JPH01199004A - Hydraulic drive device - Google Patents

Abstract

(57) [Summary] This bulletin contains application data before electronic filing, so abstract data is not recorded.

Description

[Detailed Description of the Invention] [Field of Industrial Application] The present invention relates to a hydraulic drive device for hydraulic construction machinery equipped with a plurality of hydraulic actuators, such as a hydraulic excavator or a hydraulic crane, and particularly relates to a hydraulic drive device for a hydraulic construction machine equipped with a plurality of hydraulic actuators, such as a hydraulic excavator or a hydraulic crane. The present invention relates to a hydraulic drive device that controls the flow rate of pressure oil supplied to a hydraulic actuator using a flow control valve.

[Conventional technology]

Conventionally, hydraulic drive devices for hydraulic construction machinery equipped with multiple hydraulic actuators, such as hydraulic excavators and hydraulic cranes, have been
At least one hydraulic pump, a plurality of hydraulic actuators each connected to the hydraulic pump via a main circuit and driven by pressure oil discharged from the hydraulic pump, and a plurality of flow control valves connected to respective main circuits in between.

U, S, P, 4,617,854, in such a hydraulic drive system, an auxiliary valve is disposed upstream of the main circuit of each flow control valve, and a first operation part facing the auxiliary valve is provided with an auxiliary valve. The inlet pressure and outlet pressure of the flow control valve are guided, and the discharge pressure of the hydraulic pump and the maximum load pressure of the plurality of hydraulic actuators are guided to the opposing second operation part, and the discharge pressure of the hydraulic pump is made lower than the maximum load pressure. A configuration is described in which a load-sensing type pump regulator is arranged to maintain a predetermined value high. In this configuration, by guiding the inlet pressure and outlet pressure of the flow control valve to the opposing first operating portion of the auxiliary valve, the load pressure compensation of the flow control valve is performed as is well known, and the opposing second operating portion of the auxiliary valve By guiding the discharge pressure of the hydraulic pump controlled by the pump regulator and the maximum load pressure of multiple hydraulic actuators to the operating section of the pump, the command for each hydraulic actuator can be adjusted during combined operation of multiple hydraulic actuators with different load pressures. flowing ff
Even if the total of 1 (required flow rate) exceeds the maximum discharge flow rate of the hydraulic pump, the discharge flow rate is divided according to the mutual command flow rate ratio, ensuring that pressure oil is supplied even to the hydraulic actuator on the high-load pressure side. I am trying to make it possible to flow.

On the other hand, U, S, P, 4,535,809 describes a flow control valve connected to the main circuit between a hydraulic pump and a hydraulic actuator in a hydraulic drive system intended for a single hydraulic actuator rather than multiple hydraulic actuators. is composed of a seat valve type main valve connected to the main circuit, and a pilot valve connected to a pilot circuit between the back pressure chamber and the outlet port of the main valve, and further connected to the pilot circuit. This document describes an arrangement in which an auxiliary valve is arranged, and the inlet pressure and outlet pressure of the pilot valve are guided to the opposing operation parts of the auxiliary valve to perform a pressure compensation function. The patent also includes:
For the operation of a single hydraulic actuator, a modification is also disclosed that incorporates the influence of self-load pressure and modifies the pressure compensation r11 function.

[Problem to be solved by the invention]

However, in U.S.P. 4,617.854, both the flow valve and the auxiliary valve are configured as spool valves, and since both valves are arranged in the main circuit, they are configured as relatively large spool valves. There is. Therefore, since the auxiliary valve is disposed in the main circuit through which a large flow rate flows, there is a problem in that the pressure loss in the auxiliary valve portion is large.

Additionally, in a hydraulic drive system, it is preferable that each hydraulic actuator can be supplied with a flow rate without being affected by its own load pressure or the load pressure of other hydraulic actuators. In a hydraulic drive system, it may be preferable to be influenced by the load pressure of another hydraulic actuator or the self-load pressure depending on the type of work member driven by the hydraulic actuator and the work style.

For example, when a hydraulic excavator rotates and raises the boom at the same time to load earth and sand onto a truck, the rotating body is an inertial body, so the load pressure on the swing motor becomes high at the beginning of the swing, and the swing motor is installed to protect the circuit. The pressure rises above the relief valve pressure.

On the other hand, the boom load pressure is the boom holding pressure, which is lower than the swing load pressure. In this type of work, when the swinging pressure at the beginning of swinging is high, energy waste can be reduced by supplying pressure oil to the boom as much as possible without relief, and
It is possible to automatically adjust the speed of the boom and swing by increasing the rising speed of the boom faster than the swinging speed at first, and gradually increasing the swinging speed once the boom has risen to a certain degree.

In addition, in a single swing operation or a combined operation with other hydraulic actuators, at the beginning of the swing, the swing load pressure rises above the pressure of the relief valve as described above, so as the swing load pressure increases, the swing motor If the amount of pressure oil supplied can be reduced, energy waste can be reduced.

In addition, with hydraulic excavators, work that does not accurately divide the flow rate according to the ratio of the amount of operation of the boom control lever and the arm control lever, regardless of the load pressure, such as slope formation work performed by combined operation of the boom and arm. There are also forms.

Therefore, in construction machinery such as hydraulic excavators, the characteristics of the flow control valve are not uniquely determined to perform a pressure compensation function and/or a flow diversion function, but are determined by the type and work form of the work member driven by the hydraulic actuator. It is desirable to be able to modify it to provide various functions according to the user's needs.

However, in U, S, P, 4,617.854, although the installation of the auxiliary valve achieves the pressure compensation function and flow dividing function as described above, it incorporates the influence of the load pressure of other hydraulic actuators or the self-load pressure. There was no idea of modifying these functions, and it was not possible to meet the above-mentioned demand for modifying the characteristics of the flow control valve according to the type of work member and work form.

On the other hand, in U, S, P, 4,535,809, since it is a hydraulic drive device for a single hydraulic actuator, an auxiliary valve is installed to perform a pressure compensation function regarding the operation of a single hydraulic actuator. Or, it is a technology that only modifies the pressure compensation function by incorporating the influence of the self-load pressure of the single hydraulic actuator, and is unrelated to modifying various functions regarding the combined operation of multiple hydraulic actuators. There was no idea of modifying the pressure compensation function and flow dividing function by incorporating the influence of the load pressure of other hydraulic actuators.

SUMMARY OF THE INVENTION An object of the present invention is to provide a hydraulic drive device that has low throttling loss and that can modify the characteristics of a flow control valve depending on the type of work member and work form of a hydraulic construction machine.

[Means to solve the problem]

In order to achieve the above object, the present invention includes at least one hydraulic pump: at least first and second hydraulic pumps connected to the hydraulic pump through main circuits and driven by pressure oil discharged from the hydraulic pump. two hydraulic actuators; first and second flow control valve means connected to respective main circuits between the hydraulic pump and the first and second hydraulic actuators; controlling the discharge pressure of the hydraulic pump; The first and second flow rate control valve means each include a first valve means that changes the degree of opening according to the amount of operation of the operation means; a second valve means connected in series for controlling the differential pressure between the inlet pressure and the outlet pressure of the valve means;
and for each of the second flow control valve means, the said first valve means is determined based on the inlet pressure and outlet pressure of the first valve means, the discharge pressure of the hydraulic pump, and the maximum load pressure of the first and second hydraulic actuators. In the hydraulic drive device, the first and second flow control valve means each include a valve body that controls communication between an inlet port and an outlet port connected to the main circuit. , a variable throttle that changes the degree of opening in response to the displacement of the valve body, and a back pressure that communicates with the outlet port via the variable throttle and generates a control pressure that biases the valve body in the valve opening direction. a main valve having a chamber; and a pilot circuit connected between an inlet port of the main valve and a back pressure chamber;
The second valve means is connected to the pilot circuit and configured as a pilot valve for controlling the pilot flow flowing through the pilot circuit, and the second valve means is connected to the pilot circuit and is configured to control the differential pressure between the inlet pressure and the outlet pressure of the pilot valve. The control means is configured as an auxiliary valve for controlling the first and second flow rate control valve means such that the differential pressure between the inlet pressure and the outlet pressure of the pilot valve is adjusted to the discharge of the hydraulic pump. The following formula is used for the differential pressure between the pressure and the maximum load pressure of the first and second hydraulic actuators, the differential pressure between the maximum load pressure and the self-load pressure of each hydraulic actuator, and the self-load pressure. The auxiliary valve is controlled so that the relationship expressed is Δpz = a (Ps − Plmax) + β (Plmax
x-pl) +γPl where ΔPz: Differential pressure between the inlet pressure and outlet pressure of the pilot valve Ps: Discharge pressure of the hydraulic pump P 1llaX: Maximum load pressure Pl of the first and second hydraulic actuators: The first Self-load pressures α, β, and γ of the first and second hydraulic actuators: first, second, and third constants;
and third constants α, β, and r are each set to predetermined values.

Before receiving the control pressure of the back pressure chamber, let K be the ratio of the pressure receiving area of the main valve body receiving the load pressure of the associated hydraulic actuator via the outlet port to the pressure receiving area of the main valve body. The first constant a is preferably α≦K
There is a relationship between In this case, the second and third constants β and r can each be set to zero.

Further, the first constant α is set to an arbitrary positive value corresponding to the operating amount of the operating means and the proportional gain of the flow rate passing through the main valve, and the second constant β is The third constant γ is set to an arbitrary value based on the coordination of the combined operation of the actuator and other hydraulic actuators, and the third constant γ is set to an arbitrary value based on the operating characteristics of the related hydraulic actuator.

The control means includes a plurality of hydraulic operation chambers provided in each of the auxiliary valves of the first and second flow rate control valve means, and a discharge pressure of the hydraulic pump and the maximum load in the plurality of hydraulic operation chambers. and a conduit means for directly or indirectly introducing the pressure, the inlet pressure and the outlet pressure of the pilot valve, and in this case, the pressure receiving area of each of the plurality of hydraulic operation chambers is set to the first, The second and third constants α, β, and γ are set to the predetermined values.

Examples of the configuration of the hydraulic control means will be listed below.

(1) The auxiliary valve is disposed between the inlet port of the main valve and the pilot valve, and the plurality of hydraulic operation chambers have first and second hydraulic pressures that bias the auxiliary valve in the opening direction. It has an operation chamber and third and fourth hydraulic operation chambers that bias the auxiliary valve in the valve closing direction, and the conduit means transmits the discharge pressure of the hydraulic pump to the first hydraulic operation chamber. a first conduit for guiding;
a second pipe line that leads the outlet pressure of the pilot valve to the second hydraulic operation chamber; a third pipe line that leads the maximum load pressure to the third hydraulic operation chamber; and an inlet pressure of the pilot valve. and a fourth pipe line that leads the hydraulic pressure to the fourth hydraulic operation chamber.

(2) The auxiliary valve is arranged between the back pressure chamber of the main valve and the pilot valve, and the plurality of hydraulic operation chambers include a first hydraulic operation chamber that biases the auxiliary valve in the opening direction. and second, third, and fourth hydraulic operation chambers that bias the auxiliary valve in the valve closing direction, and the conduit means transfers the outlet pressure of the pilot valve to the first hydraulic operation chamber. a first conduit leading to the pilot valve, a second conduit leading the inlet pressure of the pilot valve to the second hydraulic operating chamber, and a second conduit leading the load pressure of the associated hydraulic actuator to the third hydraulic operating chamber. and a fourth pipe line that guides the maximum load pressure to the fourth hydraulic operation chamber.

(3) The auxiliary valve is disposed between the pilot valve and the back pressure chamber of the main valve, and the plurality of hydraulic operation chambers include first and second hydraulic chambers that bias the auxiliary valve in the opening direction. It has a hydraulic operation chamber and third and fourth hydraulic operation chambers for biasing the auxiliary valve in the valve closing direction, and the conduit means transfers the load pressure of the related hydraulic actuator to the first hydraulic operation chamber. a first pipe line leading to the hydraulic operation chamber; a second pipe line guiding the outlet pressure of the pilot valve to the second hydraulic operation chamber; and a second pipe line guiding the maximum load pressure to the third hydraulic operation chamber. It has a third pipe line and a fourth pipe line that guides the control pressure of the back pressure chamber to the fourth hydraulic operation chamber.

(4) The auxiliary valve is disposed between the inlet port of the main valve and the pilot valve, and the plurality of hydraulic operation chambers have first and second hydraulic pressures that bias the auxiliary valve in the opening direction. It has an operation chamber and third and fourth hydraulic operation chambers that bias the auxiliary valve in the valve closing direction, and the conduit means transfers the load pressure of the related hydraulic actuator to the first hydraulic operation chamber. a first pipe line that leads the hydraulic pump to the second pipe line;
a second pipe line leading to the hydraulic operation chamber; a third pipe line leading the maximum load pressure to the third hydraulic operation chamber; and a third pipe line leading the inlet pressure of the pilot valve to the fourth hydraulic operation chamber. and a fourth conduit.

(5) The auxiliary valve is arranged between the pilot valve and the back pressure chamber of the main valve, and the plurality of hydraulic operation chambers include a first hydraulic operation chamber that biases the auxiliary valve in the opening direction. and second and third hydraulic operation chambers that bias the auxiliary valve in the valve closing direction, and the conduit means has a first hydraulic operation chamber that guides the outlet pressure of the pilot valve to the first hydraulic operation chamber. a second pipe line that leads the discharge pressure of the hydraulic bong 1 to the second hydraulic operation chamber, and a third pipe line that leads the maximum load pressure to the third hydraulic operation chamber. have.

The pump control means is preferably a load sensing type pump regulator that maintains the discharge pressure of the hydraulic pump at a predetermined value higher than the maximum load pressure of the first and second hydraulic actuators.

In order to achieve the above object, the present invention also provides a rotating body, each driven by a plurality of hydraulic actuators.
A hydraulic excavator is provided in which the above hydraulic drive device is applied to a hydraulic excavator having a plurality of working members including a boom, an arm, and a bucket.

In the above hydraulic excavator, the control means preferably sets the second constant β to a relatively large positive value for the flow rate control valve means related to the bottom side of the boom hydraulic actuator.

Preferably, the control means sets the second constant β to a relatively small positive value for the flow control valve means related to the bottom side of the arm hydraulic actuator.

Preferably, the control means sets the second constant β to a relatively small negative value for the flow rate control valve means related to the bottom side of the bucket hydraulic actuator.

Preferably, the control means includes the third flow control valve means related to the rotating body hydraulic actuator.
The constant γ of is set to a relatively small negative value.

Preferably, the control means sets the third constant γ to a relatively small positive value for the flow rate control valve means related to the bottom side of the bucket hydraulic actuator.

Preferably, the control means sets the second and third constants β and r to zero, respectively, for flow control valves related to rod sides of the boom and arm hydraulic actuators.

[Effect]

As a result of various studies by the present inventors on the relationship between the auxiliary valve arranged in the pilot circuit and the differential pressure across the pilot valve, the differential pressure Δpz across the pilot valve controlled by the auxiliary valve is -a, as follows: We found that it is expressed by the above formula, which is reproduced in .

ΔPz = a (Ps − Prmax) + β (Plla
X−Pl ) +yPl The meaning of the above formula is as follows. In this equation, the first term on the right side is Ps − PllaX
Since it is common to all flow rate control valves, it governs the diversion function in complex operations.The second term, Plmax-Pl, changes depending on the maximum load pressure of other actuators, so it governs the harmonization function in complex operations. The term γP1 changes depending on the self-load pressure, so it governs the self-pressure compensation function. These 31! The functions are constants α and β.

The necessity and extent of each are determined depending on the value of γ.

Here, the first term, the diversion function, is the basic function of the composite operation. Therefore, the constant α is set to a predetermined positive value regardless of the associated work member. On the other hand, the harmonic a function in the second term and the self-pressure compensation function in the third term are functions that are added depending on the type of related work member and work form. Therefore, the constants β and r are each set to predetermined values including zero. By setting α, β, and r in this way, it is possible to provide a flow dividing function, or a harmonization function and/or a self-pressure compensation function based on the flow dividing function, and it is possible to provide a flow dividing function, a harmonization function based on the flow dividing function, and/or a self-pressure compensation function. The characteristics of the flow control valve can be modified accordingly.

Furthermore, in the configuration of the present invention described above, since the auxiliary valve is installed in the pilot circuit rather than the main circuit, throttling loss in the auxiliary valve portion can be reduced even if a large flow rate flows through the main circuit.

If the first constant α is in the relationship α≦K with respect to the ratio K of the pressure receiving area receiving the load pressure of the hydraulic actuator to the pressure receiving area receiving the control pressure of the back pressure chamber, then The differential pressure obtained by α (Ps-Plaax) in term 1 is within the range of the maximum longitudinal differential pressure that can be obtained by the pilot valve on the high load pressure side, and the above equation 1 is satisfied for both the first and second flow control valves. The differential pressure between the two terminals becomes substantially the same, and the flow rate can be accurately divided in proportion to the operation amount (pilot valve opening degree) of the operating means in the above-mentioned flow division function.

Further, the first constant α has the meaning of a proportional gain between the manipulated variable (opening degree of the pilot valve) of the operating means and the pilot flow rate, and therefore a proportional gain between the manipulated variable and the main flow rate passing through the main valve, Therefore, the first constant α is set to an arbitrary positive value corresponding to the proportional gain. Here, when a= is set, the maximum proportional gain can be provided while obtaining a flow dividing function that proportionally distributes the flow rate according to the manipulated variable.

As is clear from the above description, the second constant β is set to an arbitrary value, taking into consideration the harmony of the combined operation of the related hydraulic actuator and other hydraulic actuators. Here, especially if it is preferable not to be influenced by the load pressure of other hydraulic actuators, β is set to zero.

As is clear from the above description, the third constant γ takes into account the operating characteristics of the related hydraulic actuator,
Set to any value. This is also set to zero, especially if it is preferred not to be influenced by self-load pressure.

When the pump control means is a load sensing type pump regulator, the differential pressure Ps-P1max between the discharge pressure of the hydraulic pump and the maximum load pressure of the plurality of hydraulic actuators, which is related to the first term on the right side of the above equation, is constant. becomes. Therefore, the pressure difference between the inlet pressure and the outlet pressure of the pilot valve can be controlled to be constant, and the pressure compensation function is achieved to maintain the flow rate constant regardless of the change in the pressure difference between the inlet port and the outlet port of the main valve.

In the present invention, in which the above-mentioned hydraulic drive device is applied to a hydraulic excavator, the characteristics of the flow control valves related to at least two working members, the revolving body, the boom, the arm, and the bucket, are set and modified according to the type of the working member and the working form. The above-mentioned flow dividing function or a harmonizing function and/or self-pressure compensation function can be added based on the flow dividing function.

When the second constant β is set to a relatively large positive value for the flow rate control valve means on the bottom side of the boom hydraulic actuator, the low load side during the initial acceleration of the swing and boom combined operation The flow rate control valve on the bottom side of the hydraulic actuator for the boom flows in proportion to the increase in the differential pressure between the maximum load pressure (swing pressure) and the self-load pressure (boom pressure), increasing the rising speed of the boom. be able to. As a result, even if you operate the swing and boom raise control levers at the same time to their full stroke, the boom's rising speed will initially rise faster than the swing speed, and once the boom has risen to a certain extent, the swing speed will gradually increase and A complex operation is automatically performed in which the turning speed becomes approximately constant when the robot reaches its maximum speed.

Further, in the case where the second constant β of the flow rate control valve means on the bottom side of the arm hydraulic actuator is set to a relatively small positive value, when excavation is performed by a complex operation using the arm, the arm is When the arm hydraulic actuator is on the low pressure side, the flow control valve opens in response to an increase in the differential pressure between the maximum load pressure (other hydraulic actuator pressure) and the self-load pressure (arm pressure). The flow rate is opened to reduce the degree of flow restriction. As a result, deterioration of fuel efficiency and heat balance is prevented.

When the second constant β of the flow rate control valve on the bottom side of the bucket hydraulic actuator is set to a relatively small negative value, the bucket will be subject to the excavation load during trench digging work by complex operations using the bucket. freed from
At the moment it appears on the ground, the increase in the differential pressure between the maximum load pressure (fl!! hydraulic actuator pressure) and self-load pressure (bucket pressure) reduces the flow rate passing through the relevant flow control valve, reducing the shock. .

Furthermore, in the case where the third constant γ of the flow rate control valve related to the hydraulic actuator for the rotating body is set to a relatively small negative value, when the rotating body is accelerated, the rotating body It is possible to reduce the flow rate passing through the flow control valve related to the flow rate control valve, reduce the flow rate flowing out from the relief valve, and reduce wasteful energy consumption.

When the third constant γ of the flow rate control valve on the bottom side of the bucket hydraulic actuator is set to a relatively small positive value, the bucket pressure (self-load pressure) will decrease during excavation work using the bucket. The flow rate passing through the flow control valve is increased in accordance with the increase, and a powerful digging operation feeling can be obtained.

When the second and third constants β and r are set to zero for the flow rate control valves related to the rod side of the boom and arm hydraulic actuators, slope formation of a slope using the boom and arm is possible. During work, completely eliminates the influence of load pressure of other hydraulic actuators and self-load pressure,
The flow rate can be accurately distributed in proportion to the amount of operation of the boom operation lever and the arm operation lever, and accurate slope formation can be performed.

(hereinafter referred to as the previous post) 〔Example〕 Embodiments of the present invention will be described below with reference to the drawings.

In FIG. 1, a hydraulic drive device according to an embodiment of the present invention is connected to, for example, a swash plate type variable displacement hydraulic pump 1 and a hydraulic pump 1 via a main circuit 2.3. A plurality of, for example, two hydraulic actuators 6.7 driven by pressure oil discharged from the hydraulic pump 1 are connected to each main circuit 2.3 between the hydraulic pump 1 and the hydraulic actuator 6.7, and the hydraulic The hydraulic pump 1 has a directional switching valve 8:9 that controls the drive of the actuator 6.7, and the hydraulic pump 1 has a discharge pressure of the hydraulic pump 1 higher than the maximum load pressure of the plurality of hydraulic actuators 6.7 by a predetermined value. A load-sensing type pump regulator 10 is provided.

The directional control valve 8 has four flow control valves 11.12°13.1
It consists of 4. The first flow control valve 11 is connected to a meter-in (inlet side) circuit 15 when the hydraulic actuator 6 operates to extend, and the second flow control valve 1
2 is connected to the meter-in circuit 16 when the hydraulic actuator 6 is operated to contract, and the third flow control valve 13 is connected to the hydraulic actuator 6 and the second flow control valve 12.
When the hydraulic actuator 6 is operated to extend between
The flow control valve 14 is connected to a meter-out circuit 18 between the hydraulic actuator 6 and the first flow control valve 11 when the hydraulic actuator 6 operates to contract. Between the first flow control valve 11 and the fourth flow control valve 14, a hydraulic actuator 6 is connected to the first flow control valve 1.
A check valve 19 is connected to prevent backflow of pressure oil to the second flow control valve 12 and the third flow control valve 13.
A check valve 20 is connected between the hydraulic actuator 6 and the second flow control valve 12 to prevent pressure oil from flowing back.

The first to fourth flow control valves 11 to 14 each have a main valve 21
．． 22.23.24, a pilot circuit 25.26.27.28 for controlling the main valve, and a pilot valve 29.30°31.32 connected to the pilot circuit.
and the second flow control valve 11.12 further includes a pressure compensation valve 33.1 connected in series with the pilot valve in the pilot circuit.
It is equipped with 34.

As shown in FIG. 2, the main valve 21 of the first flow control valve 11 is connected to an inlet port 31 connected to a pipe line on the hydraulic pump 1 side of the meter-in circuit 15 and a pipe line on the hydraulic actuator 6 side. valve housing 3 having an outlet port 32;
3 and a valve body 35 disposed within the valve housing 33 and having a control orifice 34, the opening degree of the control orifice 34 is adjusted according to the displacement of the valve body 35, and the inlet port 31
and the outlet port 32. A back pressure chamber 36 is formed on the side of the valve body 35 opposite to the control orifice and generates a control pressure Pc that urges the valve body 35 in the valve opening direction. Further, a chamber 38 communicating with the back pressure chamber 36 and accommodating a control piston 37 is formed at the end of the valve body 35 facing the back pressure chamber 36.
Chamber 38 also communicates with outlet port 32 via passageway 39.

One end of the control piston 37 is housed in a pressure chamber 40 within the valve body 35, and the other end is supported in contact with a plug body 41 that closes the back pressure chamber 36. Pressure chamber 40 communicates with inlet port 31 via passage 42 and holds control piston 37 in abutment position with plug body 41 .

The control piston 37 also has a tapered part 43 in its middle part, and this tapered part 43 cooperates with the wall of the opening of the chamber 38 to form a variable throttle 44 that changes the degree of opening according to the displacement of the valve body 35.
is formed. The back pressure chamber 36 thus communicates with the outlet port 32 via the variable throttle 44.

In the valve body 35, the annular end face on the upper side in the drawing facing the inlet port 31 receives the discharge pressure Ps of the hydraulic pump 1.
The bottom wall surface facing the outlet port 32 receives the load pressure P1 of the hydraulic actuator 6, and the valve body 35 is similarly urged in the valve opening direction. The top end face facing the back pressure chamber 36 receives the control pressure pc and defines a pressure receiving area AC that urges the valve body 35 downward in the figure, that is, in the valve opening direction.

These pressure receiving areas have a relationship of AC=AS+AI.

The check valve 33 mentioned above is arranged below the valve body 35, and the valve housing 33 is connected to the outlet port 45 of the check valve 19.
It further has.

Pilot circuit 25 is connected between inlet port 31 of main valve 21 and back pressure chamber 36 .

The pilot valve 29 has an inlet port 40 and an outlet port 4
a poppet-shaped valve body 42 that controls communication between the valve body 42 and the valve body 42;
The hydraulic operation chamber 44 has a spring 43 that biases the valve body 42 in the valve closing direction, and a hydraulic operation chamber 44 that biases the valve body 42 in the valve opening direction. It is connected to a pilot circuit to be generated, and the valve body 42 is opened to an opening degree corresponding to the amount of operation.

The pressure compensation valve 33 has an inlet port 50 and an outlet port 51
a seat-type valve body 52 that controls communication between the first and second hydraulic operation chambers 53 that bias the valve body 52 in the valve-opening direction;
54, and third and fourth hydraulic chambers located opposite to the first and second hydraulic operation chambers 53, 54 and urging the valve body 52 in the valve closing direction.
It consists of hydraulic operation chambers 55 and 56. The first hydraulic operating chamber 53 is formed by the inlet portion 57 of the pressure compensating valve 33 communicating with the inlet port 50, and the second hydraulic operating chamber 5
4 is connected to the pilot circuit 25 via the pilot line 58.
The third hydraulic operation chamber 55 is connected to the maximum load pressure line 61 (described later) via the pilot line 59, and the fourth hydraulic operation chamber 56 is connected to the pilot line 60 via the pilot line 59. The pilot valve 29 of the pilot circuit 25 is connected to the artificial side via the pilot valve 29 . As a result, the discharge pressure ps of the hydraulic pump 1 is introduced into the first hydraulic operation chamber 53, the outlet pressure of the pilot valve 29 is introduced into the second hydraulic operation chamber 54, and this outlet pressure is applied to the back pressure chamber 36. Control pressure P
c, and the third hydraulic operation chamber 55 has a load pressure on the high pressure side of the hydraulic actuator 6.7, that is, a maximum load pressure Plm.
ax is introduced, and the inlet pressure PZ of the pilot valve 29 is introduced into the fourth hydraulic operation chamber 56. The end surface of the valve body 52 facing the first hydraulic operation chamber 53 defines a pressure receiving area aS that receives the discharge pressure ps of the hydraulic pump 1, and the annular end surface facing the second hydraulic operation chamber 54 defines the pressure receiving area aS that receives the discharge pressure ps of the hydraulic pump 1. The end face facing the third hydraulic operation chamber 55 defines a pressure receiving area aC that receives the outlet pressure Pc, and the end face facing the third hydraulic operation chamber 55 defines a pressure receiving area all that receives the maximum load pressure pHaX of the hydraulic actuator 6.7.
The annular end face facing the fourth hydraulic operation chamber 56 defines a pressure receiving area aZ that receives the inlet pressure pz of the pilot valve 29.

In the above configuration, the portions that define the pressure receiving areas AC and AS of the first to fourth hydraulic operation chambers 53 to 56 and the pilot pipes 57 to 60 of the pressure compensating valve 33 and the valve body 35 of the main valve 21 are The differential pressure ΔPz (=Pz - Pc) between the inlet pressure and the outlet pressure of the valve 29 is the differential pressure Ps - PlnaX between the discharge pressure of the hydraulic pump 1 and the maximum load pressure of the two hydraulic actuators 6.7, and the maximum load pressure, respectively. The differential pressure between the self-load pressure of the hydraulic actuator and Plnax-
Pl and the self-load pressure P1, the control means is configured to control the pressure compensation valve 33 so that the relationship expressed by the following formula is established; Δpz = a (Ps - Plmax) + β (PlII
ax−Pl ) +γPl (1) Here α, β, γ
are first, second, and third constants, respectively, and are set to predetermined values. In this embodiment, the predetermined values of the first, second, and third constants a, β, and γ are set in the first to fourth hydraulic operation chambers 53 to 56 of the pressure compensation valve 33.
This is done by selecting the pressure receiving areas as, aC, all, and aZ. That is, the first to fourth hydraulic operation chambers 5
The pressure receiving areas as, aC, all, and aZ of 3 to 56 are as follows:
The first, second, and third constants α, β, and γ are set to obtain predetermined values. Moreover, the pressure receiving area aS of the first to fourth hydraulic operation chambers 43 to 46.

ac, atl, and az are the main valve 21 and the pilot valve 2
The valve body 52 is set to be held in the open position when the valve 9 is closed.

Main valve 21 of the first flow control valve 11 configured in this way
In the combination of the pilot valve 29 and the pilot valve 29, when the pilot valve 29 is opened by operating a control lever (not shown), at that moment the pressure oil in the hydraulic cylinder 11 flows into the pilot circuit 2.
5 to the back pressure chamber 36 of the main valve 21. and,
The pressure inside the back pressure chamber 36, that is, the 1IIIJ control pressure, increases in accordance with the opening degree of the pilot valve 29, and the pressure at the outlet port 32, which communicates with the back pressure chamber 36 via the chamber 38 and the passage 839, also increases. Valve 19 is opened. This allows the passage from the pilot circuit 25 through the back pressure chamber 36 to the outlet port 32.
A pilot flow is formed, and the control pressure in the back pressure chamber 36 is also increased in accordance with the pilot flow rate (the opening degree of the pilot valve 29) by the action of the variable throttle 44. Pilot valve 2
When the opening of the valve 9 is larger than the opening of the variable throttle 44, the control pressure Pc is correspondingly increased, and the valve body 35 begins to move toward the outlet port 32. That is, the main valve 21 is opened. When the valve body 35 moves in the valve opening direction in this way, the opening degree of the variable throttle 44 formed around the control piston 37 held under pressure by the plug body 41 increases, and the throttling action of the variable throttle 44 decreases. . Therefore, an increase in the control pressure in the back pressure chamber 36 is suppressed.

As a result, the valve body 35 is statically fixed at the point where the opening degree of the pilot valve 29 and the opening degree of the variable throttle 44 match.

In this way, the main valve body 35 opens to an opening proportional to the pilot flow rate due to the action of the variable throttle 44 and the back pressure chamber 36.
A flow rate corresponding to the operating amount of the operating lever (the opening degree of the pilot valve 31) flows out from the inlet port 31 to the outlet port 32 through the control orifice 34 of the main valve 21.

In such control of the main valve 21 by the pilot valve 29, since the pilot circuit 25 is further provided with a pressure compensation valve 33, the flow rate of the main valve 21 is further controlled by the pressure compensation valve 33. Since the control function of the pressure compensating valve 33 is an essential part of this embodiment, its operating principle will be explained later.

Main valve 221 of second flow control valve 12, pilot circuit 2
6. The pilot valve 30 and the pressure compensation valve 34 are the main valve 21 and the pilot circuit 25 of the first flow control valve 11.
, the pilot valve 29 and the pressure compensation valve 33, respectively.

As shown in FIG. 3, the main valve 23 of the third flow control valve 13 is connected to an inlet port 70 connected to a pipe line on the hydraulic actuator 6 side of the meter-out circuit 17 and a pipe line on the tank side. It has a valve housing 72 having an outlet port 71, and a valve body 74 disposed within the valve housing 72 and engaged with a valve seat 73. Controls communication of port 71. A plurality of slits 75 are formed in the axial direction on the outer periphery of the valve body 74, and the slits 75 are variable throttles 76 that cooperate with the inner wall of the valve housing 72 to change the opening degree according to the displacement of the valve body 74.
is formed. Behind the valve body 74 in the valve housing 72 is connected to the inlet port 70 via a variable throttle 76;
A back pressure chamber 77 is formed that generates a control pressure P3C.

The annular end surface on the upper side in the figure facing the inlet port 70 of the valve body 74 receives the load pressure P1 of the hydraulic actuator 6.
An annular pressure-receiving area A3 that urges 4 in the upward valve opening direction in the figure.
1, the bottom wall facing the outlet port 71 receives the tank pressure P and defines a pressure receiving area A3r that biases the valve body 74 in the valve opening direction, and the top end facing the back pressure chamber 77 receives the tank pressure P'. The area receiving the pressure P3c defines a pressure receiving area A3C that urges the valve body 74 in the downward valve closing direction in the figure.These pressure receiving areas are as follows:
The relationship is A3c=A31+A3r.

Pilot circuit 27 is connected between back pressure chamber 77 and outlet port 71 of main valve 23 .

The pilot valve 31 is configured similarly to the pilot valve 29 of the first flow jt control valve 11.

Main valve 23 of the third flow control valve 13 configured in this way
The combination of and pilot valve 31 is U.

S, P, 4,535.809, when the pilot valve 31 is opened by operating an operating lever (not shown), a pilot flow is formed in the pilot circuit 27 according to the opening degree of the pilot valve 31, and the flow rate is variable. Due to the action of the throttle 76 and the back pressure chamber 77, the main valve body 74 opens to an opening proportional to the pilot flow rate, and a flow rate corresponding to the operation amount of the operating lever (opening degree of the pilot valve 31) passes through the main valve 23 to the inlet port. 70 to exit port 71 .

Main valve 24 of the fourth flow control valve 14, pilot circuit 28
, the pilot valve 32 is configured similarly to the above-described main valve 23, pilot circuit 27, and pilot valve 31 of the third flow rate control valve 13.

The directional control valve 9 is constructed similarly to the directional control valve 8.

In the following, members of the directional switching valve 9 that are equivalent to components of the directional switching valve 8 are indicated by adding a subscript to the number given to the component of the directional switching valve 8.

The outlet ports 32 of the first and second flow control valves 11, 12 of the directional valve 8 are connected to the aforementioned pipe 61 via check valves 80, 81, respectively, and the first and second flow control valves 11, 12 of the directional valve 9 are The outlet ports of the flow control valves 11A, 12A are likewise connected to the pipe N61A via the check valves 80A, 818, and the pipes 61.6LA are interconnected via the pipe 82, which is connected to the conduit N61A. 83 to the tank. As a result, during the combined operation of the hydraulic actuator 6.7, the load pressure on the high pressure side of both, that is, the maximum load pressure is selected by the check valves 80.81, 80A, 81^, and guided to the pipes 61.61A, 82.

The lines 61, 61^, 82 thus constitute a maximum load pressure circuit.

The pump regulator 10 includes a hydraulic cylinder-type swash plate tilting device 90 and a control valve 91, and the swash plate tilting device 90 has a rod-side cylinder chamber through which the discharge pressure of the hydraulic pump 1 is introduced via a pipe 92. , the head side cylinder chamber is the control valve 91
It is connected to the tank and rod side cylinder chamber through. The discharge pressure led to the rod side cylinder chamber of the swash plate tilting device is reduced depending on the position of the control valve 91, and the difference in area between the rod side cylinder chamber and the head side cylinder chamber drives the piston, and the control valve 91 Hydraulic pump 1 corresponding to the position of
Controls the discharge amount.

The control valve 91 has opposing hydraulic operating parts 93 and 94 and a spring 95.
The hydraulic operating portion 93 is connected to the discharge line of the hydraulic pump 1 via a pilot line 96, and the hydraulic operating unit 94 is connected to the maximum load pressure circuit 82 via a pilot line 97. As a result, the discharge pressure of the hydraulic pump 1, the maximum load pressure, and the setting force of the spring 65 act on the control valve 91 in opposition to each other, and the position of the control valve 91 is adjusted according to the change in the maximum load pressure, and the swash plate The tilt device 141 is controlled to maintain the discharge pressure of the hydraulic pump 1 at a pressure higher than the maximum load pressure by a pressure corresponding to the strength of 7265.

Next, the principle of operation of the pressure compensation valves 33, 34, 33A, 34A will be explained. In addition, in the following, all pressures are 11 fl
Items common to the compensation valves 33, 34, 33A, and 34A will be explained using the pressure compensation valve 33 as a representative. Valve body 5 of pressure compensation valve 33
The pressure balance equation 2 is expressed by the following equation.

as Ps +ac Pc =al Plnax+aZ
pz Also, the pressure balance equation of the valve body 35 of the main valve 21 is expressed by the following equation.

Ac Pc = As Ps + AI Pl From the above two equations, derive the equation regarding the differential pressure across the pilot valve 29, AC=
When transformed using AS + AI, aZ(PZ-Pc) "(ac"as-aZ-all) P l Therefore, if we set, then Pz - Pc = a (Ps - Plmax) + β(P
It can be expressed as llaX−Pl ) +γPl. Here P
z −pc =ΔPz.

Therefore, the same equation as the above-mentioned equation (1) can be obtained. This formula is reproduced below.

Δpz = a (Ps - PlflaX) + β (
PlaaX−Pl ) +γPl (1) Then, the above equation (1) will be considered. In equation (1), the left side Δ
pz is the inlet pressure pz and outlet pressure Pc of the pilot valve 29
is the differential pressure of The first term on the right side is the discharge pressure P of the hydraulic pump 1
This is a term related to the differential pressure between s and the maximum load pressure Plmax, and α is a proportionality constant. The second term is the maximum load pressure P 1l
This is a term related to the differential pressure between aX and the load pressure of the hydraulic actuator 6 or 7, that is, the self-load pressure P1, and β is a proportionality constant. The third term is determined by the self-load pressure P1, and the proportionality constant is γ. Here, since the pressure balance equation of the valve body 35 of the main valve 21 is Ac Pc = As Ps + AI Pl, the load pressure P1 of the hydraulic actuator 6 is equal to the discharge pressure ps of the hydraulic pump 1 and the outlet pressure pc of the pilot valve 29. It can be expressed as Therefore, equation (1) is
The pressure compensation valve 33 has four pressures Pg, Pi, X, and Pl.
, PZ: The differential pressure Δpz between the inlet pressure pz and the outlet pressure pc of the pilot valve 29 can be controlled based on the differential pressure Δpz between the discharge pressure Ps of the hydraulic pump 1 and the maximum load pressure Plsax. P 1iax, differential pressure Plm between maximum load pressure P 1iax and self-load pressure P1
The three elements of ax-Pl and self-load pressure P1 can be controlled in proportion to each other: and the three elements Ps
-Pillax, Plmax - Pl This means that the degree of proportionality to Pl can be arbitrarily set by selecting the values of the proportionality constants α, β, and γ.

Here, the pressure compensating valve 33 is set to the differential pressure Δ before and after the pilot valve 29.
Controlling pz means controlling the pilot flow rate passing through the pilot valve 29, and as a result, controlling the main flow rate passing through the main valve 21 from the function of the combination of the main valve 21 and the pilot valve 29 described above. be.

Furthermore, in the first term on the right side, the differential pressure Pg -pHIaX is constant as long as the pump regulator 10 is functioning effectively in this embodiment, which is equipped with a load sensing type pump regulator 10 as the pump control means. , and all pressure compensation valves 33 (34), 33A
This is common to (34A).

Therefore, in the first term on the right side, controlling the differential pressure Δpz across the pilot valve 29 so that it is proportional to the differential pressure Ps - Plmax means that in the operating state where the pump regulator 10 is functioning effectively, It is to control the pressure Δpz to be constant, and if the opening degree of the pilot valve 29 is constant, the main flow rate passing through the main valve 21 can be controlled to be constant even if there is a fluctuation in the main valve inlet pressure Ps or outlet pressure Pl. It is.

That is, it performs a pressure compensation function.

In addition, during combined operation of the hydraulic actuator 6.7, the total consumed flow rate becomes larger than the maximum discharge flow rate of the hydraulic pump 1, and the discharge pressure of the hydraulic pump 1 decreases, such as when the pump regulator 10 does not function effectively. In this state, the differential pressure ΔPz becomes smaller as the differential pressure Pa − pHIaX decreases, and the main flow rate passing through the main valve 21 also decreases, but the differential pressure Ps − P I1aX decreases as the differential pressure Pa − pHIaX decreases.
Since it is common to (34) and 33A (34A),
The main flow through the main valves 21 (22), 21A (22A) decreases at the same rate. Therefore, the main valves 21 (22), 2
1A (22A) is proportionally distributed according to the operation amount of the operating lever (opening degree of pilot valve 29 (30), 29A (30A)), and the discharge flow rate of the hydraulic pump 1 is distributed to the high pressure side hydraulic actuator. will also be reliably supplied. That is, a shunt function is obtained.

In addition, in the second term on the right side, controlling the differential pressure ΔPz across the pilot valve 29 so that it is proportional to the differential pressure Plnax-Pl means that when the load pressure pHaX of another hydraulic actuator is greater than the self-load pressure P1. In addition, the differential pressure ΔPz across the pilot valve 29 is changed depending on the maximum load pressure Pl1aX of other hydraulic actuators.If the opening degree of the pilot valve 29 is constant, the pressure difference ΔPz is changed depending on the maximum load pressure pHaX. This is to change the main flow rate passing through the main valve 21.

Generally, it is preferable that the flow rate control of the flow control valve is not influenced by other hydraulic actuators, but in hydraulic construction machinery such as hydraulic excavators, the flow rate may be changed by the influence of the load pressure of other hydraulic actuators. There are some types of work that are preferable. In such a case, the second term on the right side performs a harmonizing function of changing the flow rate in harmony with other hydraulic actuators.

Furthermore, in the third term on the right side, controlling the differential pressure ΔPz across the pilot valve 29 so that it is proportional to the self-load pressure P1 means that the differential pressure ΔPz across the pilot valve 29 is controlled in accordance with changes in the self-load pressure P1. If the opening degree of the pilot valve 29 is kept constant, the main flow rate passing through the main valve 21 is changed depending on the self-load pressure P1. This provides a self-pressure compensation function that changes the flow rate in response to changes in self-load pressure.

As described above, in the above equation (1), the first term on the right side governs pressure compensation and shunt n function, the second term governs the harmonization function with other actuators, and the third term governs self-pressure compensation ff1l!
Conducts Noh. These 381 functions can be arbitrarily set by selecting the proportionality constants α, β, and r depending on the necessity and degree thereof.

By the way, among these three functions, the pressure compensation and flow dividing functions related to item 1 are basic functions in hydraulic construction machines such as hydraulic excavators, and are preferably always present regardless of the type of hydraulic actuator and the working form. , therefore, the proportionality constant α is set to an arbitrary positive value. Here, the differential pressure Δpz across the pilot valve 29 defines the pilot flow rate with respect to the opening degree of the pilot valve 29 determined by the operating amount of the operating lever, so the differential pressure Pl11 in the first term
The proportionality constant α applied to ax-Pl means the proportional gain of the pilot flow rate with respect to the operation amount (pilot valve opening degree) of the operating lever of the pilot valve 29, and therefore the proportional gain of the main flow rate passing through the main valve 21 with respect to the operation amount. Therefore, the proportionality constant α is determined corresponding to the proportional gain.

Further, if K is the ratio of the pressure receiving area AC of the main valve body 35 receiving the control pressure PC of the back pressure chamber 36 to the pressure receiving area AI of the valve body 35 receiving the load pressure P1 of the hydraulic actuator 6, then The pressure balance equation is Pc = (1-K) Ps -t-KPl. On the other hand, the discharge pressure ps of the hydraulic pump 1 and the inlet pressure pz of the pilot valve 29 have a relationship of ps≧pz, and when the pressure compensation valve 33 is completely open, ps = pz. Therefore, the differential pressure Pz−Pc(Δ
Pz ) satisfies pz-Pc≦Ps -Pc≦K (Ps - Pl ) (2), that is, the maximum differential pressure that the pilot valve 29 can take is K (Ps
-Pl), and in the above equation (1), β=0,
γ = 0, maximum load pressure side (P l1ax = P I
), pz −Pc =a (Ps −P
lnax)≦K (Ps − Plmax) (3
), that is, when the value of α is set such that α>K,
K (Ps-Pl) for the pilot valve on the maximum load pressure side
On the other hand, with the pilot valve on the low pressure side, a pressure difference of α(Ps-Pl)>K(Ps-Pl) can be obtained, so the opening degree of both pilot valves should be the same. However, the differential pressure across the pilot valve is not the same, and the pilot flow rate is different.

Therefore, it becomes impossible to proportionately distribute the flow rate according to the manipulated variable.

However, even if proportional distribution is not possible, pressure oil can be reliably supplied to the hydraulic actuator on the high pressure side.

Therefore, regarding the flow division function of the pressure compensating valve 33, when obtaining a flow division function that distributes the flow rate in proportion to the operation amount (opening degree) of the pilot valve, the proportionality constant α is set to α≦K4. In particular, when α is set, the maximum flow rate can be given for the same pilot valve opening degree, and the most efficient valve structure can be provided.

Also, as mentioned above, if the value of α is set for α>K,
For the pilot valve on the low load pressure side, α(Ps −P Im
A differential pressure of ) or more can no longer be obtained, and the differential pressure across the pilot valve decreases from α(Ps-Plsax) to K(Ps-Pl), and the pilot flow rate also decreases accordingly. As a result, the flow rate supplied to the hydraulic actuator also decreases,
The work member is decelerated, making it difficult to perform work smoothly. On the other hand, if α is set to α≦K, the differential pressure across the pilot valve on the low load pressure side will be K(P
s-Plmax), and even when switching from combined operation to single operation, the differential pressure does not fluctuate and stable work can be performed. Therefore, in this sense as well, it is preferable to set α to α≦K.

As can be seen from the above, in order to accurately proportionately distribute the flow rate according to the operating amounts of the operating levers of a plurality of hydraulic actuators, it is an essential condition to set α≦X.

Furthermore, the degree of necessity for the harmonization function related to item 2 differs depending on the type of work member driven by the hydraulic actuator 6. There are also preferred working members and working configurations.

Therefore, the proportionality constant β is set to any value including zero based on the coordination of the combined operation of the related hydraulic actuator and other hydraulic actuators.

The degree of necessity for the self-pressure compensation function related to item 3 differs depending on the type of work member driven by the hydraulic actuator 6.7, and in some cases, it is preferable for work members not to be affected by self-load pressure at all. There is also. Therefore, the proportionality constant γ is set to any value including zero depending on the type of work member driven by the related hydraulic actuator.

By setting the constants α, β, and r to predetermined values as described above, it is possible to obtain a flow dividing function, or a harmonization function and/or a self-pressure compensation function based on the flow dividing function, which can be used for hydraulic construction machinery work. The characteristics of the flow control valve can be modified depending on the type of component and the type of work.

As described above, the proportionality constants α, β, and γ are the pressure receiving areas as, ac, am, and az in the first to fourth operation chambers 53 to 56 of the pressure compensating valve 33 and the valve body 3 of the main valve 21.
It is expressed by the pressure receiving areas AC and AS of 5. Here, the pressure receiving area AC of the main valve valve 35 is.

As is determined by the conditions on the main valve 21 side. Therefore,
Once the proportionality constants α, β, and γ are determined, the pressure receiving areas as, ac
, am and az are the proportionality constants α and β.

It is set so that γ can be obtained. As a special case, if we configure as +aC=afg +aZ, then γ=
It can be set to 0, and by configuring ac=az and as=am, it can be set to β=0. Also, as=ac==ai==a
7, it is possible to set β=γ=0.

Next, we will discuss the proportionality constants α, β, and γ when the hydraulic drive system of this embodiment is applied to a backhoe type hydraulic excavator.
A specific setting example will be explained.

As shown in FIGS. 4 and 5, the hydraulic excavator includes a pair of running bodies 100, a revolving body 101 rotatably mounted on the traveling body 100, and a plane perpendicular to the revolving body 101. Funtion attachment 1 rotatably mounted inside
02, and the front attachment 102 consists of a boom 103, an arm 104, and a bucket 1'05. The traveling body 100, the revolving body 101, the boom 103, the arm 104, and the bucket 105 each have a traveling motor 106.
(plurality), swing motor 107, boom cylinder 108
, arm cylinder 109, baguette cylinder 110
driven by. Here, the swing motor 107, the boom cylinder 108, the arm cylinder 109, and the bucket cylinder 110 correspond to the hydraulic actuator 6.7 shown in FIG.

In such a hydraulic drive system for a hydraulic excavator, the proportionality constant α related to all the flow rate control valves of the swing motor 107, boom cylinder 108, arm cylinder 109, and bucket cylinder 110 is as shown in FIG. It is set to the same arbitrary positive value that takes into account the proportional gain described above. In the flow control valve related to the swing motor 107, the proportionality constant β is set to β=0 as shown in FIG. 7(A), and the proportionality constant γ is set to a small negative value as shown in FIG. 8(A). is set to In the flow control valve related to the bottom side of the boom cylinder 108, the proportionality constant β is as shown in Fig. 7 (
B) is set to an arbitrary positive value, and the proportionality constant γ
is set to γ=0 as shown in FIG. 8(B). In the flow rate control valve on the bottom side of the arm cylinder 109, the proportionality constant β is set to a small positive value as shown in FIG. 7(C), and the proportionality constant γ is set as shown in FIG. 8(B). γ=0 is set. In the flow control valve related to the bottom side of the bucket cylinder 110, the proportionality constant β is the seventh
The proportionality constant γ is set to a small negative value as shown in FIG. 8(D), and the proportionality constant γ is set to a small positive value as shown in FIG. 8(C). In addition, in the flow rate control valve related to the rod side of the boom cylinder 108, the flow rate control valve related to the rod side of the arm cylinder 109, and the flow rate control valve related to the rod side of the bucket cylinder 110, the proportionality constants β and γ are all It is set to zero as shown in FIG. 7(A) and FIG. 8(B).

Next, the operation of the hydraulic drive device configured as described above will be explained.

First, when neither of the operating levers of the directional control valve 8.9 is operated, the first and second flow control valves 11, 12
．． IIA, 12A pilot valve 29.30.29A,
30A is closed, pilot circuit 25, 26, 25A
, 16A does not have a pilot flow rate, so the main valve 2
1.22. Pressure oil does not flow through each of the variable throttles 44 of 21A and 22A, and the control pressure pc of the back pressure chamber 36 is controlled by the outlet port 32.
pressure (load pressure of hydraulic actuator 6 or 7) Pl
is the same as Furthermore, due to the above-described action of the load sensing type pump regulator 10, the discharge pressure Ps of the hydraulic pump 1 is maintained at a pressure higher than the maximum load pressure P1llaX of the hydraulic actuator 6.7 by a pressure corresponding to the set value of the spring 95. There is. Therefore, each pressure receiving area of the valve body 35 has a relationship of AC=As+AI, and Ps>
Pl, the valve body 35 is urged in the valve closing direction by the discharge pressure ps of the hydraulic pump 1, and the main valves 21, 22, 21A, 2
2A is held in the closed position. Also pressure compensation valve 33.34
．． 33A and 34A are the aforementioned pressure receiving areas aS, ac,
It is held in the open position by the settings of act and az.

Also, the third and fourth flow control valves 13.14.13A, 1
Similarly, in 4A, pilot valve 31.32.31
A, 32A is closed, pilot circuit 27.28.2
7A, 28A, and therefore no pressure oil flows to each variable throttle 76 of the main valves 23, 24, 23A, 24A, and the control pressure P3c of the back pressure chamber 77 is equal to the pressure (hydraulic load pressure of actuator 6 or 7)
It is the same as Pl. Here, the load pressure P1 is greater than the tank pressure P', therefore, each pressure receiving area of the valve body 74 has a relationship of A3C=A31+A3r, and Pl>
Pl, the valve body 71 is urged in the valve closing direction by the control pressure P3c, and the main valves 23, 24, 23A, and 24A are held in the closed position.

Therefore, no pressure oil flows into either the meter-in circuit or the meter-out circuit for the hydraulic actuator 6.7.
The hydraulic actuator 6.7 is held in the rest position.

Next, when the operation lever of the directional control valve 8 is operated alone, the pilot valve 29 of the first flow control valve 11 opens in accordance with the operation amount, a pilot flow is formed in the pilot circuit 25, and a pilot flow is formed in the pilot circuit 25. A pilot flow rate according to the opening degree of the valve 29 flows. As a result, as described above, due to the action of the variable throttle 44 and the back pressure chamber 36, the main valve body 35 opens to an opening degree proportional to the pilot flow rate, and corresponds to the operation amount of the operating lever (the opening degree of the pilot valve 29). The flow rate is the main valve 2.
1 through the inlet port 31 to the outlet port 32.45.

Similarly, in the third flow rate control valve 13, the pilot valve 31 opens according to the amount of operation of the operating lever, a pilot flow is formed in the pilot circuit 27, and the pilot valve 31 opens.
A pilot flow rate according to the opening degree of 1 flows. As a result, as described above, due to the action of the variable throttle 76 and the back pressure chamber 77, the main valve body 74 opens to an opening proportional to the pilot flow rate, and corresponds to the operating amount of the operating lever (the opening of the pilot valve 31). The flow rate flows out through the main valve 23 from the inlet port 70 to the outlet port 71.

Therefore, the meter-in circuit 1 for the hydraulic actuator 6
5 and the meter-out circuit 17 according to the amount of operation of the operating lever, and the hydraulic actuator 6 is driven at a speed corresponding to the amount of operation of the operating lever.

In this way, the first and third flow control valves 11.1
For example, when the pilot valves 29 and 31 of the third pilot valve 29 and 31 are opened by a certain amount and a certain amount of main flow is flowing, the pressure at the outlet port 32 of the first flow control valve 11 increases, and the pressure at the inlet port 31 increases.
When the differential pressure between and outlet port 32 attempts to decrease,
The discharge pressure of the hydraulic pump 1 is increased by the action of the load sensing type pump regulator 10, and the inlet port 3
1 pressure (discharge pressure of hydraulic pump 1) and outlet port 32
(the load pressure of the hydraulic actuator 6 and the maximum load pressure) is maintained constant. Therefore, a constant flow rate continues to flow through the main valve 21 in accordance with the amount of operation of the operating lever.

In addition, in such independent operation of the hydraulic actuator 6, the pressure receiving area aS, ac, am,
If az is set so that the proportionality constant γ related to the self-pressure compensation characteristic in equation (1) above is set to any value other than zero, the change in the load pressure (self-load pressure) of the hydraulic actuator unit The differential pressure Δp across the pilot valve 29 according to
z is controlled and self-load pressure compensation is provided.

For example, in the example of the hydraulic excavator described with reference to FIGS. 4 to 8, the proportionality constant γ of the flow control valve related to the swing motor 107 has a relatively small negative value as shown in FIG. 8(A). set to the value. Therefore, when the rotating body 101 is driven, since the rotating body is an inertial body, the load pressure increases and rises above the pressure of the relief valve provided for circuit protection, resulting in wasted energy, but the proportionality constant γ By setting the value to a negative value, the differential pressure Δpz is controlled to decrease as the swirling load pressure increases, and the flow rate through the flow rate control valve decreases. Therefore, even if the load pressure increases, less amount is discarded as surplus flow than the relief valve, and energy waste can be reduced.

Further, in the flow rate control valve on the bottom side of the bucket cylinder 110, the proportionality constant γ is set to a religiously small positive value as shown in FIG. 8(C). Therefore, as the self-load pressure increases during excavation work, the differential pressure Δpz increases, and the flow rate through the flow control valve increases.

That is, the excavation speed of the bucket increases. This makes it possible to obtain a powerful digging operation feeling and improve work efficiency.

Next, when both operating levers of the directional control valve 8.9 are operated, the following operation is performed. First, the directional control valve 8, for example, the first and third flow control valves 11, 13 and the directional control valve 9.
For example, in the same way as when the hydraulic actuator is independently operated in both the first and third flow control valves 11A and 13A, a pilot flow rate according to the operation amount flows, and the variable throttle 44°76 and the back pressure chamber 36. 77, the amount of operation of the operating lever (pilot valves 29, 31 and 39A).

The flow rate according to the opening degree of main valves 21, 22 and 21
A, flows through 22A. Therefore, the hydraulic actuators 7 are driven at the same time.

In such a combined operation of the hydraulic actuator 6.7, the pressure receiving area aS of the pressure compensation valve 33.33A of the first flow control valve 11, 11A.

By setting aC, all, and aZ so that the proportionality constant α in the first term on the right side of equation (1) takes an arbitrary positive value as shown in FIG. fulfilled.

Therefore, for example, in the hydraulic excavator described with reference to FIGS. 4 to 8, in an operating state in which the load sensing type pump regulator 10 is functioning effectively, each work member is controlled by the operating amount of the operating lever. It is possible to perform stable composite operations by driving at a constant flow rate. In addition, when the total current consumption of the hydraulic actuators 6.7 becomes larger than the maximum discharge flow rate of the hydraulic pump 1, and the pump regulator 10 becomes in an operating state that does not function effectively, only the hydraulic actuator on the low pressure side is activated. Instead of being supplied with pressure oil, pressure oil can be reliably supplied to the high-pressure side hydraulic actuator, and all working members can be reliably driven. In particular, when α≦K is set, even when the combined operation is switched to the individual operation, the flow rate supplied to the hydraulic actuator does not fluctuate, and the work can be continued stably.

Further, when α≦K is set, a flow rate that is accurately proportionally distributed can be supplied to each hydraulic actuator according to the operation amount of the operation lever. As a result, in particular, the pressure receiving areas as, ac, am, a of the pressure compensation valve 33.33A
When the proportionality constants β and γ in equation (1) are set to zero, the movement locus of the work member can be accurately controlled according to the amount of operation of the operating lever. As shown in FIG. 7(A) and FIG. 8(B), β=O1γ=0 in the flow rate control valve related to the rod side of the arm cylinder 108 and the flow rate control valve related to the rod side of the arm cylinder 109. is set to .

This completely eliminates the influence of the load pressure of other hydraulic actuators and self-load pressure when forming a downhill slope using the boom and arm, and reduces the amount of operation of the boom control lever and arm control lever. Accordingly, the flow rate supplied to the boom cylinder 108 and the arm cylinder 109 can be proportionally distributed accurately, thereby making it possible to form an accurate slope surface.

Further, in the configuration of the present invention described above, the auxiliary valve is installed in the pilot circuit instead of the main circuit. Therefore, even if the pressure of the hydraulic circuit is increased, there is little liquid leakage, and even if a large flow rate is passed through the main circuit, there is almost no throttling loss.

Also, the pressure receiving area aS of the pressure compensation valve 33.33A.

If ac, an, and az are set so that the proportionality constant β and/or the proportionality constant γ in the above equation (1) are arbitrary values other than zero, the above pressure compensation and flow dividing functions are used as a base. Maximum load pressure P of other hydraulic actuators
A harmonization function and/or self-load pressure compensation is performed which varies the main flow through the main valve 21 or 21A depending on Imax.

For example, in the example of the hydraulic excavator described with reference to FIGS. 4 to 8, in the flow control valve related to the swing motor 107, the proportionality constant β is β=0 as shown in FIG. 7(A).
In the flow control valve related to the bottom side of the boom cylinder 108, the proportionality constant β is set to an arbitrary positive value as shown in FIG. 7(B). - In terms of shooting, when swinging and boom raising are operated simultaneously, the swinging body 81 is an inertial body, so the load pressure of the swinging motor becomes high at the beginning of the swinging. However, when the swing reaches the maximum speed, the load pressure decreases. On the other hand, the boom load pressure becomes the boom holding pressure, so at the beginning of the swing, the pressure is lower than the swing load pressure. For example, when turning and raising the boom during excavation work with a backhoe excavator, the operator may operate the turning and boom raising control levers at the same time to their full strokes due to the ease of operation. It is preferable to be able to automatically adjust the speed of the boom and swing so that the swing speed rises quickly relative to the swing speed, and once the boom rises to a certain extent, the swing speed gradually increases.The proportionality constant β should be set as described above. Therefore, in the flow control valve on the boom side, when the swing load pressure is high at the beginning of the swing and the differential pressure pHax-Pl is large, the differential pressure Δpz across the pilot valve increases and the flow rate supplied to the boom cylinder increases. , as the differential pressure P1max-Pl decreases, the differential pressure Δpz also gradually decreases. Therefore, the speed of the boom and swing can be adjusted automatically, reducing the burden on the operator.

Further, in the flow rate control valve on the bottom side of the arm cylinder 109, the proportionality constant β is set to a relatively small positive value as shown in FIG. 7(C).

When performing excavation work using a combined operation using an arm, each hydraulic actuator must move, but at this time, more pressurized oil tends to flow to the actuator on the low pressure side. For this reason, when the pressure oil flows through the flow control valve, it is throttled.
Energy loss increases. Therefore, the fuel consumption deteriorates, and the pressure oil balance also deteriorates. As mentioned above, by setting the proportionality constant β within a range that does not disturb the balance of the complex operation, the flow control valve on the arm side can:
As the differential pressure Plnax-Pl increases, the opening degree of the main valve increases, and the degree of oil throttling becomes smaller. This can reduce fuel consumption and deterioration of heat balance.

Furthermore, in the flow rate control valve on the bottom side of the bucket cylinder 110, the proportionality constant β is set to a relatively small negative value as shown in FIG. When excavating a clearing while restricting the bucket's movement using the boom's maximum pressure, the moment the bucket reaches the surface, the load on the bucket suddenly decreases and a shock occurs. By setting the proportionality constant β as described above, the differential pressure Δpz acts as a negative element as the differential pressure Plnax−Pl increases, reducing the pilot flow rate and decelerating the bucket speed.

This reduces the occurrence of shock when the load suddenly decreases, improving work safety and work feeling.

Regarding self-load pressure compensation, substantially the same thing as described for single operation of a hydraulic actuator is performed for each actuator in combined operation.

As described above, in the hydraulic drive system of this embodiment, the pressure receiving area of the pressure compensating valve is appropriately set, and the constants α and β are set as appropriate.

By setting γ to a predetermined value, it is possible to obtain a diversion function, or a harmonization function and/or a self-pressure compensation function based on the diversion function, depending on the type of work member and work form of the hydraulic construction machine. The characteristics of the flow control valve can be modified.

Furthermore, in the hydraulic drive system of this embodiment, the pressure compensation valve as an auxiliary valve is arranged not in the main circuit but in the pilot circuit. Therefore, it is possible to provide a hydraulic circuit suitable for high pressures with little liquid leakage, and even if a large flow rate flows through the main circuit, there is little throttling loss in the auxiliary valve section, and it is economically superior.

In the above embodiments, with reference to FIGS. 6 to 8, specific flow control valves related to the revolving body, boom, arm, and bucket of the hydraulic excavator will be described using the above-mentioned type). An example is shown in which the constants β and r are set to predetermined values other than zero. However, the present invention is not limited to this, and the constants β and r can be set to zero for all flow control valves. Even in this case, the constant α in the above equation (1) can be set to a positive value, especially By setting α≦K to a positive value, the above-mentioned pressure compensation and flow dividing functions can be obtained in a circuit configuration with less liquid leakage and less pressure loss.

(Hereinafter referred to as T-Yoshun Tateryoji χ1) Next, other embodiments of the present invention will be described with reference to FIGS. 9 and 10. In these drawings, FIGS. Components that are equivalent to those in the illustrated embodiment are given the same reference numerals.

In the embodiment described above, the discharge pressure Ps of the hydraulic pump, the maximum load pressure P Inax, and the inlet pressure pz and outlet pressure Pc of the pilot valve are directly used to control the pressure compensation valve. However, these four pressures have a correlation with each other using the control pressure of the main valve back pressure chamber as a medium, and the pressure compensation valve can be controlled without directly using these four pressures, and the pressure compensation valve has the above-mentioned characteristics. can be given.

From this point of view, FIGS. 9 and 10 show an embodiment in which the four pressures described above are not directly used to control the pressure compensation valve.

That is, in FIGS. 9 and 10, the pressure compensation valve 121 disposed in the pilot circuit 25 of the flow control valve 120 is
A spool-shaped valve body 124 that controls communication between the inlet port 122 and the outlet port 123, a first hydraulic operation chamber 125 that biases the valve body 124 in the valve opening direction, and a first hydraulic operation chamber 125 that faces the first hydraulic operation chamber 125. second, third, and fourth hydraulic operation chambers 126 and 127 that are located so as to bias the valve body 124 in the valve closing direction;
．． It consists of 128. The first hydraulic operation chamber 125 is connected to the pilot valve 29 outlet side of the pilot circuit 25 via a pilot pipe 129, and the second hydraulic operation chamber 1
26 is connected to the pilot valve 29 artificial side of the pilot circuit 25 via a pilot line 130, the third hydraulic operation chamber 127 is connected to the outlet port 32 of the main valve 21 via a pilot line 131, and the fourth The hydraulic operation chamber 128 is connected to the maximum load pressure line 61 via a pilot line 132. As a result, the outlet pressure of the pilot valve 29 (control pressure of the main valve back pressure chamber 36) Pc is introduced into the first hydraulic operation chamber 125, and the inlet pressure P2 of the pilot valve 29 is introduced into the second hydraulic operation chamber 126. The load pressure P of the hydraulic actuator 6 or 7 is guided to the third hydraulic operation chamber 127.
1 is introduced into the fourth hydraulic operation chamber 128, and the maximum load pressure P l1aX of the hydraulic actuator 6.7 is introduced into the fourth hydraulic operation chamber 128.

The end face of the valve body 124 facing the first hydraulic operation chamber 125 defines a pressure receiving area aC that receives the outlet pressure pc of the pilot valve 29, and the end face of the valve body 124 facing the second hydraulic operation chamber 126 defines a pressure receiving area aC that receives the outlet pressure pc of the pilot valve 29.
The annular end face of 24 defines a pressure receiving area aZ that receives the inlet pressure pz of the pilot valve 129, and the third hydraulic operation chamber 127
The annular end surface of the valve body 124 facing the hydraulic actuator 6
or define a pressure receiving area a1 that receives the load pressure P1 of 7,
The end surface of the valve body 124 facing the fourth hydraulic operation chamber 128 defines a pressure receiving area al that receives the maximum load pressure P 1nax. These pressure receiving areas aC, aZ, at.

As in the first embodiment, am is set so that proportionality constants α, β, and γ, which will be explained below, are obtained.

The pressure balance equation of the valve body 124 in the pressure compensating valve 121 is expressed by the following equation.

ac pc =az PZ +al Pl +all
Plnax Also, the pressure balance equation of the valve body 35 in the main valve 21 is expressed by the following equation.

Ac Pc = As Ps + AI Pl From the above two formulas, derive the formula for calculating the differential pressure across the pilot valve 29, and transform it using AC = AS + AI. + (a c-a z-a II-a l) P I Therefore , then Pz −Pc =a (Ps −Plnax)+β(P
It can be expressed as lnax-Pl ) +γPl (4). Here, if the differential pressure across the pilot valve 29 is Δpz, pz - pc =Δpz.

Therefore, the same equation as equation (1) in the embodiment shown in FIG. 1 is obtained.

Therefore, in this embodiment as well, the proportionality constants α and β.

By setting γ to a predetermined value, the pilot valve 29
The differential pressure Δpz before and after is expressed as the differential pressure ps - pImax between the discharge pressure Ps of the hydraulic pump 1 and the maximum load pressure Plnax, and the differential pressure P between the maximum load pressure P1nax and the self-load pressure Pl.
It can be controlled proportionally to the three elements of 1max-Pl and self-load pressure P1, and the above-mentioned pressure compensation and flow diversion function (the first term on the right side), and the harmony in complex operations based on this pressure compensation and flow diversion function. function (second term on the right side) and self-pressure compensation function (third term on the right side). That is, in this embodiment, instead of directly using the inlet pressure pz of the pilot valve 29, the outlet pressure PC1, the discharge pressure Ps of the hydraulic pump 1, and the maximum load pressure PImax, the inlet pressure P
z, the above four pressures Pz using outlet pressure PC1 self-load pressure P1 and maximum load pressure pwax.

The same effect as using Pc, Ps, and pHax can be obtained.

Further embodiments of the present invention will be described with reference to FIGS. 11 and 12. In the fourth embodiment, the pressure compensating valve is disposed on the pilot valve 29 side of the pilot circuit. However, the pressure compensation valve can also be arranged on the pilot valve outlet side of the pilot circuit, ie between the pilot valve and the main valve backpressure chamber. Figures 11 and 12 show such an embodiment.

That is, in FIGS. 11 and 12, the flow control valve 140
is a pressure compensation valve 141 connected to the pilot circuit 25 between the pilot valve 29 and the back pressure chamber 36 of the main valve 21;
The pressure compensation valve 141 has a seat valve type valve body 14 that controls communication between the inlet port 142 and the outlet port 143.
4, first and second hydraulic operation chambers 145.146 that bias the valve body 144 in the valve opening direction, and third and fourth hydraulic operation chambers 147.146 that bias the valve body 144 in the valve closing direction. 148, the first hydraulic operating chamber 145 is connected to the outlet port 32 of the main valve 21 via a pilot line 149, and the second hydraulic operating chamber 146 is connected to the inlet port 14 of the pressure compensating valve 141.
The third hydraulic operation chamber 147 is formed in the inlet portion 150 that communicates with the pilot pipe 1 and the maximum load pressure pipe 61.
51, and the fourth hydraulic operation chamber 148 is connected to the back pressure chamber 36 of the main valve 21 via a pilot line 1'52. As a result, the load pressure P1 of the hydraulic actuator 6 or 7 is introduced into the first hydraulic operation chamber 145.
The outlet pressure pz of the pilot valve 29 is introduced into the second hydraulic operation chamber 146, the maximum load pressure pHaX is introduced into the third hydraulic operation chamber 147, and the main valve back pressure is introduced into the fourth hydraulic operation chamber 148. A control pressure PC in the pressure chamber 36 is introduced.

The annular end surface of the valve body 144 facing the first hydraulic operation chamber 145 defines a pressure receiving area al that receives the load pressure P1 of the hydraulic actuator 6 or 7.
The end face of the valve body 144 facing the defines a pressure receiving area aZ that receives the outlet pressure pz of the pilot valve 29, and the annular end face of the valve body 144 facing the third hydraulic operation chamber 147 defines a pressure receiving area aZ that receives the maximum load pressure pHlaX. The end face of the valve body 144 facing the fourth hydraulic operation chamber 148 defines the back pressure chamber 36
A pressure receiving area aC that receives a control pressure Pc is defined.

These pressure-receiving areas al, aZ, all, and aC are proportional constants α, which will be explained below, as in the above embodiment.

It is set so that β and γ can be obtained.

The pressure balance equation of the valve body 144 in the pressure compensation valve 141 is expressed by the following equation.

aCPC+all PlIlaX=al Pl +aZ
pz Also, the pressure balance equation of the valve body 35 in the main valve 21 is expressed by the following equation.

Ac Pc = As Ps + AI Pl From the above two formulas, derive the formula for calculating the differential pressure across pilot 1P29, and calculate AC
=AS When transformed using AI, +(a Z-a l-a C-am) P l Therefore, if we set, Ps - Pz = a (Ps - Plnax) + β (P
It can be expressed as lmax-p+ ) +rpl (5). Here, if the differential pressure across the pilot valve 29 is Δpz, then ps −pz =Δpz.

Therefore, the same equation as equation (1) in the embodiment shown in FIG. 1 is obtained.

Therefore, in this embodiment as well, the proportionality constants α and β.

By setting γ to a predetermined value, the pilot valve 29
The differential pressure Δpz before and after is expressed as the differential pressure Ps − Plnax between the discharge pressure Ps of the hydraulic pump 1 and the maximum load pressure P Inax,
It can be controlled in proportion to the three elements: the differential pressure PlIlax-Pl between the maximum load pressure P Inax and the self-load pressure P1, and the self-load pressure P1. Harmonic function in complex operation based on pressure compensation and flow division function (second term on the right side)
, the self-pressure compensation M function (third term on the right side) can be obtained. That is, in this embodiment, by arranging the pressure compensating valve 141 on the outlet side of the pilot valve 29, the same effect as when arranging it on the inlet side can be obtained.

Figures 13 and 14 show another embodiment in which the pressure compensating valve is placed on the pilot valve outlet side, and the inlet pressure and outlet pressure of the pilot valve, the discharge pressure and the maximum load pressure of the hydraulic pump are not directly used for control. shows.

13 and 14, a pressure compensation valve 161 disposed in the pilot circuit 25 of the flow control valve 160 includes a seat-type valve body 164 that controls communication between an inlet port 162 and an outlet port 163, and a valve body The first and second hydraulic operation chambers 165, 166 that bias the valve body 164 in the valve opening direction are located opposite to the first and second hydraulic operation chambers 165, 166, and bias the valve body 164 in the valve closing direction. It consists of third and fourth hydraulic operating chambers 167 and 168 for energizing. The first hydraulic operation chamber 165 is connected to the main valve 21 through a pilot line 169.
is connected to the outlet port 32 of the second hydraulic operating chamber 166
is formed in an inlet portion 179 communicating with the inlet port 162 of the pressure compensating valve 161, the third hydraulic operation chamber 167 is connected to the maximum load pressure line 61 via the pilot line 171, and the fourth hydraulic operation chamber 167 is connected to the maximum load pressure line 61 via the pilot line 171. Chamber 168 is pilot conduit 172
The pilot valve 29 of the pilot circuit 25 is connected to the artificial side via the pilot circuit 25. As a result, the first hydraulic operation chamber 165
The load pressure P1 of the hydraulic actuator 6 or 7 is introduced to the second hydraulic operation chamber 166, the discharge pressure ps of the hydraulic pump 1 is introduced to the third hydraulic operation chamber 167, and the load pressure P1 of the hydraulic actuator 6.7 is introduced to the second hydraulic operation chamber 166. The maximum load pressure P l1laX is introduced, and the inlet pressure pz of the pilot valve 29 is introduced into the fourth hydraulic operation chamber 168 .

The annular end surface of the valve body 164 facing the first hydraulic operation chamber 165 defines a pressure receiving area a1 that receives the load pressure P1 of the hydraulic actuator 6 or 7.
The end face of the valve body 164 facing the discharge pressure P of the hydraulic pump 1
A pressure receiving area aS receiving s is defined, and the third hydraulic operation chamber 1
If the annular end face of the valve body 164 faces 67, the maximum load pressure p
A pressure receiving area all that receives IIaX is defined, and the end face facing the fourth hydraulic operation chamber 168 defines a pressure receiving area aZ that receives the inlet pressure pz of the pilot valve 29. These pressure-receiving areas al, as, an, and aZ are proportional constants α and β, which will be explained below, as in the above embodiment.

It is set so that γ can be obtained.

The pressure balance equation of the valve body 164 in the pressure compensation valve 161 is expressed by the following equation.

az Pz +am Plmax=al Pl +as
Ps Also, the pressure balance equation of the valve body 35 in the main valve 21 is expressed by the following equation.

Ac Pc = As Ps + AI Pl From the above two equations, derive the equation for calculating the differential pressure across the pilot valve 29, and obtain AC=
When transformed using AS + AI, +(a S+a l-a m-a Z) P l Therefore, if we set, Pz - Pc = a (Ps - Plmax) + β (P
It can be expressed as lnax-Pl ) +rPl (6). Here, if the differential pressure across the pilot valve 29 is Δpz, pz - pc =ΔPz.

Therefore, the same equation as equation (1) in the embodiment shown in FIG. 1 is obtained.

Therefore, in this embodiment as well, the proportionality constants α and β.

By setting γ to a predetermined value, the pilot valve 29
The differential pressure Δpz before and after is expressed as the differential pressure Ps − Plmax between the discharge pressure Ps of the hydraulic pump 1 and the maximum load pressure P 1laX,
Differential pressure P1max-P1. between maximum load pressure Pll1aX and self-load pressure P1. The three elements of the self-load pressure P1 can be controlled in proportion to each other, and the above-mentioned pressure compensation and flow division function (the first term on the right side) and the harmonization function in complex operation based on this pressure compensation and flow division function (the first term on the right side) Section 2)
, a self-pressure compensation function (third term on the right side) can be obtained.

Still further embodiments of the invention are described with reference to FIGS. 15 and 16, all of which use four pressures for the pressure compensating valve. However, the four pressures of the hydraulic pump, the discharge pressure, the maximum load pressure, the pilot valve inlet pressure, and the outlet pressure, have a correlation with each other using the control pressure of the main valve back pressure chamber as a medium, and the four pressures are not used. The pressure compensating valve can be controlled even if the pressure compensating valve has the above-mentioned characteristics. FIGS. 15 and 16 show such an embodiment.

That is, in FIGS. 15 and 16, the flow control valve 180
has a pressure compensation valve 181 arranged in the pilot circuit 25 between the pilot valve 29 and the main valve back pressure chamber 36,
The pressure compensation valve 181 includes a seat-type valve body 184 that controls communication between the inlet port 182 and the outlet port 183, a first hydraulic operation chamber 185 that biases the valve body 184 in the valve opening direction, and a first hydraulic pressure chamber 185 that biases the valve body 184 in the valve opening direction. Located opposite the operation chamber 185, the valve body 18
4 in the valve closing direction.
It consists of 6.187. First hydraulic operation room 185
is formed in an inlet portion 188 that communicates with the inlet port 182 of the pressure compensating valve 181, and the second hydraulic operation chamber 186 is connected to the pilot valve 29 population side of the pilot circuit 25 or the main side of the meter-in circuit 15 via a pilot line 189. The third hydraulic operation chamber 187 is connected to the maximum load pressure line 61 via a pilot line 190. As a result, the outlet pressure pz of the pilot valve 29 is guided to the first hydraulic operation chamber 185, and the second hydraulic operation chamber 1
The discharge pressure ps of the hydraulic pump 1 is introduced to 86, and the maximum load pressure P1max is introduced to the third hydraulic operation chamber 187.

The end face of the valve body 184 facing the first hydraulic operation chamber 185 defines a pressure receiving area AZ that receives the outlet pressure pz of the pilot valve 29,
The annular end face of 84 defines a pressure receiving area aS that receives the discharge pressure Ps of the hydraulic pump 1, and the end face of the valve body 184 facing the third hydraulic operation chamber 187 defines a pressure receiving area a that receives the maximum load pressure PlIlaX. ing. These pressure receiving areas az
, as.

As in the above embodiment, all is set so that proportionality constants α, β, and γ, which will be explained below, are obtained.

The pressure balance equation of the valve body 184 in the pressure compensation valve 181 is expressed by the following equation.

az pz = as ps + alI pHaX Also, the pressure balance equation of the valve body 35 in the main valve 21 is expressed by the following equation.

ACPC=AS Ps +AI Pl From the above two equations, we derive the equation for calculating the differential pressure across the pilot valve 29, and when we transform it using Ac = As + AI, a z (P s - P z) = (a z - a s) ( P
s-P Inax) + (a Z-a S-a fi)
(P IIIaX-P I) + (aZ-as-all)
P l Therefore, if we set aZ, Ps - Pz = a (Ps - PlnaX) + β (P
It can be expressed as lIlaX−Pl ) +rPl (7). Here, if the differential pressure across the pilot valve 29 is Δpz, then PZ - PC = Δpz.

Therefore, the same equation as equation (1) in the embodiment shown in FIG. 1 is obtained. However, in this embodiment, β and γ have the same value, so they cannot be determined independently.

Therefore, in this embodiment as well, the proportionality constants α and β.

By setting γ to a predetermined value, the pilot valve 29
The differential pressure Δpz before and after is expressed as the differential pressure ps −pInax between the discharge pressure Ps of the hydraulic pump 1 and the maximum load pressure P1laX,
fi can be controlled in proportion to three elements: the differential pressure pHax-Pl between the large load pressure P Imax and the self-load pressure P1, and the self-load pressure P1. Based on this pressure compensation and flow division function, a harmonization function in complex operations (second term on the right side)
and a self-pressure compensation function (third term on the right side) can be obtained.

As explained above, the present invention controls the pressure compensation valve based on four pressures: the inlet pressure and outlet pressure of the pilot valve, the discharge pressure of the hydraulic pump 1, and the maximum load pressure, and performs pressure compensation and flow division functions, or A harmonization function based on pressure compensation and flow diversion function and/or a self-pressure compensation function can be selectively achieved. These four pressures have a correlation with each other using the control pressure of the main valve back pressure chamber as a medium, so there is no need to use these four pressures directly, or by using a pressure compensation valve on the inlet and outlet sides of the pilot valve. Regardless of the arrangement, the pressure compensation valve can be controlled.

The pressure compensating valve can also be controlled using a number of pressures other than four.

Next, another embodiment of the present invention regarding pump control means will be described. First, in the above embodiment, the hydraulic drive device of the present invention will be explained in combination with a load sensing type pump regulator, and the load sensing type pump regulator will be explained as controlling the discharge pressure of a variable displacement hydraulic pump. However, the hydraulic pump may be of a fixed displacement type. In this case, the pump regulator of the load sensor regulator type is configured as shown in FIG. 17. That is, in FIG. 17, the pump regulator 3
80 has a relief valve 383 having pilot chambers 381 and 382 facing each other, and a fixed displacement hydraulic pump 385 is connected to the pilot chamber 381 via a pilot line 384.
The maximum load pressure is introduced into the pilot chamber 382 via a pilot pipe line 386, and a spring 387 is arranged on the side of the pilot chamber 382. Thereby, the discharge pressure of the hydraulic pump 385 can be maintained higher than the maximum load pressure of the plurality of hydraulic actuators by a pressure corresponding to the strength of the spring 387.

Further, the hydraulic drive device of the present invention can be configured in combination with a pump regulator other than a load sensing type. Such an embodiment is shown in FIG. 18, namely:
In FIG. 18, a hydraulic pump 390 is connected to a flow control valve 391 consisting of a combination of the above-described main valve, pilot valve, and pressure compensation valve, and its discharge amount is adjusted by a pump flow control device 392. hydraulic pump 390
An unload valve 393 is connected between the flow control valve 391 and the flow control valve 391, and an operating device 394 is provided for the flow control valve 391. The operation signal of the operation device 394 is transmitted to the control device 3
95, and from there, it is further sent as a control signal to the pilot valve drive unit 396 of the flow rate control valve 391 to control the opening degree of the pilot valve. The operation signal sent to the control device 395 is also sent to the calculation device 397.
calculates the required flow rate of the flow rate control valve 391 from the map stored in advance in the storage device 398, and then calculates the required flow rate of the flow rate control valve 391.
Send a signal to 92. At the same time, the arithmetic unit 397
The set pressure of the unload valve 393 is calculated from another map stored in advance in the storage device 398, and the signal is output to the unload valve 393. As a result, the discharge pressure of the hydraulic pump 390 is changed to the memory device 39 as a function of the operation signal.
The pressure is controlled by the pressure obtained from a map stored in advance at 8.

In the hydraulic drive system of the present invention combined with such a pump control means, the differential pressure Ps - Plmax cannot be controlled to be constant in the first term on the right side of equation (1) described above, so the function of the first term on the right side is However, in a combined operation, the differential pressure is common to the flow control valves associated with a plurality of hydraulic actuators, so the flow dividing function can be achieved.

In addition, since the second and third terms on the right side of equation (1) have no relation to the pump discharge pressure Ps, if β and r are set to values other than zero, the harmonic function based on the shunt function and the /
Or it can perform self-pressure compensation function.

Although the embodiments of the present invention have been described above with reference to the drawings, the present invention is not limited to the specific embodiments described above, and various modifications and changes can be made within the spirit of the invention.

For example, in the above embodiment, an example was shown in which two hydraulic actuators are driven by a hydraulic pump, but the present invention can of course be applied to a case where there are three or more hydraulic actuators. Further, the pump control means may simply include a relief valve that maintains the hydraulic pump discharge pressure constant.

〔Effect of the invention〕

According to the present invention, by appropriately setting the constants α, β, and r to predetermined values, it is possible to perform a pressure compensation and flow division function, or a harmonic function and/or a self-pressure compensation function based on the pressure compensation and flow division function. can be selectively applied, and the characteristics of the flow control valve can be set to the optimum state according to the type of work member and work form of the hydraulic construction machine.

Therefore, in an application to a hydraulic excavator, for example, even if the swing and boom raising control levers are simultaneously operated up to the full stroke, the boom's rising speed will initially increase faster than the swinging speed, and once the boom has risen to a certain extent, When excavating with a combined operation using an arm, the flow control valve has characteristics that automatically perform a combined operation in which the rotation speed gradually increases and when the rotation reaches the maximum speed, the rotation speed becomes almost constant. The arm is driven reliably, and when the hydraulic actuator for the arm is on the low pressure side, the flow control valve characteristics prevent deterioration of fuel consumption and heat balance, and when digging trenches with complex operations using a bucket, the bucket reduces the excavation load. The flow control valve has characteristics that reduce the amount of flow passing through the flow control valve the moment it reaches the surface of the earth, reducing shock.When accelerating a turn, the flow that flows out from the relief valve is reduced, reducing wasted energy consumption. Flow control valve characteristics: Flow control valve characteristics that provide a powerful digging feeling when excavating with a bucket; Flow control valve that provides accurate slope formation when using a boom and arm to form slopes , characteristics, etc. can be obtained.

Furthermore, since the auxiliary valve is disposed in the pilot circuit, it is possible to provide a hydraulic circuit with less throttling loss and an energy-saving structure.

In addition, for the ratio K of the pressure receiving area receiving the load pressure of the hydraulic actuator to the pressure receiving area receiving the control pressure of the seat type main valve back pressure chamber, the first constant □number α is set in the relationship α≦K. In this case, the flow rate proportional to the amount of operation of the operating means (pilot valve opening degree) can be accurately divided in the flow dividing function. Here, when α is set, the maximum proportional gain can be provided while obtaining the flow dividing function of proportionally distributing the flow rate according to the manipulated variable.

[Brief explanation of the drawing]

1 is a schematic diagram showing the overall configuration of a hydraulic drive device according to an embodiment of the present invention, and FIG. 2 is a sectional view showing the structure of a flow control valve connected to a meter-in circuit of the hydraulic drive device. 3 is a sectional view showing the structure of the flow control valve connected to the meter-out circuit of the hydraulic drive device,
FIG. 4 is a side view of a hydraulic excavator to which the hydraulic drive device of the present invention is applied, FIG. 5 is a top view of the same hydraulic excavator, and FIG. 6 is a flow rate control of one of the hydraulic drive devices. 7A is a characteristic diagram showing an example of setting the proportionality constant α of a pressure compensation valve included in the valve, and FIGS. constant β
FIGS. 8(A) to 8(C) are characteristic diagrams showing examples of setting of the proportionality constant γ of a pressure compensation valve included in one flow control valve of the same hydraulic drive device. 9 is a schematic diagram of a flow control valve connected to a meter-in circuit of a hydraulic drive device according to another embodiment of the present invention, and FIG. 10 is a sectional view showing the structure of the flow control valve. FIG. 11 is a schematic diagram of a flow control valve connected to a meter-in circuit of a hydraulic drive device according to still another embodiment of the present invention, and FIG.
Figure 2 is a sectional view showing the structure of the flow control valve.
3 is a schematic diagram of a flow control valve connected to a meter-in circuit of a hydraulic drive device according to still another embodiment of the present invention, and FIG. 14 is a sectional view showing the structure of the flow control valve. FIG. 15 is a schematic diagram of a flow control valve connected to a meter-in circuit of a hydraulic drive device according to yet another embodiment of the present invention, and FIG. 16 is a cross-sectional view showing the structure of the flow control valve. FIG. 17 is a circuit diagram showing an embodiment of a load sensing type pump regulator when a constant displacement hydraulic pump is used in the hydraulic drive device of the present invention.
The figure is a circuit diagram showing an embodiment of a non-load sensing type pump control means used in the hydraulic drive system of the present invention. Explanation of symbols 1.385; 389...Hydraulic pump 2.3...Main circuit 6.7.107-110...Hydraulic actuator 8.
9.11, 12, 11A, 12A; 120; 140; 1
60; 180...Flow rate control valve means 10.380.392...Pump control means 21.22
, 21A, 22^... Main valve 25.26, 25A, 26
A...Pilot circuit 29.30, 29A, 30A・
... Pilot valve 31 ... Inlet port 32 ... Outlet port 33. 34, 33A, 34A; 121; 141
;161 ;181... Auxiliary valve 36... Back pressure chamber
44...Variable aperture 53-60; 125-1
32;145-152;165-172;185-19
0...control means

Claims (1)

[Scope of Claims] (1) At least one hydraulic pump; at least first and second hydraulic pumps each connected to the hydraulic pump via a main circuit and driven by pressure oil discharged from the hydraulic pump. a hydraulic actuator; first and second flow control valve means connected to respective main circuits between the hydraulic pump and the first and second hydraulic actuators; controlling the discharge pressure of the hydraulic pump; a pump control means; the first and second flow rate control valve means each have a first valve means that changes the degree of opening according to the amount of operation of the operation means; and a pump control means that is connected in series with the first valve means. a second valve means connected to the valve means for controlling a differential pressure between the inlet pressure and the outlet pressure of the valve means; A hydraulic drive device comprising a control means for controlling the second valve means based on an inlet pressure and an outlet pressure of the valve means, a discharge pressure of the hydraulic pump, and a maximum load pressure of the first and second hydraulic actuators, The first and second flow control valve means each include a valve body that controls communication between an inlet port and an outlet port connected to the main circuit, and a variable valve that changes the degree of opening in response to displacement of the valve body. a main valve having a back pressure chamber that communicates with the outlet port via the variable throttle and generates a control pressure that biases the valve body in the valve opening direction; and an inlet port of the main valve and the back pressure. a pilot circuit connected between the chamber; the first valve means being configured as a pilot valve connected to the pilot circuit and controlling the pilot flow flowing through the pilot circuit; The valve means is
connected to the pilot circuit and configured as an auxiliary valve for controlling the differential pressure between the inlet pressure and the outlet pressure of the pilot valve; the control means for each of the first and second flow control valve means; , the differential pressure between the inlet pressure and the outlet pressure of the pilot valve is the differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure of the first and second hydraulic actuators, and the maximum load pressure and the self of each hydraulic actuator are The auxiliary valve is controlled so that the differential pressure with the load pressure and the self-load pressure have the relationship expressed by the following formula, ΔPz=α(Ps-Plmax) +β(Plmax-Pl)+γPl where: ΔPz: Differential pressure between the inlet pressure and outlet pressure of the pilot valve Ps: Discharge pressure of the hydraulic pump Plmax: Maximum load pressure Pl of the first and second hydraulic actuators: Pl of the first and second hydraulic actuators Respective self-load pressures α, β, γ: first, second, and third constants The first, second, and third constants α, β, and γ are each set to predetermined values. Hydraulic drive. (2) If K is the ratio of the pressure receiving area of the main valve body receiving the load pressure of the associated hydraulic actuator via the outlet port to the pressure receiving area of the main valve body receiving the control pressure of the back pressure chamber, The hydraulic drive device according to claim 1, wherein the first constant α has a relationship of α≦K. (3) The hydraulic drive device according to claim 2, wherein the second and third constants β and γ are each set to zero. (4) The first constant α is set to an arbitrary positive value corresponding to the operation amount of the operation means and the proportional gain of the main flow rate passing through the main valve. Hydraulic drive. (5) The hydraulic drive device according to claim 1, wherein the second constant β is set to an arbitrary value based on coordination of combined operations of the related hydraulic actuator and other hydraulic actuators. (6) The hydraulic drive device according to claim 1, wherein the third constant γ is set to an arbitrary value based on the operating characteristics of the related hydraulic actuator. (7) The control means includes a plurality of hydraulic operation chambers provided in each of the auxiliary valves of the first and second flow control valve means, and a discharge pressure of the hydraulic pump in the plurality of hydraulic operation chambers. pipe means for directly or indirectly introducing the maximum load pressure, the inlet pressure and outlet pressure of the pilot valve, and the pressure receiving area of each of the plurality of hydraulic operation chambers is 2. The hydraulic drive device according to claim 1, wherein constants α, β, and γ of 3 are set to the predetermined values. (8) The auxiliary valve is disposed between the inlet port of the main valve and the pilot valve, and the plurality of hydraulic operation chambers include:
the auxiliary valve has first and second hydraulic operation chambers that bias the auxiliary valve in the opening direction, and third and fourth hydraulic operation chambers that bias the auxiliary valve in the valve closing direction; The means includes a first conduit for guiding the discharge pressure of the hydraulic pump to the first hydraulic operation chamber, a second conduit for guiding the outlet pressure of the pilot valve to the second hydraulic operation chamber, and A claim characterized by comprising a third pipe line that leads maximum load pressure to the third hydraulic operation chamber, and a fourth pipe line that leads the inlet pressure of the pilot valve to the fourth hydraulic operation chamber. Item 7. Hydraulic drive device. (9) The auxiliary valve is arranged between the back pressure chamber of the main valve and the pilot valve, and the plurality of hydraulic operation chambers include a first hydraulic operation chamber that biases the auxiliary valve in the opening direction. and second, third, and fourth hydraulic operation chambers that bias the auxiliary valve in the valve closing direction, and the conduit means transfers the outlet pressure of the pilot valve to the first hydraulic operation chamber. a first conduit leading to the pilot valve, a second conduit leading the inlet pressure of the pilot valve to the second hydraulic operating chamber, and a second conduit leading the load pressure of the associated hydraulic actuator to the third hydraulic operating chamber. 8. The hydraulic drive device according to claim 7, further comprising a third pipe line and a fourth pipe line that guides the maximum load pressure to the fourth hydraulic operation chamber. (10) The auxiliary valve is disposed between the pilot valve and the back pressure chamber of the main valve, and the plurality of hydraulic operation chambers include first and second hydraulic chambers that bias the auxiliary valve in the opening direction. It has a hydraulic operation chamber, and third and fourth hydraulic operation chambers that bias the auxiliary valve in the valve closing direction, and the conduit means controls the load pressure of the related hydraulic actuator to the first hydraulic operation. a first pipe line that leads the outlet pressure of the pilot valve to the second hydraulic operation chamber, and a third pipe line that leads the maximum load pressure to the third hydraulic operation chamber. 8. The hydraulic drive device according to claim 7, further comprising a conduit and a fourth conduit that guides the control pressure of the back pressure chamber to the fourth hydraulic operation chamber. (11) The auxiliary valve is disposed between the inlet port of the main valve and the pilot valve, and the plurality of hydraulic operation chambers have first and second hydraulic pressures that bias the auxiliary valve in the opening direction. an operation chamber, and a third and fourth auxiliary valve that urges the auxiliary valve in the valve-closing direction.
a hydraulic operation chamber, and the conduit means has a first conduit that guides the load pressure of the associated hydraulic actuator to the first hydraulic operation chamber, and a first conduit that guides the discharge pressure of the hydraulic pump to the second hydraulic operation chamber. a second pipe line leading to the hydraulic operation chamber; a third pipe line leading the maximum load pressure to the third hydraulic operation chamber; and a third pipe line leading the inlet pressure of the pilot valve to the fourth hydraulic operation chamber. 8. The hydraulic drive device according to claim 7, further comprising a fourth conduit. (12) The auxiliary valve is arranged between the pilot valve and the back pressure chamber of the main valve, and the plurality of hydraulic operation chambers include a first hydraulic operation chamber that biases the auxiliary valve in the opening direction. and second and third hydraulic operation chambers that bias the auxiliary valve in the valve closing direction, and the conduit means has a first hydraulic operation chamber that guides the outlet pressure of the pilot valve to the first hydraulic operation chamber. 1 pipe line, and a second pipe line that leads the discharge pressure of the hydraulic pump to the second hydraulic operation chamber.
8. The hydraulic drive device according to claim 7, further comprising: a pipe line and a third pipe line that guides the maximum load pressure to the third hydraulic operation chamber. (13) The pump control means is a load sensing type pump regulator that maintains the discharge pressure of the hydraulic pump higher than the maximum load pressure of the first and second hydraulic actuators by a predetermined value. Item 1
Hydraulic drive as described. (14) at least one hydraulic pump; a plurality of hydraulic actuators each connected to this hydraulic pump via a main circuit and driven by pressure oil discharged from the hydraulic pump; each driven by the plurality of hydraulic actuators; a plurality of work members including a revolving structure, a boom, an arm, and a bucket; a plurality of flow control valve means connected to respective main circuits between the hydraulic pump and the plurality of hydraulic actuators; and the hydraulic pump; a pump control means for controlling the discharge pressure of the plurality of flow control valve means; each of the plurality of flow rate control valve means includes a first valve means for changing the opening degree according to an operation amount of the operation means; a second valve means connected in series with the valve means for controlling the differential pressure between the inlet pressure and the outlet pressure of the valve means;
Furthermore, for each of the plurality of flow rate control valve means, based on the inlet pressure and outlet pressure of the first valve means, the discharge pressure of the hydraulic pump, and the maximum load pressure of the first and second hydraulic actuators, the In a hydraulic drive device having a control means for controlling a second valve means, each of the plurality of flow rate control valve means includes a valve body for controlling communication between an inlet port and an outlet port connected to the main circuit, and a valve body for controlling communication between an inlet port and an outlet port connected to the main circuit; a variable throttle whose opening degree changes in response to the displacement of the body; and a variable throttle that communicates with the outlet port via the variable throttle;
The main valve has a back pressure chamber that generates a control pressure that biases the valve body in the valve opening direction; and a pilot circuit connected between the inlet port of the main valve and the back pressure chamber; The first valve means is connected to the pilot circuit and configured as a pilot valve for controlling the pilot flow flowing through the pilot circuit, and the second valve means is connected to the pilot circuit and is configured as a pilot valve for controlling the pilot flow flowing through the pilot circuit. The control means is configured as an auxiliary valve for controlling the pressure difference between the pressure and the outlet pressure; The differential pressure between the inlet pressure and outlet pressure of the valve is the differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure of the plurality of hydraulic actuators,
The auxiliary valve is controlled so that the relation between the maximum load pressure and the self-load pressure of each hydraulic actuator and the self-load pressure is expressed by the following formula, ΔPz=α(Ps- Plmax) +β(Plmax-Pl)+γPl where ΔPz: Differential pressure Ps between the inlet pressure and outlet pressure of the pilot valve: Discharge pressure Plmax of the hydraulic pump: Maximum load pressure Pl of the plurality of hydraulic actuators: Each self-load pressure α, β, γ of the hydraulic actuator: first, second, and third constants The first, second, and third constants α, β, and γ are each set to a predetermined value. Hydraulic drive device. (15) If K is the ratio of the pressure receiving area of the main valve body receiving the load pressure of the associated hydraulic actuator via the outlet port to the pressure receiving area of the main valve body receiving the control pressure of the back pressure chamber, The hydraulic drive device according to claim 14, wherein the first constant α has a relationship of α≦K. (16) The hydraulic pressure according to claim 14, wherein the control means sets the second constant β to a relatively large positive value for the flow rate control valve means for the bottom side of the boom hydraulic actuator. shovel. (17) The hydraulic pressure according to claim 14, wherein the control means sets the second constant β to a relatively small positive value for the flow rate control valve means for the bottom side of the arm hydraulic actuator. shovel. (18) The hydraulic pressure according to claim 14, wherein the control means sets the second constant β to a relatively small negative value for the flow rate control valve on the bottom side of the bucket hydraulic actuator. shovel. (19) The hydraulic excavator according to claim 14, wherein the control means sets the third constant γ to a relatively small negative value for the flow rate control valve for the rotating body hydraulic actuator. (20) The control means includes:
15. The hydraulic excavator according to claim 14, wherein the third constant γ of the flow rate control valve related to the bottom side of the bucket hydraulic actuator is set to a relatively small positive value. (21) The control means sets the second and third constants β and r to zero, respectively, for the flow control valves on the rod sides of the boom and arm hydraulic actuators. 15. The hydraulic excavator according to 15.