selector 76, the maximum load pressure driver determination unit 77, the directional switching valve meter-in opening calculation unit 78, the post-correction required flow rate calculation unit 79, and the target differential pressure calculation unit 80, calculates the pressure loss of the meter-in of a specific one of the plurality of directional switching valves 6a, 6b, and 6c based on the opening area of the meter-in and the required flow rate, and outputs the pressure loss as the target differential pressure Δ Psd to control the set pressure of the unload valve 15.

The controller 90 calculates, among the plurality of directional control valves 6a, 6b, 6c, the meter-in pressure loss of the directional control valve associated with the driver of the highest load pressure detected by the highest load pressure detection device (shuttle valves 9a, 9b, 9c) as the meter-in pressure loss of the specific directional control valve, in the maximum value selector 76, the highest load pressure driver determination unit 77, the directional control valve meter-in opening calculation unit 78, the post-correction required flow rate calculation unit 79, and the target differential pressure calculation unit 80, and outputs the pressure loss as the target differential pressure Δ Psd to control the set pressure of the unload valve 15.

Fig. 14 is a functional block diagram of the maximum load voltage driver determining unit 77.

In the determination unit 77, the load pressures Pl1, Pl2, and Pl3 of the respective drivers input from the pressure sensors 40a, 40b, and 40c are introduced to the negative side of the differentiators 77a, 77b, and 77c, the highest load pressure Plmax from the maximum value selector 76 is introduced to the positive side of the differentiators 77a, 77b, and 77c, and the differentiators 77a, 77b, and 77c output Plmax-Pl1, Plmax-Pl2, and Plmax-Pl3 to the determination devices 77d, 77e, and 77f, respectively. The judgers 77d, 77e, and 77f are switched to the ON state and to the upper side of the figure when each judgment text is true, and are switched to the OFF state and to the lower side of the figure when each judgment text is false.

Fig. 14 shows a case where Plmax is Pl1, that is, a case where Plmax-Pl1 is 0, and in this case, the arithmetic unit 77g is selected to output i as the identifier i of 1 to the adder 77 m. On the other hand, when the judgment text is false in the judgers 77e and 77f, the calculators 77j and 77l are selected, respectively, and i is introduced as the identifier i, 0, to the adder 77 m. The adder 77m adds the outputs of the arithmetic units 77g, 77j, and 77l, and outputs i equal to 1.

Thus, when Plmax is Pl1, the output i is 1. Similarly, when Plmax is Pl2, the output i is 2, and when Plmax is Pl3, the output i is 3.

Fig. 15 is a functional block diagram of the direction switching valve meter-in opening calculation unit 78 of the maximum load pressure actuator.

In the arithmetic unit 78, the identifier i input from the maximum load pressure driver determining unit 77 is introduced into the determining devices 78a, 78b, and 78c, and the opening areas Am1, Am2, and Am3 input from the meter-in opening arithmetic unit 74 are introduced into the arithmetic devices 78d, 78f, and 78h, respectively. Fig. 15 shows a case where i is 1.

Since i is 1, the determiner 78a is turned ON, and switches to the upper side in the figure, and the arithmetic unit 78d is selected to introduce Am1 as the meter-in opening area Ami to the adder 78 j. In addition, when the determination devices 78b and 78c are OFF, switching is performed to the lower side in the figure, and the arithmetic devices 78g and 78i are selected, respectively, and 0 is introduced as the meter-in opening area Ami to the adder 78 j. In the adder 78j, Am1+0+0 is output as an inlet throttle opening area Ami 1.

Similarly, when i is 2, Am2 is output as the opening area Ami, and when i is 3, Am3 is output as the opening area Ami.

Fig. 16 is a functional block diagram of the post-correction required flow rate calculation unit 79 of the maximum load pressure driver.

In the arithmetic unit 79, the identifier i input from the maximum load pressure driver determination unit 77 is introduced into the determination units 79a, 79b, and 79c, and the corrected required flow rates Qr1 ', Qr2 ', and Qr3 ' input from the required flow rate correction unit 73 are introduced into the arithmetic units 79d, 79g, and 79h, respectively. Fig. 16 shows a case where i is 1.

Since i is 1, the determiner 79a is turned ON, and switches to the upper side in the figure, and selects the calculator 79d, and introduces Qr1 'to the adder 79j as the corrected required flow rate Qri'. The determination devices 79b and 79c are in the OFF state, switched to the lower side in the figure, and select the arithmetic devices 79g and 79i, respectively, and both introduce 0 to the adder 79j as the corrected required flow rate Qri'. The adder 79j outputs Qr1 '+ 0+0 as the corrected required flow rate Qri'.

Similarly, when i is 3, Qr2 'is output as the corrected required flow rate Qri', and when i is 3, Qr3 'is output as the corrected required flow rate Qri'.

Fig. 17 is a functional block diagram of the target differential pressure calculation unit 80.

In the arithmetic unit 80, the corrected required flow rate Qri' inputted from the corrected required flow rate arithmetic unit 79 of the maximum load pressure actuator is introduced into the arithmetic unit 80a, the meter-in opening area Ami inputted from the direction switching valve meter-in opening arithmetic unit 78 of the maximum load pressure actuator is introduced into the arithmetic unit 80a via the limiter 80c, the arithmetic unit 80a calculates the meter-in pressure loss of the direction switching valve of the maximum load pressure actuator as the target differential pressure Δ Psd (the adjustment pressure for variably controlling the set pressure of the unload valve 15) by the following equation, and the target differential pressure Δ Psd passed through the limiter 80d is outputted to the table 80b and the external adder 81. Here, C is a predetermined contraction coefficient, and ρ is the density of the hydraulic oil.

[ number 4 ]

In the table 80b, the target differential pressure Δ Psd is converted into the command pressure Pi _ ul of the pair of electromagnetic proportional pressure reducing valves 22, and is output in accordance with the command value.

Work ^ E

In the first embodiment, the meter-in pressure losses Δ Psd1, Δ Psd2, and Δ Psd3 of the direction switching valves 6a, 6b, and 6c associated with the boom cylinder 3a, the arm cylinder 3b, and the rotation motor 3c are calculated, respectively, and the maximum value of these is calculated as the total target differential pressure Δ Psd, whereas in the target differential pressure calculation unit 80 of the second embodiment, the highest load pressure driver determination unit 77 determines the highest load pressure driver, and the target differential pressure calculation unit 80 calculates the meter-in pressure loss of the highest load pressure driver as the total target differential pressure Δ Psd.

The unloading valve 15 is controlled to a set pressure determined by the target differential pressure Δ Psd, the maximum load pressure Plmax, and the spring force, as in the first embodiment. The adder 81 adds the maximum load pressure Plmax, which is the output of the maximum selector 76, to the target differential pressure Δ Psd, calculates the target pump pressure Psd, and outputs the target pump pressure Psd to the differentiator 82.

Effect E

1. The present embodiment can obtain the same effects as effects 1, 3, 4, and 5 of the first embodiment, and can obtain the following effects similar to effect 2.

2. In the present embodiment, the controller 790 calculates the opening areas of the meter-in ports of the plurality of directional control valves 6a, 6b, and 6c based on the input amounts of the respective operation levers, calculates the pressure loss of the meter-in port of the directional control valve (the specific directional control valve) based on the opening area of the directional control valve (the specific directional control valve) associated with the highest load pressure actuator among the plurality of directional control valves 6a, 6b, and 6c and the required flow rate of the directional control valve (the specific directional control valve), and outputs the pressure loss as the target differential pressure Δ Psd to control the set pressure (Plmax + Δ Psd + spring force) of the unload valve 15. Thus, the set pressure of the unload valve 15 is controlled to a value obtained by adding the target differential pressure Δ Psd and the spring force to the maximum load pressure, and therefore, when the meter-in opening of the directional switching valve (specific directional switching valve) associated with the maximum load pressure actuator is throttled by a half operation or the like of the directional switching valve, the set pressure of the unload valve 15 can be finely controlled. As a result, even when the required flow rate changes rapidly and the pump pressure rises rapidly due to insufficient responsiveness of the pump flow rate control when the composite operation including the lever operation of the directional control valve associated with the highest load pressure actuator shifts to the semi-independent operation, for example, the release loss of the pressure oil from the unloading valve 15 to the tank is suppressed to the minimum, and the rapid change in the actuator speed due to the rapid change in the flow rate of the pressure oil supplied to each actuator is suppressed, thereby achieving excellent composite operability.

< third embodiment >

Hereinafter, a hydraulic drive system for a construction machine according to a third embodiment of the present invention will be described mainly focusing on differences from the first embodiment.

Structure ^ E

Fig. 18 is a diagram showing a configuration of a hydraulic drive device of a construction machine according to a third embodiment.

In fig. 18, the third embodiment is configured to eliminate the pressure sensor 42 for detecting the pressure of the pressure oil supply passage 5, that is, the pump pressure, and to provide a controller 95 instead of the controller 70, as compared with the first embodiment.

Fig. 19 is a functional block diagram of the controller 95 according to the present embodiment.

Fig. 19 differs from the first embodiment shown in fig. 5 in that a required flow rate calculation unit 91 and a main pump target tilt angle calculation unit 93 are provided instead of the required flow rate calculation unit 72 and the main pump target tilt angle calculation unit 83, and the adder 81 and the differentiator 82 are eliminated.

The controller 95 calculates the sum of the required flow rates of the actuators 3a, 3b, and 3c based on the input amounts of the operation levers of the plurality of operation lever devices 60a, 60b, and 60c in the required flow rate calculation unit 91 and the main pump target tilt angle calculation unit 93, calculates a command value Pi _ fc for making the discharge flow rate of the main pump 2 (hydraulic pump) equal to the sum of the required flow rates, and outputs the command value Pi _ fc to the regulator 11 (pump control device) to control the discharge flow rate of the main pump 2.

Fig. 20 is a functional block diagram of the required flow rate calculation unit 91.

In fig. 20, the operation pressures Pi _ a1, Pi _ b1, and Pi _ c input from the pressure sensors 41a1, 41b1, and 41c are converted into the required tilt angles (capacities) Qr1, Qr2, and Qr3 by tables 91a, 91b, and 91c, the required flow rates Qr1, Qr2, and Qr3 are calculated by multipliers 91d, 91e, and 91f together with the input Nm from the rotation speed sensor 51, the adder 91g calculates qra (Qr 1+ Qr2+ Qr3), and the sum qra of the required tilt angles is output to the main pump target tilt angle calculation unit 93.

Fig. 21 is a functional block diagram of the main pump target tilt angle calculation unit 93.

The input qra (qr 1+ qr2+ qr3) from the required flow rate calculation unit 91 is limited by the limiter 93a to a value between the minimum value and the maximum value of the tilting of the main pump 2, and is converted into the command pressure Pi _ fc of the pair electromagnetic proportional pressure reducing valve 21 by the table 93b and output as a command value.

Work ^ E

While the first embodiment performs so-called load sensing control in which the discharge flow rate of the main pump 2 is controlled so that the pressure of the hydraulic oil supply passage 5, that is, the pump pressure becomes the maximum load pressure Plmax + the meter-in pressure loss of the maximum load pressure actuator, in the second embodiment, the discharge flow rate of the main pump 2 is determined only by the required tilt angle qra by the main pump target tilt angle calculation unit 93, and the required tilt angle qra is determined only by the input amount of each operation lever.

Effect E

1. In the present embodiment, the same effects as effects 1 to 3 and 6 of the first embodiment can be obtained, and the following effects can be obtained.

2. In the present embodiment, the main pump 2 performs flow rate control for calculating the sum of the required flow rates of the plurality of directional control valves 6a, 6b, and 6c based on the input amount of each control lever and determining the target flow rate, and thus a more stable hydraulic system can be realized than in the case where load sensing control, which is one type of feedback control, is performed as described in the first embodiment. In addition, a pressure sensor for detecting the pump pressure can be omitted, and the cost of the hydraulic system can be reduced.

< others >

In the above embodiment, the spring 15b is provided to stabilize the operation of the unload valve 15, but the spring 15b may be eliminated. Instead of providing the spring 15b in the unloading valve 15, the controller 70, 90, or 95 may calculate the value of "Δ Psd + spring force" as the target differential pressure.

In the second embodiment, as in the first embodiment, a device that performs load sensing control may be used as the pump control device, and in the first embodiment, a device that performs flow control by calculating the sum of the flow rates requested by the plurality of directional control valves 6a, 6b, and 6c may be used as the pump control device, as in the second embodiment.

Further, the above-described embodiment has described the case where the construction machine is a hydraulic excavator having a crawler on the lower traveling body, but similar effects can be obtained even in the case of other construction machines such as a wheel type hydraulic excavator, a hydraulic crane, and the like.

Description of the symbols

1-power machine, 2-variable capacity type main pump (hydraulic pump), 3 a-3 h-driver, 4-control valve block, 5-pressure oil supply path (main), 6 a-6 c-directional switching valve (control valve device), 7 a-7 c-pressure compensating valve (control valve device), 9 a-9 c-shuttle valve (highest load pressure detecting device), 11-regulator (pump control device), 14-overflow valve, 15-unload valve, 15a, 15 c-pressure receiving part, 15 b-spring, 21, 22-electromagnetic proportional pressure reducing valve, 30-pilot pump, 31 a-pressure oil supply path (pilot), 32-pilot overflow valve, 40, 41a 1-41 h2, 42-pressure sensor, 40 a-40 c-pressure sensor, 60 a-60 c-operating lever means, 70, 90, 95-controller.