JPH0830481B2 - Hydraulic drive - Google Patents

Description

The present invention relates to a hydraulic drive device for a hydraulic construction machine having a plurality of hydraulic actuators, such as a hydraulic shovel and a hydraulic crane, and more particularly to a hydraulic drive device having a pressure compensation function. The present invention relates to a hydraulic drive device that controls a flow rate of pressure oil supplied to a hydraulic actuator by a flow control valve.

[Conventional technology]

Conventionally, a hydraulic drive device of a hydraulic construction machine equipped with a plurality of hydraulic actuators such as a hydraulic shovel and a hydraulic crane,
Generally, 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 the respective main circuits therebetween.

USP 4,617,854, in such a hydraulic drive device, an auxiliary valve is arranged on the upstream side of the main circuit of each flow control valve,
While guiding the inlet pressure and the outlet pressure of the flow control valve to the opposing first operating portion of this auxiliary valve, and guiding the discharge pressure of the hydraulic pump and the maximum load pressure of the plurality of hydraulic actuators to the opposing second operating portion, There is described a configuration in which a load-sensing type pump regulator that holds the discharge pressure of the hydraulic pump higher than the maximum load pressure by a predetermined value is arranged. In this configuration, by guiding the inlet pressure and the outlet pressure of the flow control valve to the opposing first operation unit of the auxiliary valve,
As is well known, the load pressure of the flow control valve is compensated. In addition, by guiding the discharge pressure of the hydraulic pump controlled by the pump regulator and the maximum load pressure of the plurality of hydraulic actuators to the opposing second operating portion of the auxiliary valve, the combined operation of the plurality of hydraulic actuators having different load pressures is performed. In this case, even if the sum of the command flow rates (required flow rates) of the respective hydraulic actuators exceeds the maximum discharge flow rate of the hydraulic pump, the discharge flow rates are divided according to the mutual command flow rate ratios, and the high load pressure side The pressure oil is also reliably supplied to the hydraulic actuator.

On the other hand, USP 4,535,809 discloses a hydraulic drive device for a single hydraulic actuator instead of a plurality.
The flow control valve connected to the main circuit between the hydraulic pump and the hydraulic actuator is connected to the seat valve type main valve connected to the main circuit and the pilot circuit between the back pressure chamber and the outlet port of the main valve. In addition to being configured in combination with the connected pilot valve, an auxiliary valve is further arranged in the pilot circuit, and the inlet pressure and the outlet pressure of the pilot valve are guided to the opposing operation section of the auxiliary valve so as to perform a pressure compensation function. Is described. The patent also discloses a modified example in which the effect of self-load pressure is taken into consideration and the pressure compensation function is modified regarding the operation of a single hydraulic actuator.

[Problems to be Solved by the Invention]

However, in US Pat. No. 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. Therefore, when the hydraulic circuit is made to have a high pressure for energy saving, there is a problem that a large amount of liquid leaks from these spool valves. Further, since the auxiliary valve is arranged in the main circuit in which a large flow rate flows, there is a problem that the pressure loss in the auxiliary valve portion is large.

In general, in a hydraulic drive device, it is preferable that a flow rate can be supplied to each hydraulic actuator without being affected by a self-load pressure and a load pressure of another hydraulic actuator. However, in a hydraulic drive system for a construction machine such as a hydraulic excavator, it may be preferable to be influenced by the load pressure or the self-load pressure of another hydraulic actuator depending on the type and working mode of the working member driven by the hydraulic actuator. .

For example, in a hydraulic shovel, when turning and boom raising are performed at the same time and earth and sand are loaded on a truck, since the revolving body is an inertial body, the load pressure of the revolving motor becomes high in the initial stage of the revolving, and is provided for circuit protection. It rises above the pressure of the relief valve. On the other hand, the load pressure of the boom is the boom holding pressure, and thus is lower than the turning load pressure. In this type of work, when the swing pressure at the beginning of swing is high, energy can be reduced by supplying pressure oil to the boom as much as possible without relief, and at the same time, the boom rising speed can be increased at first. The speed of the boom and the turning can be automatically adjusted such that the turning speed is raised faster than the turning speed, and the turning speed gradually increases when the boom rises to some extent.

In addition, in a swing operation alone or in a combined operation with another hydraulic actuator, at the initial stage of the swing, the swing load pressure rises above the pressure of the relief valve as described above. If the supply of pressurized oil can be reduced, energy waste can be reduced.

There is also a work mode such as slope formation work performed by combined operation of the boom and arm where it is desired to divide the flow rate accurately according to the operation amount ratio of the boom operation lever and the arm operation lever regardless of the load pressure. .

Therefore, in a construction machine such as a hydraulic excavator, the characteristics of the flow control valve are not uniquely determined so as to perform a pressure compensation function and / or a diversion function, but are determined by the type and working mode of the working member driven by the hydraulic actuator. It is desirable to be able to exert the corresponding function.

However, in USP 4,617,854, although the pressure compensation function and the flow dividing function are achieved by installing the auxiliary valve as described above, the control incorporating the influence of the load pressure or the self-load pressure of other hydraulic actuators is not performed,
It has not been possible to set the characteristics of the flow control valve so that the function according to the type of work member and work form can be exerted.

On the other hand, in USP4,535,809, because it is a hydraulic drive device for a single hydraulic actuator, it only performs a pressure compensation function that incorporates the effect of self-load pressure of a single hydraulic actuator, and It was not possible to set the characteristics of the flow control valve so that it could exhibit the function according to the combined operation of the actuators.

An object of the present invention is to provide a hydraulic drive device that has a small amount of liquid leakage and throttle loss, and that can set the characteristics of a flow control valve so as to exhibit a function according to the type and working form of a working member of a hydraulic construction machine. is there.

[Means for solving the problem]

In order to achieve the above object, the present invention provides at least one
Two hydraulic pumps; at least first and second hydraulic actuators connected to the hydraulic pumps via main circuits and driven by pressure oil discharged from the hydraulic pumps; the hydraulic pumps, the first and second hydraulic pumps, respectively. First and second flow rate control valve means connected to respective main circuits between the second hydraulic actuators; and pump control means for controlling the discharge pressure of the hydraulic pump; the first
The second flow control valve means and the first valve means for changing the opening degree according to the operation amount of the operation means, and the first valve means are connected in series, and the inlet pressure of the valve means A second valve means for controlling the differential pressure of the outlet pressure, an inlet pressure and an outlet pressure of the first valve means, a discharge pressure of the hydraulic pump, and a maximum load pressure of the first and second hydraulic actuators. And a control unit for controlling the second valve unit, the first and second flow rate control valve units respectively connect the inlet port and the outlet port connected to the main circuit. A valve body to be controlled, a variable throttle for changing the opening degree according to the displacement of the valve body, and a control pressure communicating with the inlet port via the variable throttle to urge the valve body in the valve closing direction. A sheet type main valve having a back pressure chamber for generating Has connected the pilot circuit between the back pressure chamber and the outlet port of the first valve means,
A pilot valve connected to the pilot circuit for controlling a pilot flow flowing through the pilot circuit;
Valve means is an auxiliary valve that is connected to the pilot circuit and controls the differential pressure between the inlet pressure and the outlet pressure of the pilot valve, and the control means is each of the first and second flow rate control valve means. , The differential pressure between the inlet pressure and the outlet pressure of the pilot valve is ΔPz, the discharge pressure of the hydraulic pump is Ps, the maximum load pressure of the first and second hydraulic actuators is Plmax, and the first and second hydraulic pressures are When the self-load pressure of the actuator is Pl, ΔPz = α (Ps-Plmax) + β (Plmax-Pl) + γPl where α, β, γ satisfy the relation of the first, second and third constants. The auxiliary valve is controlled as described above, and the first constant α is set to a positive value by the first and second flow rate control valve means, respectively, and the first flow rate control valve means and the second flow rate control valve means are set. With the flow control valve means, the second and third constants β and γ can be reduced. Kutomo one of those set to different values.

Preferably, when the ratio of the pressure receiving area of the main valve valve body receiving the discharge pressure of the hydraulic pump via the inlet port to the pressure receiving area of the main valve valve body receiving the control pressure of the back pressure chamber is K, In each of the control means of the first and second flow rate control valve means, the first constant α is α ≦ K
Is set to the relationship.

Further, in one of the control means of the first and second flow rate control valve means, the second and third constants β and γ may be set to zero.

Further, as an example, the control means includes a plurality of hydraulic operation chambers provided in the auxiliary valve in each of the first and second flow rate control valve means, and the hydraulic pumps in the plurality of hydraulic operation chambers. And a pipe line means for directly or indirectly introducing the discharge pressure, the maximum load pressure, the inlet pressure and the outlet pressure of the pilot valve, and the pressure receiving area of the plurality of hydraulic operation chambers is ΔPz = α (Ps -Plmax) + β (Plmax-P
l) + γPl, and the first constant and the second flow control valve means respectively have a positive value, and the first flow control valve means and the first flow control valve means have a positive value. And the second and third constants β,
The setting relationship of the pressure receiving areas of the plurality of hydraulic operation chambers is made different so that at least one of γ has a different value.

In this case, for example, the auxiliary valve includes the first and second auxiliary valves.
In each of the flow control valve means, the plurality of hydraulic operation chambers are arranged between the back pressure chamber of the main valve and the pilot valve, and the plurality of hydraulic operation chambers are control means of the first and second flow control valves, respectively. In a first hydraulic operating chamber for urging the auxiliary valve in the valve opening direction, and second, third and fourth hydraulic operating chambers for urging the auxiliary valve in the valve closing direction, The pipe line means includes a first pipe line for guiding the control pressure of the main valve back pressure chamber to the first hydraulic operation chamber and a second pipe line for guiding the inlet pressure of the pilot valve to the second hydraulic operation chamber. A structure having a pipeline, a third pipeline that guides the maximum load pressure to the third hydraulic operation chamber, and a fourth pipeline that guides the discharge pressure of the hydraulic pump to the fourth hydraulic operation chamber. And

Further, in this case, the main valve and the auxiliary valve may be integrally formed in each of the first and second flow rate control valve means.

Further, as another example, the control means includes, in each of the first and second flow rate control valve means, an electromagnetic operation portion provided in the auxiliary valve, a discharge pressure of the hydraulic pump, and a maximum load pressure. , Pressure detection means for directly or indirectly detecting the inlet pressure and the outlet pressure of the pilot valve, and the above ΔPz = based on the detection signal of the pressure detection means.
The target differential pressure between the inlet pressure and the outlet pressure of the pilot valve that satisfies the relationship of α (Ps-Plmax) + β (Plmax-Pl) + γPl is calculated, and this target differential pressure is commanded by the electromagnetic operating unit of the auxiliary valve. A first value and a second value in the first and second flow rate control valve means, and the first flow rate control valve means and the second flow rate control valve means. The first, second and third constants α, β, γ are stored in the arithmetic means so that at least one of the second and third constants β, γ is different from the flow control valve means. Set.

Also, preferably, the pump control means is a load-sensing type pump regulator that holds the discharge pressure of the hydraulic pump higher than the maximum load pressure of the first and second hydraulic actuators by a predetermined value.

Further, in order to achieve the above object, the present invention provides at least one hydraulic pump; a plurality of hydraulic actuators connected to the hydraulic pump via a main circuit and driven by pressure oil discharged from the hydraulic pump. Driven by the plurality of hydraulic actuators,
A plurality of working members including a swing 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 a discharge pressure of the hydraulic pump. A plurality of flow control valve means, each of the plurality of flow control valve means is provided with a first valve means for changing an opening degree according to an operation amount of the operation means, and a first valve means in series. Second valve means connected to control the differential pressure between the inlet pressure and the outlet pressure of the valve means, the inlet pressure and the outlet pressure of the first valve means, the discharge pressure of the hydraulic pump, and the plurality of hydraulic actuators In the hydraulic excavator having a control means for controlling the second valve means based on the maximum load pressure of each of the plurality of flow control valve means, each of the plurality of flow control valve means includes an inlet port and an outlet port connected to the main circuit. For controlling the communication of the valve body, a variable throttle for changing the opening degree according to the displacement of the valve body, and a valve for communicating with the inlet port through the variable throttle to urge the valve body in the valve closing direction. A seat type main valve having a back pressure chamber for generating a control pressure; and a pilot circuit connected between the back pressure chamber of the main valve and an outlet port, wherein the first valve means comprises: A pilot valve that is connected to the pilot circuit and controls the pilot flow that flows through the pilot circuit.
The second valve means is connected to the pilot circuit,
It is an auxiliary valve for controlling the differential pressure between the inlet pressure and the outlet pressure of the pilot valve, and the control means includes the flow control valve means for at least two working members of the swing body, the boom, the arm, and the bucket. When the differential pressure between the inlet pressure and the outlet pressure of the pilot valve is ΔPz, the discharge pressure of the hydraulic pump is Ps, the maximum load pressure of the plurality of hydraulic actuators is Plmax, and the self-load pressure of the plurality of hydraulic actuators is Pl, ΔPz = α (Ps−Plmax) + β (Plmax−Pl) + γPl where α, β, γ control the auxiliary valve so as to satisfy the relationship of the first, second and third constants, and The first constant α is set to a positive value by each of the plurality of flow rate control valve means, and two of the plurality of flow rate control valve means are used.
The second and third constants β and γ between the two flow control valve means.
Is set to a different value.

Also in this case, preferably, the ratio of the pressure receiving area of the main valve valve body receiving the discharge pressure of the hydraulic pump via the inlet port to the pressure receiving area of the main valve valve body receiving the control pressure of the back pressure chamber is K. Then, in each of the control means of the plurality of flow rate control valve means, the first constant α is set to α
≦ K is set.

Further, preferably, in the control means of the flow rate control valve means for the bottom side of the boom and arm hydraulic actuators, the second constant β is set to a positive value, respectively, and for the bottom side of the boom hydraulic actuator. The second constant β is set to be larger than the second constant β related to the bottom side of the arm hydraulic actuator.

Further, preferably, the second constant β is set to a negative value in the control means of the flow control valve means on the bottom side of the bucket hydraulic actuator.

Further preferably, in the control means of the flow rate control valve means relating to the hydraulic actuator for revolving structure,
The constant γ of is set to a negative value.

Preferably, the third constant γ is set to a positive value in the control means of the flow control valve means on the bottom side of the bucket hydraulic actuator.

Further, preferably, in the control means of the flow rate control valve means on the rod side of the boom or arm hydraulic actuator, the second and third constants β and γ are set to zero.

[Action]

The present inventors have variously studied the relationship between the auxiliary valve arranged in the pilot circuit and the differential pressure across the pilot valve in the hydraulic drive system having the first and second flow rate control valve means as described above. It has been found that the differential pressure ΔPz across the pilot valve controlled by the auxiliary valve is generally represented by the above-mentioned formula re-illustrated below.

[Delta] Pz = [alpha] (Ps-Plmax) + [beta] (Plmax-Pl) + [gamma] Pl The inventors analyzed the meaning of the above equation as follows.
In this equation, the first term on the right side, Ps-Plmax, is common to all the flow control valve means, so it is related to the flow dividing function in the combined operation, and the second term, Plmax-Pl, changes depending on the maximum load pressure of other actuators. Therefore, it relates to the harmony function in the composite operation, and γPl in the third term changes according to the self-load pressure, and therefore relates to the self-pressure compensation function. The selection and degree of each of these three functions are determined by the values of constants α, β, and γ. Here, the diversion function of the first term is the basic function of the composite operation. Therefore, the constant α is set to a predetermined positive value in all the flow control valve means. Meanwhile, the second
The harmony function of the item and the self-pressure compensation function of the item 3 are functions added according to the type and work form of the related work member. Therefore, the constants β and γ are set to predetermined values including zero, and at least one of the constants α and β is set to different values between the first flow rate control valve means and the second flow rate control valve means. . In this way, α is set to a positive value in each of the flow rate control valve means, and at least one of β and γ is set to a different value in the first end control valve means and the second flow rate control valve means. By doing so, the characteristics of the flow control valve are set so as to exert a flow dividing function, or a harmonization function based on the flow dividing function and / or a self-pressure compensation function according to the type and work mode of the working member of the hydraulic construction machine. be able to.

Further, in the above-described configuration of the present invention, the auxiliary valve is installed in the pilot circuit instead of the main circuit, and the main valve arranged in the main circuit is configured as a seat valve. Therefore, it is possible to provide a hydraulic circuit with little liquid leakage and high pressure. Further, since the auxiliary valve is arranged in the pilot circuit, even if a large flow rate is supplied to the main circuit, there is almost no throttling loss in the auxiliary valve portion.

When the first constant α has a relationship of α ≦ K with respect to the ratio K of the pressure receiving area for receiving the discharge pressure of the hydraulic pump to the pressure receiving area for receiving the control pressure of the back pressure chamber, if The differential pressure obtained by α (Ps-Plmax) in the first term is within the range of the maximum front-rear differential pressure that can be taken by the pilot valve on the high load pressure side, and the above equation 1 is applied to both the first and second flow rate control valves. The differential pressures of the terms become substantially the same, and in the flow dividing function, the flow rate can be accurately divided in proportion to the operation amount (pilot valve opening) of the operation means.

Further, the first constant α is a proportional gain between the operation amount of the operation means (opening of the pilot valve) and the pilot flow rate,
Therefore, it has the meaning of 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. Where α = K
When set to, the maximum proportional gain can be given while obtaining the flow dividing function of proportionally distributing the flow rate according to the operation amount.

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

As is apparent from the above description, the third constant γ is set to an arbitrary value in consideration of the operating characteristics of the associated hydraulic actuator. This is also set to zero, especially if it is preferable not to be affected by the self-load pressure.

When the control means is hydraulically constructed and the main valve and the auxiliary valve are integrated, a compact and rational arrangement valve structure can be obtained.

When the control means is electrically configured, the first, second and third constants α, β and γ can be easily set and changed.

When the pump control means is a load-sensing type pump regulator, the differential pressure Ps-Plmax between the discharge pressure of the hydraulic pump according to the first term on the right side of the above equation and the maximum load pressure of the plurality of hydraulic actuators is constant. Becomes Accordingly, the pressure difference between the inlet pressure and the outlet pressure of the pilot valve can be controlled to be constant, and a pressure compensation function for maintaining the flow rate constant regardless of the change in the pressure difference between the inlet port and the outlet port of the main valve is performed.

In the present invention in which the above hydraulic drive device is applied to a hydraulic excavator, the characteristics of the flow control valve relating to at least two working members of the revolving structure, the boom, the arm, and the bucket are set according to the type and working form of the working member. Can
The above-mentioned diversion function, or a harmony function and / or a self-pressure compensation function based on the diversion function can be exerted.

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, a low load side is set at the time of initial swing acceleration in the combined swing and boom operation. The flow rate control valve on the bottom side of the boom hydraulic actuator is a flow rate according to the increase in the differential pressure between the maximum load pressure (swing pressure) and the self load pressure (boom pressure), increasing the boom rising speed. be able to. As a result, even if the operating levers for turning and raising the boom are operated at the same time up to the full stroke, the speed at which the boom rises initially increases faster than the turning speed, and when the boom rises to a certain extent, the turning speed gradually increases, When the vehicle reaches the maximum speed, the compound operation that the turning speed becomes substantially constant is automatically performed.

When the second constant β is set to a relatively small positive value with respect to the flow control valve means on the bottom side of the hydraulic actuator for the arm, when excavating by a combined operation using the arm, When the arm hydraulic actuator is on the low pressure side while being driven reliably, the flow control valve is opened according to the increase in the differential pressure between the maximum load pressure (other hydraulic actuator pressure) and the self-load pressure (arm pressure). The degree is widened, and the degree of restriction of the flow rate is reduced.
As a result, deterioration of fuel efficiency and heat balance is prevented.

When the second constant β is set to a relatively small negative value for the flow control valve on the bottom side of the bucket hydraulic actuator, when the bucket is used for excavation work by a combined operation, the bucket has an excavating load. Is released from the ground, and at the moment when it comes to the surface of the earth, the flow rate of the flow control valve is reduced by increasing the differential pressure between the maximum load pressure (other hydraulic actuator pressure) and the self-load pressure (bucket pressure) to reduce the shock. You can

Further, when the third constant γ is set to a relatively small negative value for the flow control valve related to the hydraulic actuator for the swinging body, the swing is increased according to the increase of the swing pressure (self-load pressure) during the swing acceleration. The flow rate through the flow control valve, and the flow rate out of the relief valve,
Energy consumption can be reduced.

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

When the second and third constants β and γ are set to zero for the flow rate control valve on the rod side of the hydraulic actuator for the boom and the arm, the slope of the inclined surface is formed using the boom and the arm. During operation, the influence of the load pressure of other hydraulic actuators and the self-load pressure are completely eliminated, and the flow rate is accurately proportionally distributed according to the operation amount of the operating lever for the boom and the operating lever for the arm to form an accurate slope. It can be performed.

〔Example〕

 Embodiments of the present invention will be described below with reference to the drawings.

Basic Embodiment Referring to FIG. 1, a hydraulic drive system according to an embodiment of the present invention includes, for example, a swash plate type variable displacement hydraulic pump 1 and main pipelines 2, 3 and 4, which form a main circuit in the hydraulic pump 1, respectively. Five
Between the hydraulic pump 1 and the hydraulic actuators 6, 7 and a plurality of hydraulic actuators 6, 7 that are connected via the hydraulic pump 1 and are driven by the pressure oil discharged from the hydraulic pump 1. Load sensing which has flow control valves 8 and 9 connected to the hydraulic pump 1 and holds the discharge pressure of the hydraulic pump 1 higher than the maximum load pressure of the hydraulic actuators 6 and 7 by a predetermined value. A type pump regulator 10 is provided.

The flow control valve 8 is a seat valve type main valve 11 connected to the main pipes 2 and 3 between the hydraulic pump 1 and the hydraulic actuator 6.
And pilot lines 12, 13, 14 constituting a pilot circuit for the main valve 11, a pilot valve 15 connected to the pilot lines 13, 14, and a pilot valve 15 connected to the pilot lines 12, 13 in series with the pilot valve 15. And a pressure compensating valve 16 as an auxiliary valve.

The main valve 11 includes a valve housing 19 having an inlet port 17 and an outlet port 18 connected to the main pipelines 2 and 3, and a valve housing
A valve body 21 that is disposed inside the valve seat 19 and that engages with the valve seat 20; and the inlet port according to the displacement (opening) of the valve body 21 with respect to the valve seat 20.
Controls communication between 17 and exit port 18. A plurality of slits 22 are formed in the outer periphery of the valve body 21 in the axial direction, and the slits 22 cooperate with the inner wall of the valve housing 19 to form a variable throttle 23 that changes the opening in accordance with the displacement of the valve body 21. doing. Behind the valve body 21 inside the valve housing 21, a back pressure chamber 24 that communicates with the inlet port 17 via a variable throttle 23 and generates a control pressure Pc is formed.

As shown in FIG. 2, an upper annular end surface of the valve body 21 facing the inlet port 17 defines an annular pressure receiving area As for receiving the discharge pressure Ps of the hydraulic pump 1, and a bottom portion facing the outlet port 18. The wall surface defines a pressure receiving area Al for receiving the load pressure Pl of the hydraulic actuator 6, and the top end surface facing the back pressure chamber 24 defines a pressure receiving area Ac for receiving the control pressure Pc. Here, each pressure receiving area has a relationship of Ac = As + Al.

In the pilot circuit, the pilot line 12 is the main valve 11
The pilot line 14 is connected to the back pressure chamber 24
Connected to 18.

The pilot valve 15, as shown in FIG.
Needle-type valve body that controls communication between the inlet port 31 that is driven by the operating lever 30 and that is connected to the pilot conduit 13 and the outlet port 32 that is connected to the pilot conduit 14
It consists of 33.

The pressure compensation valve 16 has an inlet port 40 connected to the pilot line 12 and an outlet port connected to the pilot line 13.
A spool type valve body 42 for controlling communication between 41 and first and second hydraulic operation chambers 43, 44 for urging the valve body 42 in a valve opening direction.
And third and fourth hydraulic operating chambers 45, 46 which are located opposite to the first and second hydraulic operating chambers 43, 44 and urge the valve element 42 in the valve closing direction. The first hydraulic operating chamber 43 is connected to the main pipeline 2 via the pilot pipeline 47, and the second hydraulic operating chamber 44 is connected to the pilot pipeline via the pilot pipeline 48.
That is, the third hydraulic operating chamber 45 is connected to the outlet side of the pilot valve 15, the third hydraulic operating chamber 45 is connected to the maximum load pressure pipeline 50 described later via the pilot pipeline 49, and the fourth hydraulic operating chamber 46 is the pilot pipeline. Pilot line 13 or 51 through 51
Is connected to the entrance side of. The pilot conduit 51 is formed in the valve body 42 as an internal passage. This makes the first
The discharge pressure Ps of the hydraulic pump 1 is introduced into the hydraulic operation chamber 43 of the second hydraulic operation chamber 43, and the outlet pressure of the pilot valve 15 is introduced into the second hydraulic operation chamber 44.
Pl is introduced and the pilot valve 15 is installed in the third hydraulic operating chamber 45.
The inlet pressure Pz is introduced, and the load pressure on the high pressure side of the hydraulic actuators 6 and 7, that is, the maximum load pressure Plmax is introduced into the fourth hydraulic operation chamber 46. The end surface of the valve body 42 facing the first hydraulic operation chamber 43 defines the pressure receiving area as for receiving the discharge pressure Ps of the hydraulic pump 1, and the annular end surface of the valve body 42 facing the second hydraulic operation chamber 44 has the pilot valve 15 The pressure receiving area al for receiving the outlet pressure Pl is defined, and the end surface facing the third hydraulic operating chamber 45 defines the pressure receiving area am for receiving the maximum load pressure Plmax of the hydraulic actuators 6, 7, and the fourth hydraulic operating chamber 46 is provided. The facing annular end surface defines a pressure receiving area az that receives the inlet pressure Pz of the pilot valve 15.

The first to fourth hydraulic operation chambers 43 to 46 configured in this way
In the pilot lines 47 to 49, 51, the differential pressure ΔPz (= Pz−Pl) between the inlet pressure and the outlet pressure of the pilot valve 15 is the discharge pressure of the hydraulic pump 1 and the maximum load pressure of the two hydraulic actuators 6, 7. Differential pressure Ps-Plmax, differential pressure Plmax-P between maximum load pressure and self-load pressure of each hydraulic actuator
The control means is configured to control the pressure compensating valve so as to have a relationship represented by the following equation with respect to l and the self-load pressure Pl: ΔPz = α (Ps-Plmax) + β (Plmax-Pl ) + ΓPl (1) where α, β and γ are the first, second and third constants, respectively, and α is set to a positive value by the flow control valves 8 and 9, respectively, and the flow control valve 8 and the flow rate are set. At least one of β and γ is set to a different value from the control valve 9. In the present embodiment, the setting of the predetermined values of the first, second and third constants α, β, γ is performed by setting the pressure receiving areas as, al, am of the first to fourth hydraulic operation chambers 43 to 46. , az is selected. That is,
Pressure receiving areas as, al, am, az of the first to fourth hydraulic operating chambers 43 to 46
Is set so that predetermined values of the first, second and third constants α, β and γ can be obtained. Further, the pressure receiving areas as, al, am, az of the first to fourth hydraulic operation chambers 43 to 46 are such that the valve body 42 is held in the open position when the main valve 11 and the pilot valve 15 are closed. It is set.

In the flow control valve 8 thus configured, the combination of the seat valve type main valve 11 and the pilot valve 15 is USP
Known from 4,535,809, as described in the patent specification, when the operation lever 30 of the pilot valve 15 is operated, a pilot flow according to the opening degree of the pilot valve 15 is formed in the pilot circuits 12 to 14, Due to the actions of the variable throttle 23 and the back pressure chamber 24, the main valve valve body 21 opens to an opening degree proportional to the pilot flow rate, and the operation amount of the operation lever 30 (opening degree of the pilot valve 15)
A flow rate according to the above flows out from the inlet port 17 to the outlet port 18 through the main valve 11.

The flow rate control valve 9 is also configured similarly to the flow rate control valve 8 and has a seat valve type main valve 70, pilot conduits 71, 72, 73 forming a pilot circuit, a pilot valve 74 and a pressure compensation valve 75. .

The pilot lines 14, 73 of the flow control valves 8, 9 are connected to the maximum load pressure line 50 via load pressure introduction lines 54, 55 having check valves 52, 53, respectively, and hydraulic actuators 6, 7
The load pressure on the high pressure side is guided to the maximum load pressure line 50 as the maximum load pressure. The maximum load pressure line 50 is connected to the tank 57 via a throttle 56.

Further, in the main pipe lines 3 and 5 on the downstream side of the main valves 11 and 70 of the flow rate control valves 8 and 9, check valves 58 and 5 that block the flow of pressure oil from the hydraulic actuators 6 and 7 toward the main valves 11 and 70, respectively. 59 is connected.

The pump regulator 10 includes an auxiliary pump 60 and an auxiliary pump.
It is composed of a hydraulic cylinder type swash plate tilting device 61 driven by pressure oil discharged from 60, and a control valve 62 connected between the auxiliary pump 60 and the tank 56 and the swash plate tilting device 61. There is. The control valve 62 has first and second pilot chambers 63, 64 at opposite ends, and a pressure setting spring 65 is installed at the end where the second pilot chamber 64 is located. The first and second pilot chambers 63 and 64 are connected to the main pipeline 2 and the maximum load pressure pipeline 50 via pilot pipelines 66 and 67, respectively. As a result, the control valve 62 receives the discharge pressure of the hydraulic pump 1 and the maximum load pressure and the force of the spring 65 in opposition to each other, and supplies and discharges the pressure oil to and from the swash plate tilting device 61 in response to changes in the maximum load pressure. Is controlled, and the discharge pressure of the hydraulic pump 1 is maintained at a pressure higher than the maximum load pressure by a set pressure corresponding to the strength of the spring 65.

Operating Principle Next, the operating principle of the pressure compensation valves 16 and 75 will be described. For each of the pressure compensation valves 16 and 75, the pressure balance equation of the valve element 42 is represented by the following equation.

asPs + alPl = amPlmax + azPz When this formula is transformed Then, Pz−Pl = α (Ps−Plmax) + β (Plmax−Pl) + γPl can be expressed. Here, Pz−Pl = ΔPz. Therefore, the same formula as the formula (1) described above is obtained. This formula is reproduced below.

ΔPz = α (Ps−Plmax) + β (Plmax−Pl) + γPl (1) Then, the above formula (1) will be considered. In the formula (1), the left side is the inlet pressure Pz and the outlet pressure Pl of the pilot valve 15.
Is the differential pressure ΔPz. The first term on the right side relates to the pressure difference between the discharge pressure Ps of the hydraulic pump 1 and the maximum load pressure Plmax,
α is a proportionality constant. The second term is the maximum load pressure Plmax and the load pressure of the hydraulic actuator 6 or 7, that is, the self-load pressure.
It is a term related to the pressure difference with Pl, and β is a proportional constant. The third term is determined by the self-load pressure Pl, and γ is a proportional constant. That is, the equation (1) shows that the pressure compensating valves 16 and 75 have four pressures P.
Inlet pressure Pz of pilot valve 15 based on s, Plmax, Pl, Pz
And the outlet pressure Pl, the differential pressure ΔPz can be controlled; at that time, the differential pressure ΔPz is calculated by comparing the discharge pressure Ps and the maximum load pressure Pl of the hydraulic pump 1.
It is possible to control proportionally to each of the three elements of the differential pressure Ps-Plmax with respect to the max, the differential pressure Plmax-Pl with the maximum load pressure Plmax and the self-load pressure Pl, and the self-load pressure Pl;
The degree proportional to the two elements Ps-Plmax, Plmax-Pl, Pl,
It means that the values can be set arbitrarily by selecting the values of the proportional constants α, β, and γ.

Controlling the differential pressure ΔPz across the pilot valves 15 and 74 by the pressure compensating valves 16 and 75 here is controlling the pilot flow rate through the pilot valves 15 and 74, and as a result, as described above, the sheet type Controlling the main flow rate through the main valves 11,70 from a known function of the combination of the main valves 11,70 and the pilot valves 15,74.

In the first term on the right side, the differential pressure Ps-Plmax is a load-sensing type pump regulator as pump control means.
In this embodiment including 10, the pump regulator 10 is constant as long as it functions effectively, and is common to the two pressure compensation valves 16 and 75.

Therefore, in the first term on the right-hand side, controlling the differential pressure ΔPz across the pilot valves 15 and 74 so as to have a proportional relationship with the differential pressure Ps−Plmax means that the pump regulator 10 is operating effectively. Is to control the differential pressure ΔPz to be constant, and if the opening of the pilot valves 15 and 74 is constant, even if the main valve inlet pressure Ps or the outlet pressure Pl fluctuates,
It is to control the main flow rate through 11,70 to be constant. That is, it fulfills a pressure compensation function.

Further, in the operating state in which the pump regulator 10 does not function effectively, such as the case where the total of the required consumption flow rates of the hydraulic actuators 6 and 7 becomes larger than the maximum discharge flow rate of the hydraulic pump 1 and the discharge pressure of the hydraulic pump 1 decreases. Shows that the differential pressure ΔPz decreases as the differential pressure Ps-Plmax decreases, and the main flow rate through the main valves 11 and 70 also decreases, but the differential pressure Ps-Plmax is 2
Since it is common to the two pressure compensation valves 16 and 75, the main valve 11 and
The main flow rate through 70 decreases at the same rate. Therefore, the main valve 1
The main flow rate passing through 1,70 is proportionally distributed according to the operation amount of each operation lever 30 of the pilot valves 15,74 (opening degree of the pilot valves 15,74), and the discharge flow rate of the hydraulic pump 1 on the high pressure side is It is also reliably supplied to the hydraulic actuator. That is, a shunt function is obtained.

In the second term on the right side, controlling the differential pressure ΔPz across the pilot valves 15 and 74 so as to have a proportional relationship with the differential pressure Plmax-Pl is to control the load pressure Pl of another hydraulic actuator.
When max is larger than the self-load pressure Pl, the differential pressure ΔPz across the pilot valve 15 or 74 is changed depending on the maximum load pressure Plmax of other hydraulic actuators. If it is constant, the main flow rate through the main valves 11 and 70 is changed depending on the maximum load pressure Plmax. In general, it is preferable that the flow control of the flow control valve is not affected by other hydraulic actuators.However, in a hydraulic construction machine such as a hydraulic shovel, the flow is controlled by the load pressure of another hydraulic actuator. Is also preferable in some work modes. In such a case, the second term on the right side performs a harmonic function of changing the flow rate in harmony with another hydraulic actuator.

Further, in the third item on the right side, the pilot valves 15, 74
Controlling the differential pressure ΔPz across the pilot valve 15 so that it has a proportional relationship with the self-load pressure Pl changes the differential pressure ΔPz across the pilot valve 15 or 74 according to the change in the self-load pressure Pl. If the opening of 15 or 74 is kept constant, the main flow rate through the main valves 11 and 70 is changed depending on the self-load pressure Pl. This provides a self-pressure compensation function that changes the flow rate in accordance with the change in the self-load pressure.

As described above, in the above equation (1), the first term on the right-hand side relates to the pressure compensation and flow dividing function, the second term relates to the harmony function with other actuators, and the third term relates to the self-pressure compensation function. These three functions can be arbitrarily set by selecting the proportional constants α, β, and γ for the necessity and degree of the functions.

By the way, among these three functions, the pressure compensation and branching functions according to the first term are basic functions in a hydraulic construction machine such as a hydraulic shovel, and are preferably always present irrespective of the type and working form of the hydraulic actuator. Therefore, the proportional constant α is set to an arbitrary positive value. Here, since the differential pressure ΔPz across the pilot valves 15 and 74 determines the pilot flow rate with respect to the opening of the pilot valves 15 and 74 which is determined by the operation amount of the operating lever 30, the differential pressure Plmax of the first term is set.
-The proportional constant α for Pl is the proportional gain of the pilot flow rate with respect to the operation amount (pilot valve opening) of the operation lever 30 of the pilot valves 15 and 74, and thus the main valve with respect to the operation amount.
It has the meaning of proportional gain of the main flow rate through 11,70. Therefore, the proportionality constant α is determined according to the proportional gain.

Further, if the ratio of the pressure receiving area As receiving the pressure Pc of the back pressure chamber 24 of the main valve body 21 to the pressure receiving area As receiving the discharge pressure Ps of the hydraulic pump 1 of the valve body 21 is K, the pressure of the valve body 21 is The equilibrium formula is Pc = KPs + (1-K) Pl. On the other hand, the control pressure Pc and the inlet pressure Pz of the pilot valves 15 and 74 have a relationship of Pc ≧ Pz, and Pc = Pz when the pressure compensation valves 16 and 75 are completely open. Therefore, the differential pressure Pz-Pl (ΔPz) across the pilot valves 15 and 74 is Pz-Pl≤Pc-Pl = K (Ps-Pl) (2). That is, the maximum differential pressure that the pilot valves 15 and 74 can take is K (Ps-Pl). In the above equation (1), β =
When 0 and γ = 0 and the maximum load pressure side (Plmax = Pl) during combined operation of the hydraulic actuators 6 and 7 is considered, Pz−Pl = α (Ps−Plmax) ≦ K (Ps−Plmax) (3 ). Therefore, when the value of α of α> K is set,
The pilot valve on the maximum load pressure side cannot obtain a differential pressure of K (Ps-Plmax) or more, while the pilot valve on the low pressure side has a differential pressure of α (Ps-Plmax)> K (Ps-Plmax). Therefore, even if the pilot valve opening is the same, the differential pressure across the pilot valve is not the same, and the pilot flow rate is different. Therefore, the flow rate cannot be proportionally distributed according to the manipulated variable. However, even if the proportional distribution cannot be performed, the pressure oil can be reliably supplied to the high-pressure side hydraulic actuator.

Therefore, regarding the flow dividing function of the pressure compensating valves 16 and 75, in order to obtain the flow dividing function of distributing the flow rate in proportion to the operation amount (opening degree) of the pilot valve, the proportional constant α is set to α ≦ K.
In particular, when α = K is set, the maximum flow rate can be given for the same pilot valve opening, and the most efficient valve structure can be provided.

Also, as described above, when α is set to α> K, the pilot valve on the low load pressure side has α (Ps−Plmax)
> K (Ps-Plmax) is obtained, but when the composite operation is switched to the single operation of the hydraulic actuator on the low load pressure side, the K is also applied to the pilot valve on the low load pressure side.
(Ps-Pl) or more, and the differential pressure across the pilot valve changes from α (Ps-Plmax) to K (Ps
−Pl) and the pilot flow is correspondingly reduced. As a result, the flow rate supplied to the hydraulic actuator also decreases, the work member is decelerated, and smooth work becomes difficult to perform. On the other hand, when α is set to α ≦ K,
Even in combined operation, the differential pressure across the pilot valve on the low load pressure side is limited to K (Ps-Plmax), and even when the combined operation is switched to independent operation, the differential pressure does not fluctuate and stable work is performed. be able to. Therefore, in this sense, it is preferable to set α to α ≦ K.

As can be understood from the above, when the flow rate is accurately proportionally distributed according to the operation amounts of the operation levers of the plurality of hydraulic actuators, it is an essential condition to set α ≦ K.

The necessity of the harmony function according to the second term differs depending on the type and working form of the working member driven by the hydraulic actuators 6 and 7, and in some cases, it is better not to be affected by the load pressure of other actuators at all. There are also preferred working members and working configurations. Accordingly, the proportionality constant β is set to an arbitrary value including zero based on the harmony of the combined operation of the relevant hydraulic actuator and another hydraulic actuator.

The necessity of the self-pressure compensation function according to the third term differs depending on the type of the working member driven by the hydraulic actuators 6 and 7. In some cases, it is preferable that the working member not receive the influence of the self-load pressure at all. is there. Therefore, the proportionality constant γ is set to an arbitrary value including zero according to the type of the working member driven by the associated hydraulic actuator.

As described above, the constant α is set to a positive value by the flow rate control valves 8 and 9, and at least one of the constants β and γ is set to a different value between the flow rate control valve 8 and the flow rate control valve 9. The characteristics of the flow control valve can be set so as to exert a diversion function, or a harmonization function and / or a self-pressure compensation function based on the diversion function according to the type and work form of the work member of the hydraulic construction machine. .

Then, as described above, the proportional constants α, β, γ are represented by the pressure receiving areas as, al, am, az in the first to fourth operation chambers 43 to 46 of the pressure compensation valves 16 and 17, respectively. Therefore, if the proportional constants α, β, γ are determined, the pressure receiving areas as, al, am, az are set so that the proportional constants α, β, γ can be obtained. As a special case, if as + a1 = am + az,
γ = 0 can be set, and β = 0 can be set by configuring as = am. If as = al = am = az, then β = γ
It can be set to = 0.

Next, when the hydraulic drive device of the present embodiment is applied to a backhoe type excavator, the proportional constants α, β,
A specific setting example of γ will be described.

As shown in FIGS. 3 and 4, a hydraulic excavator generally has a pair of traveling bodies 80, a revolving structure 81 rotatably mounted on the traveling structure 80, and a plane perpendicular to the revolving structure 81. Has a front attachment 82 that is rotatably mounted. The front attachment 82 includes a boom 83, an arm 84, and a bucket.
It consists of 85. The traveling body 80, the swinging body 81, the boom 83, the arm 84, and the bucket 85 are the traveling motors 86 (plurality), the swinging motor 87, the boom cylinder 88, and the arm cylinder 8 respectively.
9. Driven by bucket cylinder 90. Here, the turning motor 87, the boom cylinder 88, the arm cylinder 89, and the bucket cylinder 90 correspond to the hydraulic actuators 6, 7 shown in FIG.

In such a hydraulic excavator hydraulic drive, a swing motor 87, a boom cylinder 88, an arm cylinder 89
Further, the proportional constant α relating to all the flow rate control valves of the bucket cylinder 90 is set to the same arbitrary positive value in consideration of the above-mentioned proportional gain, as shown in FIG. In the flow control valve related to the swing motor 87, the proportional constant β is the sixth
As shown in FIG. 7A, β = 0 is set, and the proportional constant γ is set to a relatively small negative value as shown in FIG. In the flow control valve related to the bottom side of the boom cylinder 88, the proportional constant β is set to an arbitrary positive value as shown in FIG. 6 (B), and the proportional constant γ is shown in FIG. 7 (B). Is set to γ = 0. In the flow control valve on the bottom side of the arm cylinder 89, the proportional constant β is set to a relatively small positive value as shown in FIG. 6 (C), and the proportional constant γ is shown in FIG. 7 (B). Is set to γ = 0. In the flow control valve on the bottom side of the bucket cylinder 90, the proportional constant β is set to a relatively small negative value as shown in FIG. 6 (D), and the proportional constant γ is shown in FIG. 7 (C). Is set to a relatively small positive value. Further, in the flow control valve related to the rod side of the boom cylinder 88, the flow control valve related to the rod side of the arm cylinder 89, and the flow control valve related to the rod side of the bucket cylinder 90, the proportional constants β and γ are all Fig. 6 (A)
And is set to zero as shown in FIG. 7 (B).

Operation of the Embodiment Next, the operation of the hydraulic drive system configured as described above will be described.

First, when neither of the operation levers 30 of the flow control valves 8 and 9 is operated, the pilot valve 1574 is closed and the pilot flow does not flow into the pilot circuits 12 to 14 and 71 to 73. Therefore, the pressure oil does not flow through the variable throttles 23 of the main valves 11 and 70, and the control pressure Pc of the back pressure chamber 24 is the same as the pressure of the inlet port 17 (the discharge pressure of the hydraulic pump 1) Ps. Further, due to the above-described operation of the load sensing type pump regulator 10, the discharge pressure Ps of the hydraulic pump 1 is changed to the hydraulic actuator.
The pressure is kept higher than the maximum load pressure Plmax of 6 and 7 by a pressure corresponding to the set value of the spring 65. Therefore, the valve element 21
Are in a relationship of Ac = As + Al and Ps> Pl, the valve element 21 is urged in the valve closing direction by the control pressure Pc, and the main valves 11, 70 are held in the closed position. The pressure compensating valves 16 and 17 are held in the open position by setting the pressure receiving areas as, al, am and az.

Next, when the operation lever 30 of the flow rate control valve 8 is operated independently, the pilot valve 15 opens according to the operation amount, a pilot flow is formed in the pilot circuits 12 to 14, and the opening degree of the pilot valve 15 increases. A corresponding pilot flow rate flows. As a result, as described above, due to the action of the variable throttle 23 and the back pressure chamber 24, the main valve valve body 21 opens to an opening degree proportional to the pilot flow rate, and the operation amount of the operation lever 30 (opening degree of the pilot valve 15) increases. A corresponding flow rate flows out from the inlet port 17 to the outlet port 18 through the main valve 11.

Then, in such a state that the pilot valve 15 is opened by a certain amount and a certain amount of main flow rate is flowing from the inlet port 17 to the outlet port 18, for example, the pressure of the outlet port 18 is increased and the inlet port 17 and the outlet port 18 are increased. When the differential pressure of 18 is about to decrease, the discharge pressure of the hydraulic pump 1 is increased by the action of the load sensing type pump regulator 10.
The differential pressure between the pressure at the inlet port 17 (discharge pressure of the hydraulic pump 1) and the pressure at the outlet port 18 (load pressure of the hydraulic actuator 6; maximum load pressure) is kept constant. Therefore the main valve
A constant flow rate corresponding to the operation amount of the operation lever 30 continues to flow through 11.

Further, in such a single operation of the hydraulic actuator 6, the pressure receiving areas as, al, am, az of the pressure compensating valve 16 are set to the proportional constant γ concerning the self-pressure compensating characteristic in the equation (1).
If is set to any value other than zero,
The differential pressure ΔPz across the pilot valve 15 is controlled according to the change in the load pressure (self-load pressure) of the hydraulic actuator 6,
Self-load pressure compensation is provided.

For example, in the example of the hydraulic excavator described with reference to FIGS. 3 to 7, in the flow control valve related to the swing motor 87, the proportional constant γ is a relatively small negative value as shown in FIG. 7 (A). It is set to a value. Therefore, when the revolving structure 81 is driven, since the revolving structure is an inertial body, the load pressure increases, and the pressure rises above the pressure of the relief valve provided for circuit protection, resulting in a waste of energy. Is set to a negative value, the differential pressure ΔPz is controlled to decrease as the turning load pressure increases, and the flow rate through the flow rate control valve decreases. For this reason, even if the load pressure rises, the excess flow rate is less discarded than the relief valve,
The waste of energy can be reduced.

Further, in the flow rate control valve on the bottom side of the bucket cylinder 90, the proportionality constant γ is set to a relatively small positive value as shown in FIG. 7 (C). Therefore, the differential pressure ΔPz increases as the self-load pressure increases during excavation work, and the flow rate through the flow rate control valve increases. That is, the excavation speed of the bucket increases. As a result, a strong feeling of excavation operation can be obtained and workability is improved.

Next, when both the operation levers 30 of the flow control valves 11 and 70 are operated, the following operation is performed. First, the flow control valve 11,7
As in the case where the flow control valve 11 is operated independently in both of 0, the pilot flow according to the operation amount flows, and the variable throttle
By the action of 23 and the back pressure chamber 24, the flow rate according to the operation amount of the operation lever 30 (opening degree of the pilot valves 15, 74) flows out from the inlet port 17 to the outlet port 18 through the main valves 11, 70.

In such combined operation of the hydraulic actuators 6 and 7, the pressure receiving areas as, al, am and az of the pressure compensating valves 16 and 17 are as follows.
The proportional constant α in the first term on the right side of the equation (1) is the fifth
As shown in the figure, the pressure compensation and the diversion function are fulfilled by setting the value to be an arbitrary positive value.

Therefore, for example, in the hydraulic excavator described with reference to FIGS. 3 to 7, in the operating state in which the load sensing type pump regulator 10 is effectively functioning, each working member is set to the operation amount of the operation lever. It is possible to carry out stable composite operation by driving at a constant constant flow rate. Further, when the total required flow rate of the hydraulic actuators 6 and 7 becomes larger than the maximum discharge flow rate of the hydraulic pump 1 and the pump regulator 10 is in an operating state in which it does not function effectively, only the low pressure side hydraulic actuator is operated. Instead of supplying the pressure oil, the pressure oil can be surely supplied to the high pressure side hydraulic actuator, and all the working members can be surely driven. In particular, when α ≦ K is set, the flow supplied to the hydraulic actuator does not fluctuate even when the operation is switched from the combined operation to the single operation, and the work can be stably continued.

If α ≦ K is set, it is possible to supply a flow rate accurately proportionally distributed to each hydraulic actuator in accordance with the operation amount of the operation lever. As a result, particularly when the pressure receiving areas as, al, am, az of the pressure compensating valve 16 are set such that the proportional constants β and γ in the equation (1) are set to zero, the movement locus of the working member is set to the operation lever. It can be controlled accurately according to the operation amount of. For example, in the flow rate control valve related to the rod side of the boom cylinder 88 and the flow rate control valve related to the rod side of the arm cylinder 89, as shown in FIGS. 6 (A) and 7 (B), β = It is set to 0, γ = 0. This completely eliminates the effects of the load pressure of other hydraulic actuators and the self-load pressure during the work of forming the downward slope using the boom and arm, and reduces the amount of operation of the boom operation lever and arm operation lever. Accordingly, the flow rates supplied to the boom cylinder 88 and the arm cylinder 89 can be precisely proportionally distributed, and accurate slope formation can be performed.

When the pressure receiving areas as, al, am, az of the pressure compensating valve 16 are set so that the proportional constant β and / or the proportional constant γ in the equation (1) is any value other than zero, On the basis of the pressure compensation and diversion functions described above, a harmonization function and / or a self-load pressure compensation of varying the main flow rate through the main valves 11, 70 depending on the maximum load pressure Plmax of another hydraulic actuator is performed.

For example, in the example of the hydraulic excavator described with reference to FIGS. 3 to 7, in the flow control valve related to the swing motor 87, the proportional constant β is β = 0 as shown in FIG. 6 (A). In the flow rate control valve related to the bottom side of the boom cylinder 88, the proportional constant β is set to an arbitrary positive value as shown in FIG. 6 (B). Generally, when the swing and the boom raising are simultaneously operated, the swing body 81 is an inertial body, so that the load pressure of the swing motor becomes high at the beginning of the swing. However, when the turn reaches the maximum speed, the load pressure decreases. On the other hand, since the load pressure of the boom is the boom holding pressure, the load pressure is lower than the load pressure of the turning at the beginning of turning. In addition, for example, when performing turning and boom raising in digging work with a backhoe shovel, even if the operator simultaneously operates the turning and boom raising operation levers to full stroke for the sake of simplicity of operation, the boom rising speed is initially It is preferable that the boom and the turning speed can be automatically adjusted so that the speed rises faster than the turning speed and the turning speed gradually increases when the boom rises to some extent. By setting the proportionality constant β as described above, in the boom-side flow control valve, when the load pressure at the start of turning is high and the differential pressure Plmax−Pl is large, the pressure difference ΔPz across the pilot valve increases. The flow rate supplied to the boom cylinder increases, and the differential pressure ΔPz gradually decreases as the differential pressure Plmax−Pl decreases. Therefore, the boom and the turning speed can be automatically adjusted, and the burden on the operator can be reduced.

Further, in the flow rate control valve on the bottom side of the arm cylinder 89, the proportionality constant β is set to a relatively small positive value as shown in FIG. 6 (C). When performing excavation work by a combined operation using an arm, each hydraulic actuator must move without fail, but at this time, a large amount of pressure oil tends to flow to the actuator on the low pressure side. For this reason, the pressure oil is throttled when flowing through the flow control valve, resulting in a large energy loss. Therefore, the fuel efficiency is also deteriorated, and the heat balance of the pressure oil is also deteriorated. As described above, by setting the proportional constant β within the range where the balance of the composite operation is not hindered,
In the flow control valve on the arm side, the opening degree of the main valve increases in accordance with the increase of the differential pressure Plmax-Pl, and the degree of restricting the oil decreases. Thereby, deterioration of fuel efficiency and heat balance can be reduced.

Further, in the flow rate control valve on the bottom side of the bucket cylinder 90, the proportionality constant β is set to a relatively small negative value as shown in FIG. 6 (D). For example, when excavating a ditch while restricting the movement of the bucket with the maximum pressure of the boom by the combined operation of the boom and the bucket, the moment the bucket comes to the ground, the load on the bucket suddenly decreases and a shock occurs. . By setting the proportional constant β as described above, the differential pressure ΔPz acts as a negative element in accordance with the increase of the differential pressure Plmax-Pl, reducing the pilot flow rate and decelerating the bucket speed. As a result, the occurrence of a shock when the load suddenly decreases is reduced, and work safety and work feeling are improved.

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

As described above, in the hydraulic drive system of the present embodiment, the pressure receiving area of the pressure compensation valve is appropriately set, and the constant α is set to the flow control valve 8,
By setting positive values in 9 and at least one of the constants β and γ to different values in the flow control valve 8 and the flow control valve 9, the type and work mode of the work members of the hydraulic construction machine are set. The characteristics of the flow control valve can be set so as to exert the diversion function, or the harmonization function and / or the self-pressure compensation function based on the diversion function.

Further, in the hydraulic drive system according to the present embodiment, the pressure compensating valve as the auxiliary valve is arranged not in the main circuit but in the pilot circuit, and the main valve arranged in the main circuit is configured as a seat valve. Therefore, it is possible to provide a hydraulic circuit suitable for increasing the pressure with little liquid leakage. Further, since the pressure compensating valve, that is, the compensating valve is arranged in the pilot circuit, even if a large flow rate is caused to flow in the main circuit, there is little throttling loss in the auxiliary valve section, which is economically excellent.

Other Embodiment 1 Next, another embodiment of the present invention will be described with reference to FIGS. 8 and 9. In these drawings, the same members as those in the embodiment shown in FIG. 1 are designated by the same reference numerals.

In the above-described embodiment, since the pressure compensation valves 16 and 75 are controlled, the discharge pressure Ps of the hydraulic pump 1, the maximum load pressure Plmax,
The inlet pressure Pz and the outlet pressure Pl of the pilot valves 15 and 74 are directly used. However, these four pressures are
The 24 control pressures have a correlation with each other as a medium, and the pressure compensation valve can be controlled without directly using these four pressures, and the above-mentioned characteristics can be given to the pressure compensation valve. FIG. 8 and FIG. 9 show an embodiment in which the above four pressures are not directly used for controlling the pressure compensating valve from such a viewpoint. Further, in FIG. 1, the flow control valves 8 and 9 arranged in the meter-in (inlet side) circuit when the hydraulic actuators 6 and 7 operate so as to extend or rotate in one direction.
Although only shown, the flow control valves 8 and 9 function as a part of the directional control valve in the actual circuit, and in order to clarify that point, the entire structure of the directional control valve is shown in FIG. Shows.

That is, in FIG. 8, the hydraulic pump 1 and the hydraulic cylinder 6,
Direction control valves 100, 101 for controlling the drive of the hydraulic cylinders 6, 7 are arranged between the seven, and the direction control valve 100 is composed of four seat valve type flow rate control valves 102, 103, 104, 105. The first flow control valve 102 is connected to the meter-in (inlet side) circuit 106 when the hydraulic cylinder 6 operates so as to extend, and corresponds to the flow control valve 8 of the embodiment shown in FIG. The second flow control valve 103 is connected to the meter-in circuit 107 when the hydraulic cylinder 6 operates so as to contract, and the third flow control valve 104 includes the hydraulic cylinder 6 and the second flow control valve 104.
Connected to the meter-out (outlet side) circuit 108 when operating to extend the hydraulic cylinder 6 between the flow control valves 103 of the first and second flow control valves 103.
Is connected to the meter-out circuit 109 when the hydraulic cylinder 6 operates to contract between the flow control valves 102 of FIG. A check valve 110 for preventing backflow of pressure oil to the first flow control valve is connected between the first flow control valve 102 and the fourth flow control valve 105, and is connected to the second flow control valve. 103 and the third
A check valve 111 for preventing backflow of pressure oil to the second flow control valve is connected between the check valve 111 and the flow control valve 104.

Each of the first to fourth flow control valves 102 to 105 has a seat valve type main valve 112, 113, 114, 115, a pilot circuit 116, 117, 118, 119 for the main valve, and pilot valves 120, 121, 122, 123 connected to the pilot circuit. The second flow control valves 102 and 103 further include pressure compensation valves 124 and 125 connected to the pilot circuit in series with the pilot valves. The structures and functions of the main valves 112 to 115 are similar to those of the main valves 11 and 70 of the embodiment shown in FIG. That is, when the pilot valves 120 to 123 are operated, a pilot flow according to the opening of the pilot valves is formed in the pilot circuits 116 to 119,
Due to the action of the variable throttle 23 and the back pressure chamber 24, the main valve valve body 21 opens to an opening proportional to the pilot flow rate, and the pilot valves 120 to 1
The flow rate according to the opening degree of 23 flows out from the inlet port 17 to the outlet port 18 through the main valve 11.

The pilot valves 120 to 123 are basically the same as the pilot valves 15 and 74 shown in FIG. 1, except that the pilot valves 120 to 123 have a hydraulic operating portion 126 as shown in FIG.

As shown in FIG. 9, the pressure compensating valve 124 includes a spool type valve element 130, a first hydraulic operating chamber 131 for biasing the valve element 130 in a valve opening direction, and a first hydraulic operating chamber 131. Located opposite each other,
It is composed of second, third and fourth hydraulic operating chambers 132, 133 and 134 for urging the valve body 125 in the valve closing direction. The first hydraulic operating chamber 131 is connected to the back pressure chamber 24 of the main body 112 via the pilot line 135, and the second hydraulic operating chamber 132 is connected to the pressure compensating valve 124.
Is formed at a position communicating with the outlet port 41 of the third hydraulic operation chamber 133, and the maximum load pressure line via the pilot line 136.
The fourth hydraulic operating chamber 134 is connected to the pilot line 1
It is connected to the main circuit 106 on the inlet port 17 side of the main valve 112 via 37. As a result, the control pressure Pc of the back pressure chamber 24 is introduced into the first hydraulic operating chamber 131, the inlet pressure Pz of the pilot valve 120 is introduced into the second hydraulic operating chamber 132, and the third hydraulic operating chamber 133. The maximum load pressure Plmax is introduced into, and the discharge pressure Ps of the hydraulic pump 1 is introduced into the fourth hydraulic operation chamber 134. The end surface of the valve body 130 facing the first hydraulic operation chamber 131 defines a pressure receiving area ac for receiving the control pressure Pc of the back pressure chamber 24, and the annular end surface of the valve body 130 facing the second hydraulic operation chamber 132. Defines a pressure receiving area az for receiving the inlet pressure Pz of the pilot valve 120, and the annular end surface of the valve body 130 facing the third hydraulic operating chamber 133 defines a pressure receiving area am for receiving the maximum load pressure Plmax, The end surface of the valve element 130 facing the hydraulic operation chamber 134 defines a pressure receiving area as for receiving the discharge pressure Ps of the hydraulic pump 1.
These pressure receiving areas as, ac, am, az are set so that predetermined values of proportional constants α, β, γ described later can be obtained. Further, the pressure receiving areas as, ac, am, az are set so as to hold the valve element 130 in the open position when the main valve 112 and the pilot valve 120 are closed.

The pressure compensation valve 125 has the same structure as the pressure compensation valve 124.

The directional control valve 101 for the hydraulic cylinder 7 is also configured similarly to the directional control valve 100.

The hydraulic pump 1 is provided with a load-sensing type pump regulator 140 that keeps the discharge pressure of the hydraulic pump 1 higher than the maximum load pressure of the hydraulic actuators 6, 7 by a predetermined value.

The pump regulator 140 has a hydraulic cylinder type swash plate tilting device 141 and a control valve 142. The swash plate tilting device 141 divides the rod side cylinder chamber and the head side cylinder chamber according to the position of the control valve 142. Driven by the area difference, the discharge amount of the hydraulic pump 1 is controlled. The drive system of the control valve 142 is the same as that of the control valve 62 shown in FIG. That is, the control valve 142 opposes the discharge pressure of the hydraulic pump 1, the maximum load pressure, and the set force of the spring 65, controls the swash plate tilting device 141 according to the change of the maximum load pressure, and controls the hydraulic pump. The discharge pressure of 1 is maintained at a pressure higher than the maximum load pressure by a pressure corresponding to the strength of the spring 65.

In the hydraulic drive system configured as described above, for example, the pressure balance equation of the valve element 130 in the pressure compensation valve 124 is expressed by the following equation.

acPc = asPs + amPlmax + azPz Further, the pressure balance equation of the valve body 21 in the main valve 102 is expressed by the following equation.

AcPc = AsPs + AlPl From the above two equations, derive the equation to obtain the differential pressure across the pilot valve 120, If this equation is transformed using Ac = As + Al, , Pz-Pl = α (Ps-Plmax) + β (Plmax-Pl) + γPl
It can be expressed as (4). Where the differential pressure across pilot valve 120 is ΔP
If z, then Pz−Pl = ΔPz. Therefore, the same equation as equation (1) in the embodiment shown in FIG. 1 is obtained.

Therefore, also in the present embodiment, by setting the proportional constants α, β, γ to predetermined values, the differential pressure ΔPz across the pilot valve 120 is set to the difference between the discharge pressure Ps of the hydraulic pump 1 and the maximum load pressure Plmax. Pressure Ps-Plmax, differential pressure Plmax-Pl between maximum load pressure Plmax and self-load pressure Pl, and proportional control to the three elements of self-load pressure Pl can be performed in proportion to each of them. Item), or a harmony function (the second item on the right side) and / or a self-pressure compensation function (the third item on the right side) in a combined operation based on the pressure compensation and diversion function.
Can be obtained. That is, in this embodiment, instead of directly using the inlet pressure Pz and the outlet pressure Pl of the pilot valve 120, the discharge pressure Ps of the hydraulic pump 1, and the maximum load pressure Plmax,
The control pressure Pc, the inlet pressure Pz, the maximum load pressure Plmax, and the discharge pressure Ps of the hydraulic pump 1 are introduced to introduce the above four pressures Pz, Pl, P.
The same effect as using s and Plmax is obtained.

FIG. 10 shows an example of changing the arrangement order of the hydraulic operation chambers of the pressure compensating valve shown in FIG. That is, in this embodiment, the pressure compensating valve 150 positions the first hydraulic operating chamber 151, which receives the control pressure Pc of the back pressure chamber 24, adjacent to the back pressure chamber 24 and omits the pilot line 135. , First hydraulic operating room 1
The three hydraulic operating chambers facing 51 are a hydraulic operating chamber 152 that receives the inlet pressure Pz of the pilot valve 120, a hydraulic operating chamber 153 that receives the discharge pressure Ps of the hydraulic pump 1, and a maximum load pressure Plmax.
The hydraulic operation chamber 154 is arranged in this order. Even if the hydraulic operation chamber is arranged in this way, the above formula (4) is established and the same effect as that of the embodiment shown in FIG. 9 can be obtained.

FIG. 11 shows a modification of the structure of the seat type main valve. In this embodiment, the seat-type main valve 160 has a through hole 161 that allows the inlet port 17 to communicate with the back pressure chamber 24 instead of the valve body having the slit 22 as the variable throttle as in the above-described embodiments. The through hole 161 is configured as a variable throttle that changes the throttle amount according to the movement of the valve body 162. In addition, in the above-described embodiment, the inlet port
Although it is configured such that the drilling direction of 17 and the moving direction of the valve body 21 are orthogonal, and the drilling direction of the outlet port 18 and the moving direction of the valve body 21 are the same, in the present embodiment, Port 17
In order that the drilling direction of the valve body 162 and the moving direction of the valve body 162 may coincide with each other, the drilling direction of the outlet port 18 and the moving direction of the valve body 162 may be orthogonal to each other.

In this embodiment, the bottom end surface of the valve element 162 defines the pressure receiving area As that receives the pump discharge pressure Ps. Also,
The flow direction of the pressure oil flowing from the inlet port 17 to the outlet port 18 is opposite to that in the above-described embodiment.

Also in this embodiment, the main valve 160 functions similarly to the main valve 11 of the above-described embodiment, and the action of the variable throttle provided by the through hole 161 and the back pressure chamber 24 causes the main flow rate according to the pilot flow rate. Can be drained. As a result, the pressure compensation valve
124 functions similarly to the embodiment of FIG. 9 and can obtain the same effect.

Still another embodiment of the present invention will be described with reference to FIGS. 12 and 13. In the figure, the same members as those shown in FIGS. 2 and 12 are designated by the same reference numerals.

In this embodiment, the directional control valves are represented by 170 and 171, and these directional control valves 170 and 171 are the same as the embodiment shown in FIG. 8 except for the configuration relating to the pressure compensating valves 172 and 173.

First, the pressure compensating valve 172 (173) is connected to the pilot circuit 116 (1
The arrangement position in 17) is different from that in the above-described embodiment. That is, the pressure compensation valve 172 (173) is the pilot valve 120.
It is connected to the pilot circuit 116 (117) between the outlet side of (121) and the outlet port 18 of the main valve 102 (103). There is also a difference in the pressure introduced to control the pressure compensation valve. That is, the pressure compensating valve 172 (173) includes a spool type valve body 174, a first hydraulic pressure operation chamber 175 that biases the valve body 174 in the valve opening direction, and a first hydraulic pressure operation chamber 175 that faces the first hydraulic pressure operation chamber 175. Located, valve body
Second and third hydraulic operating chambers 17 for urging 174 in the valve closing direction
6 and 177. The first hydraulic operating chamber 175 is formed in communication with the inlet port 178 of the pressure compensation valve, and the second hydraulic operating chamber 176 is connected to the main valve 102 (103) via the pilot line 179.
Is connected to the outlet port 18 and the third hydraulic operating chamber 177 is connected to the maximum load pressure line 50 via the pilot line 180. As a result, the outlet pressure Pz of the pilot valve 120 (121) is introduced into the first hydraulic operation chamber 175, and the outlet pressure (load pressure) P of the main valve 102 (103) is introduced into the second hydraulic operation chamber 176.
l is introduced, and the maximum load pressure Pl is applied to the third hydraulic operating room 177.
max is introduced. And the first hydraulic operating chamber 17 of the valve body 174
The end surface facing 5 defines a pressure receiving area az for receiving the pilot valve outlet pressure Pz, and the valve body 17 facing the second hydraulic operating chamber 176
The annular end surface of 4 defines a pressure receiving area a1 that receives the main valve outlet pressure Pl, and the end surface of the valve body 174 that faces the third hydraulic operating chamber 177 defines a pressure receiving area am that receives the maximum load pressure Plmax.
These pressure receiving areas az, al, am are proportional constants α,
It is set so that predetermined values of β and γ can be obtained. The pressure receiving areas az, al, am are the main valve 102 (103) and the pilot valve 1
The valve element 174 is set to be held in the open position when the 20 (121) is closed.

In the hydraulic drive system configured as described above, the pressure balance equation of the valve element 174 of the pressure compensation valve 172 (173) is represented by the following equation.

azPz = amPlmax + alPl Further, the pressure balance equation of the valve body 21 in the main valve 102 is expressed by the following equation.

AcPc = AsPs + AlPl From the above two equations, derive the equation to obtain the differential pressure across the pilot valve 120, If this equation is transformed using Ac = As + Al, , Pc-Pz = α (Ps-Plmax) + β (Plmax-Pl) + γPl
It can be expressed as (5). Where the differential pressure across pilot valve 120 is ΔP
If z, then Pc−Pz = ΔPz. Therefore, the same formula as the formula (1) in the above-described embodiment can be obtained.

Therefore, also in this embodiment, by setting the proportional constants α, β, γ to predetermined values, the pilot valve 120 (12
The differential pressure Pz across 1) is the differential pressure Ps-Plmax between the discharge pressure Ps of the hydraulic pump 1 and the maximum load pressure Plmax, and the maximum load pressure Plmax.
Pressure difference between the self-load pressure Pl and the self-load pressure Pl
Can be controlled in proportion to each of the three elements, and the above-mentioned pressure compensation and diversion function (first term on the right side), or harmony function in complex operation based on this pressure compensation and diversion function (second term on the right side) And / or the self-pressure compensation function (the third term on the right side) can be obtained.

Further, in the above equation (5), the self-load pressure Pl on the right side is
From the relationship of AcPc = AsPs + AlPl described above, pilot valve 120
Inlet pressure Pc (= control pressure) and hydraulic pump discharge pressure Ps
Can be expressed by and, after all, formula (5) is the pilot valve
It can be represented by four pressures, that is, the inlet pressure Pc and the outlet pressure Pz of 120, the discharge pressure Ps of the hydraulic pump 1, and the maximum load pressure Plmax. Therefore, also in this embodiment, the inlet pressure Pc and the outlet pressure Pz of the pilot valve 120, the discharge pressure P of the hydraulic pump 1
s, outlet pressure instead of directly using maximum load pressure Plmax
The same effect as that obtained by introducing three pressures of Pz, the main valve outlet pressure Pl, and the maximum load pressure Pmax and using the above four pressures Pc, Pz, Ps, Plmax is obtained.

As described above, the present invention controls the pressure compensating valve based on the four pressures of the inlet pressure and the outlet pressure of the pilot valve, the discharge pressure of the hydraulic pump 1, and the maximum load pressure, and the pressure compensating and diversion function, or A harmony function and / or a self-pressure compensation function based on the pressure compensation and the diversion function can be selectively achieved. The four pressures are the back pressure chamber 24
Since the control pressure Pc of 1 is used as a medium and has a correlation with each other, the pressure compensating valve can be controlled without directly using these four pressures. The position of the pressure compensation valve may be either before or after the pilot valve. In the following, further modifications of this point will be listed. In the following description, the main valve is represented by reference numeral 11, and the pilot valve is represented by reference numeral 15.

FIG. 14 shows the pressure compensating valve 190 including the back pressure chamber 24 and the pilot valve 1
5 is arranged in the pilot circuit between 5 and the control pressure Pc of the back pressure chamber and the pilot valve outlet pressure Pl are set in the hydraulic operating chamber in the valve opening direction having the pressure receiving area ac, al, and the pilot valve inlet pressure Pz,
This is an embodiment in which the maximum load pressure Plmax is guided to the hydraulic operating chamber in the valve closing direction having the pressure receiving areas az and am.

The pressure balance equation of the pressure compensation valve 190 in this configuration is represented by the following equation.

acPc + alPl = amPlmax + azPz From this formula and the pressure balance formula of the main valve 11, the formula for obtaining the differential pressure across the pilot valve 15 is derived as in the above-described embodiment. Therefore, if the right-hand side constants are set as α, β, and γ, respectively, Pz−Pl = α (Ps−Plmax) + β (Plmax−Pl) + γPl
(6) is required.

FIG. 15 shows that the pressure compensating valve 191 includes a pilot valve 15 and a main valve 11.
It is arranged between the outlet ports of the hydraulic pump 1 and the discharge pressure P of the hydraulic pump 1.
s, pilot valve outlet pressure Pz to the hydraulic operating chamber in the valve opening direction with pressure receiving areas as and az, and pilot valve inlet pressure Pc and maximum load pressure Plmax to the hydraulic operating chamber in the valve closing direction with pressure receiving areas ac and am It is a guided example.

The pressure balance equation of the pressure compensating valve 191 in this configuration is: azPz + asPs = acPc + amPlmax The equation of the differential pressure across the pilot valve 15 is: If the right-hand side constants are set as α, β, γ, respectively: Pc-Pz = α (Ps-Plmax) + β (Plmax-Pl) + γPl
(7) In FIG. 16, the pressure compensating valve 192 includes a pilot valve 15 and a main valve 11.
It is arranged between the outlet ports of the hydraulic pump 1 and the discharge pressure P of the hydraulic pump 1.
s, the pilot valve outlet pressure Pz is introduced into the hydraulic operating chamber in the valve opening direction having the pressure receiving areas as and az, and the maximum load pressure Plmax is introduced into the hydraulic operating chamber in the valve closing direction having the pressure receiving area am.

The pressure balance equation of the pressure compensation valve 192 in this configuration is: azPz + asPs = amPlmax The equation of the differential pressure across the pilot valve 15 is: If the right-hand side constants are set as α, β, γ, respectively: Pc-Pz = α (Ps-Plmax) + β (Plmax-Pl) + γPl
(8) In FIG. 17, the pressure compensating valve 193 includes a pilot valve 15 and a main valve 11.
It is arranged between the outlet ports of the hydraulic pump 1 and the discharge pressure P of the hydraulic pump 1.
s, pilot valve inlet pressure Pc, pilot valve outlet pressure P
z in the hydraulic operating room in the valve opening direction with pressure receiving areas as, ac, az,
This is an embodiment in which the maximum load pressure Plmax is introduced to the hydraulic operating chamber in the valve closing direction having the pressure receiving area am.

The pressure balance equation of the pressure compensating valve 193 in this configuration is: azPz + acPc + asPs = amPlmax The equation of the differential pressure across the pilot valve 15 is: If the right-hand side constants are set as α, β, γ, respectively: Pc-Pz = α (Ps-Plmax) + β (Plmax-Pl) + γPl
(9) In FIG. 18, the pressure compensating valve 194 includes a pilot valve 15 and a main valve 11.
The pilot valve outlet pressure Pz in the opening direction hydraulic operation chamber having the pressure receiving area as, the pilot valve inlet pressure Pc, the main valve 11 outlet pressure Pl, and the maximum load pressure Plmax. This is an embodiment in which the hydraulic pressure operation chamber having a pressure receiving area ac, al, am in the valve closing direction is introduced.

The pressure balance equation of the pressure compensating valve 194 in this configuration is: azPz = acPc + alPl + amPlmax The equation of the differential pressure across the pilot valve 15 is: If the right-hand side constants are set as α, β, and γ, respectively: Pc−Pz = α (Ps−Plmax) + β (Plmax−Pl) + γPl (1
0) FIG. 19 shows the pressure compensating valve 195 including the pilot valve 15 and the main valve 11
Located between the outlet ports of the pilot valve 15 inlet pressure
Pc, pilot valve outlet pressure Pz in the valve-opening direction hydraulic operation chamber with pressure-receiving areas ac, as, main valve 11 outlet pressure Pl, maximum load pressure Plmax in the valve-closing direction hydraulic operation with pressure-receiving areas al, am It is the example which led to the room.

The pressure balance equation of the pressure compensating valve 195 in this configuration is: azPz + acPc + alPl + amPlmax The equation of the differential pressure across the pilot valve 15 is: If the right-hand side constants are set as α, β, and γ, respectively: Pc−Pz = α (Ps−Plmax) + β (Plmax−Pl) + γPl (1
1) In Fig. 20, the pressure compensating valve 196 has a pilot valve 15 and a main valve 11
Between the outlet ports of the pilot valve outlet pressure Pz,
The discharge pressure Ps of the hydraulic pump 1 and the outlet pressure Pl of the main valve 11 are operated in the valve-opening direction hydraulic operation chamber having the pressure receiving areas az, as, al, and the maximum load pressure Plmax is operated in the valve closing direction having the pressure receiving area am. It is the example which led to the room.

The pressure balance equation of the pressure compensating valve 196 in this configuration is: azPz + asPs + alPl = amPlmax The equation for the differential pressure across the pilot valve 15 is: If the right-hand side constants are set as α, β, and γ, respectively: Pc−Pz = α (Ps−Plmax) + β (Plmax−Pl) + γPl (1
2) Other Embodiment 2 Still another embodiment of the present invention will be described with reference to FIGS. 21 to 23. In the figure, the same members as those in the embodiment shown in FIG. 1 are designated by the same reference numerals.

In the above embodiment, as the control means of the pressure compensating valve, hydraulic means for directly or indirectly introducing the discharge pressure of the hydraulic pump, the maximum load pressure, the inlet pressure and the outlet pressure of the pilot valve into the plurality of hydraulic operating chambers. It is an example configured with
The control means can be electrically configured. 21 to 23 show such an embodiment.

That is, in FIG. 21, flow rate control valves for controlling the hydraulic actuators 6 and 7 are represented by reference numerals 200 and 201, respectively.
The flow control valves 200 and 201 have pressure compensation valves 202 and 203, which are electromagnetic proportional valves 202 and 203 having electromagnetic operating portions 202A and 203A, respectively, and other configurations are flow control valves 8 and 9 in the embodiment of FIG.
Is the same as A pressure detector 204 for detecting the discharge pressure Ps of the hydraulic pump 1 is connected to the discharge line of the hydraulic pump 1 connected to the main lines 2 and 3, and the pilot valve 15 and 74 of the pilot valve 15 and 74 are connected to the lines 13 and 72 of the pilot circuit. Pressure detector 205,2 to detect inlet pressure Pz
06 is connected, and pilot valves 1 and 2 are connected to pilot lines 14 and 73.
Pressure detectors 207, 208 for detecting the outlet pressure Pl of 5,74 are connected, and hydraulic actuators 6, 7 are connected to the maximum load pressure line 50.
A pressure detector 209 for detecting the maximum load pressure Plmax of is connected. Further, the hydraulic pump 1 is provided with an angle meter 210 for detecting the tilt angle of a displacement volume varying mechanism such as a swash plate. The discharge flow rate of the hydraulic pump 1 is controlled by the discharge amount control device 212 driven by the pressure oil from the auxiliary pump 211.

Pressure signals from pressure detectors 204 to 209 Ps, Pz1, Pz2, Pl1, P
The inversion angle signal Qr from l2, Plmax and the angle meter 210 is input to the control unit 213, and the control unit 213 controls the control signal Q ₀ of the hydraulic pump 1 and the pressure compensation valve based on these input signals.
Control signals I1 ₀ and I2 ₀ of 202 and 203 are calculated, and these signals are output to the discharge amount control device 212 and the electromagnetic operation unit 202A and 203A of the pressure compensation valve.
Output to.

The control unit 213 is composed of a microcomputer, and as shown in FIG. 22, an A / D converter 214 for converting the pressure signal and the inversion angle signal into a digital signal, and a central processing unit.
215 and a memory 216 storing a program of control procedure
, D / A converter 217 for output, interface 218 for output, amplifiers 219, 220 connected to electromagnetic operating units 202A, 203A of the pressure compensation valve, and two input terminals 212A of the discharge amount control device 212. And amplifiers 221, 222 connected to 212B.

The control unit 213 uses the pressure signal Ps of the pressure detector 204 that detects the discharge pressure of the hydraulic pump 1 and the pressure signal Plmax of the pressure detector 209 that detects the maximum load pressure of the hydraulic actuators 6 and 7 to determine the memory 216. The discharge amount target value Q _{0 of} the hydraulic pump 1 for increasing the pump discharge pressure by a predetermined value from the maximum load pressure based on the control procedure program stored in
Is calculated and the target value signal Q ₀ is calculated by the I / O interface 218
From the amplifiers 221, 222 via the input terminal of the discharge amount control device 212
Output to 212A and 212B. As a result, the discharge amount control device 212 controls the tilt angle of the swash plate of the hydraulic pump 1 so that the tilt angle Qr detected by the angle meter 210 becomes equal to the target value Q ₀ . As a result, the pump discharge pressure is kept higher than the maximum load pressure by a predetermined value, and the same function as the hydraulic load-sensing type pump regulator of the above-described embodiment is achieved.

Further, the control unit 213 uses the pressure signals Ps, Pz1, Pz2, Pl1, Pl2, Plmax of the pressure detectors 204 to 209 to detect the pressure compensation valves 202, 203.
The control amount of is calculated and the pressure compensation valve is controlled. FIG. 23 is a flow chart showing the control contents of such a pressure compensation valve. In step 230, the pressure signals Ps, Pz1, Pz2, Pl1, Pl2, Plmax detected by the pressure detectors 204 to 209 are read. Next, in step 231, the target inlet pressures Pz10 and Pz20 of the pilot valves 15 and 74 are calculated by the following equation.

Pz1 ₀ = α (Ps-Plmax) + β (Plmax-Pl1) + γPl1 + Pl1 Pz2 ₀ = α (Ps-Plmax) + β (Plmax-Pl2) + γPl2 + Pl2 This formula is the same as the formula (1) of the first embodiment described above. The constants α, β, and γ are predetermined values as shown in FIGS. 5 to 7, for example, depending on the selection of the three functions of the pressure compensation and diversion function, the harmony function and the self-pressure compensation, and the function. Is set to. Next, in step 232, I10 = G (Pz10−Pz1) I20 = G (Pz20−Pz2) is calculated, and the control 5 signals I1 ₀ and I2 of the pressure compensation valves 202 and 203 are calculated.
_{Get 0} . Finally, in step 233, this calculated control signal I1
₀ and I2 ₀ are output from the amplifiers 219 and 220 to the electromagnetic operation units 202A and 203A of the pressure compensation valves 202 and 203 via the D / A converter 217.

Even in the embodiment in which the pressure compensation valves 202 and 203 are electrically controlled as described above, the above-described (1) shown in the procedure 231 is performed.
By setting an expression equivalent to the expression in the program,
Similar to the embodiment shown in FIG. 1, performing a pressure compensation and diversion function, or a harmony function and / or a self-pressure compensation function based on the pressure compensation and diversion function according to the set values of α, β, γ. You can

In the embodiment for electrically controlling the pressure compensating valve described above, the constants α, β, γ are set as a part of the program, but they are operated from the outside as shown by the phantom line in FIG. Connect the possible regulator 240 to the control unit 213,
The constants α, β, γ may be variably set according to the type of hydraulic construction machine, the type of working member, and the like.

Other Embodiment 3 Next, an embodiment relating to the valve structure of the present invention will be described with reference to FIG. FIG. 24 shows an embodiment in which the seat type main valve and the pressure compensating valve of the flow control valve shown in FIG. 9 are integrated.

That is, in FIG. 24, the flow control valve 270 is the main valve portion 2
71 and a pressure compensation valve portion 272, and the main valve portion 271 has a valve housing 27 having an inlet port 273 and an outlet port 274.
It has a valve element 276 of the seat valve type, which is arranged in the container 5 and controls communication between the inlet port 273 and the outlet port 274. Disc 2
A passage 277 that constitutes a variable throttle is provided in the valve 76, and the valve body 276
A back pressure chamber 278 that communicates with the inlet port 273 via a variable throttle 277 is formed behind the. The pressure compensation valve portion 272 is
A spool type valve element 280 which is arranged in the valve housing 275 and narrows the passage between the back pressure chamber 278 and the pilot outlet auto 279.
This valve body 280 is engaged with a piston 281 which is axially movably inserted into the main valve body 276. The pressure compensating valve portion 272 includes a first hydraulic operating chamber 282 facing the end surface of the valve body 280 on the side opposite to the piston, and a second hydraulic operating chamber 283 facing the first annular end surface of the valve body 280. A third hydraulic operating chamber 284 facing the second annular end face of the valve body 280 and a fourth hydraulic operating chamber 285 formed in the main valve body 276 facing the end face of the piston 281 are provided. . The first hydraulic operating room 282 has a passage 286
Through the back pressure chamber 278, the second hydraulic operating chamber 283 communicates with the pilot outlet port 279, and the third hydraulic operating chamber 284
Communicates with the maximum load pressure port 287, and the fourth hydraulic operating room 2
85 communicates with main valve inlet port 273 via passage 288. Pilot outlet port 279 is connected to pilot valve 290 via pilot line 289, and maximum load pressure port 2
87 is connected to a maximum load pressure line (not shown). Accordingly, the back pressure chamber 27 is provided in each of the first to fourth hydraulic operation chambers.
The control pressure Pc of 8, the inlet pressure Pz of the pilot valve 290, the maximum load pressure Plmax, and the discharge pressure Ps of the hydraulic pump are introduced.
In this way, the first to fourth hydraulic operating chambers 282 to 285 correspond to the first to fourth hydraulic operating chambers 131 of the flow control valve shown in FIG.
Corresponds to ~ 134.

By integrally combining the main valve and the pressure compensation valve in this way, a compact and rational valve structure can be obtained.

Next, another embodiment of the present invention relating to the pump control means will be described. First, in the above embodiments, the hydraulic drive system of the present invention is described in combination with a load-sensing type pump regulator, and the load-sensing type pump regulator controls the discharge pressure of a variable displacement hydraulic pump. However, the hydraulic pump may be a fixed displacement type. In this case, the load sensing regulator type pump regulator is configured as shown in FIG. That is, in FIG. 25, the pump regulator 380
Is a relief valve 383 having opposed pilot chambers 381,382
The discharge pressure of the fixed displacement hydraulic pump 385 is guided to the pilot chamber 381 via the pilot pipe 384, and the maximum load pressure is guided to the pilot chamber 382 via the pilot pipe 386. The spring 387 is arranged on the side of.
As a result, the discharge pressure of the hydraulic pump 385 can be held 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 also be configured in combination with a pump regulator other than the load sensing type. Such an embodiment is shown in FIG. That is,
In FIG. 26, a hydraulic pump 390 is connected to a flow control valve 391 composed of a combination of the above-mentioned main valve, pilot valve and pressure compensating valve, and its discharge amount is adjusted by a pump flow control device 392. Hydraulic pump 390 and flow control valve
Unload valve 393 is connected between 391 and flow control valve 39
An operating device 394 is provided for 1. Operation device 39
The operation signal of 4 is sent to the control device 395, and from there is further sent as a control signal to the pilot valve drive unit 396 of the flow control valve 391 to control the opening degree of the pilot valve. Control device 395
The operation signal sent to is also sent to the arithmetic unit 397, the arithmetic unit 397 calculates the required flow rate of the flow control valve 391 from the map stored in the storage device 398 in advance,
Send a signal to 92. At the same time, the arithmetic unit 397 calculates the set pressure of the unload valve 393 from another map stored in the storage device 398 in advance, and outputs the signal to the unload valve 393. Thus, the discharge pressure of the hydraulic pump 390 is controlled as a function of the operation signal to a pressure obtained from a map stored in the storage device 398 in advance.

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 the above-mentioned equation (1). Therefore, the pressure compensation function cannot be obtained among the functions of the first term on the right side. However, in the combined operation, since the differential pressure is still common to the flow control valves related to the plurality of hydraulic actuators, the flow dividing function can be performed. Since the second and third terms on the right side of the equation (1) are not related to the pump discharge pressure Ps, when β and γ are set to values other than zero, the harmony function based on the branching function and And / or perform a self-pressure compensation function.

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

For example, in the above embodiment, an example in which two hydraulic actuators are driven by a hydraulic pump has been described. However, the present invention is naturally applicable to a case where three or more hydraulic actuators are used. Further, the pump control means may be provided with a simple relief valve for keeping the hydraulic pump discharge pressure constant.

〔The invention's effect〕

According to the present invention, the constant α is set to a positive value in each of the first and second flow control valves, and at least one of the constants β and γ is set in the first flow control valve means and the second flow control. By setting different values for the valve means, depending on the type and working mode of the working member of the hydraulic construction machine, pressure compensation and diversion function, or harmony function based on pressure compensation and diversion function and / or self pressure compensation The characteristics of the flow control valve can be set to exert its function.

Therefore, for example, in an application example to a hydraulic excavator,
Even if the turning and boom raising operation levers are simultaneously operated to full stroke, the boom rising speed increases faster than the turning speed at first, and when the boom rises to a certain degree, the turning speed gradually increases, and the turning speed reaches the maximum speed. When the excavation is carried out by the combined operation using the arm, the arm is driven reliably and the hydraulic pressure for the arm is adjusted. When the actuator is on the low pressure side, the characteristics of the flow control valve that prevent the deterioration of fuel consumption and heat balance, during the trench excavation work by the combined operation using the bucket, the bucket is released from the excavation load, the flow rate control at the moment when it comes to the surface Flow control valve characteristics that reduce the flow rate through the valve to reduce shock, reduce the flow rate from the relief valve during swing acceleration, and reduce energy consumption. Characteristics of flow control valve to reduce waste of consumption, characteristics of flow control valve to obtain strong feeling of excavation operation during excavation work using buckets, accurate slope preparation for slope slope formation work using boom and arm It is possible to obtain the characteristics of the flow rate control valve that performs the formation.

Also, since the auxiliary valve is arranged in the pilot circuit and the seat-type main valve is arranged in the main circuit, it is possible to provide a hydraulic circuit with less liquid leakage, which is suitable for high pressure, and which has less throttling loss and energy saving structure. You can

Further, when the first constant α is set in a relation of α ≦ K with respect to the ratio K of the pressure receiving area receiving the discharge pressure of the hydraulic pump to the pressure receiving area receiving the control pressure of the seat type main valve back pressure chamber, Can accurately divert the flow rate proportional to the operation amount (pilot valve opening) of the operation means in the flow dividing function. Here, when α = K is set, the maximum proportional gain can be given while obtaining the flow dividing function of proportionally distributing the flow rate according to the operation amount.

When the control means is hydraulically constructed and the main valve and the auxiliary valve are integrated, a compact and rational arrangement valve structure can be obtained.

When the control means is electrically configured, the first, second and third constants α, β and γ can be easily set and changed.

[Brief description of drawings]

FIG. 1 is a schematic view showing the overall structure of a hydraulic drive system according to an embodiment of the present invention, FIG. 2 is a sectional view showing the structure of a flow control valve of the hydraulic drive system, and FIG. FIG. 4 is a side view of a hydraulic excavator to which the hydraulic drive system of the invention is applied, FIG. 4 is a top view of the hydraulic excavator, and FIG.
FIG. 6 is a characteristic diagram showing an example of setting the proportional constant α of the pressure compensation valve included in one flow control valve of the hydraulic drive system, and FIGS. 6 (A) to 6 (D) show one example of the hydraulic drive system. FIG. 7 is a characteristic diagram showing a setting example of a proportional constant β of a pressure compensation valve included in the flow control valve, and FIGS. 7A to 7C are pressure compensation valves included in one flow control valve of the hydraulic drive system. FIG. 8 is a characteristic diagram showing a setting example of a proportional constant γ of FIG. 8, and FIG. 8 is a schematic diagram showing an overall configuration of a hydraulic drive system according to another embodiment of the present invention.
FIG. 9 is a sectional view showing the structure of the flow control valve of the hydraulic drive system, FIG. 10 is a sectional view showing a modification of the flow control valve, and FIG. 11 is another flow control valve of the same. FIG. 12 is a cross-sectional view showing a modified example, FIG. 12 is a schematic view showing the overall configuration of a hydraulic drive system according to still another embodiment of the present invention, and FIG. 13 shows the structure of the flow control valve of the hydraulic drive system. 14 to 20 are schematic cross-sectional views each showing a flow control valve of a hydraulic drive system according to still another embodiment of the present invention, and FIG. 21 is still another embodiment of the present invention. FIG. 22 is a schematic diagram showing an overall configuration of a hydraulic drive system according to an example, and FIG. 22 is a schematic diagram showing a configuration of a control unit of the hydraulic drive system,
FIG. 23 is a flow chart showing a control signal producing procedure in the control unit, and FIG. 24 is a sectional view showing an embodiment in which a main valve and a pressure compensating valve of a flow control valve used in the hydraulic drive system of the present invention are integrated. FIG. 25 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 system of the present invention, and FIG. 26 is a hydraulic drive system of the present invention. FIG. 3 is a circuit diagram showing an embodiment of a pump control means that is not a load sensing type used in the above. Explanation of symbols 1; 385; 390 ...... hydraulic pump 2-5 ...... main circuit 6,7; 87-90 ...... first and second hydraulic actuators 8,9; 100,101; 170,171; 200,201 ...... first and second Second flow control valve means 10; 140; 212; 380; 392 ...... Pump control means 11,70; 102,103; 160; 271 ...... Main valve 12-14,71-73; 116,117; 289 ...... Pilot circuit 15 , 74; 120,121; 290 ...... Pilot valve 16,75; 124,125; 150; 172,173; 190-196; 202,203; 242,24
3; 272 …… Auxiliary valve 17; 273 …… Inlet port 21; 162; 276 …… Valve element 18; 274 …… Outlet port 22; 163; 277 …… Variable throttle 24; 278 …… Back pressure chamber 30 Operating means 43-49,51; 131-137; 151-154; 175-180; 202A, 203A, 21
3; 282-286 …… Control means

─────────────────────────────────────────────────── ─── Continuation of the front page (51) Int.Cl. ⁶ Identification code Internal reference number FI Technical display location F15B 11/16 (72) Inventor Yusaku Nozawa 650 Jinmachi-cho, Tsuchiura-shi, Ibaraki Hitachi Construction Machinery Co., Ltd. Tsuchiura Factory (56) References JP-A-60-11706 (JP, A) Special table S58-501781 (JP, A)

Claims (15)

[Claims] 1. At least one hydraulic pump; at least first and second hydraulic actuators connected to the hydraulic pump via a main circuit and driven by pressure oil discharged from the hydraulic pump; First and second flow rate control valve means connected to respective main circuits between the hydraulic pump and the first and second hydraulic actuators; and pump control means for controlling discharge pressure of the hydraulic pump. The first and second flow rate control valve means are respectively connected in series to the first valve means for changing the opening degree according to the operation amount of the operation means, and the valve means. Second valve means for controlling the differential pressure between the inlet pressure and the outlet pressure of the means, the inlet pressure and the outlet pressure of the first valve means, the discharge pressure of the hydraulic pump, and the first and second hydraulic actuators. Maximum negative A hydraulic drive having a control means for controlling the second valve means based on pressure, wherein the first and second flow control valve means are respectively an inlet port and an outlet port connected to the main circuit. For controlling the communication of the valve body, a variable throttle for changing the opening degree according to the displacement of the valve body, and a valve for communicating with the inlet port through the variable throttle to urge the valve body in the valve closing direction. A seat type main valve having a back pressure chamber for generating a control pressure; and a pilot circuit connected between the back pressure chamber of the main valve and an outlet port, wherein the first valve means comprises: A pilot valve connected to a pilot circuit for controlling a pilot flow flowing through the pilot circuit, wherein the second valve means is connected to the pilot circuit and is an auxiliary device for controlling a differential pressure between an inlet pressure and an outlet pressure of the pilot valve. Is a valve, The control means, in each of the first and second flow rate control valve means, the differential pressure between the inlet pressure and the outlet pressure of the pilot valve is ΔPz, the discharge pressure of the hydraulic pump is Ps, the first and second pressure control means. The maximum load pressure of the hydraulic actuator of
Plmax, where Pl is the self-load pressure of the first and second hydraulic actuators, ΔPz = α (Ps−Plmax) + β (Plmax−Pl) + γPl where α, β, γ are the first and second And the third constant, the auxiliary valve is controlled so as to satisfy the relationship, and the first and second flow rate control valve means set the first constant α to a positive value, respectively. A hydraulic drive system, wherein at least one of the second and third constants β and γ is set to different values between the first flow control valve means and the second flow control valve means. 2. The ratio of the pressure receiving area of the main valve valve body receiving the discharge pressure of the hydraulic pump via the inlet port to the pressure receiving area of the main valve valve body receiving the control pressure of the back pressure chamber is K.
2. The hydraulic drive system according to claim 1, wherein in each of the control means of the first and second flow rate control valve means, the first constant α is set to satisfy a relation of α ≦ K. . 3. The hydraulic pressure according to claim 2, wherein one of the control means of the first and second flow rate control valve means sets the second and third constants β and γ to zero. Drive. 4. The control means, in each of the first and second flow rate control valve means, a plurality of hydraulic operation chambers provided in the auxiliary valve, and the hydraulic pumps in the plurality of hydraulic operation chambers. And a pipe line means for directly or indirectly introducing the discharge pressure, the maximum load pressure, the inlet pressure and the outlet pressure of the pilot valve, and the pressure receiving area of the plurality of hydraulic operation chambers is ΔPz = α (Ps -Plmax) + β (Plmax-Pl) + γPl
Is satisfied so that the first constant becomes a positive value in each of the first and second flow rate control valve means, and the first flow rate control valve means and the second flow rate are satisfied. The setting relationship of the pressure receiving areas of the plurality of hydraulic operation chambers is made different so that at least one of the second and third constants β and γ has a different value from the control valve means. The hydraulic drive described. 5. The auxiliary valve is arranged between the back pressure chamber of the main valve and the pilot valve in each of the first and second flow rate control valve means, and the plurality of hydraulic operation chambers are provided. In each of the control means of the first and second flow rate control valves, a first hydraulic operating chamber for urging the auxiliary valve in the opening direction and a first hydraulic operation chamber for urging the auxiliary valve in the closing direction. And second, third and fourth hydraulic operation chambers, wherein the conduit means guides the control pressure of the main valve back pressure chamber to the first hydraulic operation chamber, and the pilot. The inlet pressure of the valve is the second
Second hydraulic line to the hydraulic operating chamber, a third hydraulic line to guide the maximum load pressure to the third hydraulic operating chamber, and a discharge pressure of the hydraulic pump to the fourth hydraulic operating chamber. The hydraulic drive system according to claim 4, further comprising a fourth pipeline. 6. The hydraulic drive system according to claim 5, wherein in each of the first and second flow rate control valve means, the main valve and the auxiliary valve are integrated. 7. The control means, in each of the first and second flow rate control valve means, an electromagnetic operation portion provided in the auxiliary valve, a discharge pressure of the hydraulic pump, the maximum load pressure, and the Based on the pressure detection means for directly or indirectly detecting the inlet pressure and the outlet pressure of the pilot valve and the detection signal of the pressure detection means, the above ΔPz = α (Ps-Plma
x) + β (Plmax−Pl) + γPl The target differential pressure between the inlet pressure and the outlet pressure of the pilot valve that satisfies the relationship is calculated, and this target differential pressure is used as the command value of the electromagnetic operating unit of the auxiliary valve. And a first constant in each of the first and second flow rate control valve means has a positive value, and the first flow rate control valve means and the second flow rate control valve means. And the first, second and third constants α, β, γ are set in the arithmetic means so that at least one of the second and third constants β, γ has a different value. The hydraulic drive system according to claim 1. 8. The pump control means is a load-sensing type pump regulator that holds the discharge pressure of the hydraulic pump higher by a predetermined value than the maximum load pressure of the first and second hydraulic actuators. The hydraulic drive system according to claim 1. 9. 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; A plurality of working members each of which is driven, 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; Pump control means for controlling the discharge pressure of the hydraulic pump; each of the plurality of flow control valve means has a first valve means for changing an opening degree according to an operation amount of the operation means, and a first valve means. Second valve means connected in series to the valve means for controlling the differential pressure between the inlet pressure and the outlet pressure of the valve means, and the inlet pressure and the outlet pressure of the first valve means, In the hydraulic excavator having a control means for controlling the second valve means based on the discharge pressure of the hydraulic pump and the maximum load pressure of the plurality of hydraulic actuators, each of the plurality of flow rate control valve means includes the main circuit. A valve body for controlling communication between the inlet port and the outlet port connected to the valve, a variable throttle for changing the opening degree according to the displacement of the valve body, and a valve for communicating with the inlet port through the variable throttle, A seat type main valve having a back pressure chamber for generating a control pressure for urging the body in the valve closing direction, and a pilot circuit connected between the back pressure chamber of the main valve and the outlet port, The first valve means is a pilot valve connected to the pilot circuit to control a pilot flow flowing through the pilot circuit, and the second valve means is connected to the pilot circuit, and the pilot valve is connected to the pilot circuit. Is an auxiliary valve for controlling the differential pressure between the inlet pressure and the outlet pressure of the pilot valve, wherein the control means is the flow control valve means for at least two working members of the revolving structure, the boom, the arm, and the bucket. The pressure difference between the inlet pressure and the outlet pressure of the valve is ΔPz, and the discharge pressure of the hydraulic pump is P
s, the maximum load pressure of the hydraulic actuators
ax, Pl self load pressure of the hydraulic actuator
Where ΔPz = α (Ps−Plmax) + β (Plmax−Pl) + γPl where α, β, γ control the auxiliary valve so as to satisfy the relationship of the first, second and third constants. In addition, the first constant α is set to a positive value by each of the plurality of flow rate control valve means, and two of the plurality of flow rate control valve means are used.
The second and third constants β and γ between the two flow control valve means.
Is set to a different value. 10. The ratio of the pressure receiving area of the main valve valve body receiving the discharge pressure of the hydraulic pump via the inlet port to the pressure receiving area of the main valve valve body receiving the control pressure of the back pressure chamber is defined as K. 10. The hydraulic excavator according to claim 9, wherein the first constant α is set to satisfy α ≦ K in each of the control means of the plurality of flow rate control valve means. 11. The second constant β is set to a positive value in the control means of the flow control valve means on the bottom side of the boom and arm hydraulic actuators, respectively.
The hydraulic excavator according to claim 9, wherein the second constant β for the bottom side of the boom hydraulic actuator is set to be larger than the second constant β for the bottom side of the arm hydraulic actuator. 12. The hydraulic excavator according to claim 9, wherein the second constant β is set to a negative value in the control means of the flow control valve means on the bottom side of the bucket hydraulic actuator. 13. A hydraulic excavator according to claim 9, wherein the third constant γ is set to a negative value in the control means of the flow rate control valve means relating to the hydraulic actuator for the revolving structure. 14. The hydraulic excavator according to claim 9, wherein the third constant γ is set to a positive value in the control means of the flow control valve means on the bottom side of the bucket hydraulic actuator. 15. The control means of the flow control valve means on the rod side of the boom and arm hydraulic actuator sets the second and third constants β and γ to zero. Hydraulic excavator described in 10.