JP2592502B2 - Hydraulic drive and hydraulic construction machinery - 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,
In general, at least one hydraulic pump, a plurality of hydraulic actuators connected to the hydraulic pumps via respective main circuits, and driven by hydraulic oil discharged from the hydraulic pumps; And a plurality of flow control valves connected to the respective main circuits therebetween.

In USP 4,617,854, in such a hydraulic drive device, an auxiliary valve is arranged upstream of the main circuit of each flow control valve,
The inlet pressure and the outlet pressure of the flow control valve are guided to the opposing first operating portion of the auxiliary valve, and the discharge pressure of the hydraulic pump and the maximum load pressure of the plurality of hydraulic actuators are guided to the opposing second operating portion. It describes a configuration in which a load sensing type pump regulator that keeps the discharge pressure of a hydraulic pump higher by a predetermined value than the maximum load pressure 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 modification of the operation of a single hydraulic actuator that incorporates the effect of self-load pressure and modifies the pressure compensation function.

[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, there is a problem that the pressure loss at the auxiliary valve portion is large because the auxiliary valve is arranged in the main circuit through which a large flow rate flows.

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 device of a construction machine such as a hydraulic shovel, it may be preferable to be affected by the load pressure of another hydraulic actuator or the self-load pressure depending on the type and working form of a 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 such a working mode, when the turning pressure at the beginning of turning is high, if the hydraulic pressure can be supplied to the boom as much as possible without relief, the waste of energy can be reduced and the rising speed of the boom can be reduced at first. It is possible to automatically adjust the boom and the turning speed such that the turning speed is increased quickly with respect to the turning speed, and the turning speed gradually increases when the boom is raised 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.

Also in hydraulic excavators, such as a slope forming operation performed by a combined operation of a boom and an arm, we want to accurately divide the flow rate according to the operation amount ratio of the boom operation lever and the arm operation lever regardless of the load pressure. There are also work forms.

Therefore, in a construction machine such as a hydraulic shovel, the characteristics of the flow control valve are not uniquely determined so as to perform the pressure compensating function and / or the flow dividing function. It would be desirable to be able to modify it to provide the corresponding functions.

However, in U.S. Pat.No.4,617,854, although the pressure compensation function and the shunt function are achieved by the installation of the auxiliary valve as described above, the idea of modifying these functions by incorporating the influence of the load pressure of other hydraulic actuators or the self-load pressure is not considered. However, the above-mentioned demand for correcting the characteristics of the flow control valve in accordance with the type of the working member and the working form cannot be met.

On the other hand, in USP 4,535,809, since the hydraulic drive device is intended for a single hydraulic actuator, it is possible to perform a pressure compensation function relating to the operation of the single hydraulic actuator by installing an auxiliary valve, or This is a technology that only modifies the pressure compensation function by taking into account the effect of the self-load pressure of the hydraulic actuator, and is independent of modifying the functions for the combined operation of multiple hydraulic actuators. There was no idea to modify the pressure compensation and shunt functions to incorporate the effects.

SUMMARY OF THE INVENTION An object of the present invention is to provide a hydraulic drive device and a hydraulic construction machine capable of correcting the characteristics of a flow control valve in accordance with the type and work form of a working member of the hydraulic construction machine with a small throttle loss. .

[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 respectively connected to the hydraulic pumps via a main circuit and driven by hydraulic oil discharged from the hydraulic pumps; and the hydraulic pumps and the first and second hydraulic actuators. First and second flow control valve means connected to respective main circuits between second hydraulic actuators; and pump control means for controlling discharge pressure of the hydraulic pump;
And the second flow 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 inlet pressure of the valve means. Second valve means for controlling the differential pressure of the outlet pressure; further comprising, for each of said first and second flow control valve means, an inlet pressure and an outlet pressure of said first valve means; A hydraulic drive unit having control means for controlling said second valve means based on a discharge pressure of a pump and a maximum load pressure of said first and second hydraulic actuators, wherein said first and second flow control valve means Are respectively a valve body that controls the communication between an inlet port and an outlet port connected to the main circuit, a variable throttle that changes an opening in accordance with the displacement of the valve body, and the variable throttle that is connected to the outlet port. And urges the valve body in the valve opening direction. A main valve having a back pressure chamber for generating a control pressure, and a pilot circuit connected between an inlet port of the main valve and the back pressure chamber; the first valve means is provided in the pilot circuit. The second valve means is connected to the pilot circuit and configured to control a differential pressure between an inlet pressure and an outlet pressure of the pilot valve. Being configured as a valve; wherein, for each of the first and second flow control valve means, a differential pressure between an inlet pressure and an outlet pressure of the pilot valve is equal to a discharge pressure of the hydraulic pump and the second pressure control valve means. With respect to the differential pressure between the maximum load pressure of the first and second hydraulic actuators, the differential pressure between the maximum load pressure and the respective self-load pressures of the hydraulic actuators, and the self-load pressure, The auxiliary valve is controlled so as to have a relationship represented by the following formula: ΔPz = α (Ps−Plmax) + β (Plamx−Pl) + γPl where ΔPz: differential pressure between the inlet pressure and the outlet pressure of the pilot valve. Ps: discharge pressure of the hydraulic port P1max: maximum load pressure of the first and first hydraulic actuators P1: self-load pressures of the first and second hydraulic actuators α, β, γ: first and second Second and third constants There is provided a hydraulic drive device wherein the first, second and third constants α, β, γ are set to predetermined values, respectively.

Assuming that the ratio of the pressure receiving area of the main valve element receiving the load pressure of the associated hydraulic actuator via the outlet port to the pressure receiving area of the main valve element receiving the control pressure of the back pressure chamber is K, Is preferably α ≦ K
In a relationship. In this case, the second and third constants β and γ can be set to zero.

The first constant α is set to a positive value corresponding to the proportional gain between the operation amount of the operation means and the flow rate passing through the main valve, and the second constant β is set to a value other than the related hydraulic actuator. Is set to a value based on the operating characteristics of both actuators when the combined operation with the hydraulic actuator is performed, and the third constant γ is set to a value based on the operating characteristics of the associated hydraulic actuator.

A plurality of hydraulic operation chambers provided in the auxiliary valve of each of the first and second flow control valve means,
The plurality of hydraulic operation chambers may have a pipeline 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, The pressure receiving areas of the plurality of hydraulic operation chambers are set so that the first, second, and third constants α, β, and γ become the predetermined values.

 The configuration examples of the hydraulic control means will be listed.

(1) The auxiliary valve is disposed between the inlet port of the main valve and the pilot valve, and the plurality of hydraulic operation chambers are configured to urge the auxiliary valve in a valve opening direction with first and second hydraulic pressures. An operating chamber, and third and fourth hydraulic operating chambers for urging the auxiliary valve in a valve closing direction, wherein the conduit means sends the discharge pressure of the hydraulic pump to the first hydraulic operating chamber. A first conduit for guiding;
A second conduit for guiding the outlet pressure of the pilot valve to the second hydraulic operating chamber, a third conduit for guiding the maximum load pressure to the third hydraulic operating chamber, and an inlet pressure of the pilot valve. To the fourth hydraulic operation chamber.

(2) The auxiliary valve is disposed between the back pressure chamber of the main valve and the pilot valve, and the plurality of hydraulic operation chambers are a first hydraulic operation chamber that biases the auxiliary valve in a valve opening direction. And second, third, and fourth hydraulic operating chambers for urging the auxiliary valve in a valve closing direction, wherein the conduit means reduces the outlet pressure of the pilot valve to the first hydraulic operating chamber. A second conduit for guiding the inlet pressure of the pilot valve to the second hydraulic operating chamber, and a second conduit for guiding the load pressure of the associated hydraulic actuator to the third hydraulic operating chamber. And a fourth conduit for guiding the maximum load pressure to the fourth hydraulic operation chamber.

(3) The auxiliary valve is disposed between the pilot valve and a back pressure chamber of the main valve, and the plurality of hydraulic operation chambers are configured to urge the auxiliary valve in a valve opening direction. A hydraulic chamber, and third and fourth hydraulic operating chambers for urging the auxiliary valve in a valve closing direction, wherein the conduit means adjusts a load pressure of an associated hydraulic actuator to the first hydraulic operating chamber. A first conduit for guiding the outlet pressure of the pilot valve to the second hydraulic operating chamber, and a third conduit for guiding the maximum load pressure to the third hydraulic operating chamber. A fourth conduit for guiding a control pressure of the back pressure chamber to the fourth hydraulic operation chamber.

(4) The auxiliary valve is disposed between an inlet port of the main valve and the pilot valve, and the plurality of hydraulic operation chambers are configured to urge the auxiliary valve in a valve opening direction with first and second hydraulic pressures. An operating chamber, and third and fourth hydraulic operating chambers for urging the auxiliary valve in a valve closing direction, wherein the conduit means adjusts a load pressure of an associated hydraulic actuator to the first hydraulic operating chamber. And a discharge pressure of the hydraulic pump to the second pipe.
A second conduit for guiding the maximum load pressure to the third hydraulic operating chamber, and a third conduit for guiding the maximum load pressure to the third hydraulic operating chamber, and guiding the inlet pressure of the pilot valve to the fourth hydraulic operating chamber. And a fourth conduit.

(5) The auxiliary valve is disposed between the pilot valve and the back pressure chamber of the main valve, and the plurality of hydraulic operation chambers are a first hydraulic operation chamber that biases the auxiliary valve in a valve opening direction. And second and third hydraulic operation chambers for urging the auxiliary valve in a valve closing direction, wherein the conduit means guides the outlet pressure of the pilot valve to the first hydraulic operation chamber. A first pipeline, a second pipeline that guides the discharge pressure of the hydraulic pump to the second hydraulic operating chamber, and a third pipeline that guides the maximum load pressure to the third hydraulic operating chamber. Have.

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, according to the present invention, in order to achieve the above object, a revolving structure driven by a plurality of hydraulic actuators,
A hydraulic construction machine having a plurality of working members including a boom, an arm, and a bucket, to which the above hydraulic drive device is applied is provided.

In the hydraulic construction machine, the control means preferably sets the second constant β to a positive value for the flow control valve means on the bottom side of the boom hydraulic actuator.

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

More preferably, the control means sets the second constant β to a negative value for the flow control valve means related to the bottom side of the bucket hydraulic actuator.

More preferably, the control means sets the third constant γ to a negative value for the flow control valve means relating to the hydraulic actuator for the rotating body.

More preferably, the control means sets the third constant γ to a positive value for the flow control valve means related to the bottom side of the bucket hydraulic actuator.

Preferably, the control means sets the second and third constants β and γ to zero for each of the flow control valves on the rod side of the boom and arm hydraulic actuators.

[Action]

The present inventors have conducted various studies on the relationship between the auxiliary valve arranged in the pilot circuit and the differential pressure across the pilot valve.
Pilot valve differential pressure ΔPz controlled by auxiliary valve
Has generally been found to be represented by the above formula, which is reproduced below.

ΔPz = α (Ps−Plmax) + β (Plamx−Pl) + γPl The above expression has the following meaning. In this equation, Ps-Plmax of the first term on the right side is common to all the flow control valves, and thus controls the flow dividing function in the combined operation.
Since max-Pl changes depending on the maximum load pressure of another actuator, it controls the harmonic function in the combined operation.
Since the term γPl changes according to the self-load pressure, it controls the self-pressure compensation function. The necessity and the degree of each of these three functions are determined according to the values of the constants α, β, and γ. Here, the branch function of the first term is a basic function of the complex operation. Therefore, the constant α is set to a predetermined positive value irrespective of the associated work member. On the other hand, the harmony function of the second term and the self-pressure compensation function of the third term are functions added according to the type and work form of the related work member. Therefore, the constant β,
γ is set to a predetermined value including zero. By setting α, β, and γ in this manner, it becomes possible to provide a shunt function or a harmony function based on the shunt function and / or a self-pressure compensation function, and the types and working forms of working members of the hydraulic construction machine , The characteristics of the flow control valve can be modified.

In the configuration of the present invention, since the auxiliary valve is installed in the pilot circuit instead of the main circuit, the throttle loss in the auxiliary valve portion can be reduced even if a large flow rate flows through the main circuit.

The ratio of the pressure receiving area receiving the load pressure of the hydraulic actuator to the pressure receiving area receiving the control pressure of the back pressure chamber, K
On the other hand, when the first constant α has a relation of α ≦ K, the differential pressure obtained by α (Ps−Plmax) in the first term on the right side of the above equation can be taken by the pilot valve on the high load pressure side. Within the range of the maximum differential pressure, the differential pressure of the first term in both the first and second flow control valves becomes substantially the same, and the operating amount of the operating means (pilot valve opening The flow rate can be accurately divided in proportion to (degree).

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 a meaning of a proportional gain between the manipulated variable and the main flow rate passing through the main valve, and accordingly, the first constant α is set to any positive value corresponding to the proportional gain. Where α = K
In this case, 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 especially when it is preferable not to be affected by the load pressure of another hydraulic actuator.

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

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 and the maximum load pressure of the plurality of hydraulic actuators according to the first term of the above equation 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 hydraulic drive device is applied to a hydraulic construction machine, for example, a hydraulic excavator, the characteristics of the flow control valve relating to at least two working members of a revolving unit, a boom, an arm, and a bucket are determined according to the type and working form of the working member. It can be set and modified accordingly, and a shunt function or a harmonic function and / or a self-pressure compensation function based on the shunt function described above can be added.

When the second constant β is set to a relatively large positive value for the flow rate control valve means for the bottom side of the hydraulic actuator for boom, when the initial acceleration of the swing in the combined operation of the swing and the boom, the low load side The flow rate corresponding to the increase in the differential pressure between the maximum load pressure (swing pressure) and the self-load pressure (boom pressure) flows through the flow control valve on the bottom side of the hydraulic actuator for the boom, thereby 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 with respect to the flow control valve on the bottom side of the hydraulic actuator for bucket, when the digging operation is performed by the combined operation using the bucket, Is released from the ground, and at the moment when it comes to the surface, the differential pressure between the maximum load pressure (other hydraulic actuator pressure) and the self-load pressure (bucket pressure) is increased to reduce the flow rate through the flow control valve and reduce shock. Can be.

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〕

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

1. Basic Embodiment In FIG. 1, a hydraulic drive device according to an embodiment of the present invention includes a variable displacement hydraulic pump 1 of, for example, a swash plate type, and is connected to the hydraulic pump 1 via main circuits 2 and 3, respectively. A plurality of, for example, two hydraulic actuators 6, 7 driven by pressure oil discharged from the hydraulic pump 1, and respective main circuits between the hydraulic pump 1 and the hydraulic actuators 6, 7
Directional switching valves 8, 9 connected to the hydraulic pumps 6, 7 for controlling the driving of the hydraulic actuators 6, 7; There is provided a load sensing type pump regulator 10 which keeps the load pressure higher than the load pressure by a predetermined value.

The directional control valve 8 includes four flow control valves 11, 12, 13, and 14. The first flow control valve 11 is a hydraulic actuator 6
Meter-in when the valve operates to extend (entrance side)
The second flow control valve 12 is connected to the meter-in circuit 16 when the hydraulic actuator 6 operates so as to contract, and the third flow control valve 13 is connected to the hydraulic actuator 6 and the second flow rate. The fourth flow control valve 14 is connected to a meter-out (outlet side) circuit 17 when the hydraulic actuator 6 operates to extend between the control valves 12. It is connected to a meter-out circuit 18 when the hydraulic actuator 6 operates so as to contract between them. The first flow control valve 11 and the fourth
A check valve for preventing backflow of pressure oil from the hydraulic actuator 6 to the first flow control valve 11 between the first flow control valve 11 and the flow control valve 14
A check valve 20 is connected between the second flow control valve 12 and the third flow control valve 13 to prevent a backflow of hydraulic oil from the hydraulic actuator 6 to the second flow control valve 12. Is connected.

The first to fourth flow control valves 11 to 14 are main valves 21, 22,
23, 24 and pilot circuits 25, 26, 27, 28 for controlling the main valve
And pilot valves 29, 30, 3 connected to the pilot circuit
The first and second flow control valves 11 and 12 further include pressure compensating valves 33 and 34 connected to the pilot circuit in series with the pilot valve.

As shown in FIG. 2, the main valve 21 of the first flow control valve 11 is connected to an inlet port 31 connected to a line on the hydraulic pump 1 side of the meter-in circuit 15 and to a line on the hydraulic actuator 6 side. A valve housing 33 having an outlet port 32 and a valve body 35 disposed in the valve housing 33 and having a control orifice 34.
The opening of 34 is adjusted to control the communication between the inlet port 31 and the outlet port 32. A back pressure chamber 36 for generating a control pressure Pc for urging the valve body 35 in the valve opening direction is formed on the side opposite to the control orifice of the valve body 35. A chamber 38 communicating with the back pressure chamber 36 and accommodating the control piston 37 is formed at an end of the valve body 35 facing the back pressure chamber 36, and the chamber 38 also communicates with the outlet port 32 via a passage 39. doing. One end of the control piston 37 is housed in the pressure chamber 40 in the valve body 35, and the other end is supported by a stopper 41 closing the back pressure chamber 36. The pressure chamber 40 is an inlet port via a passage 42
The control piston 37 is held at a position where the control piston 37 is in contact with the stopper 41 in communication with the plug 31. The control piston 37 also has a tapered portion 43 at an intermediate portion thereof, and the tapered portion 43 cooperates with the wall of the opening of the chamber 38 to change the opening degree according to the displacement of the valve body 35.
Is formed. Thus, the back pressure chamber 36 communicates with the outlet port 32 via the variable throttle 44.

In the valve element 35, the upper annular end face facing the inlet port 31 defines an annular pressure receiving area As that receives the discharge pressure Ps of the hydraulic pump 1 and urges the valve element 35 upward, that is, in the valve opening direction.
The bottom wall facing the outlet port 32 is a hydraulic actuator 6
A pressure receiving area Al is also defined which receives the negative pressure Pl and urges the valve body 35 in the valve opening direction in the same manner, and the top end face facing the back pressure chamber 36 receives the control pressure Pc to open the valve body 35 downward, i.e. The pressure receiving area Ac urged in the valve direction is defined. These pressure receiving areas are: Ac = As
+ Al.

The check valve 33 described above is arranged below the valve body 35,
The valve housing 33 further has an outlet port 45 for the check valve 19.

The pilot circuit 25 includes a main valve 21, an inlet port 31, and a back pressure chamber 36.
Connected between

The pilot valve 29 includes a poppet type valve body 42 that controls communication between the inlet port 40 and the outlet port 41, a spring 43 that urges the valve body 42 in the valve closing direction, and urges the valve body 42 in the valve opening direction. The hydraulic operation chamber 44 is connected to a pilot circuit that generates a pilot pressure according to the operation amount of an operation lever (not shown), and opens the valve body 42 to an opening degree according to the operation amount. It has become.

The pressure compensating valve 33 includes a seat-type valve body 52 that controls communication between the inlet port 50 and the outlet port 51, and first and second hydraulic operation chambers 53 and 54 that urge the valve body 52 in the valve opening direction. And third and fourth hydraulic operation chambers 55 and 56 which are located opposite to the first and second hydraulic operation chambers 53 and 54 and urge the valve body 52 in the valve closing direction. The first hydraulic operation chamber 53 is provided at the above-mentioned inlet port.
Formed by the inlet 57 of the pressure compensating valve 33 communicating with 50,
The second hydraulic operation chamber 54 is connected to the pilot valve 25 outlet side of the pilot circuit 25 via a pilot line 58,
The hydraulic operating chamber 55 is connected to a later-described maximum load pressure pipe 61 via a pilot pipe 59, and the fourth hydraulic operating chamber 56 is connected to the pilot valve 29 inlet side of the pilot circuit 25 via the pilot pipe 60. It is connected. As a result, the discharge pressure Ps of the hydraulic pump 1 is introduced into the first hydraulic operation chamber 53,
The outlet pressure of the pilot valve 29 is led to the hydraulic operating chamber 54, which is equal to the control pressure Pc of the back pressure chamber 36, and the third hydraulic operating chamber 55 is connected to the high-pressure side of the hydraulic actuators 6, 7. The load pressure, ie the maximum load pressure Plmax is introduced and the fourth
The inlet pressure Pz ′ of the pilot valve 29 is introduced into the hydraulic operation chamber 56. The end face of the valve body 52 facing the first hydraulic operation chamber 53 defines a pressure receiving area as for receiving the discharge pressure Ps of the hydraulic pump 1, and the annular end face facing the second hydraulic operation chamber 54 is The pressure receiving area ac for receiving the outlet pressure Pc is defined, and the end face facing the third hydraulic operation chamber 55 is provided with the hydraulic actuator 6,
The pressure receiving area am for receiving the maximum load pressure Plmax of 7 is defined, and the annular end face facing the fourth hydraulic operation chamber 56 defines the pressure receiving area az for receiving the inlet pressure Pz of the pilot valve 29.

In the above configuration, the first to fourth hydraulic operation chambers 53 to 56 of the pressure compensating valve 33, the pilot lines 57 to 60, and the main valve 21
The part that defines the pressure receiving areas Ac and As of the valve element 35 is that the differential pressure ΔPz = (Pz−Pc) between the inlet pressure and the outlet pressure of the pilot valve 29 is the discharge pressure of the hydraulic pump 1 and the two hydraulic actuators 6 and
7 Differential pressure Ps-Plmax from the maximum load pressure, differential pressure Plax between the maximum load pressure and the self-negative pressure of each hydraulic actuator
-Pl and the self-load pressure Pl constitute control means for controlling the pressure compensating valve 33 so as to have a relationship represented by the following equation: ΔPz = α (Ps-Plmax + β (Plmax-Pl ) + ΓPl (1) where α, β, and γ are first, second, and third constants, respectively, and are set to predetermined values, respectively. And third constants α, β,
The setting of the predetermined value of γ is performed by selecting the pressure receiving areas as, ac, am, az of the first to fourth hydraulic operation chambers 53 to 56 of the pressure compensating valve 33. That is, the first to fourth hydraulic operation chambers 53 to 56
Are selected so as to obtain predetermined values of the first, second and third constants α, β, γ. The pressure receiving area a of the first to fourth hydraulic operation chambers 43 to 46
s, ac, am, az are set so as to hold the valve body 52 in the open position when the main valve 21 and the pilot valve 29 are closed.

In the combination of the main valve 21 and the pilot valve 29 of the first flow control valve 11 thus configured, the pilot valve 29
Is opened by operating an operation lever (not shown), the hydraulic oil of the hydraulic pump 1 is guided to the back pressure chamber 36 of the main valve 21 via the pilot circuit 25 at that moment. Then, the pressure in the back pressure chamber 36, that is, the control pressure, increases according to the opening degree of the pilot valve 29, and the pressure of the outlet port 32 communicating with the back pressure chamber 36 via the chamber 38 and the passage 39 also increases. Valve 19 is opened. As a result, a pilot flow from the pilot circuit 25 to the outlet port 32 through the back pressure chamber 36 is formed.
By the operation of 44, the control pressure of the back pressure chamber 36 is also increased according to the pilot flow rate (opening of the pilot valve 29). When the opening of the pilot valve 29 becomes larger than the opening of the variable throttle 44,
The control pressure Pc is correspondingly increased, and the valve element 35 starts to move toward the outlet port 32. That is, the main valve 21 opens.
When the valve body 35 moves in the valve opening direction in this way, the opening degree of the variable throttle 44 formed around the control piston 37 pressed and held by the plug 41 increases, and the throttle action of the variable throttle 44 decreases. . Therefore, the increase in the control pressure of the back pressure chamber 36 is controlled. As a result, the valve body 35 is settled at the point where the opening of the pilot valve 29 and the opening of the variable throttle 44 match.

As described above, the main valve valve body 35 opens to an opening proportional to the pilot flow rate by the action of the variable throttle 44 and the back pressure chamber 36, and the flow rate corresponding to the operation amount of the operation lever (the opening degree of the pilot valve 31) is mainly controlled. It flows from the inlet port 31 to the outlet port 32 through the control orifice 34 of the valve 21.

In such control of the main valve 21 by the pilot valve 29, since the pressure compensating valve 33 is further disposed in the pilot circuit 25, the flow rate of the main valve 21 is further controlled by the pressure compensating valve 33. Since the control function of the pressure compensating valve 33 is an essential part of the present embodiment, an operation principle will be described below.

A pilot circuit 26 of the main valve 22 of the second flow control valve 12,
The pilot valve 30 and the pressure compensating valve 34 are a first flow control valve.
11, the main valve 21, the pilot circuit 25, and the pilot valve described above.
The configuration is the same as that of the pressure compensation valve 29 and the pressure compensation valve 33, respectively.

As shown in FIG. 3, the main valve 23 of the third flow control valve 13 is connected to an inlet port 70 connected to a line on the hydraulic actuator 6 side of the meter-out circuit 17 and a line on the tank side. A valve housing 72 having an outlet port 71, and a valve element 74 disposed in the valve housing 72 and engaging with the valve seat 73. The inlet port 70 and the outlet port are provided in accordance with the displacement (opening) of the valve element 74. Controls communication with port 71. A plurality of slits 75 are formed in the outer periphery of the valve body 74 in the axial direction, and the slits 75 form a variable throttle 76 that changes the opening degree according to the displacement of the valve body 74 in cooperation with the inner wall of the valve housing 72. ing. Behind the valve element 74 in the valve housing 72, an inlet port 70 is provided via a variable throttle 76.
And a back pressure chamber 77 that generates a control pressure P3c.

The upper annular end face facing the inlet port 70 of the valve body 74 defines an annular pressure receiving area A3l that receives the load pressure Pl of the hydraulic actuator 6 and urges the valve body 74 in the upward opening direction in the figure. The bottom wall surface facing the port 71 defines a pressure receiving area A3r that receives the tank pressure Pr and urges the valve body 74 in the valve opening direction, and the top end face that faces the back pressure chamber 77 receives the control pressure P3c and opens the valve body 74. The pressure receiving area A3c that is urged in the valve closing direction below the figure is defined. These pressure receiving areas have a relationship of A3c = A31 + A3r.

The pilot circuit 27 has a back pressure chamber 77 of the main valve 23 and an outlet port.
Connected between 71.

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

A combination of the main valve 23 and the pilot valve 31 of the third flow control valve 13 configured as described above is known from US Pat. No. 4,535,809. When the pilot valve 31 is opened by operating an operation lever (not shown), a pilot circuit is opened. In 27, a pilot flow is formed according to the opening of the pilot valve 31, and the main valve body 74 opens to an opening proportional to the pilot flow by the action of the back pressure chamber 77 of the variable throttle 76, and the operation amount of the operating lever The flow rate corresponding to the (opening degree of the pilot valve 31) flows from the inlet port 70 to the outlet port 71 through the main valve 23.

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

The direction switching valve 9 is configured similarly to the direction switching valve 8. In the following, members equivalent to the components of the direction switching valve 8 of the direction switching valve 9 are represented by adding the letter A to the numbers assigned to the components of the direction switching valve 8.

The outlet ports 32 of the first and second flow control valves 11 and 12 of the directional control valve 8 are connected to the above-mentioned conduit 61 via check valves 80 and 81, respectively, and the first and second directional control valves 9 and Check the flow control valves 11A and 12A at the outlet ports.
The line 61A is connected to the line 61A via 1A, and the lines 61 and 61A are connected to each other via a line 82. The line 82 reaches a tank via a throttle 83. Thus, during the combined operation of the hydraulic actuators 6 and 7, the load pressure on the high pressure side of both, that is, the maximum load pressure is selected by the check valves 80, 81 and 80A, 81A, and the line 6
It is led to 1,61A, 82. Thus, the lines 61, 61A and 82 constitute a maximum load pressure circuit.

The pump regulator 10 includes a hydraulic cylinder type swash plate tilting device 90 and a control valve 91. The swash plate tilting device 90 receives the discharge pressure of the hydraulic pump 1 through a pipe 92 into a rod-side cylinder chamber. The head side cylinder chamber is connected to the tank and the rod side cylinder chamber via a control valve 91. The discharge pressure guided to the rod side cylinder chamber of the swash plate tilting device is reduced according to the position of the control valve 91, and drives the piston by the area difference between the rod side cylinder chamber and the head side cylinder chamber,
The discharge amount of the hydraulic pump 1 is controlled in accordance with the position of the control valve 91.

The control valve 91 has opposed hydraulic operating units 93 and 94 and a spring 95. The hydraulic operating unit 93 is connected to the discharge line of the hydraulic pump 1 via a pilot line 96, and the hydraulic operating unit 94 is connected to the pilot line. It is connected to the maximum load pressure circuit 82 via 97. As a result, the discharge pressure of the hydraulic pump 1 and the maximum load pressure and the set force of the spring 65 act on the control valve 91 in opposition to adjust the position of the control valve 91 according to the change in the maximum load pressure. By controlling the tilting device 141, the discharge pressure of the hydraulic pump 1 is maintained at a pressure higher than the maximum load pressure by a pressure corresponding to the strength of the spring 65.

Next, the operation principle of the pressure compensating valves 33, 34, 33A, 34A will be described. In the following, all the pressure compensating valves 33, 34, 33A, 34
Items common to A will be described using the pressure compensating valve 33 as a representative. The equation for balancing the pressure of the valve element 52 of the pressure compensating valve 33 is represented by the following equation.

asPs + acPc = amPlmax + azPz The pressure balance equation of the valve element 35 of the main valve 21 is represented by the following equation.

AcPc = AsPs + AlPl Using the above two equations, derive the equation for the differential pressure across pilot valve 29, and deform using Ac = As + Al Therefore, Can be expressed as Pz-Pc = α (Ps-Plmax) + β (Plmax-Pl) + γPl. Here, Pz−Pc = ΔPz.

Therefore, the same equation as the above equation (1) is obtained. This equation is reproduced below.

ΔPz = α (Ps−Plmax) + β (Plmax−Pl) + γPl (1) Therefore, the above equation (1) will be considered. In the equation (1), ΔPz on the left side is a differential pressure between the inlet pressure Pz and the outlet pressure Pc of the pilot valve 29. 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.
This is a term relating to the pressure difference from Pl, and β is a proportional constant. The third term is determined by the self-load pressure Pl, and the proportionality constant is γ. Here, the pressure balance equation of the valve element 35 of the main valve 21 is AcPc =
Since AsPs + AlPl, the load pressure Pl of the hydraulic actuator 6 is equal to the discharge pressure Ps of the hydraulic pump 1 and the pilot valve 29.
At the outlet pressure Pc. Therefore, in the equation (1), the pressure compensating valve 33 determines the differential pressure ΔPz between the inlet pressure Pz and the outlet pressure Pc of the pilot valve 29 based on the four pressures Ps, Plmax, Pl, and Pz.
Can be controlled at that time;
Differential pressure Ps-Plmax between the discharge pressure Ps and the maximum load pressure Plmax,
Differential pressure Plmax-P between maximum load pressure Plmax and self-load pressure Pl
l, controllable in proportion to each of the three elements of the self-load pressure Pl; and the three elements Ps-Plmax, Plmax
-Pl, means that the degree proportional to Pl can be arbitrarily set by selecting the values of the proportional constants α, β, γ.

Here, the pressure compensating valve 33 is the differential pressure ΔPz across the pilot valve 29.
Is to control the pilot flow through the pilot valve 29, and consequently to control the main flow through the main valve 21 from the function of the combination of the main valve 21 and the pilot valve 29 described above. .

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 the present embodiment including the pressure regulator 10, the pressure regulator is constant as long as the pump regulator 10 functions effectively, and is common to all the pressure compensating valves 33 (34) and 33A (34A).

Therefore, in the first term on the right side, controlling the pressure difference ΔPz before and after the pilot valve 29 so as to be proportional to the pressure difference Ps−Plmax is a difference in the operating state where the pump regulator 10 is functioning effectively. It is to control the pressure ΔPz to be constant, and to keep the main flow rate through the main valve 21 constant even if the main valve inlet pressure Ps or the outlet pressure Pl fluctuates if the opening degree of the pilot valve 29 is fixed. It is. That is, it fulfills a pressure compensation function.

Also, when the combined consumption of the hydraulic actuators 6 and 7 is larger than the maximum discharge flow rate of the hydraulic pump 1 when the combined operation of the hydraulic actuators 6 and 7 is performed, as in the case where the discharge pressure of the hydraulic pump 1 decreases,
In an operation state in which the pump regulator 10 does not function effectively, the differential pressure ΔPz decreases as the differential pressure Ps−Plmax decreases, and the main flow through the main valve 21 also decreases, but the differential pressure Ps−Pl
Since max is common to the two pressure compensating valves 33 (34) and 33A (34A), the main flow rate through the main valves 21 (22) and 21A (22A) decreases at the same rate. Therefore, the main valves 21 (22) and 21A (22
The main flow passing through A) is proportionally distributed according to the operation amount of the operation lever (opening of the pilot valves 29 (30) and 29A (30A)), and the discharge flow of the hydraulic pump 1 is also transmitted to the high-pressure side hydraulic actuator. Surely supplied. That is, a shunt function is obtained.

In the second term on the right side, controlling the differential pressure ΔPz across the pilot valve 29 to be proportional to the differential pressure Plmax−Pl is equivalent to controlling the load pressure Plmax of another hydraulic actuator.
Is greater than the self-load pressure Pl, it is to change the differential pressure ΔPz across the pilot valve 29 depending on the maximum load pressure Plmax of the other hydraulic actuators.If the opening of the pilot valve 29 is constant, The purpose is to vary the main flow rate through the main valve 21 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 may be changed by the load pressure of the other 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 term on the right side, controlling the pressure difference ΔPz across the pilot valve 29 so as to be proportional to the self-load pressure Pl is based on the change in the pressure difference ΔPz across the pilot valve 29 according to the change in the self-load pressure Pl. In other words, if the opening of the pilot valve 29 is kept constant, the main flow through the main valve 21 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 side governs the pressure compensation and branching functions, the second term governs the harmony function with other actuators, and the third term governs 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 proportionality constant α is set to any positive value. Here, the differential pressure ΔPz before and after the pilot valve 29 defines the pilot flow rate with respect to the opening degree of the pilot valve 29 determined by the operation amount of the operation lever. Therefore, the proportional constant α applied to the differential pressure Plmax−Pl in the first term is Has a meaning of a proportional gain of the pilot flow to the operation amount of the operation lever of the pilot valve 29 (opening degree of the pilot valve), that is, a proportional gain of the main flow through the main valve 21 to the operation amount. Therefore, the proportionality constant α is determined according to the proportional gain.

Further, assuming that the ratio of the pressure receiving area Al of the main valve body 35 receiving the control pressure Pc of the back pressure chamber 36 to the pressure receiving area Al of the valve body 35 receiving the load pressure Pl of the hydraulic actuator 6 is K, the valve body 35
Is Pc = (1−K) Ps + KPl. On the other hand, the discharge pressure Ps of the pump 1 and the pilot valve 29
Of the pressure compensating valve 33
Is fully open, Ps = Pz. Therefore, the differential pressure Pz-Pc (ΔPz) before and after the pilot valve 29 is as follows: Pz−Pc ≦ Ps−Pc ≦ K (Ps−Pl) (2) That is, the maximum differential pressure that the pilot valve 29 can take is K
(Ps-Pl). In the above equation (1), β = 0,
When γ = 0 and the maximum load pressure side (Plmax = Pl) is considered, Pz−Pc = α (Ps−Plmax) ≦ K (Ps−Plmax) (3) That is, when the value of α>α> K is set, the pilot valve on the maximum load pressure side cannot obtain a differential pressure of K (Ps−Pl) or more, while the pilot valve on the low pressure side has α. (Ps-Pl)> K (Ps-Pl)
Even if the pilot valve opening degree of both is the same, the differential pressure across the pilot valve will not be the same, and the pilot flow rate will be different. Therefore, the flow rate cannot be proportionally distributed according to the operation amount. However, even if the proportional distribution cannot be performed, the pressure oil can be reliably supplied to the high-pressure side hydraulic actuator.

Accordingly, in the case of obtaining the flow dividing function of distributing the flow in proportion to the operation amount (opening) of the pilot valve with respect to the flow dividing function of the pressure compensating valve 33, the proportionality 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) differential pressure is obtained. However, when the composite operation is switched to the single operation of the hydraulic actuator on the low load pressure side, 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 is also reduced, the working member is decelerated, and it becomes difficult to perform a smooth operation. On the other hand, when α is set to α ≦ K,
Even in the combined operation, the differential pressure across the pilot valve on the low load pressure side is limited to K (Ps-Plmax), and even when switching from the combined operation to the independent operation, the differential pressure does not fluctuate and stable work is performed. Can be. Therefore, in this sense, it is preferable to set α to α ≦ K.

As can be seen from the above, in the case where 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.

Further, the degree of 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 self-pressure compensation function according to the third item has a different degree of necessity 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 be completely unaffected by the self-load pressure. 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.

By setting the constants α, β, and γ to predetermined values as described above, a shunt function, a harmony function based on the shunt function, and / or a self-pressure compensation function can be obtained. The characteristics of the flow control valve can be modified according to the type of the member and the working mode.

As described above, the proportional constants α, β, and γ are the pressure receiving areas a in the first to fourth operation chambers 53 to 56 of the pressure compensating valve 33.
s, ac, am, az and the pressure receiving area Ac, As of the valve element 35 of the main valve 21 are represented. Here, the pressure receiving areas Ac and As of the main valve body 35 are determined by the conditions on the main valve 21 side. Therefore, if the proportional constants α, β, γ are determined, the pressure receiving areas as, ac, am, az are set so that the proportional constants α, β, γ can be obtained. Here, as a special case, if as + ac = am + az, γ =
It can be set to 0, and can be set to β = 0 if ac = az and as = am. If as = ac = am = az, β = γ
= 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.

Generally, as shown in FIGS. 4 and 5, a hydraulic excavator includes a pair of traveling bodies 100, a revolving body 101 mounted on the traveling body 100 so as to be pivotable, and a vertical plane with respect to the revolving body 101. It has a front attachment 102 that is rotatably mounted, and the front attachment 102 includes a boom 103, an arm 104,
It consists of a bucket 105. Traveling structure 100, revolving structure 101,
The boom 103, the arm 104, and the bucket 105 are respectively composed of a traveling motor 106 (plural), a swing motor 107, and a boom cylinder 10.
8. Driven by arm cylinder 109 and bucket cylinder 110. Here, the swing motor 107, the boom cylinder 108, the arm cylinder 109, the bucket cylinder
110 corresponds to the hydraulic actuators 6, 7 shown in FIG.

In such a hydraulic excavator hydraulic drive device, a swing motor 107, a boom cylinder 108, an arm cylinder
As shown in FIG. 6, the proportional constant α for all the flow control valves of the bucket cylinder 110 and the bucket cylinder 110 is set to the same arbitrary positive value in consideration of the above-described proportional gain. In the flow control valve related to the swing motor 107, the proportional constant β
Is set to β = 0 as shown in FIG. 7 (A), and the proportionality constant γ is set to a small negative value as shown in FIG. 8 (A). In the flow control valve on the bottom side of the boom cylinder 108, the proportionality constant β is set to an arbitrary positive value as shown in FIG. 7 (B), and the proportionality constant γ is set in FIG. 8 (B).
Is set to γ = 0 as shown in FIG. Arm cylinder 10
In the flow control valve on the bottom side of 9, the proportional constant β
Is set to a small positive value as shown in FIG.
The proportionality constant γ is set to γ = 0 as shown in FIG. 8 (B). In the flow control valve on the bottom side of the bucket cylinder 110, the proportionality constant β is set to a small negative value as shown in FIG. 7 (D), and the proportionality constant γ is set as shown in FIG. 8 (C). Set to a small positive value. In the flow control valve related to the rod side of the boom cylinder 108, the flow control valve related to the rod side of the arm cylinder 109, and the flow control valve related to the rod side of the bucket cylinder 110, all of the proportional constants β and γ It is set to zero as shown in FIG. 7 (A) and FIG. 8 (B).

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

First, when none of the operation levers of the direction switching valves 8 and 9 is operated, the first and second flow control valves 11, 12, 11A,
The pilot valves 29, 30, 29A, 30A of 12A are closed, and the pilot flow does not flow through the pilot circuits 25, 26, 25A, 16A.
Therefore, no pressure oil flows to the respective variable throttles 44 of the main valves 21, 22, 21A, 22A, and the control pressure Pc of the back pressure chamber 36 is the same as the pressure (load pressure of the hydraulic actuator 6 or 7) Pl of the outlet port 32. It has become. Load sensing type pump regulator 10
The discharge pressure Ps of the hydraulic pump 1 is higher than the maximum load pressure Plmax of the hydraulic actuators 6 and 7 by a spring 95
Is maintained at a higher pressure by a pressure corresponding to the set value of Accordingly, each pressure receiving area of the valve element 35 has a relationship of Ac = As + Al, and Ps> Pl.
The main valves 21, 22, 21A, 22A are urged in the valve closing direction by Ps, and are maintained in the closed position. Further, the pressure compensating valves 33, 34, 33A, 34A are held at the open positions by setting the pressure receiving areas as, ac, am, az described above.

Similarly, in the third and fourth flow control valves 13, 14, 13A, 14A, the pilot valves 31, 32, 31A, 32A are closed,
The pilot flow does not flow through the pilot circuits 27, 28, 27A, 28A. Accordingly, no pressure oil flows to the respective variable throttles 76 of the main valves 23, 24, 23A, and 24A, and the control pressure P3c of the back pressure chamber 77 is changed to the inlet port 70.
(The load pressure of the hydraulic actuator 6 or 7) Pl. Here, the load pressure Pl is larger than the tank pressure Pr. Therefore, each pressure receiving area of the valve element 74 is A3c = A31 + A3
Because of the relationship r, Pl> Pl, the valve body 71 is urged in the valve closing direction by the control pressure P3c, and the main valves 23, 24, 23A, 24A are held in the closed position.

Therefore, pressure oil does not flow to either the meter-in circuit or the meter-out circuit for the hydraulic actuators 6 and 7,
The hydraulic actuators 6, 7 are held at the stop position.

Next, when the operating lever of the direction switching valve 8 is operated alone, for example, the pilot valve 29 of the first flow control valve 11 is opened according to the operation amount, and a pilot flow is formed in the pilot circuit 25, and the pilot flow is formed. A pilot flow according to the opening of the valve 29 flows. This allows the variable aperture
By the action of 44 and the back pressure chamber 36, the main valve body 35 opens to an opening proportional to the pilot flow, and the flow corresponding to the operation amount of the operation lever (opening of the pilot valve 29) flows through the main valve 21 through the inlet port. Outflow from 31 to exit ports 32,45.

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

Therefore, the meter-in circuit for the hydraulic actuator 6
A flow rate flows through 15 and the meter-out circuit 17 according to the operation amount of the operation lever, and the hydraulic actuator 6 is driven at a speed according to the operation amount of the operation lever.

And the first and third flow control valves 11 and 13
In a state where the pilot valves 29 and 31 are opened by a fixed amount and a constant main flow is flowing, for example, the pressure at the outlet port 32 of the first flow control valve 11 increases, and the inlet port 31 and the outlet port
When the pressure difference of 32 is going to decrease, the hydraulic pump 1
Is increased, and the pressure difference between the pressure at the inlet port 31 (discharge pressure of the hydraulic pump 1) and the pressure at the outlet port 32 (load pressure of the hydraulic actuator 6; maximum load pressure) is kept constant. Therefore, a constant flow rate according to the operation amount of the operation lever continues to flow through the main valve 21.

Further, in such an independent operation of the hydraulic actuator 6, the pressure receiving area as, ac, am, az of the pressure compensating valve 33 is set to an arbitrary value other than zero with the proportionality constant γ relating to the self-pressure compensating characteristic in the equation (1). In such a case, the differential pressure ΔPz across the pilot valve 29 is controlled in accordance with a change in the load pressure (self-load pressure) of the hydraulic actuator 6, and self-load pressure compensation is performed.

For example, in the example of the hydraulic excavator described with reference to FIGS. 4 to 8, in the flow control valve related to the swing motor 107, the proportionality constant γ is relatively small negative as shown in FIG. Is set to a value. Therefore, when the revolving unit 101 is driven, the load pressure increases because the revolving unit is an inertial body,
Although the pressure rises above the pressure of the relief valve provided for circuit protection and wastes energy, the differential pressure ΔPz decreases as the turning load pressure increases by setting the proportional constant γ to a negative value. And the flow through the flow control valve is reduced. For this reason, even if the load pressure rises, the amount discarded as a surplus flow rate from the relief valve is reduced, and waste of energy can be reduced.

In the flow control valve for the bottom side of the bucket cylinder 110, the proportionality constant γ is set to a relatively small positive value as shown in FIG. 8 (C). Therefore, as the self-load pressure increases in the excavation work, the differential pressure ΔPz increases, and the flow rate through the flow control valve increases. That is, the excavation speed of the bucket increases. Thereby, a strong excavation operation feeling can be obtained, and workability is improved.

Next, when both of the operation levers of the direction switching valves 8 and 9 are operated, the following operation is performed. First, the case where the hydraulic actuator is operated alone at both the first and third flow control valves 11 and 13 of the directional control valve 8 and the first and third flow control valves 11A and 13A of the directional control valve 9, for example. Similarly, the pilot flow according to the operation amount flows, and the operation of the variable throttles 44, 76 and the back pressure chambers 36, 77 corresponds to the operation amount of the operation lever (opening of the pilot valves 29, 31 and 39A, 39A). Flow is main valve 2
Flow through 1,22 and 21A, 22A. Therefore, the hydraulic actuators 6 and 7 are simultaneously driven.

In such a combined operation of the hydraulic actuators 6 and 7, the pressure compensating valves 33 and 3 of the first flow control valves 11 and 11A are used.
By setting the pressure receiving areas as, ac, am, and az of 3A such that the proportionality constant α in the first term on the right side of the above equation (1) becomes an arbitrary positive value as shown in FIG. Pressure compensation and diversion functions are performed.

Therefore, for example, in the hydraulic shovel described with reference to FIGS. 4 to 8, in the operating state in which the load sensing type pump regulator 10 is functioning effectively, each working member is controlled by the operation amount of the operation lever. Driving is performed at a constant flow rate corresponding to the operation, and a stable composite operation can be performed. The sum of the consumption flow rates of the hydraulic actuators 6 and 7 is equal to the hydraulic pump 1
Pump flow rate is larger than the maximum discharge flow rate of
When the operating state becomes such that the 10 does not function effectively, the hydraulic oil can be reliably supplied not only to the low-pressure side hydraulic actuator but also to the high-pressure side hydraulic actuator. The working member can be reliably 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. Accordingly, especially when the pressure receiving areas as, ac, am, and az of the pressure compensating valves 33 and 33A are set to the proportional constants β and γ in the above equation (1), the movement locus of the working member Can be accurately controlled according to the operation amount of the operation lever. For example, in the flow control valve related to the rod side of the boom cylinder 108 and the flow control valve related to the rod side of the arm cylinder 109, FIG.
As shown in FIG. 8 (B), β = 0 and 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 108 and the arm cylinder 109 can be proportionally distributed accurately, and accurate slope formation can be performed.

In the configuration of the present invention, the auxiliary valve is installed in the pilot circuit instead of the main circuit. Therefore, even if the pressure of the hydraulic circuit is increased, liquid leakage is small, and even if a large flow rate is supplied to the main circuit, almost no throttle loss occurs.

Further, the pressure receiving areas as, ac, am and az of the pressure compensating valves 33 and 33A are set so that the proportional constant β and / or the proportional constant γ in the above-mentioned equation (1) become any value other than zero. In the case, the main valve 21 or 21A depending on the maximum load pressure Plmax of another hydraulic actuator based on the above-described pressure compensation and branch function.
A harmonic function and / or a self-load pressure compensation to change the main flow through the pump is performed.

For example, in the example of the hydraulic excavator described with reference to FIGS. 4 to 8, in the flow control valve related to the swing motor 107, the proportionality constant β is set to β = 0 as shown in FIG.
In the flow control valve on the bottom side of the boom cylinder 108, the proportionality constant β is set to an arbitrary positive value as shown in FIG. 7 (B). 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. Also, for example, when turning and raising the boom in a trench excavation work with a backhoe shovel, even if the operator simultaneously operates the operating levers for turning and raising the boom to a full stroke for the sake of simplicity of operation, initially the rising speed of the boom is reduced. It is preferable that the speed of the boom and the swing can be automatically adjusted so that the swing speed rises quickly with respect to the swing speed and 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 speed of the boom and the turning can be automatically adjusted, and the burden on the operator can be reduced.

In the flow control valve for the bottom side of the arm cylinder 109, the proportionality constant β is set to a relatively small positive value as shown in FIG. 7 (C). When performing excavation work 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. Therefore, the pressure oil is throttled when flowing through the flow control valve, and the energy loss increases. Therefore, the fuel efficiency is also deteriorated, and the heat balance of the pressure oil is also deteriorated. As described above, by setting the proportionality constant β within a range in which the balance of the combined operation is not hindered, the differential pressure Plmax−Pl
The opening of the main valve is increased in accordance with the increase of the oil pressure, and the degree of restricting the oil is reduced. Thereby, deterioration of fuel efficiency and heat balance can be reduced.

Further, in the flow control valve on the bottom side of the bucket cylinder 110, the proportionality constant β is set to a relatively small negative value as shown in FIG. 7 (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 proportionality 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, thereby reducing the pilot flow rate and reducing 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 device of the present embodiment, by appropriately setting the pressure receiving area of the pressure compensating valve and setting the constants α, β, and γ to predetermined values, the flow dividing function or the flow dividing function is based. A harmonizing function and / or a self-pressure compensating function can be obtained, and the characteristics of the flow control valve can be modified according to the type and working form of the working member of the hydraulic construction machine.

Further, in the hydraulic drive device of the present embodiment, the pressure compensating valve as the auxiliary valve is arranged not in the main circuit but in the pilot circuit. Therefore, it is possible to provide a hydraulic circuit suitable for increasing the pressure with a small amount of liquid leakage, and even when a large flow rate is supplied to the main circuit, there is little throttle loss at the auxiliary valve portion, and the system is economically excellent.

In the above embodiment, referring to FIG. 6 to FIG.
An example is shown in which the constants β and γ in the above equation (1) are set to predetermined values other than zero for specific ones of the flow control valves relating to the swing body, the boom, the arm, and the bucket of the hydraulic excavator. However, the present invention is not limited to this, and the constants β and γ can be set to zero for all the flow control valves. Even in this case, the constant α in the above equation (1) is set to a positive value, especially By setting α ≦ K to a positive value, the above-described pressure compensation and shunt function can be obtained in a circuit configuration with less liquid leakage and less pressure loss.

Another Embodiment Next, another embodiment of the present invention will be described with reference to FIGS. 9 and 10. FIG. In these drawings, the same members as those in the embodiment shown in FIGS. 1 and 2 are denoted by the same reference numerals.

In the embodiment described above, the discharge pressure Ps of the hydraulic pump, the maximum load pressure Plmax, the inlet pressure Pz and the outlet pressure Pc of the pilot valve are directly used for controlling the pressure compensating valve. However, these four pressures have a correlation with each other using the control pressure of the main valve back pressure chamber as a medium, and the pressure compensating valve can be controlled without directly using these four pressures. Can be given. FIGS. 9 and 10 show an embodiment in which the above four pressures are not directly used for controlling the pressure compensating valve from such a viewpoint.

That is, in FIGS. 9 and 10, the pressure compensating valve 121 disposed in the pilot circuit 25 of the flow control valve 120 includes a spool type valve element 124 for controlling communication between the inlet port 122 and the outlet port 123, and a valve. A first hydraulic operating chamber 125 for urging the body 124 in the valve opening direction, and second, third, and third hydraulic chambers located opposite to the first hydraulic operating chamber 125 for urging the valve body 124 in the valve closing direction. 4th
And hydraulic operating chambers 126, 127, and 128. The first hydraulic operation chamber 125 is connected to the pilot valve 29 outlet side of the pilot circuit 25 via a pilot line 129, and the second hydraulic operation chamber 126 is connected to the pilot line The third hydraulic operating chamber 127 is connected to the outlet port 32 of the main valve 21 via a pilot line 131, and the fourth hydraulic operating chamber 128 is connected to the maximum load pressure via a pilot line 132. It is connected to conduit 61.
As a result, the outlet pressure of the pilot valve 29 (the control pressure of the main valve back pressure chamber 36) Pc is guided to the first hydraulic operation chamber 125,
The inlet pressure Pz of the pilot valve 29 is guided to the hydraulic operating chamber 126 of
Alternatively, the load pressure Pl of 7 is guided, and the maximum load pressure Plmax of the hydraulic actuators 6, 7 is guided to the fourth hydraulic operation chamber 128.

The end face of the valve element 124 facing the first hydraulic operation chamber 125 defines a pressure receiving area ac for receiving the outlet pressure Pc of the pilot valve 29, and the annular end face of the valve element 124 facing the second hydraulic operation chamber 126 has The pressure receiving area az for receiving the inlet pressure Pz of the pilot valve 129 is defined, and the annular end face of the valve element 124 facing the third hydraulic operation chamber 127 defines the pressure receiving area al for receiving the load pressure Pl of the hydraulic actuator 6 or 7. , The valve element 1 facing the fourth hydraulic operating chamber 128
The end face 24 defines a pressure receiving area am for receiving the maximum load pressure Plmax. These pressure receiving areas ac, az, al, am are set so as to obtain the following proportional constants α, β, γ, as in the first embodiment.

The pressure balance formula of the valve element 124 in the pressure compensating valve 121 is represented by the following formula.

acPc = azPz + alPl + amPlmax The pressure balance equation of the valve element 35 in the main valve 21 is represented by the following equation.

AcPc = AsPs + AlPl From the above two formulas, a formula for obtaining the differential pressure across pilot valve 29 is derived, and when deformed using Ac = As + Al Therefore, Then, Pz−Pc = α (Ps−Plmax) + β (Plmax−Pl) + γPl (4) Here, the differential pressure across pilot valve 29 is ΔPz
Then, Pz−Pc = ΔPz. Therefore, the same equation as equation (1) in the embodiment shown in FIG. 1 is obtained.

Therefore, also in the present embodiment, by setting the proportional constants α, β, γ to predetermined values, the pressure difference ΔPz before and after the pilot valve 29 can be reduced by the discharge pressure Ps of the hydraulic pump 1 and the maximum load pressure.
The pressure can be controlled in proportion to the three elements of the differential pressure Ps-Plmax from Plmax, the differential pressure Plmax-Pl between the maximum load pressure Plmax and the self-load pressure Pl, and the self-load pressure Pl. (The first term on the right side), a harmony function (second term on the right side) and a self-compensation function (third term on the right side) in a combined operation based on the pressure compensation and branch functions can be obtained. That is, in this embodiment, instead of directly using the inlet pressure Pz and the outlet pressure Pc of the pilot valve 29, the discharge pressure Ps of the hydraulic pump 1, and the maximum load pressure Plmax, the inlet pressure Pz, the outlet pressure Pc, the self-load pressure Pl, The same effect as using the above four pressures Pz, Pc, Ps, and Plmax is obtained by using the maximum load pressure Pmax.

Still another embodiment of the present invention will be described with reference to FIG. 11 and FIG. In the above embodiment, the pressure compensating valve is disposed on the pilot circuit 29 inlet side of the pilot circuit. However, the pressure compensating valve can also be arranged at the pilot valve outlet side of the pilot circuit, that is, between the pilot valve and the main valve back pressure chamber.
FIG. 11 and FIG. 12 show such an embodiment.

That is, in FIGS. 11 and 12, the flow control valve 140 has a pressure compensating valve 141 connected to the pilot circuit 25 between the pilot valve 29 and the back pressure chamber 36 of the main valve 21. A valve body 144 of a seat valve type for controlling communication between the inlet port 142 and the outlet port 143; first and second hydraulic operation chambers 145 and 146 for urging the valve body 144 in the valve opening direction;
And fourth hydraulic operating chambers 147, 14 for urging the valve in the valve closing direction.
8 and the first hydraulic operation chamber 145 is an outlet port of the main valve 21.
32, a second hydraulic operating chamber 146 is formed in an inlet portion 150 communicating with an inlet port 142 of the pressure compensating valve 141, and the third hydraulic operating chamber 147 has a maximum load pressure. Connected to the pipe 61 via the pilot pipe 151,
The fourth hydraulic operation chamber 148 is connected to the back pressure chamber 36 of the main valve 21 via a pilot line 152. As a result, the load pressure Pl of the hydraulic actuator 6 or 7 is guided to the first hydraulic operation chamber 145, and the pilot valve 2 is supplied to the second hydraulic operation chamber 146.
The 9 outlet pressure Pz is led, the maximum load pressure Plmax is led to the third hydraulic operating chamber 147, and the control pressure Pc of the main valve back pressure chamber 36 is led to the fourth hydraulic operating point 148.

The annular end face of the valve body 144 facing the first hydraulic operation chamber 145 defines a pressure receiving area al for receiving the load pressure Pl of the hydraulic actuator 6 or 7, and the valve body 144 of the valve body 144 facing the second hydraulic operation chamber 146. The end face defines a pressure receiving area az for receiving the outlet pressure Pz of the pilot valve 29, and the valve body 1 facing the third hydraulic operation chamber 147.
44 annular end face receives the maximum load pressure Plmax
And the end face of the valve body 144 facing the fourth hydraulic operating chamber 148 defines a pressure receiving area ac for receiving the control pressure Pc of the back pressure chamber 36. These pressure receiving areas al, az, am, and ac are set so as to obtain the following proportional constants α, β, and γ, as in the above embodiment.

The pressure balance equation of the valve element 144 in the pressure compensating valve 141 is represented by the following equation.

acPc + amPlmax = alPl + azPz The pressure balance equation of the valve element 35 in the main valve 21 is represented by the following equation.

AcPc = AsPs + AlPl From the above two formulas, a formula for obtaining the differential pressure across pilot valve 29 is derived, and when deformed using Ac = As + Al Therefore, Then, Ps−Pz = α (Ps−Plmax) + β (Plmax−Pl) + γPl (5) Here, the differential pressure across pilot valve 29 is ΔPz
Then, Ps−Pz = ΔPz. Therefore, the same equation as equation (1) in the embodiment shown in FIG. 1 is obtained.

Therefore, also in the present embodiment, by setting the proportional constants α, β, γ to predetermined values, the pressure difference ΔPz before and after the pilot valve 29 can be reduced by the discharge pressure Ps of the hydraulic pump 1 and the maximum load pressure.
Plmax-Plmax, the pressure difference between the maximum load pressure Plmax and the self-load pressure Pl, Plmax-Pl, and the self-load pressure Pl. The first term on the right side), the harmony function in the composite operation based on the pressure compensation and the branching function (second on the right side)
Term), a self-pressure compensation function (third term on the right side) can be obtained. That is, in the present embodiment, by arranging the pressure compensating valve 141 on the outlet side of the pilot valve 29, it is possible to obtain the same effect as that provided on the inlet side.

FIGS. 13 and 14 show another embodiment in which a pressure compensating valve is arranged at the outlet side of the pilot valve, and the control thereof does not directly use the inlet pressure and outlet pressure of the pilot valve, the discharge pressure of the hydraulic pump and the maximum load pressure. Is shown.

13 and 14, the pressure compensating valve 161 arranged in the pilot circuit 25 of the flow control valve 160 has an inlet port 1
A seat-type valve body 164 for controlling the communication between the valve body 62 and the outlet port 163; first and second hydraulic operation chambers 165 and 166 for urging the valve body 164 in the valve opening direction; Room 165,1
The third and fourth hydraulic operation chambers 167 and 168 urge the valve body 164 in the valve closing direction. The first hydraulic operation chamber 165 is connected to the main valve 21 via the pilot line 169.
The second hydraulic operating chamber 166 is formed in an inlet portion 179 communicating with the inlet port 162 of the pressure compensating valve 161, and the third hydraulic operating chamber 167 is connected via a pilot line 171. The fourth hydraulic operating chamber 168 is connected to the maximum load pressure line 61, and is connected to the pilot valve 29 inlet side of the pilot circuit 25 via the pilot line 172. Thereby, the load pressure Pl of the hydraulic actuator 6 or 7 is guided to the first hydraulic operation chamber 165, the discharge pressure Ps of the hydraulic pump 1 is guided to the second hydraulic operation chamber 166, and the third hydraulic operation chamber The maximum load pressure Plmax of the hydraulic actuators 6 and 7 is guided to 167, and the inlet pressure Pz of the pilot valve 29 is guided to the fourth hydraulic operation chamber 168.

The annular end face of the valve body 164 facing the first hydraulic operating chamber 165 defines a pressure receiving area al for receiving the load pressure Pl of the hydraulic actuator 6 or 7, and the valve body 164 of the valve body 164 facing the second hydraulic operating chamber 166. The end surface defines a pressure receiving area as for receiving the discharge pressure Ps of the hydraulic pump 1, and the valve body 164 facing the third hydraulic operation chamber 167.
The annular end face defines the pressure receiving area am receiving the maximum load pressure Plmax, and the end face facing the fourth hydraulic operating chamber 168 defines the pressure receiving area az receiving the inlet pressure Pz of the pilot valve 29. These pressure receiving areas al, as, am, and az are set so as to obtain the following proportional constants α, β, and γ, as in the above-described embodiment.

The pressure balance formula of the valve body 164 in the pressure compensating valve 161 is represented by the following formula.

azPz + amPlmax = alPl + asPs The pressure balance equation of the valve element 35 in the main valve 21 is expressed by the following equation.

AcPc = AsPs + AlPl From the above two formulas, a formula for obtaining the differential pressure across pilot valve 29 is derived, and when deformed using Ac = As + Al Therefore, Then, Pz−Pc = α (Ps−Plmax) + β (Plmax−Pl) + γPl (6) Here, the differential pressure across pilot valve 29 is ΔPz
Then, Pz−Pc = ΔPz. Therefore, the same equation as equation (1) in the embodiment shown in FIG. 1 is obtained.

Therefore, also in the present embodiment, by setting the proportional constants α, β, γ to predetermined values, the pressure difference ΔPz before and after the pilot valve 29 can be reduced by the discharge pressure Ps of the hydraulic pump 1 and the maximum load pressure.
The pressure can be controlled in proportion to the three elements of the differential pressure Ps-Plmax from Plmax, the differential pressure Plmax-Pl between the maximum load pressure Plmax and the self-load pressure Pl, and the self-load pressure Pl. (The first term on the right side), a harmony function (second term on the right side) and a self-compensation function (third term on the right side) in a combined operation based on the pressure compensation and branch functions can be obtained.

Still another embodiment of the present invention will be described with reference to FIG. 15 and FIG. All of the above embodiments are 4
One pressure is used. However, the discharge pressure of the hydraulic pump, the maximum load pressure, and the four pressures of the inlet pressure and the outlet pressure of the pilot valve have a correlation with each other using the control pressure of the main valve back pressure chamber as a medium. Thus, the pressure compensating valve can be controlled, and the above-described characteristics can be given to the pressure compensating valve. FIG. 15 and FIG. 16 show such an embodiment.

That is, in FIGS. 15 and 16, the flow control valve 180 has a pressure compensating valve 181 disposed in the pilot circuit 25 between the pilot valve 29 and the main valve back pressure chamber 36.
Is a seat-type valve body 184 for controlling communication between the inlet port 182 and the outlet port 183, a first hydraulic operating chamber 185 for urging the valve body 184 in the valve opening direction, and a first hydraulic operating chamber 185. It comprises second and third hydraulic operation chambers 186 and 187 which are opposed to each other and bias the valve body 184 in the valve closing direction. The first hydraulic operating chamber 185 is formed in an inlet portion 188 communicating with the inlet port 182 of the pressure compensating valve 181, and the second hydraulic operating chamber 186 is connected to an inlet of the pilot valve 29 of the pilot circuit 25 through a pilot line 189. Side or the main valve 21 inlet side of the meter-in circuit 15,
The third hydraulic operating chamber 187 is connected to the maximum load pressure line 61 via a pilot line 190. As a result, the outlet pressure Pz of the pilot valve 29 is guided to the first hydraulic operating chamber 185, and the discharge pressure P of the hydraulic pump 1 is applied to the second hydraulic operating chamber 186.
s is led to the third hydraulic operating chamber 187 and the maximum load pressure Plma
x is derived.

The end face of the valve body 184 facing the first hydraulic operation chamber 185 defines a pressure receiving area Az for receiving the outlet pressure Pz of the pilot valve 29, and the annular end face of the valve body 184 facing the second hydraulic operation chamber 186 has The pressure receiving area as for receiving the discharge pressure Ps of the hydraulic pump 1 is defined, and the end face of the valve body 184 facing the third hydraulic operation chamber 187 defines the pressure receiving area am for receiving the maximum load pressure Plmax. These pressure receiving areas az, as, am are set so as to obtain the following proportional constants α, β, γ, similarly to the above-described embodiment.

The pressure balance formula of the valve element 184 in the pressure compensating valve 181 is represented by the following formula.

azPz = asPs + amPlmax The pressure balance equation of the valve element 35 in the main valve 21 is represented by the following equation.

AcPc = AsPs + AlPl A formula for obtaining the differential pressure across the pilot valve 29 is derived from the above two equations, and when deformed using Ac = Ac + Al, az (Ps−Pz) = (az−as) (Ps−Plmax) + (az−as) −am) (Plmax−Pl) + (az−as−am) Pl Then, Ps−Pz = α (Ps−Plmax) + β (Plmax−Pl) + γPl (7) Here, the differential pressure across pilot valve 29 is ΔPz
Then, Pz−Pc = ΔPz. Therefore, the same equation as equation (1) in the embodiment shown in FIG. 1 is obtained. However, in this embodiment, since β and γ have the same value, they cannot be determined independently.

Therefore, also in the present embodiment, by setting the proportional constants α, β, γ to predetermined values, the pressure difference ΔPz before and after the pilot valve 29 can be reduced by the discharge pressure Ps of the hydraulic pump 1 and the maximum load pressure.
The pressure can be controlled in proportion to the three elements of the differential pressure Ps-Plmax from Plmax, the differential pressure Plmax-Pl between the maximum load pressure Plmax and the self-load pressure Pl, and the self-load pressure Pl. (The first term on the right side), a harmony function (the second term on the right side) and a self-pressure compensation function (the third term on the right side) in a combined operation based on the pressure compensation and branching functions can be 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 performs the pressure compensating and branching function, or A harmonic function and / or a self-pressure compensation function based on a pressure compensation and a flow dividing function can be selectively achieved. Since the four pressures have a correlation with each other using the control pressure of the main valve back pressure chamber as a medium, the pressure compensating valve can be used without using the four pressures directly, and the pressure compensating valve can be provided on the inlet side and the outlet side of the pilot valve. In any case, the pressure compensating valve can be controlled. Also, the pressure compensating valve can be controlled using a pressure other than four.

Next, another embodiment of the present invention relating to the pump control means will be described. First, in the above embodiments, the hydraulic drive device of the present invention will be described in combination with a load sensing type pump regulator, and the load sensing type pump regulator will be described as controlling the discharge pressure of a variable displacement type hydraulic pump. However, the hydraulic pump may be of a fixed displacement type. In this case, the load sensing regulator type pump regulator is configured as shown in FIG. That is, in FIG.
Is a relief valve 383 having opposed pilot chambers 381,382
The pilot chamber 381 guides the discharge pressure of the fixed displacement hydraulic pump 385 via a pilot line 348, the pilot chamber 382 guides the maximum load pressure via a pilot line 386, The spring 387 is arranged on the side of.
Accordingly, the discharge pressure of the hydraulic pump 385 can be maintained higher than the maximum load pressure of the plurality of hydraulic actuators by a pressure corresponding to the strength of the spring 387.

Further, the hydraulic drive device of the present invention can 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. 18, a hydraulic pump 390 is connected to a flow control valve 391 composed of a combination of the above-described main valve, pilot valve, and pressure compensating valve, and the 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
For one, an operating device 394 is provided. Operation device 39
The operation signal of 4 is sent to the control device 395, from which it is further sent as a control signal to the pilot valve drive section 396 of the flow control valve 391 to control the opening 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 apparatus 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 include a simple relief valve for keeping the discharge pressure of the hydraulic pump constant.

〔The invention's effect〕

According to the present invention, by appropriately setting the constants α, β, and γ to predetermined values, the pressure compensation and branching function, or the harmonic function and / or the self-pressure compensation function based on the pressure compensation and branching function can be selected. The characteristics of the flow control valve can be set to an optimum state according to the type and working form of the working member of the hydraulic construction machine.

Therefore, for example, in an application example to a hydraulic excavator,
Even if the operating levers for turning and raising the boom are simultaneously operated up to the full stroke, the boom ascending speed initially increases faster than the turning speed, and when the boom rises to a certain extent, the turning speed gradually increases, and the turning speed reaches the maximum speed. When excavation is performed by the combined operation using the arm, the arm is reliably driven, and the hydraulic pressure for the arm is increased. When the actuator is on the low pressure side, the flow control valve characteristics to prevent the deterioration of fuel efficiency and heat balance, the flow control at the moment when the basket is released from the excavation load and goes out to the surface during the trenching work by the combined operation using the bucket, Flow control valve characteristics that reduce the flow rate through the valve and reduce shock, reduce the flow rate from the relief valve during turning acceleration, -Flow control valve characteristics to reduce waste of consumption, flow control valve characteristics to obtain a strong excavation operation feeling during excavation work using buckets, accurate method for slope slope formation work using booms and arms It is possible to obtain characteristics of a flow control valve for performing surface formation.

Also, since the auxiliary valve is located in the pilot circuit,
A hydraulic circuit having a small throttle loss and an energy-saving structure can be provided.

Further, when the first constant α is set in a relation of α ≦ K with respect to a ratio K of a pressure receiving area receiving the load pressure of the hydraulic actuator to a pressure receiving area receiving the control pressure of the seat type main valve back pressure chamber. In the flow dividing function, the flow rate proportional to the operation amount of the operating means (the pilot valve opening) can be accurately divided. 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.

[Brief description of the drawings]

FIG. 1 is a schematic diagram showing the overall configuration of a hydraulic drive device according to one embodiment of the present invention, and FIG. 2 is a cross-sectional view showing the structure of a flow control valve connected to a meter-in circuit of the hydraulic drive device. FIG. 3 is a sectional view showing a structure of a flow control valve connected to a meter-out circuit of the hydraulic drive device, and FIG. 4 is a side view of a hydraulic shovel to which the hydraulic drive device of the present invention is applied. FIG. 5 is a top view of the hydraulic excavator, and FIG. 6 is a characteristic diagram showing an example of setting a proportionality constant α of a pressure compensating valve included in one flow control valve of the hydraulic drive device. 7 (A) to 7 (D) are characteristic diagrams showing examples of setting a proportionality constant β of a pressure compensating valve included in one flow control valve of the hydraulic drive device, and FIGS. 8 (A) to 8 (C). ) Is the setting of the proportionality constant γ of the pressure compensating valve included in one flow control valve of the hydraulic drive. FIG. 9 is a schematic diagram of a flow control valve connected to a meter-in circuit of a hydraulic drive device according to another embodiment of the present invention, and FIG. 10 shows a structure of the flow control valve. FIG. 11 is a cross-sectional view, FIG. 11 is a schematic view of a flow control valve connected to a meter-in circuit of a hydraulic drive device according to still another embodiment of the present invention, and FIG. 12 is a cross-sectional view showing the structure of the flow control valve. FIG. 13 is a schematic view of a flow control valve connected to a meter-in circuit of a hydraulic drive device according to still another embodiment of the present invention, and FIG.
Figure is a cross-sectional view showing the structure of the flow control valve, FIG. 15 is a schematic diagram of a flow control valve connected to the meter-in circuit of the hydraulic drive device according to still another embodiment of the present invention,
FIG. 16 is a sectional view showing the structure of the flow control valve,
FIG. 17 is a circuit diagram showing an embodiment of a load sensing type pump regulator when a constant displacement type hydraulic pump is used in the hydraulic drive device of the present invention, and FIG. 18 is a circuit diagram showing a load used in the hydraulic drive device of the present invention. It is a circuit diagram showing an embodiment of a pump control means which is not a sensing type. Description of reference numerals 1; 385; 389: Hydraulic pump 2, 3 ... Main circuit 6, 7; 107-110: Hydraulic actuator 8, 9, 11, 12, 11, 11A, 12A; 120; 140; 160; 180 ... Flow control valve Means 10; 380; 392 Pump control means 21, 22, 21A, 22A Main valve 25, 26, 25A, 26A Pilot circuit 29, 30, 29A, 30A Pilot valve 31 Inlet port, 32 Outlet port 33 , 34,33A, 34A; 121; 141; 161; 181 ... Auxiliary valve 36 ... Back pressure chamber, 44 ... Variable throttle 53-60; 125-132; 145-152; 165-172; 185-190 ... Control means

 ──────────────────────────────────────────────────続 き Continued on the front page (31) Priority claim number Japanese Patent Application No. 62-249902 (32) Priority date October 5, 1987 (33) (33) Priority claim country Japan (JP) (56) References JP-A-60-11706 (JP, A) JP-A-60-172707 (JP, A) JP-A-61-165428 (JP, A) JP-A-2-31003 (JP, A) JP-B-59-8684 (JP, B2)

Claims (21)

(57) [Claims] 1. at least one hydraulic pump; at least first and second hydraulic actuators respectively connected to the hydraulic pump via a main circuit and driven by hydraulic oil discharged from the hydraulic pump; First and second flow 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 a discharge pressure of the hydraulic pump. The first and second flow control valve means are respectively connected in series with the first valve means for changing an opening degree in accordance with the operation amount of the operation means, and the valve is connected to the first valve means; Second valve means for controlling the differential pressure between the inlet pressure and the outlet pressure of the means; and for each of said first and second flow control valve means, the inlet pressure of said first valve means and Outlet pressure, the hydraulic pump On the basis of the discharge pressure and the maximum load pressure of said first and second hydraulic actuators second
A hydraulic drive device having control means for controlling the valve means, wherein the first and second flow control valve means each control a communication between an inlet port and an outlet port connected to the main circuit; A variable throttle that changes the opening in accordance with the displacement of the valve body, and a back pressure chamber that communicates with the outlet port through the variable throttle to generate a control pressure for urging the valve body in the valve opening direction. And a pilot circuit connected between an inlet port of the main valve and a back pressure chamber; the first valve means is connected to the pilot circuit and flows through a pilot circuit. And the second valve means is connected to the pilot circuit and is configured as an auxiliary valve for controlling a differential pressure between an inlet pressure and an outlet pressure of the pilot valve. And the control means is arranged such that, for each of the first and second flow control valve means, the differential pressure between the inlet pressure and the outlet pressure of the pilot valve is the discharge pressure of the hydraulic pump and the first and second hydraulic pressures. The auxiliary valve so that the differential pressure between the maximum load pressure of the actuator, the differential pressure between the maximum load pressure and the self-load pressure of each hydraulic actuator, and the self-load pressure have a relationship represented by the following equation. ΔPz = α (Ps−P1max) + β (P1max−P1) + γP1 where ΔPz: differential pressure between the inlet pressure and the outlet pressure of the pilot valve Ps: discharge pressure of the hydraulic pump P1max: the first and the second Maximum load pressure of the first hydraulic actuator P1: Self-load pressure of each of the first and second hydraulic actuators α, β, γ: First, second, and third constants First, second, and third constants Where the constants α, β, and γ of 3 A hydraulic drive device characterized by being set to a fixed value. 2. The ratio of the pressure receiving area of the main valve body receiving the load pressure of the associated hydraulic actuator through the outlet port to the pressure receiving area of the main valve body receiving the control pressure of the back pressure chamber is represented by K. 2. The hydraulic drive device according to claim 1, wherein the first constant α has a relationship of α ≦ K. 3. The hydraulic drive system according to claim 2, wherein said second and third constants β and γ are each set to zero. 4. The hydraulic drive according to claim 1, wherein the first constant α is set to a positive value corresponding to a proportional gain of an operation amount of the operation means and a main flow rate passing through the main valve. apparatus. 5. The apparatus according to claim 1, wherein the second constant β is set to a value based on the operation characteristics of the related hydraulic actuator and another hydraulic actuator when the combined operation is performed. Hydraulic drive. 6. The hydraulic drive device according to claim 1, wherein the third constant γ is set to a value based on an operation characteristic of an associated hydraulic actuator. 7. The control means comprises: a plurality of hydraulic operation chambers provided in the auxiliary valve of each of the first and second flow control valve means; and a discharge of the hydraulic pump to the plurality of hydraulic operation chambers. Pipeline means for directly or indirectly introducing the pressure, the maximum load pressure, the inlet pressure and the outlet pressure of the pilot valve, and the pressure receiving areas of the plurality of hydraulic operation chambers are respectively defined by the first and second hydraulic operation chambers. The hydraulic drive device according to claim 1, wherein the third constants α, β, and γ are set to the predetermined values. 8. The auxiliary valve is disposed between an inlet port of the main valve and the pilot valve, and the plurality of hydraulic operation chambers are configured to urge the auxiliary valve in a valve opening direction. And a third and a fourth hydraulic operation chamber for urging the auxiliary valve in a valve closing direction, wherein the conduit means controls the discharge pressure of the hydraulic pump to the first hydraulic operation. A first conduit leading to the chamber, a second conduit leading the outlet pressure of the pilot valve to the second hydraulic operating chamber, and a
8. The hydraulic drive device according to claim 7, further comprising a third conduit for guiding the hydraulic pressure to the hydraulic operating chamber, and a fourth conduit for guiding the inlet pressure of the pilot valve to the fourth hydraulic operating chamber. . 9. The auxiliary valve is disposed between the back pressure chamber of the main valve and the pilot valve, and the plurality of hydraulic operation chambers are
A first hydraulic operation chamber for urging the auxiliary valve in a valve opening direction;
There are second, third and fourth hydraulic operating chambers for urging the auxiliary valve in the valve closing direction, and the conduit means guides the outlet pressure of the pilot valve to the first hydraulic operating chamber. A first conduit, a second conduit for directing inlet pressure of the pilot valve to the second hydraulic operating chamber, and a third conduit for directing load pressure of an associated hydraulic actuator to the third hydraulic operating chamber. Pipes,
The hydraulic drive device according to claim 7, further comprising a fourth conduit for guiding the maximum load pressure to the fourth hydraulic operation chamber. 10. The auxiliary valve is disposed between the pilot valve and a back pressure chamber of the main valve, and the plurality of hydraulic operation chambers are configured to urge the auxiliary valve in a valve opening direction. And a third and a fourth hydraulic operation chamber for urging the auxiliary valve in the valve closing direction, wherein the conduit means reduces a load pressure of an associated hydraulic actuator to the first hydraulic operation chamber. A first conduit leading to a hydraulic operating chamber, a second conduit leading outlet pressure of the pilot valve to the second hydraulic operating chamber, and a second conduit leading the maximum load pressure to the third hydraulic operating chamber. 8. The hydraulic drive device according to claim 7, further comprising: a third conduit; and a fourth conduit for guiding a control pressure of the back pressure chamber to the fourth hydraulic operation chamber. 11. The auxiliary valve is disposed between an inlet port of the main valve and the pilot valve, and the plurality of hydraulic operation chambers are configured to urge the auxiliary valve in a valve opening direction. And a third and a fourth hydraulic operation chamber for urging the auxiliary valve in a valve closing direction, wherein the conduit means reduces a load pressure of an associated hydraulic actuator to the first hydraulic pressure. A first conduit leading to an operation chamber, a second conduit leading discharge pressure of the hydraulic pump to the second hydraulic operation chamber, and a second conduit leading the maximum load pressure to the third hydraulic operation chamber. The hydraulic drive device according to claim 7, further comprising a third pipeline and a fourth pipeline for guiding an inlet pressure of the pilot valve to the fourth hydraulic operation chamber. 12. The auxiliary valve is disposed between the pilot valve and a back pressure chamber of the main valve, and the plurality of hydraulic operation chambers are provided with a first hydraulic pressure for urging the auxiliary valve in a valve opening direction. An operating chamber, and second and third hydraulic operating chambers for urging the auxiliary valve in a valve closing direction, wherein the conduit means sends an outlet pressure of the pilot valve to the first hydraulic operating chamber. A first conduit for guiding;
A second conduit for guiding the discharge pressure of the hydraulic pump to the second hydraulic operation chamber; and a third conduit for guiding the maximum load pressure to the third hydraulic operation chamber. The hydraulic drive device according to claim 7. 13. The pump control means is a load sensing type pump regulator for maintaining a discharge pressure of a hydraulic pump higher than a maximum load pressure of the first and second hydraulic actuators by a predetermined value. The hydraulic drive device according to claim 1. 14. At least one hydraulic pump; a plurality of hydraulic actuators each connected to the hydraulic pump via a main circuit and driven by hydraulic oil discharged from the hydraulic pump; A plurality of working members each including a rotating body, a boom, an arm and a bucket, each driven; 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 a discharge pressure of a hydraulic pump; the plurality of flow control valve means each having a first valve means for changing an opening degree according to an operation amount of an operation means; A second valve means connected in series with the valve means for controlling a differential pressure between an inlet pressure and an outlet pressure of the valve means; and the plurality of flow control valve means. Control means for controlling the second valve means based on the inlet pressure and the outlet pressure of the first valve means, the discharge pressure of the hydraulic pump and the maximum load pressure of the first and second hydraulic actuators, respectively In the hydraulic construction machine having, the plurality of flow control valve means, respectively, a valve body for controlling the communication of the inlet port and the outlet port connected to the main circuit, the opening degree corresponding to the displacement of the valve body A main valve having a variable throttle to be changed, a back pressure chamber communicating with the outlet port through the variable throttle, and generating a control pressure for urging the valve body in a valve opening direction; and an inlet port of the main valve. A pilot circuit connected between the pilot circuit and the back pressure chamber; the first valve means is configured as a pilot valve connected to the pilot circuit and controlling a pilot flow flowing through the pilot circuit. And the second valve means is connected to the pilot circuit and is configured as an auxiliary valve for controlling a differential pressure between an inlet pressure and an outlet pressure of the pilot valve; For each of the flow control valve means for at least two working members of a boom, an arm and a bucket, a differential pressure between an inlet pressure and an outlet pressure of the pilot valve is a discharge pressure of the hydraulic pump and a maximum load pressure of the plurality of hydraulic actuators. And the differential pressure between the maximum load pressure and the self-load pressure of each hydraulic actuator, and the self-load pressure, by controlling the auxiliary valve so as to have a relationship represented by the following equation: ΔPz = Α (Ps−P1max) + β (P1amx−P1) + γP1 where ΔPz: differential pressure between the inlet pressure and the outlet pressure of the pilot valve Ps: discharge pressure of the hydraulic pump P1max: the plurality of pressures Maximum load pressure of pressure actuator P1: Self-load pressure of each of the plurality of hydraulic actuators α, β, γ: first, second, and third constants First, second, and third constants α, β, γ Is set to a predetermined value. 15. The ratio of the pressure receiving area of the main valve element receiving the load pressure of the associated hydraulic actuator via the outlet port to the pressure receiving area of the main valve element receiving the control pressure of the back pressure chamber is represented by K. 15. The hydraulic construction machine according to claim 14, wherein the first constant α has a relation of α ≦ K. 16. The flow control valve means for the bottom side of the boom hydraulic actuator,
15. The hydraulic construction machine according to claim 14, wherein the second constant β is set to a positive value. 17. The flow control valve means for a bottom side of the arm hydraulic actuator, wherein:
15. The hydraulic construction machine according to claim 14, wherein the second constant β is set to a positive value. 18. The hydraulic construction machine according to claim 14, wherein the control means sets the second constant β to a negative value for the flow control valve on the bottom side of the hydraulic actuator for bucket. . 19. The hydraulic construction machine according to claim 14, wherein said control means sets the third constant γ to a negative value for the flow control valve for the hydraulic actuator for the swing body. 20. The hydraulic construction machine according to claim 14, wherein the control means sets the third constant γ to a positive value for the flow control valve on the bottom side of the hydraulic actuator for bucket. . 21. The control means sets the second and third constants β and γ to zero for the flow control valve on the rod side of the boom and arm hydraulic actuator, respectively. 14. The hydraulic construction machine according to 14 or 15.