JP2644533B2 - Hydraulic drive - Google Patents

Description

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

[Conventional technology]

Conventionally, a hydraulic drive device of a hydraulic construction machine equipped with a plurality of hydraulic actuators such as a hydraulic shovel and a hydraulic crane,
Generally, at least one hydraulic pump, a plurality of hydraulic actuators each connected to the hydraulic pump via a main circuit, and driven by pressure oil discharged from the hydraulic pump; And a plurality of flow control valves connected to the respective main circuits therebetween.

In USP 4,614,854, 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]

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 releap 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 pressurized oil can be supplied to the boom as much as possible without relieving, the waste of energy can be reduced, and the rising speed of the boom can be reduced at first. Can be automatically increased with respect to the turning speed, and the speed of the boom and turning can be automatically adjusted such that the turning speed gradually increases when the boom rises to some extent.

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

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 compensation 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 It is only a technique to correct the pressure compensation function by taking into account the effect of the self-load pressure of the hydraulic actuator, and it is irrelevant to modify the functions for the combined operation of multiple hydraulic actuators, especially the effect of the load pressure of other hydraulic actuators. There was no idea at all to modify the pressure compensation function and the diversion function by incorporating the pressure.

An object of the present invention is to provide a hydraulic drive device capable of modifying the characteristics of a flow control valve in accordance with the type and work form of a working member of a hydraulic construction machine.

An object of the present invention is to provide a hydraulic shovel provided with a hydraulic drive device capable of correcting the characteristics of a flow control valve in accordance with the type and working form of a working member.

[Means for solving the problem]

In order to achieve the above object, the present invention provides at least one
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 the second hydraulic actuators, and the discharge pressure of the hydraulic pump is set to be lower than the maximum load pressure of the first and second hydraulic actuators. Pump control means for maintaining the pressure at a predetermined value higher, the first and second 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 to the first valve means for controlling a differential pressure between an inlet pressure and an outlet pressure of the valve means, wherein the second valve means is provided with the first valve means. Means inlet pressure is closed The introduced as load 1
And a second control pressure chamber which is introduced so that the outlet pressure of the second valve means is loaded in the valve opening direction, wherein the first and second hydraulic pressures are provided. First and second pressure detecting means for respectively detecting a load pressure of the actuator, and a first differential for detecting a differential pressure between a discharge pressure of the hydraulic pump and a maximum load pressure of the first and second hydraulic actuators. Pressure detecting means, second differential pressure detecting means for detecting a differential pressure between the maximum load pressure and the load pressure of the first hydraulic actuator, and the maximum load pressure and the load pressure of the second hydraulic actuator. A third differential pressure detecting means for detecting a differential pressure between the first and second flow rate control valve means, and a third differential pressure detecting means provided in the second valve means of each of the first and second flow control valve means.
And the pressure signals of the first and second pressure detecting means and the differential pressure signals of the first, second and third differential pressure detecting means are inputted, Control means for calculating each control amount and outputting a control pressure based on the control amount to each of the third control pressure chambers of the second valve means to control the second valve means. Then, the control means introduces a control pressure based on the control amount into the third control pressure chamber for each of the second valve means,
The differential pressure between the inlet pressure and the outlet pressure of the first valve means controlled by the second valve means is the differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure, The second valve means is controlled as expressed by the following equation with respect to the differential pressure from the self-load pressure of the hydraulic actuator and the self-load pressure, and ΔPz = α (Ps−Plmax) + B (Plmax−Pl ) + ΓPl where ΔPz: differential pressure between the inlet pressure and the outlet pressure of the first valve means Ps: discharge pressure of the hydraulic pump Plmax: maximum load pressure of the first and second hydraulic actuators Pl: the first Self-load pressures α, β, γ of first and second hydraulic actuators: first, second, and third constants The first, second, and third constants α, β, γ are provided to the control means, respectively. Provided is a hydraulic drive device which is set as a predetermined value.

In the above configuration, 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 first valve means, and the second constant β is
The third constant γ is set to a value based on the operation characteristics of the related hydraulic actuator when the related hydraulic actuator and another hydraulic actuator are operated in combination with each other when the combined operation is performed. . The second and third constants β and γ can be set to zero.

The second valve means preferably has a spring acting in the valve opening direction opposite to the third control pressure chamber.

The first and second flow control valve means each include a valve body that controls communication between an inlet port and an outlet port connected to the main circuit, and a variable valve that changes an opening corresponding to a displacement of the valve body. A restrictor and a seat-type main valve having a back pressure chamber communicating with the inlet port via the variable restrictor and generating a control pressure for urging the valve body in a valve closing direction; and a back pressure of the main valve. A pilot circuit connected between the chamber and an outlet port, wherein the first valve means is arranged as a pilot valve connected to the pilot circuit and controlling a pilot flow through the pilot circuit; May be arranged as an auxiliary valve connected to the pilot circuit and controlling a differential pressure between an inlet pressure and an outlet pressure of the pilot valve.

In this case, preferably, the ratio of the pressure receiving area of the main valve valve body receiving the discharge pressure of the hydraulic pump via the inlet port to the pressure receiving area of the main valve valve body receiving the control pressure of the back pressure chamber is represented by K. Then, the first constant α has a relation of α ≦ K.

The first valve means may be arranged in the main circuit as a spool-type main valve, and the second valve means may be arranged as an auxiliary valve in the main circuit upstream of the main valve.

Further, according to the present invention, in order to achieve the above object, a revolving structure driven by a plurality of hydraulic actuators,
Provided is a hydraulic shovel in which the hydraulic drive circuit is applied to a hydraulic shovel having a plurality of working members including a boom, an arm, and a bucket.

In the above-described hydraulic excavator, the control means preferably sets the second constant β to a positive value for the flow control valve means related to 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 a first valve means for changing an opening in accordance with an operation amount of an operation means in a flow control valve means, and an inlet pressure and an outlet of the valve means which are connected in series to the first valve means. As a result of various studies on the relationship between the pressure and the second valve means for controlling the differential pressure, the differential pressure ΔPz before and after the first valve means controlled by the second valve means is generally expressed by the above equation. Was found to be represented. The formula is reprinted below.

ΔPz = α (Ps−Plmax) + β (Plmax−Pl) + γPl The above expressions have the following meanings. In this equation, Ps-Plmax in the first term on the right side is constant as long as the pump control means is functioning effectively and is common to all flow control valve means. Controls the pressure compensation function to keep the flow rate constant irrespective of the change in the differential pressure and the branch function to surely supply the hydraulic oil to each hydraulic actuator during the combined operation. Plmax-Pl of the second term is the maximum load of other actuators. Since it changes depending on the pressure, it governs the harmony function in the combined operation, and γPl of the third term controls the self-pressure control function because it changes according to the self-load pressure. 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 pressure compensation and branching functions of the first term are the basic functions of the combined 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 third
The self-pressure compensation function in the item is a function added according to the type and work form of the related work member. Therefore, the constants β and γ are set to predetermined values including zero, respectively. By setting α, β, and γ in this manner, it is possible to provide a pressure compensation and branching function, or a harmonic function and / or a self-pressure compensation function based on the pressure compensation and branching function. The characteristics of the flow control valve can be modified according to the type of the member and the working mode.

The first constant α has a meaning of a proportional gain between the operation amount of the operating means (opening degree of the first valve means) and the flow rate. Therefore, the first constant α corresponds to the proportional gain. Set to any positive value.

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 is set to an arbitrary value in consideration of the operation characteristics of the related hydraulic actuator. This is also set to zero, especially if it is preferable not to be affected by the self-load pressure.

If a spring acting in the valve opening direction is installed on the second valve means, the second valve means can be reliably opened even if the control means is broken and the second valve means cannot be controlled. Thus, normal pressure compensation control can be performed.

The first and second flow control valve means are each configured to have a seat-type main valve and a pilot circuit, the first valve means is arranged as a pilot valve, and the second valve means is When it is arranged as an auxiliary valve that controls the differential pressure between the inlet pressure and the outlet pressure of the pilot valve, the auxiliary valve is installed in the pilot circuit instead of the main circuit, and the main valve installed in the main circuit is a seat valve Therefore, it is possible to provide a hydraulic circuit that has little liquid leakage and is suitable for high pressure. Further, since the auxiliary valve is arranged in the pilot circuit, the throttle loss due to the auxiliary valve striking can be reduced even if a large flow rate flows through the main circuit.

When the ratio of the pressure receiving area receiving the discharge pressure of the hydraulic pump to the pressure receiving area receiving the control pressure of the main valve back pressure chamber is K, and the first constant α is set to α ≦ K, The differential pressure obtained by α (Ps−Plmax) falls within the range of the maximum front-back differential pressure that can be taken by the pilot valve on the high load pressure side. The differential pressure becomes substantially the same, and in the flow dividing function, a 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.

In the flow control valve means, wherein the first valve means is disposed as a spool-type main valve in the main circuit, and the second valve means is disposed as an auxiliary valve on the upstream side of the main valve of the main circuit. The present invention is applicable and can provide the pressure compensation and shunt function described above, or a harmonic function and / or a self-pressure compensation function based on the pressure compensation and shunt function.

Further, in the hydraulic shovel of the present invention, the characteristics of the flow control valve relating to at least two working members of the revolving superstructure, the boom, the arm and the bucket can be set and corrected according to the type and working form of the working member. It is possible to add a harmony function and / or a self-pressure compensation function based on the pressure compensation and branching functions described above.

The control means may be configured such that the flow control valve means on the bottom side of the boom hydraulic actuator
Is preferably set to a relatively large positive value. As a result, the maximum flow pressure (swing pressure) and the self-load pressure (boom pressure) are applied to the flow control valve on the bottom side of the hydraulic actuator for the boom, which is the low load side, during the initial acceleration of the swing in the combined operation of the swing and the boom. The flow rate according to the increase in the differential pressure flows, and the rising speed of the boom can be increased.
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.

The control means may be a second flow control valve means associated with a bottom side of the arm hydraulic actuator.
Is preferably set to a relatively small positive value. This allows the arm to be reliably driven when performing digging by a combined operation using the arm, and when the arm hydraulic actuator is on the low pressure side, the maximum load pressure (other hydraulic actuator pressure) and the self-load pressure (arm The opening degree of the flow control valve is increased in accordance with an increase in the pressure difference between the pressure and the pressure, and the degree of restriction of the flow is reduced. As a result, deterioration of fuel efficiency and heat balance is prevented.

Furthermore, it is preferable that the control means sets the second constant β to a relatively small negative value for the flow control valve means related to the bottom side of the bucket hydraulic actuator. As a result, the maximum load pressure (other hydraulic actuator pressure) at the moment when the bucket is released from the excavation load and comes out to the surface during the trenching work by the combined operation using the bucket,
By increasing the pressure difference between the pressure and the self-load pressure (bucket pressure), the flow rate through the flow rate control valve can be reduced, and the shock can be reduced.

Preferably, the control means sets the third constant γ to a relatively small negative value for the flow control valve means relating to the revolving superstructure. Thereby, at the time of turning acceleration, the passing flow rate of the flow control valve related to turning is reduced according to the increase of the turning pressure (self-load pressure), the flow rate flowing out of the relief valve is reduced, and waste of energy consumption can be reduced.

Further, it is preferable that the control means sets the third constant γ to a relatively small positive value for the flow control valve means related to the bottom side of the hydraulic actuator for bucket. Thereby, at the time of excavation work using the bucket, the flow rate through the flow control valve can be increased according to the increase of the bucket pressure (self-load pressure), and a strong excavation operation feeling can be obtained.

Further, it is preferable that the control means sets the second and third constants β and γ to zero for the flow control valve means related to the rod side of the boom and arm hydraulic actuators. This completely eliminates the effects of the load pressure of other hydraulic actuators and the self-load pressure during the slope forming work using the boom and arm, and reduces the amount of operation of the boom operation lever and arm operation lever. Accordingly, the flow rate can be accurately proportionally distributed, and accurate slope formation can be performed.

〔Example〕

First Embodiment In FIG. 1, a hydraulic drive device according to an embodiment of the present invention includes, for example, a variable displacement hydraulic pump 1 of a swash plate type, and main pipelines 2, 3 and 4,5
, And a plurality of hydraulic actuators 6, 7 driven by pressure oil discharged from the hydraulic pump 1, and respective main pipelines 2, 3 and 4, between the hydraulic pump 1 and the hydraulic actuators 6, 7. And flow control valves 8 and 9 connected to the control valve 5.

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

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

In the main valve 21 configured as described above, the upper annular end face facing the inlet port 17 of the valve body 21 is
An annular pressure receiving area As for receiving the discharge pressure Ps is defined, a bottom wall surface facing the outlet port 18 defines a pressure receiving area Al for receiving the load pressure Pl of the hydraulic actuator 6, and a top end face facing the back pressure chamber 24 is The pressure receiving area Ac for receiving the control pressure Pc is defined. Here, each pressure receiving area has a relationship of Ac = As + Al.

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

The pilot valve 15 is driven by an operation lever (not shown), and an opening degree is set according to the operation amount of the operation lever.

The pressure compensating valve 16 is configured as a spool-type variable throttle valve, and is disposed opposite to the first control pressure chamber 30 for biasing the spool in the valve closing direction. And third control pressure chambers for urging the valve in the valve opening direction
31, 32. The first control pressure chamber 30 is connected to an inlet side of the pilot valve 15 via a pilot line 33,
The second control pressure chamber is connected to an outlet side of the pilot valve 15 via a pilot line 34, and the third control pressure chamber is connected to an electromagnetic proportional valve 250 described later via a pilot line 35. As a result, the inlet pressure Pz of the pilot valve 15 is introduced into the first control pressure chamber 30, and the outlet pressure Pl of the pilot valve 15 (= the hydraulic actuator 6,
7), and the control pressure Pm of the solenoid proportional valve 250 is introduced into the third control pressure chamber 32. Each of the first to third control pressure chambers has a pressure receiving area of the same size.

In the flow control valve 8 thus configured, the combination of the seat valve type main valve 11 and the pilot valve 15 is USP
4,535,809, as described in the patent specification, when the operating lever of the pilot valve 15 is operated, a pilot flow corresponding to the opening of the pilot valve 15 is formed in the pilot circuits 12 to 14, and the pilot flow is variable. By the action of the throttle 23 and the back pressure chamber 24, the main valve body 21 opens to an opening proportional to the pilot flow, and a flow corresponding to the operation amount of the operation lever (opening of the pilot valve 15) flows through the main valve 11 into the inlet. Exit port from port 17
Outflow to 18.

The flow control valve 9 is also configured similarly to the flow control valve 8, and includes a seat valve type main valve 35, pilot lines 36, 37, and 38 constituting a pilot circuit, a pilot valve 39, and a pressure compensating valve 40. . The pressure compensating valve 40 is similar to the pressure compensating valve 16,
It has the third control pressure chambers 41, 42, 43, and the first control pressure chamber 41
Is connected to the inlet side of the pilot valve 39 via the pilot line 44, and the second control pressure chamber 42 is connected to the outlet side of the pilot valve 39 via the pilot line 45, and is connected to the third control pressure chamber 43. Is an electromagnetic proportional valve to be described later via a pilot line 46.
Connected to 251.

The pilot lines 14, 38 of the flow control valves 8, 9 are connected to the maximum load pressure line 50 via load pressure introduction lines 54, 55 having check valves 52, 53, respectively, and the hydraulic actuators 6, 7
The load pressure on the high pressure side is guided to the maximum load pressure line 50 as the maximum load pressure. The maximum load pressure line 50 is connected to a tank 57 via a throttle 56.

Check valves 58, which prevent the flow of hydraulic oil from the hydraulic actuators 6, 7 to the main valves 11, 70 are respectively provided in the downstream main pipelines 3, 5 of the main valves 11, 35 of the flow control valves 8, 9, respectively. 59 is connected.

A pressure detector 204 for detecting the discharge pressure Ps of the hydraulic pump 1 is connected to a discharge line of the hydraulic pump 1 connected to the main lines 2 and 3, and pilot valves 15, 38 are connected to lines 14 and 38 of the pilot circuit.
39 pressure detectors 207 and 208 for detecting the outlet pressure Pl are connected. In the following description, pilot valves 15, 39
Outlet pressure (or load pressure of hydraulic actuators 6 and 7)
If it is clearer to express the former distinctively, the former should be Pl
Expressed as 1, the latter as Pl2. Therefore, the pressure detector 20
7,208 is the outlet pressure Pl1, Pl2 of pilot valve 15,39 respectively
Is detected. Further, a differential pressure gauge 209 for detecting a differential pressure ΔPs between the discharge pressure Ps of the hydraulic pump 1 and the maximum load pressure Plmax of the hydraulic actuators 6 and 7 is connected to the discharge pipe and the maximum load pressure pipe 50 of the hydraulic pump 1. The differential pressure gauges 252 and 253 for detecting the differential pressures ΔPl1 and Pl2 between the self-load pressures Pl1 and Pl2 and the maximum load pressure Plmax are connected to the pilot lines 14 and 38 and the maximum load pressure line 50 of the flow control valves 8 and 9, respectively. ing.

Further, the hydraulic pump 1 is provided with an angle meter 210 for detecting the tilt angle Qr of the displacement capacity variable mechanism such as a swash plate, and the discharge flow rate of the hydraulic pump 1 is controlled by hydraulic oil from the auxiliary pump 211. It is controlled by the quantity control device 212.

The pressure signals Ps, Pl1, Pl2 from the pressure detectors 204, 207, 208, the differential pressure signals ΔPs, ΔPl1, ΔPl2 from the differential pressure gauges 209, 252, 253 and the tilt angle signal Qr from the angle meter 210 are input to the control unit 213, and the control unit 213 Based on these input signals,
The control amount of the displacement mechanism of the hydraulic pump 1 and the control amount of the pressure compensating valves 202 and 203 are calculated, and the control signal Qo and the control signals Im1 and Im2 are output to the discharge amount control device 212 and the electromagnetic proportional valves 250 and 251 respectively.

The differential pressure gauges 209, 252, 253 are configured as shown in FIG. 2 as an example. FIG. 2 shows a differential pressure gauge 252 as a representative. That is, the differential pressure gauge 252 includes a body 350 having pressure oil supply ports 347 and 348 connected to the pipelines 50 and 14 and a pressure oil discharge port 349 connected to the tank 57, respectively.
A cylinder 351 mounted on the cylinder 350; a piston 352 housed in the cylinder 351 and having opposed pressure receiving portions 352a and 352b having the same area and receiving two pressures from the supply ports 347 and 348; The shaft 353 that transmits the displacement and the force of the 352 and the piston 3 that is housed in the cylinder 351 and receives the force of the piston 352 and changes the displacement in proportion to the force
52, a spring 354, a case 355 made of a non-magnetic material and attached to the cylinder 351 and a magnetic material, attached to the tip of the shaft 353 and housed in the case 355. A core 356 having the same displacement, a displacement sensor 357 fixed to the outer periphery of the case 355 and converting the displacement of the core 356 into an electric signal, and a cylinder 3
An amplifier 5 housed in a cover 358 attached to 51, amplifies the electric signal from the displacement sensor 357, and outputs the amplified signal to the outside.
39, and a spring 360 disposed between the piston 352 and the body 350.

In the differential pressure gauge configured as described above, the supply port 34
Maximum load pressure Plmax and self-load pressure Pl1 through 7,348
Acts on the pressure receiving portions 352a and 352b of the piston 352. At this time, assuming that the pressure receiving area is A, Plmax ≧ Pl.
, A force of A × (Plmax−Pl) acts on the upper side of the figure. By this force, the piston 352 is displaced against the springs 354 and 360 that elastically support the piston in a pre-compressed state, and the core 356 is displaced similarly. Assuming that the spring constants of the springs 354 and 360 are K1 and K2, the displacement S is as follows: S = A × (Plmax−Pl) / (K1 + K2) The displacement sensor 357 converts this displacement into an electric signal, and amplifies it with an amplifier 359 and outputs it. The displacement sensor 357 is preferably of a non-contact type because oil is present in the displaced portion of the core 356, and is, for example, a differential transformer type or a magnetic resistance element type. For this reason, shaft 353 and case 3
55 is made of a non-magnetic material. In any of these types of displacement sensors, the relationship between the electric signal level E and the displacement S has good linearity and is in a linear proportional relationship. Therefore, assuming that the proportionality constant is K, the electric signal level E is as follows: E = K · S = {K · A / (K1 + K2)} (Plmax−Pl) Here, since A, K1, and K2 are all constants, the electric signal level E is the differential pressure between the maximum load pressure and the self-load pressure (Plmax
−Pl), and a differential pressure signal ΔPl1 can be obtained.

As described above, in order to apply the differential pressure between the two pressures to the opposing pressure receiving portions of the piston 352, the respective pressures are led to separate pressure sensors to obtain respective electric signals, and then the difference between them is obtained to obtain the differential pressure. As in the case of obtaining an electrical signal, there is no error based on the non-linearity of the output with respect to the pressure in the pressure sensor and the hysteresis with respect to the increase and decrease of the pressure, and the differential pressure can be measured with high accuracy even when the pressure is increased. .

In the illustrated embodiment, since the differential pressure gauge only needs to be able to measure the differential pressure at the time of Plmax ≧ Pl, the spring 360 may not be necessary, and in this case, the structure is simplified, and the output electric signal level E at this time is reduced. The relation with the differential pressure is as follows: E = {K · A / K2} (Plmax−Pl)

The electromagnetic proportional valves 250 and 251 are configured as shown in FIG. 3 as an example. FIG. 3 shows an electromagnetic proportional valve 250 as a representative. This electromagnetic proportional valve is an example constituted by an electromagnetic proportional pressure reducing valve, and includes a proportional solenoid portion 362 and a pressure reducing valve portion 363. The proportional solenoid unit 362 has a known structure including a proportional solenoid and an iron core (both not shown), and the proportional solenoid has terminals 364a and 364b. This terminal 364
Control signals Im1 and Im2 from the control unit 213 are input to a and 362b.

The pressure reducing valve portion 363 includes a pressure oil supply port 367 connected to the auxiliary pump 211 via a supply line 366, a pressure oil discharge port 369 connected to the tank 57 via a return line 368, and a pressure compensation valve. A body 371 having a pressure oil output port 370 connected to the 16 control pressure chambers 32, and a spool 372 having opposed end faces 372a, 372b and formed with an internal passage 372c disposed in the body 371; At one end engages the iron core of the proportional solenoid portion 362 and at the other end the end face of the spool 372
The push bar 373 is in contact with the push bar 372a.

When a current is supplied from the terminals 364a and 364b to the proportional solenoid, a force proportional to the current is applied to the iron core of the proportional solenoid portion 362, and this force is applied to the end face 372a of the spool 372 via the push rod 373 engaged with the iron core. Conveyed to. As a result, the spool 372 moves to the right from the position shown in FIG.
c and the supply port 367 are communicated, and the supply port 367 and the output port 370 are communicated via the internal passage 372c. As a result, the hydraulic pressure in the output port 370 increases, and the force acting on the end surface 372b of the spool 372 also increases. This force is the pressing force of the push rod 373 (the force applied to the iron core of the proportional solenoid 362).
When it becomes larger, the spool 372 moves to the left, the internal passage 372c communicates with the discharge port 369, and the output port 370 communicates with the discharge port 369 via the internal passage 372c. As a result, the oil pressure at the output port 370 decreases, and the end face 372
The force received by b also decreases. When this force becomes smaller than the pressing force of the push bar 373, the spool 372 moves to the right again in the drawing.

As described above, since the spool 372 of the pressure reducing valve portion 363 operates by receiving the force given to the iron core of the proportional solenoid portion 362, the output port 370 eventually has a pressure proportional to the level of the current supplied to the proportional solenoid. Is generated, and this pressure is output to the control pressure chamber 32 of the pressure compensating valve 16 described above.

The pressure in the supply line 366 is configured to be a constant pressure always set by the relief valve 311.

The control unit 213 is composed of a microcomputer, and as shown in FIG. 4, the pressure signals Ps, Pl1, Pl2 from the pressure detectors 204, 207, 208, the differential pressure signals ΔPs, ΔPl1, ΔPl2 from the differential pressure gauges 209, 252, 253, and the angle meter 210. Tilt angle signal Qr
A / D converter 214 that converts the digital signal into a digital signal, a central processing unit 215, and a memory 2 that stores a control procedure program.
16, an output D / A converter 217, an output interface 218, and an amplifier 219 connected to the electrode proportional valves 250 and 251.
220 and two input terminals 212A and 212B of the discharge amount control device 212.
And the amplifiers 221 and 222 connected to the power supply.

The control unit 213 controls the discharge pressure Ps of the hydraulic pump 1
From the differential pressure signal ΔPs of the differential pressure gauge 209 which detects the differential pressure between the pressure and the maximum load pressure of the hydraulic actuators 6 and 7 based on the control procedure program stored in the memory 216.
A discharge amount target value Qo of the hydraulic pump 1 that holds Ps higher than the maximum load pressure by a predetermined value is calculated, and the target value signal Qo is calculated.
The signals are output from the amplifiers 221 and 222 to the input terminals 212A and 212B of the discharge amount control device 212 via the I / O interface 218. Accordingly, the discharge amount control device 212 controls the tilt angle of the swash plate of the hydraulic pump 1 so that the tilt angle Qr detected by the goniometer 210 becomes equal to the target value Qo, and increases the pump discharge pressure Ps to the maximum. The function as a load-sensing type pump regulator that maintains the load pressure higher than the load pressure Plmax by a predetermined value can be achieved.

Further, the control unit 213 includes the pressure signals Pl1 and Pl2 of the pressure detectors 207 and 208 and the differential pressure signals ΔPs and ΔPl of the differential pressure gauges 209, 252 and 253.
A control amount of the pressure compensating valve 16 (40) is calculated from 1, ΔPl2, and a control signal is output from the amplifiers 219 and 220 to the electromagnetic proportional valves 250 and 251. Hereinafter, this processing content will be described with reference to FIG. FIG. 5 is a flowchart showing the control contents in this regard.

In step 260, the pressure signals Pl1, Pl2 detected by the pressure detectors 207, 208 and the differential pressure signals ΔPs (= Ps−Plmax), ΔPl1 (= Plmax−Pl1), ΔPs detected by the differential pressure gauges 209, 252, 253
Read Pl2 (= Plmax-Pl2). Next, in step 261, the control pressures Pm1 and Pm2 for controlling the differential pressure between the pilot valves 15 and 39 are calculated by the following equation.

Pm1 = α (Ps−Plmax) + β (Plmax−Pl1) + γPl1 Pm2 = α (Ps−Plmax) + β (Plmax−Pl2) + γPl2 where α, β, and γ are constants and are set to predetermined values, respectively. I have. The meaning of this equation will be described later in the section of the operation principle.

Next, in step 262, based on a predetermined functional relationship determined by the characteristics of the electromagnetic proportional valves 250 and 251, an operation of Im1 = f (Pm1) Im2 = f (Pm1) is performed to obtain control signals Im1 and Im2 of the electromagnetic proportional valves 250 and 251. Finally, in step 263, the calculated control signals Im1, I
m2 from D / A converter 217 to amplifier 219,220 from solenoid proportional valve 2
Output to 50,251.

Upon receiving this control signal, the electromagnetic proportional valves 250 and 251 receive the control pressures Pm1 and Pm2 proportional to the control signals Im1 and Im2 as described above.
And the control pressure is introduced into the third control pressure chambers 32, 43 of the pressure compensating valves 16, 40 via the hydraulic lines 35, 46.

In the pressure compensating valves 16, 40, the first and second control pressure chambers 30,
Pilot valves 15, 39 introduced to 31 and 41, 42
The differential pressure control of the pilot valves 15, 39 is performed based on Pz1, Pl1, Pz2, Pl2 and the control pressures Pm1, Pm2.

Operation principle Next, as described above, the pressure P
The operating principle of z1, Pl1, Pz2, Pl2 and the pressure compensating valves 16, 40 controlled by the control pressures Pm1, Pm2 generated by the electromagnetic proportional valves 250, 251 will be described.

A pilot valve is provided in the first control pressure chamber 30 of the pressure compensating valve 16.
The 15 inlet pressure Pz1 acts in the valve closing direction, the outlet pressure Pl1 of the pilot valve 15 acts on the second control pressure chamber 31 in the valve opening direction, and the electromagnetic proportional valve 250 acts on the third control pressure chamber 32. Is acting in the valve opening direction. Accordingly, the balance of the pressure of the pressure compensating valve 16 is expressed by the following equation.

Pz1−Pl1 = Pm1 (1) That is, the differential pressure Pz1−Pl1 across the pilot valve 15 is controlled by the output Pm1 of the proportional solenoid valve 250. Here, the control pressure Pm1 is calculated as the following value in the control unit 213 as shown in a procedure 261 in FIG.

Pm1 = α (Ps−Plmax) + β (Plmax−Pl1) + γPl1 (2) Therefore, from the expressions (1) and (2), Pz1−Pl1 = α (Ps−Plmax) + β (Plmax−Pl1) + γPl1 (3) Similarly, the following pressure-balancing formula holds for the pressure compensating valve 40 as well.

Pz2−Pl2 = α (Ps−Plmax) + β (Plmax−Pl2) + γPl2 (4) Therefore, the differential pressure between the pilot valves 15, 39 is ΔPz, and the above equations (3) and (4) are expressed in the general formula. When expressed, ΔPz = α (Ps−Plmax) + β (Plmax−Pl) + γPl (5)

Therefore, the above equation (5) will be considered. In the equation (5), the left side is a differential pressure ΔPz between the inlet pressure Pz and the outlet pressure Pl of the pilot valve 15 or 39. The first term on the right side is a term relating to the pressure difference between the discharge pressure Ps of the hydraulic pump 1 and the maximum load pressure Plmax, and α is a proportional constant. The second term is the maximum load pressure Plma
This is a term relating to the differential pressure between x and the load pressure of the hydraulic actuator 6 or 7, that is, the self-load pressure Pl, and β is a proportional constant. The third term is determined by the self-load pressure Pl, and γ is a proportional constant. That is, the equation (5) shows that the pressure compensating valves 16 and 40 are controlled by the pilot valve 15 or 39 based on the four pressures Ps, Plmax, Pl and Pz.
The differential pressure ΔPz between the inlet pressure Pz and the outlet pressure Pl of the hydraulic pump 1 can be controlled; at this time, the differential pressure ΔPz can be controlled by the differential pressure Ps−Plmax between the discharge pressure Ps of the hydraulic pump 1 and the maximum load pressure Plmax, and the maximum load pressure Pl.
Differential pressure Plmax-Pl between max and self-load pressure Pl, self-load pressure
Controllable in proportion to each of the three elements of Pl;
This means that the degree proportional to the three elements Ps-Plmax, Plmax-Pl, and Pl can be arbitrarily set by selecting the values of the proportional constants α, β, and γ.

Here, that the pressure compensating valves 16 and 40 control the differential pressure ΔPz before and after the pilot valves 15 and 39 is to control the pilot flow rate passing through the pilot valves 15 and 40, and as a result, as described above, the seat type The known function of the combination of the main valves 11, 35 and the pilot valves 15, 39 is to control the main flow through the main valves 11, 35.

Further, in the first term on the right side, the differential pressure Ps−Plmax is, as described above, in the present embodiment in which the discharge amount control device 212 controlled by the control unit 213 functions as a load sensing type pump regulator, the pump regulator 212 is As long as it works effectively, it is constant and is common to the two pressure compensating valves 16,40.

Accordingly, in the first term on the right side, controlling the differential pressure ΔPz before and after the pilot valves 15 and 39 so as to be proportional to the differential pressure Ps−Plmax is performed in an operating state where the pump regulator 212 is effectively functioning. Is to control the differential pressure ΔPz to be constant, and if the opening degree of the pilot valves 15, 39 is constant, the main flow rate through the main valves 11, 35 even if the main valve inlet pressure Ps or the outlet pressure Pl fluctuates. Is controlled to be constant. That is, it fulfills a pressure compensation function.

Further, in an operation state in which the pump regulator 212 does not function effectively, such as when the sum of the consumption flow rates of the hydraulic actuators 6 and 7 becomes larger than the maximum discharge flow rate of the hydraulic pump 1 and the discharge pressure of the hydraulic pump 1 decreases. , The differential pressure ΔPz decreases as the differential pressure Ps−Plmax decreases, and the main valve 1
Although the main flow passing through the pressure compensating valves 1 and 35 also decreases, the differential pressure Ps−Plmax is common to the two pressure compensating valves 16 and 40.
The main flow through is reduced at the same rate. Therefore, main valve 11,3
The main flow passing through 5 is proportionally distributed according to the operation amount of each operating lever of pilot valves 15 and 39 (opening of pilot valves 15 and 39), and the discharge flow of hydraulic pump 1 is supplied to the hydraulic actuator on the high pressure side. Is also supplied reliably. That is, a shunt function is obtained.

In the second term on the right side, controlling the differential pressure ΔPz between the pilot valves 15 and 39 to be proportional to the differential pressure Plmax−Pl is equivalent to the load pressure Pl of another hydraulic actuator.
When max is greater than the self-load pressure Pl, the differential pressure ΔPz across the pilot valve 15 or 39 is changed depending on the maximum load pressure Plmax of other hydraulic actuators, and the opening of the pilot valve 15 or 39 is changed. If it is constant, the main flow rate through the main valves 11, 35 is changed depending on the maximum load pressure Plmax. In general, it is preferable that the flow control of the flow control valve is not affected by other hydraulic actuators.However, in a hydraulic construction machine such as a hydraulic shovel, the flow is controlled by the load pressure of another hydraulic actuator. Is also preferable in some work modes. In such a case, the second term on the right side performs a harmonic function of changing the flow rate in harmony with another hydraulic actuator.

Further, in the third term on the right side, the pilot valves 15, 39
To make the front-rear differential pressure ΔPz proportional to the self-load pressure Pl is to change the front-rear differential pressure ΔPz of the pilot valve 15 or 39 according to the change of the self-load pressure Pl. Assuming that the opening of the valve 15 or 39 is constant, the main flow rate through the main valves 11, 35 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 (5), the first term on the right-hand 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, since the differential pressure ΔPz between the pilot valves 15 and 39 determines the pilot flow with respect to the opening of the pilot valves 15 and 39 determined by the operation amount of the operation lever, the differential pressure Ps−Plma
x is a proportional gain of the pilot flow rate to the operation amount (pilot valve opening) of the operation lever of the pilot valves 15 and 39, and therefore the main valve 11 and 35 to the operation amount.
Has the meaning of the proportional gain of the main flow passing through. Therefore, the proportionality constant α is determined according to the proportional gain.

Further, assuming that the ratio of the pressure receiving area As for receiving the discharge pressure Ps of the hydraulic pump 1 of the valve body 21 to the pressure receiving area Ac for receiving the pressure Pc of the back pressure chamber 24 of the main valve body 21 is K, the pressure of the valve body 21 The balance equation is as follows: Pc = KPs + (1-K) Pl. On the other hand, the control pressure Pc and the inlet pressure Pz of the pilot valves 15 and 39 have a relationship of Pc ≧ Pz, and Pc = Pz when the pressure compensating valves 16 and 40 are completely open. Therefore, the differential pressure Pz-Pl (ΔPz) before and after the pilot valves 15 and 39 is as follows: Pz-Pl ≦ Pc-Pl = K (Ps-Pl) (6) That is, the maximum differential pressure that can be taken by the pilot valves 15, 39 is K (Ps-Pl). In the above equation (5), β =
0, γ = 0, and considering the maximum load pressure side (Plmax = Pl) in the combined operation of the hydraulic actuators 6 and 7, Pz−Pl = α (Ps−Plmax) ≦ K (Ps−Plmax) (7) ). Therefore, when a value of α <α <K is set,
The pilot valve on the maximum load pressure side cannot obtain a differential pressure of K (Ps-Plmax) or more, while the pilot valve on the low pressure side has a differential pressure of α (Ps-Plmax)> K (Ps-Plmax). Therefore, even if the pilot valve opening is the same, the differential pressure across the pilot valve is not the same, and the pilot flow rate is different. Therefore, the flow rate cannot be proportionally distributed according to the manipulated variable. However, even if the proportional distribution cannot be performed, the pressure oil can be reliably supplied to the high-pressure side hydraulic actuator.

Therefore, 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 valves 16 and 40, 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) is obtained, but when the composite operation is switched to the single operation of the hydraulic actuator on the low load pressure side, the K is also applied to the pilot valve on the low load pressure side.
(Ps-Pl) or more, and the differential pressure across the pilot valve changes from α (Ps-Plmax) to K (Ps
−Pl) and the pilot flow is correspondingly reduced. As a result, the flow rate supplied to the hydraulic actuator 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 the combined operation is switched to the single operation, the differential pressure does not fluctuate and stable work is performed. be able to. Therefore, in this sense, it is preferable to set α to α ≦ K.

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

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

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

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.

Then, as described above, the proportional constants α, β, and γ can be arbitrarily set in the control unit 213.

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

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

In such a hydraulic excavator hydraulic drive, a swing motor 87, a boom cylinder 88, an arm cylinder 89
As shown in FIG. 8, the proportional constant α for all the flow control valves of the bucket cylinder 90 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 87, the proportional constant β
As shown in FIG. (A), β = 0 is set, and the proportionality constant γ is set to a negative value close to zero as shown in FIG. 10 (A). In the flow control valve related to the bottom side of the boom cylinder 88, the proportional constant β is set to any positive value as shown in FIG. 9 (B), and the proportional constant γ is shown in FIG. 10 (B). Is set to γ = 0. In the flow control valve for the bottom side of the arm cylinder 89, the proportional constant β is set to a positive value close to zero as shown in FIG. 9 (C), and the proportional constant γ is shown in FIG. 10 (B). Is set to γ = 0.
In the flow control valve for the bottom side of the bucket cylinder 90, the proportional constant β is set to a negative value close to zero as shown in FIG. 9 (D), and the proportional constant γ is shown in FIG. 10 (C). Is set to a positive value close to the spirit. In the flow control valve related to the rod side of the boom cylinder 88, the flow control valve related to the rod side of the arm cylinder 89, and the flow control valve related to the rod side of the bucket cylinder 90, the proportional constants β and γ are all It is set to zero as shown in FIGS. 9 (A) and 10 (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 flow control valves 8, 9 is operated, the pilot valves 15, 39 are closed, and the pilot flow does not flow through the pilot circuits 12 to 14, 36 to 38. Therefore, no pressure oil flows into each of the variable throttles 23 of the main valves 11 and 35, and the control pressure Pc of the back pressure chamber 24 is equal to the pressure (discharge pressure of the hydraulic pump 1) Ps of the inlet port 17. In addition, the discharge pressure of the hydraulic pump 1 is controlled by the function of the discharge amount control device 212 as the load sensing type pump regulator described above.
Ps is maintained at a pressure higher than the maximum load pressure Plmax of the hydraulic actuators 6 and 7 by a predetermined value. Therefore, the valve element 21
Are in a relationship of Ac = As + Al and Ps> Pl, the valve element 21 is urged in the valve closing direction by the control pressure Pc, and the main valves 11, 70 are held in the closed position. The pressure compensating valves 16 and 40 are held at the open positions by setting the constant α described above.

Next, when the operation lever of the flow control valve 8 is operated alone, the pilot valve 15 is opened according to the operation amount, and a pilot flow is formed in the pilot circuits 12 to 14, and the pilot flow is formed according to the opening degree of the pilot valve 15. Pilot flow. As a result, as described above, the main valve valve element 21 opens to an opening proportional to the pilot flow rate by the action of the variable throttle 23 and the back pressure chamber 24, and is operated according to the operation amount of the operation lever (opening of the pilot valve 15). Flow from the inlet port 17 through the main valve 11 to the outlet port
Outflow to 18.

Then, in a state where the pilot valve 15 is opened by a certain amount and a certain amount of the main flow is flowing from the inlet port 17 to the outlet port 18, for example, the pressure at the outlet port 18 increases, and the inlet port 17 and the outlet When the pressure difference at the port 18 is about to decrease, the discharge pressure of the hydraulic pump 1 is increased by the function of the discharge amount control device 212 as a load sensing type pump regulator, and the pressure at the inlet port 17 (the hydraulic pump 1) is increased. ) And the pressure at the outlet port 18 (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 11.

Also, in such a single operation of the hydraulic actuator 6, if the control unit 213 is set so that the proportional constant γ relating to the self-pressure compensation characteristic in the equation (5) becomes an arbitrary value other than zero, The pilot valve 15 according to the change of the load pressure (self-load pressure) of the actuator 6
Is controlled, and self-load pressure compensation is performed.

For example, in the example of the hydraulic excavator described with reference to FIGS. 6 to 10, in the flow control valve related to the swing motor 87, the proportionality constant γ is a negative value close to zero as shown in FIG. Is set to a value. Therefore, when the revolving unit 81 is driven, the revolving unit is an inertial body, so that the load pressure is increased and rises above the pressure of a relief valve provided for circuit protection, causing waste of energy. Is a negative value, the flow rate through the flow control valve, which is controlled so that the differential pressure ΔPz decreases as the turning load pressure increases, decreases. 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 on the bottom side of the bucket cylinder 90, the proportionality constant γ is set to a positive value close to zero as shown in FIG. 10 (C). Therefore, as the self-load pressure increases in the excavation work, the differential pressure ΔPz
And the flow 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 flow control valves 11 and 35 are operated, the following operation is performed. First, the flow control valve 11,35
As in the case where the flow control valve 11 is independently operated in both cases, the pilot flow according to the operation amount flows, and the variable throttle 23
The flow according to the operation amount of the operation lever 30 (the opening degree of the pilot valves 15 and 39) flows from the inlet port 17 to the outlet port 18 through the main valves 11 and 35 by the action of the back pressure chamber 24.

In such a combined operation of the hydraulic actuators 6 and 7, the control unit 213 sets the proportional constant α in the first term on the right side of the above equation (5) to an arbitrary positive value as shown in FIG. With the setting, the pressure compensation and the flow dividing function are performed.

Accordingly, for example, in the hydraulic excavator described with reference to FIGS. 6 to 10, in the operation state in which the discharge amount control device 212 is effectively functioning as a load sensing type pump regulator, each work member is operated. Driving is performed at a constant flow rate according to the operation amount of the lever, and a stable combined operation can be performed. Also, when the sum of the consumption flow rates of the hydraulic actuators 6 and 7 becomes larger than the maximum discharge flow rate of the hydraulic pump 1 and the operation becomes such that the discharge amount control device 212 cannot function effectively as a load sensing type pump regulator. In addition, not only the hydraulic oil is supplied to the low-pressure side hydraulic actuator, but also the hydraulic oil to the high-pressure side hydraulic actuator can be reliably supplied, so that all the working members 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. Thereby, in particular, in the control unit 213, the proportionality constant β,
When γ is set to zero, the movement trajectory of the working member can be accurately controlled according to the operation amount of the operation lever. For example, in a flow control valve related to the rod side of the boom cylinder 88 and a flow control valve related to the rod side of the arm cylinder 89, as shown in FIGS. 9A and 10B, β = 0, γ = 0. This completely eliminates the effects of the load pressure of other hydraulic actuators and the self-load pressure during the work of forming the downward slope using the boom and arm, and reduces the amount of operation of the boom operation lever and arm operation lever. Accordingly, the flow rates supplied to the boom cylinder 88 and the arm cylinder 89 can be precisely proportionally distributed, and accurate slope formation can be performed.

When the control unit 213 sets the proportional constant β and / or the proportional constant γ to an arbitrary value other than zero in the equation (5), the control unit 213 performs the above-described functions based on the pressure compensation and the flow dividing function. A harmonic function and / or self-load pressure compensation is performed in which the main flow through the main valves 11, 35 is varied depending on the maximum load pressure Plmax of the hydraulic actuator.

For example, in the example of the hydraulic excavator described with reference to FIGS. 6 to 10, in the flow control valve related to the swing motor 87, the proportionality constant β becomes β = 0 as shown in FIG. 9 (A). The proportional constant β is set to an arbitrary positive value as shown in FIG. 9 (B) in the flow control valve associated with the bottom side of the boom cylinder 88. 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 89, the proportionality constant β is set to a relatively small positive value as shown in FIG. 9 (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 in a range where the balance of the composite operation is not inhibited,
In the flow control valve on the arm side, the opening degree of the main valve increases in accordance with the increase of the differential pressure Plmax-Pl, and the degree of restricting the oil decreases. Thereby, deterioration of fuel efficiency and heat balance can be reduced.

Further, in the flow control valve on the bottom side of the bucket cylinder 90, the proportionality constant β is set to a small negative value as shown in FIG. 9 (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 according to the present embodiment, the control unit 213 sets the constants α, β, and γ to predetermined values, so that the shunt function, the harmonic function based on the shunt function, and / or the self-pressure A compensation 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 detection of the differential pressure Ps-Plmax between the discharge pressure of the hydraulic pump 1 and the maximum load pressure, which is the basis of the above control, and the differential pressure Plmax between the maximum load pressure and the self-load pressure -Instead of detecting Pl by detecting the individual pressure with a pressure detector and then taking the difference between the detected values,
Since the measurement is directly performed using the differential pressure gauges 209, 252, and 253, even if the pressure changes, a minute change in the differential pressure can be accurately detected, and the above control can be performed accurately.

In the hydraulic drive device of the present embodiment, the pressure compensating valves 16 and 40 as auxiliary valves are arranged in the pilot circuit instead of the main circuit, and the main valves 11 and 35 arranged in the main circuit are seat valves. It is configured. Therefore, it is possible to provide a hydraulic circuit suitable for increasing the pressure with little liquid leakage. Further, since the pressure compensating valve, that is, the auxiliary valve, is arranged in the pilot circuit, even if a large flow rate flows through the main circuit, the throttle loss at the auxiliary valve portion is small, and the cost is excellent.

In the above embodiment, referring to FIG. 8 to FIG.
For at least some of the flow control valves related to the swing body, boom, arm, and bucket of the hydraulic excavator,
An example has been shown in which the constants β and γ in the above equation (5) are set to predetermined values other than zero. 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 (5) is set to a positive value, particularly 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.

In the above embodiment, the constants α, β, γ are
In this example, an externally operable adjuster 240 is connected to the control unit 213, as shown by the imaginary line in FIG. 1, and constants α, β,
γ may be variably set.

In the above embodiment, the pressure compensating valves 16 and 40 are arranged on the pilot valve inlet side of the pilot circuit. Can be controlled, and a similar effect can be obtained.

Further, in the above embodiment, the discharge amount control device 21 is controlled as a pump control means for maintaining the discharge pressure of the hydraulic pump 1 at a pressure higher than the maximum load pressure of the first and second hydraulic actuators 6 and 7 by a predetermined value. Although an example of the electrical configuration controlled by the unit 213 has been described, it goes without saying that a load sensing type regulator using ordinary hydraulic control may be used.

Second Embodiment A second embodiment of the present invention will be described with reference to FIG. 11 and FIG. In the figure, the same reference numerals are given to members equivalent to the members shown in FIG.

In the above embodiment, when the control unit 213 is broken and the output pressures of the electromagnetic proportional valves 250 and 251 become 0 kg / cm ² , the pressure compensating valves 16 and 40 are connected to the first and second control pressure chambers 30, 31 and 4.
The pilot valves 15, 39 acting in opposition to the valves 1, 39 are driven in the valve closing direction by the pressure difference ΔPz, so that the operation of the hydraulic drive device becomes completely impossible. In the present embodiment, in consideration of such a point, the hydraulic drive device can operate even when the control unit 213 is damaged and the output pressure of the electromagnetic proportional valves 250 and 251 becomes 0 kg / cm ² .

That is, in the flow control valve 8 of this embodiment, the pressure compensating valve 24
4 is the same as the first embodiment in having first and second control pressure chambers 30 and 31, except that the third control pressure chamber 244 is provided with a spool on the same side as the first control pressure chamber 30. A spring 246 is provided so as to urge the spool in the valve closing direction, and is arranged on the side of the second control pressure chamber 31 to urge the spool in the valve opening direction. The third control pressure chamber 244 is connected to the solenoid proportional valve 2 via the pilot line 35.
Connected to 50.

The pressure compensating valve 243 of the flow control valve 9 is similarly configured, and the third control pressure chamber 248 is installed on the same side as the first control pressure chamber 41 so as to urge the spool in the valve closing direction. A spring for biasing the spool in the valve opening direction is provided on the side of the control pressure chamber 42.
249 are located. The third control pressure chamber 248 is connected to the proportional solenoid valve 251 via the pilot line 46.

On the other hand, in the control unit 213, the pressure detectors 207, 208
And the differential pressure signal Δ of the differential pressure gauges 209, 252 and 253.
The control contents of the pressure compensating valves 242 and 243 are calculated from Ps, ΔPl1 and ΔPl2, and the control contents for driving the pressure compensating valves are as shown in the flowchart of FIG.

That is, in step 270, the pressure signals Pl1 and Pl2 detected by the pressure detectors 207 and 208 and the differential pressure signals ΔPs (= Ps−Plmax) and ΔPl1 (= Plmax−Pl) detected by the differential pressure gauges 209, 252 and 253.
1), ΔPl2 (= Plmax-Pl2) is read. Then the procedure
At 271, control pressures P′m 1 and P′m 2 for controlling the differential pressure between the pilot valves 15 and 39 are calculated by the following equation.

P′m1 = Ks− {α (Ps−Plmax) + β (Plmax−Pl1) + γPl1} P′m2 = Ks− {α (Ps−Plmax) + β (Plmax−Pl2) + γPl2} where Ks is the strength of the springs 247,249. Are converted into pressure values, and α, β, and γ are constants as in the first embodiment, and are set to predetermined values.

Next, in step 272, based on a predetermined functional relationship determined by the characteristics of the electromagnetic proportional valves 250 and 251, an operation of I'm1 = f (P'm1) I'm2 = f (P'm1) is performed to control the electromagnetic proportional valves 250 and 251. Signal I'm1,
Obtain I'm2. Finally, in step 273, the calculated control signals I'm1 and I'm2 are output to the electromagnetic proportional valves 250 and 251.

Electromagnetic proportional valves 250 and 251 receive these control signals and generate control pressures P'm1 and P'm2 proportional to control signals I'm1 and I'm2, and transmit the control pressures via hydraulic lines 35 and 46. To the third control pressure chambers 244, 248 of the pressure compensating valves 242, 243.

In the pressure compensating valves 242 and 243, the first and second control pressure chambers 3
The pressures Pz1, Pl1, Pz2, Pl2 of the pilot valves 15, 39 introduced to 0, 31 and 41, 42, the control pressures P'm1, P'm2, and the spring 2
The differential pressure control of the pilot valves 15, 39 is performed from 47, 249.

In the present embodiment configured as described above, the inlet pressure Pz1 of the pilot valve 15 acts on the first control pressure chamber 30 of the pressure compensating valve 242 in the valve closing direction, and the second control pressure chamber 31 acts on the second control pressure chamber 31. The outlet pressure Pl1 of the pilot valve 15 acts in the valve opening direction, and the control pressure P'm1 from the electromagnetic proportional valve 250 acts in the third control pressure chamber 244 in the valve closing direction. Further, the set force Ks of the spring 247 acts in the valve opening direction. Therefore, the pressure compensating valve
The balance of the pressure of 242 is expressed by the following equation.

Pz1-Pl1 = Ks-P'm1 (8) Here, the control pressure P'm1 is determined by the control unit as described above.
Since the value is calculated in step 213, if the value of P′m1 shown in step 271 is substituted into equation (8), Pz1−Pl1 = α (Ps−Plmax) + β (Plmax−Pl1) + γPl1 (9) Can be In the pressure compensating valve 243, the following pressure balance equation is obtained in the same manner.

Pz2−Pl2 = α (Ps−Plmax) + β (Plmax−Pl2) + γPl2 (10) Therefore, the differential pressure across the pilot valves 15, 39 is ΔPz, and the above equations (9) and (10) are expressed in the general formula. When expressed, ΔPz = α (Ps−Plmax) + β (Plmax−Pl) + ΔPl (11) Therefore, the same equation as the above equation (5) is obtained.

Therefore, also in this embodiment, by setting the proportional constants α, β, and γ to predetermined values in the control unit 213, the differential pressure ΔPz across the pilot valves 15, 39 can be set to a value that is equal to the discharge pressure Ps of the hydraulic pump 1 Differential pressure Ps-Plma from load pressure Plmax
x, differential pressure Plmax between maximum load pressure Plmax and self-load pressure Pl
-Pl, the self-load pressure Pl can be controlled in proportion to each of the three elements.
), Or a harmony function (right-hand side second term) and / or a self-pressure compensation function (right-hand side third term) in a combined operation based on this pressure compensation and branch function.

Further, in this embodiment, since the springs 247, 249 acting in the valve opening direction are installed on the pressure compensating valves 242, 243, even if the control unit 213 is broken, the force acting on the pressure compensating valve in the valve opening direction is obtained, The operation of the drive system does not become impossible, and normal pressure compensation control can be performed by the action of the spring.

Third Embodiment A third embodiment of the present invention will be described with reference to FIG.
In the figure, the same reference numerals are given to members equivalent to the members shown in FIG.

The embodiment described above uses USP 4,535,80 as a flow control valve.
This is an embodiment constituted by a combination of a seat-type main valve as described in 9 and a pilot valve arranged in a pilot circuit. However, the present invention is not limited to the flow control valve having such a special structure, and is applicable to a general spool-type flow control valve.
FIG. 13 shows such an embodiment.

That is, in FIG. 13, the first flow control valve 100 includes a spool-type main valve 101 disposed in the main circuits 2, 3A and 3B, and an auxiliary valve disposed upstream of the main valve 101 in the main circuit 2. The second flow control valve 103 is also provided with a spool-type main valve 104 similarly disposed in the main circuits 4, 5A, and 5B, and an upstream side of the main valve 104 of the main circuit 4. And a pressure control valve 105 as an auxiliary valve. Check valves 106 and 107 for preventing backflow of pressure oil are arranged between the main valves 101 and 104 and the pressure compensating valves 102 and 105, respectively. The main valves 101 and 104 are configured as direction switching valves, and the outlet side is connected to the tank 57 via return lines 108 and 109. The main valves 101 and 104 are hydraulic pilot operated valves, and the pilot lines 110A and 1
It is driven by the pilot pressure oil transmitted via 10B and 111A, 111B according to the operation direction and operation amount of an operation lever (not shown).

The main valves 101 and 104 are provided with load pressure ports 112 and 113. The load pressure ports 112 and 113 communicate with the pressure oil supply side of the main pipes 3A and 3B in the left and right switching operation positions, and communicate with the return pipes 108 and 109 in the neutral position. . In this way, the load pressure is transmitted to the load pressure ports 112, 113 when the hydraulic actuators 6, 7 operate. Load pressure port 112,11
3 is connected to a high pressure selection valve 116 via pilot lines 114 and 115, and the high pressure selection valve 116 is connected to a maximum load pressure line 117.

Pressure detectors 204, 207, 208 and differential pressure gauges 209, 252, 253 are connected to the pilot lines 114, 115, the maximum load pressure line 117, and the discharge line of the hydraulic pump 1 as in the first embodiment.

The pressure compensating valves 102 and 105 are the pressure compensating valves 16 and
It is configured similarly to 40. That is, it has first control pressure chambers 118, 121 acting in the valve closing direction and second and third control pressure chambers 119, 120 and 122, 123 acting in the valve opening direction.
The inlet pressures of the main valves 101, 104 are introduced into the control pressure chambers 118, 121 through pilot lines 124, 125, and the main valves 101, 104 are introduced into the second control pressure chambers 119, 122 through pilot lines 126, 127.
The outlet pressure (load pressure) of 104 is introduced, and the output pressures of the proportional solenoid valves 250 and 251 are introduced into the third control pressure chambers 120 and 123 via the pilot lines 128 and 129. Each of the first to third control pressure chambers has a pressure receiving area of the same size.

 The configuration other than the above is the same as that of the first embodiment.

Also in the present embodiment configured as described above, the control procedure shown in the flowchart of FIG. 5 is stored in the control unit 213, so that the pressure compensating valves 102 and 105 have the above-described pressure balancing type (5). Holds. Therefore,
By setting the proportional constants α, β, γ to predetermined values in the control unit 213, the differential pressure ΔPz across the main valves 101, 104
Is the differential pressure Ps-Plmax between the discharge pressure Ps of the hydraulic pump 1 and the maximum load pressure Plmax, the maximum load pressure Plmax, and the self-load pressure Pl.
And the pressure difference Plmax-Pl and the self-load pressure Pl can be controlled in proportion to each of the three factors, and the above-described pressure compensation and branching function (the first term on the right side) or based on this pressure compensation and branching function Harmony function in compound operation (second term on right side)
And / or a self-pressure compensation function (the third term on the right side) 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.

The third embodiment shown in FIG. 13 is also hydraulically driven even when the control unit 213 is damaged and the output pressure of the solenoid proportional valves 250 and 251 reaches 0 kg / cm ² , similarly to the embodiment shown in FIG. Configuration and control unit of the pressure compensating valve so that the device operates
The control contents of 213 can be changed. FIG. 14 shows a modified example thereof. That is, in FIG. 14, the pressure compensating valves 140, 14
1, the third control pressure chambers 142, 143 are respectively installed on the same side as the first control pressure chambers 118, 121 so as to act in the valve closing direction, and the second control pressure chambers 119, 122 are arranged in the valve opening direction. Acting springs 145, 146 are arranged, and the control unit 213 has a second
The control procedure shown in FIG. 12 is stored, and the control pressures P′m1 and P′m2 are obtained by the calculation shown in the procedure 271 in FIG. Thus, in the flow control valves 100 and 103 having the spool-type main valve shown in FIG. 13, the same effects as those of the embodiment shown in FIG. 11 can be obtained.

〔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, in an example of application to a hydraulic excavator, for example, even if the operating levers for turning and boom raising are simultaneously operated up to a full stroke, initially, the boom ascending speed rises faster than the turning speed, and the boom rises to some extent. When the excavation is performed by the compound operation using the arm, the flow control valve characteristics such that the swivel speed gradually increases and the swivel speed becomes almost constant when the swivel reaches the maximum speed When the arm is reliably driven and the arm hydraulic actuator is on the low pressure side,
Flow control valve characteristics to prevent deterioration of fuel economy and heat balance, during trenching work by combined operation using buckets, the bucket is released from the excavation load, the moment it comes out on the ground, the flow rate through the flow control valve is reduced, Flow control valve characteristics to reduce shock, flow control valve characteristics to reduce the amount of flow out of the relief valve during turning acceleration and reduce waste of energy consumption, and obtain a strong excavation operation feeling during excavation work using buckets It is possible to obtain the flow control valve characteristics, the flow control valve characteristics for performing accurate slope formation, and the like at the time of forming an inclined surface using a boom and an arm.

Further, in the present invention, the detection of the differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure, and the detection of the differential pressure between the maximum load pressure and the self-load pressure, which are the basis of the above control, are individually performed by pressure detection means. Instead of taking the difference between the detected values after detecting the pressure, it is done directly using the differential pressure detection means, so even if the pressure changes, it is possible to accurately detect minute changes in the differential pressure The above control can be performed with high accuracy.

If a spring acting in the valve opening direction is installed in the second valve means, even if the control means is damaged and the second valve means cannot be controlled, the second valve means can be reliably opened. In addition to enabling the operation of the hydraulic drive device, normal pressure compensation control can be performed.

When each of the first and second flow control means is constituted by a combination of a sheet-type main valve and a pilot valve connected thereto, the liquid leakage is small and suitable for high pressure,
A hydraulic circuit having a small throttle loss and an energy-saving structure can be provided.

When the ratio of the pressure receiving area receiving the discharge pressure of the hydraulic pump to the pressure receiving area receiving the control pressure of the seat-type main valve back pressure chamber is set to K, and the first constant α is set to a relation of α ≦ K. Thus, the flow rate proportional to the operation amount of the operating means (opening of the pilot valve) in the flow dividing function 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 view of a hydraulic drive device according to an embodiment of the present invention, FIG. 2 is a sectional view of a differential pressure gauge used in the hydraulic drive device, and FIG. 3 is used in the hydraulic drive device. FIG. 4 is a cross-sectional view of the electromagnetic proportional valve, FIG. 4 is a schematic diagram of a control unit used in the hydraulic drive device, and FIG. 5 is a flowchart showing control contents in the control unit;
FIG. 6 is a side view of a hydraulic shovel to which the hydraulic drive device of the present invention is applied, FIG. 7 is a top view of the hydraulic shovel, and FIG. 8 is a flow control of one of the hydraulic drive devices. 9 (A) to 9 (D) are characteristic diagrams showing an example of setting a constant α of a pressure compensating valve included in the valve.
10A to 10C are characteristic diagrams showing setting examples of a constant β of a pressure compensating valve included in one flow control valve, and FIGS. FIG. 11 is a characteristic diagram showing a setting example of a constant γ of FIG. 11, FIG. 11 is a schematic diagram of a hydraulic drive device according to another embodiment of the present invention, and FIG. FIG. 13 is a flowchart showing the contents, and FIG. 13 is a schematic view of a hydraulic drive device according to still another embodiment of the present invention,
FIG. 14 is a schematic diagram of a hydraulic drive device according to still another embodiment of the present invention. DESCRIPTION OF SYMBOLS 1 ... Hydraulic pump, 2-5 ... Main circuit 6,7 ... Hydraulic actuator 8,9; 100,103 ... Flow control valve 11,35 ... Seat type main valve 12 ... Back pressure chamber 12- 14, 36 to 38: Pilot circuit 15, 39: Pilot valve (first valve means) 16, 40; 242, 243; 102, 105; 140, 141
30, 41; 118, 121 ... first control pressure chamber 31, 42; 119, 122 ... second control pressure chamber 32, 43; 244, 248; 120, 123; 142, 143 ... third control pressure chamber 213 ... Control unit 247,249; 145,146 ... Spring 204,207,208 ... Pressure detector 209,252,253 ... Differential pressure gauge 101,104 ... Spool type main valve (first valve means)

 ──────────────────────────────────────────────────続 き Continuation of the front page (72) Inventor Toichi Hirata 650 Kandamachi, Tsuchiura-shi, Ibaraki Pref. Tsuchiura Works, Hitachi Construction Machinery Co., Ltd. 61-165428 (JP, A) JP-A-60-172707 (JP, A) JP-A-60-11706 (JP, A) JP-B-59-8684 (JP, B2)

Claims (16)

(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 a hydraulic pump and the first and second hydraulic actuators;
Pump control means for maintaining a pressure higher than the maximum load pressure of the hydraulic actuator by a predetermined value,
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. A second valve means for controlling a differential pressure of the outlet pressure, wherein the second valve means is introduced such that an inlet pressure of the first valve means is loaded in a valve closing direction. And a second control pressure chamber which is introduced so that the outlet pressure of the second valve means is loaded in the valve opening direction, wherein the first and second hydraulic pressures are provided. First and second pressure detecting means for respectively detecting a load pressure of the actuator, and a first differential for detecting a differential pressure between a discharge pressure of the hydraulic pump and a maximum load pressure of the first and second hydraulic actuators. Pressure detection means, the maximum load pressure and the load pressure of the first hydraulic actuator A second differential pressure detecting means for detecting a differential pressure between the first and second hydraulic actuators; a third differential pressure detecting means for detecting a differential pressure between the maximum load pressure and the load pressure of the second hydraulic actuator; A third control pressure chamber provided in each of the second valve means of the second flow control valve means;
And the pressure signal of the second pressure detecting means and the differential pressure signal of the first, second and third differential pressure detecting means are inputted, and the control amount of each of the second valve means is calculated. Control means for controlling the second valve means by outputting a control pressure based on a control amount to each of the third control pressure chambers of the second valve means, wherein the control means comprises: As a result of introducing a control pressure based on the control amount into the third control pressure chamber for each of the second valve means, an inlet pressure and an outlet of the first valve means controlled by the second valve means The differential pressure of the pressure, the differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure, the differential pressure between the maximum load pressure and the self-load pressure of each hydraulic actuator, and the self-load pressure, the following: The second valve means is controlled as expressed by the following equation: ΔPz = α (Ps−Plmax + Β (Plmax−P1) + γP1 where ΔPz: differential pressure between the inlet pressure and the outlet pressure of the first valve means Ps: discharge pressure of the hydraulic pump Plmax: maximum load pressure of the first and second hydraulic actuators 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 α, β, γ A hydraulic drive device wherein each of the control means is set as a predetermined value. 2. The system according to claim 1, wherein the first constant α is set to a positive value corresponding to a proportional gain between an operation amount of the operation means and a flow rate passing through the first valve means. Hydraulic drive. 3. 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 a combined operation is performed. Hydraulic drive. 4. The hydraulic drive device according to claim 1, wherein said third constant γ is set to a value based on an operation characteristic of an associated hydraulic actuator. 5. The hydraulic drive device according to claim 1, wherein said second and third constants β and γ are each set to zero. 6. The hydraulic drive device according to claim 1, wherein said second valve means has a spring which is opposed to said third control pressure chamber and acts in a valve opening direction. 7. The first and second flow control valve means are each a valve body for controlling communication between an inlet port and an outlet port connected to the main circuit, and are opened corresponding to displacement of the valve body. A variable throttle that changes the degree, and a seat-type main valve having a back pressure chamber that communicates with the inlet port through the variable throttle and generates a control pressure that urges the valve body in a valve closing direction; A pilot circuit connected between the back pressure chamber of the main valve and an outlet port, wherein the first valve means is arranged as a pilot valve connected to the pilot circuit and controlling a pilot flow flowing through the pilot circuit. 2. The hydraulic system according to claim 1, wherein said second valve means is connected to said pilot circuit, and is arranged as an auxiliary valve for controlling a differential pressure between an inlet pressure and an outlet pressure of said pilot valve. Drive. 8. A ratio of a pressure receiving area of the main valve valve body receiving the discharge pressure of the hydraulic pump through the inlet port to a pressure receiving area of the main valve valve body receiving the control pressure of the back pressure chamber is represented by K.
The hydraulic drive device according to claim 7, wherein the first constant α has a relation of α ≦ K. 9. The first valve means is disposed in the main circuit as a spool-type main valve, and the second valve means is disposed as an auxiliary valve in the main circuit upstream of the main valve. The hydraulic drive device according to claim 1, wherein: 10. A hydraulic pump, comprising: at least one hydraulic pump; a plurality of hydraulic actuators respectively connected to the hydraulic pump via a main circuit and driven by pressure oil discharged from the hydraulic pump; A plurality of working members each including a revolving structure, a boom, an arm and a bucket to be 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 maintaining a discharge pressure of a hydraulic pump at a pressure higher than a maximum load pressure of the plurality of hydraulic actuators by a predetermined value, wherein the plurality of flow control valve means each include:
A first valve means for changing an opening degree according to an operation amount of the operation means, and a second valve connected in series to the first valve means for controlling a differential pressure between an inlet pressure and an outlet pressure of the valve means Means, wherein the second valve means is provided with a first control pressure chamber introduced so that an inlet pressure of the first valve means is loaded in a closing direction, and a second control means of the second valve means. In a hydraulic shovel having a second control pressure chamber and a valve which are introduced so that an outlet pressure is loaded in a valve opening direction, a plurality of pressure detecting means for respectively detecting a load pressure of the plurality of hydraulic actuators; First differential pressure detecting means for detecting a differential pressure between a discharge pressure of a pump and a maximum load pressure of the plurality of hydraulic actuators, and a differential pressure between the maximum load pressure and a load pressure of each of the plurality of hydraulic actuators. A plurality of second differential pressure detecting means for detecting; A third control pressure chamber provided in the second valve means of each of the first and second flow control valve means, a pressure signal of the plurality of pressure detection means, and a first and second difference; The differential pressure signal of the pressure detecting means is input, the control amount of each of the second valve means is calculated, and the control pressure based on the control amount is changed to the third control pressure chamber of each of the second valve means. And control means for controlling the second valve means to output to the second valve means, wherein the control means comprises a second one of the flow control valve means for at least two working members of the revolving unit, the boom, the arm and the bucket. As a result of introducing a control pressure based on the control amount into the third control pressure chamber for each of the valve means, the inlet pressure and the outlet pressure of the first valve means controlled by the second valve means Is the differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure. For the pressure difference between the maximum load pressure and the self-load pressure of each hydraulic actuator, and for the self-load pressure, the second valve means is controlled as represented by the following equation: ΔPz = α (Ps− Plmax) + β (Plmax−P1) + γP1 where ΔPz: differential pressure between the inlet pressure and the outlet pressure of the first valve means Ps: discharge pressure of the hydraulic pump Plmax: maximum load pressure of the plurality of hydraulic actuators P1: Self-load pressures α, β, γ of the plurality of hydraulic actuators: first, second, and third constants The first, second, and third constants α, β, γ are predetermined for the control means, respectively. A hydraulic excavator characterized in that it is set as a value of: 11. The flow control valve means associated with a bottom side of the boom hydraulic actuator, wherein:
11. The hydraulic shovel according to claim 10, wherein the second constant β is set to a positive value. 12. The control means according to claim 11, wherein said flow control valve means on the bottom side of said arm hydraulic actuator comprises:
11. The hydraulic shovel according to claim 10, wherein the second constant β is set to a positive value. 13. The hydraulic system according to claim 10, wherein said control means sets the second constant β to a negative value for the flow control valve means on the bottom side of the bucket hydraulic actuator. Excavator. 14. The hydraulic excavator according to claim 10, wherein said control means sets the third constant γ to a negative value for the flow control valve means related to the hydraulic actuator for the revolving unit. 15. The hydraulic control system according to claim 10, wherein said control means sets said third constant γ to a positive value for flow control valve means related to the bottom side of said bucket hydraulic actuator. Excavator. 16. The control means sets the second and third constants .beta. And .gamma. To zero for the flow control valves on the rod side of the boom and arm hydraulic actuators, respectively. Item 10. The hydraulic excavator according to item 10.