JP2706564B2 - Hydraulic drive for tracked vehicles - Google Patents

Description

Description: TECHNICAL FIELD The present invention relates to a hydraulic drive device for a tracked vehicle such as a hydraulic shovel, and more particularly, to a pair of left and right traveling motors using hydraulic oil of a hydraulic pump through a plurality of pressure compensating valves. TECHNICAL FIELD The present invention relates to a hydraulic drive device capable of separately supplying a plurality of actuators including a plurality of actuators and performing a combined operation of traveling and other operations.

BACKGROUND ART A hydraulic excavator is an example of a tracked vehicle in which a plurality of operations including traveling and other operations are performed by a plurality of actuators including a pair of left and right traveling motors. Hydraulic excavator
A lower traveling body including a pair of right and left footwear for moving the hydraulic shovel, an upper revolving body pivotally mounted on the lower traveling body, and a front mechanism including a boom, an arm, and a bucket. I have. Various equipment such as a driver's cab, a prime mover, and a hydraulic pump are mounted on the upper swing body, and a front mechanism is attached.

By the way, as described in Japanese Patent Application Laid-Open No. 60-11706, for example, a hydraulic drive device used for this kind of tracked vehicle has a discharge pressure of a hydraulic pump that is a fixed value that is higher than a maximum load pressure of a plurality of actuators. There is a load sensing system provided with a pump regulator that controls the pump discharge flow rate so as to increase the flow rate only to increase the flow rate required for driving the actuator from the hydraulic pump. In this load sensing system, generally, a pressure compensating valve is generally disposed upstream of each flow control valve, whereby a differential pressure across the flow control valve is set to a specified value determined by a spring of the pressure compensating valve. Is held. By arranging the pressure compensating valve in this way and maintaining the differential pressure across the flow control valve at a specified value, when a plurality of actuators are driven simultaneously, the differential pressure across the flow control valves for all of the actuators increases. The value is held at the specified value, so that the flow control can be accurately performed by all the flow control valves irrespective of the fluctuation of the load pressure, and the combined driving of the actuator can be stably performed at a desired driving speed.

Further, in the load sensing system described in Japanese Patent Application Laid-Open No. 60-11706, a means for opposing the pump discharge pressure and the maximum load pressure is provided instead of the spring of the pressure compensating valve, and the differential pressure between the two is provided. Is used to set the specified value. As described above, the differential pressure between the pump discharge pressure and the maximum load pressure is maintained at a constant value by the pump regulator. Therefore, it is possible to use the differential pressure between the pump discharge pressure and the maximum load pressure instead of the spring, and to set the specified value of the differential pressure across the flow control valve, thereby enabling a stable combined drive of the actuator as described above. Become.

When the differential pressure is used instead of the spring, the hydraulic pump saturates, and when the discharge flow rate becomes insufficient with respect to the required flow rate, the differential pressure between the pump discharge pressure and the maximum load pressure decreases, and this drop Since the same differential pressure is applied to all the pressure compensating valves, the differential pressure across all the flow control valves is uniformly maintained at a value smaller than the normal specified value. As a result, when the pump discharge flow rate is insufficient, it is avoided that a large flow rate is preferentially supplied to the actuator on the low load side,
The pump discharge flow is divided at a ratio corresponding to the ratio of the required flow. That is, the pressure compensating valve exerts a branch flow compensating function even when the hydraulic pump is saturated. By this shunt compensation function, even when the hydraulic pump is saturated, the drive speed ratio of the plurality of actuators is appropriately controlled, and stable combined driving of the actuators is enabled.

In addition, the pressure compensating valve installed so as to exert the shunt compensation function even when the hydraulic pump is saturated,
In this specification, it is referred to as a “shunt compensating valve” for convenience.

However, the above-described conventional hydraulic drive has the following problems.

In the conventional hydraulic drive device, the traveling speed is basically controlled by operating the operation levers of the left and right traveling motors. However, when the traveling speed is reduced by the same operation amount of the operation lever, the discharge of the hydraulic pump is controlled. It has been practiced to reduce the flow rate to reduce the flow rate of the pressure oil supplied to the left and right traveling motors. The decrease in the discharge flow rate of the hydraulic pump is performed, for example, by reducing the rotation speed of a prime mover that drives the hydraulic pump. That is, by reducing the rotation speed of the prime mover, the maximum discharge flow rate of the hydraulic pump is reduced. If this maximum discharge flow rate is smaller than the required flow rate when the operation lever is operated, the hydraulic pump becomes saturated and travels left and right. The flow rate of the pressure oil supplied to the motor decreases, and the traveling speed decreases.

However, in the related art in which the rotation speed of the prime mover is reduced and the traveling speed is reduced by decreasing the discharge flow rate of the hydraulic pump in this manner, during a combined operation including traveling, for example, during a combined operation of traveling and boom, When the traveling speed is reduced and the vehicle is to be traveled slowly, the reduced discharge flow rate of the hydraulic pump is divided at a ratio corresponding to the ratio of the required flow rate by the above-described operation of the diversion compensating valve. And the operating speed of the boom cylinder is also reduced, resulting in a reduction in work efficiency. Conversely, if the running speed is increased to increase the running speed during the combined operation of the running and the boom, the operating speed of the boom cylinder also increases, and there is a concern that the safety may be degraded depending on the type of work. As described above, in the conventional hydraulic drive device, there is a problem that the work efficiency and the safety decrease due to the change in the traveling speed in the combined operation including the traveling.

An object of the present invention is to provide a hydraulic drive device for a tracked vehicle that can change the traveling speed without depending on the discharge flow rate of a hydraulic pump.

DISCLOSURE OF THE INVENTION In order to achieve the above object, according to the present invention, a hydraulic pump, a plurality of left and right traveling motors driven by pressure oil discharged from the hydraulic pump, and a plurality of other motors including at least one other actuator Actuator and
A plurality of flow control valves that respectively control the flow of the pressure oil supplied to these actuators, and a plurality of diversion compensating valves that respectively control differential pressures before and after these flow control valves, wherein the plurality of diversion compensating valves are provided. A drive signal for setting a target value of a differential pressure across the corresponding flow control valve, a hydraulic drive device for a tracked vehicle, wherein a switching signal for changing an operation speed of the pair of traveling motors is output. And means for controlling a driving means of a shunt compensation valve related to the pair of traveling motors in accordance with a switching signal output from the first means and a target value of a differential pressure across the corresponding flow control valve. And a second means for changing. The hydraulic drive device of the tracked vehicle is provided.

Since the present invention is configured as described above, it is sufficient to operate the first means when it is desired to change the traveling speed,
As a result, the second means functions to control the driving means of the flow compensating valve relating to the pair of traveling motors, thereby changing only the differential pressure across the corresponding flow control valve. Due to the change, the flow supplied to the pair of traveling motors varies, so that the traveling speed can be varied without depending on the discharge flow of the hydraulic pump.

Preferably, the first means is a means having a plurality of switching positions for operating speeds of the pair of traveling motors and outputting a switching signal corresponding to each switching position.

Also preferably, the second means includes a third means for obtaining a control force according to a switching signal output from the first means, and a shunt compensating valve for controlling the control force with respect to the pair of traveling motors. To control the driving means so as to be applied to the fourth
Means.

Here, the third means is a means for storing a functional relationship between a switching signal output from the first means and a control force to be applied to a shunt compensation valve related to the pair of traveling motors, The present invention can be configured to include a means for obtaining a control force according to the switching signal from the switching signal output from the first means and the functional relationship.

The third means includes means for detecting a differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure of the plurality of actuators, and a shunt compensation valve related to the differential pressure and the pair of traveling motors. Means for storing a plurality of functional relationships with a control force to be applied, and selecting one of the plurality of functional relationships in accordance with a switching signal output from the first means;
Means for obtaining a control force corresponding to the differential pressure from the detected differential pressure and the selected functional relationship.

Also preferably, the second means calculates a control force to be applied to the shunt compensating valve and outputs a corresponding signal, and a control pressure corresponding to the calculated control force based on the signal. Control pressure generating means for generating pressure. The control force generating means includes a pilot hydraulic pressure source,
An electromagnetic proportional valve that generates the control pressure based on the hydraulic pressure source may be included.

Also preferably, the driving means of the shunt compensating valve includes means for driving the shunt compensating valve in a valve opening direction with a constant force, and a driving unit for generating a control force for driving the shunt compensating valve in a valve closing direction. The second means is configured to increase the control force when the switching signal output from the first means is a switching signal for decreasing the operating speed of the pair of traveling motors. Control the unit.

The driving means of the shunt compensation valve may be configured to include a single drive unit that generates a control force for driving the shunt compensation valve in the valve opening direction. In this case, the second means may include the When the switching signal output from the first means is a switching signal for decreasing the operating speed of the pair of traveling motors, the driving unit is controlled so that the control force is reduced.

BRIEF DESCRIPTION OF THE DRAWINGS FIGS. 1 and 2 are a side view and a top view, respectively, of a hydraulic excavator provided with a hydraulic drive device according to a first embodiment of the present invention.

FIG. 3 is a schematic diagram of a hydraulic drive device according to the first embodiment.

FIG. 4 is a schematic view of a controller which forms a part of the hydraulic drive device.

FIG. 5, FIG. 6, and FIG. 7 are diagrams respectively showing a first functional relationship, a second functional relationship, and a third functional relationship set in the controller.

FIG. 8 is a flowchart showing a processing procedure of control of the branch flow compensation valve performed by the controller.

FIG. 9 is a view for explaining the balance of the forces acting on the shunt compensating valve relating to the left running motor.

FIG. 10 and FIG. 11 are diagrams each showing characteristics obtained by controlling the flow dividing compensation valve of the hydraulic drive device shown in FIG.

FIG. 12 is a flowchart showing a processing procedure for controlling the discharge flow rate of the main pump performed by the controller.

FIGS. 13, 14, and 15 show a first functional relationship, a second functional relationship, and a third functional relationship set in the controller of the hydraulic drive device according to the second embodiment of the present invention, respectively. FIG.

FIG. 16 is a flowchart showing a processing procedure of control of the branch flow compensation valve performed by the controller.

FIG. 17 and FIG. 18 are schematic diagrams of modified examples of the discharge flow rate control means of the main pump, respectively.

FIG. 19 and FIG. 20 are diagrams each showing a modified example of the diversion compensating valve.

FIG. 21 and FIG. 22 are diagrams showing a functional relationship instead of the functional relationship shown in FIG. 6 and FIG. 7, which is set in the controller when the shunt compensating valve shown in FIG. 20 is adopted.

FIG. 23 is a view showing still another modified example of the branch flow compensating valve.

BEST MODE FOR CARRYING OUT THE INVENTION Hereinafter, preferred embodiments of the present invention will be described with reference to the drawings.

First Embodiment First, a first embodiment of the present invention will be described with reference to FIGS. In this embodiment, the present invention is applied to a hydraulic drive device of a hydraulic shovel.

Configuration As shown in FIGS. 1 and 2, a hydraulic excavator includes a lower traveling body 102 including left and right crawler tracks 100 and 101 and a lower traveling body 10.
2 includes an upper revolving body 103 rotatably mounted on the upper revolving body 103, a boom 104, an arm 105, and a bucket 106 constituting a front attachment mounted on the upper revolving body 103.
By operating a work member such as the boom 104 while driving the 0, 101, a composite operation including traveling can be performed.

The hydraulic excavator is equipped with the hydraulic drive device of the present embodiment. As shown in FIG. 3, the hydraulic drive device includes a prime mover 1, a variable displacement hydraulic pump driven by the prime mover 1, that is, a main pump 2 and a main pump 2.
Driven by pressure oil discharged from the left and right crawler described above
A pair of left and right traveling motors 3 and 4 for driving 100 and 101 and a boom cylinder 5 for driving a boom 104 which is one of the above-mentioned front attachments. A flow control valve for controlling the oil flow, that is, a left traveling directional control valve 6, a pressure compensating valve for controlling the pressure difference between the front and rear of the left traveling directional control valve 6, that is, a branch flow compensating valve 7, and a main pump 2 is a flow control valve for controlling the flow of the pressure oil supplied to the right traveling motor 4, that is, a right traveling direction control valve 8, and a pressure compensating valve for controlling a differential pressure across the right traveling direction control valve 8. That is, the shunt compensation valve 9,
A flow control valve for controlling the flow of pressurized oil supplied from the main pump 2 to the boom cylinder 5, that is, a boom directional control valve 10
And a pressure compensating valve for controlling the pressure difference between the front and rear of the boom directional control valve 10, that is, a diversion compensating valve 11.

The flow control valves 6, 8, and 10 have left and right traveling motors 3,
Detection pipes 6a, 8a and 10a for extracting the load pressure of the boom cylinder 5 and the boom cylinder 5 are connected, and the higher one of the load pressures transmitted to the detection pipes 6a and 8a is selected by the shuttle valve 12. The higher one of the load pressures output to the detection pipeline 12a and transmitted to the detection pipelines 10a and 12a, that is, the maximum load pressure is selected by the shuttle valve 13 and output to the detection pipeline 13a.

The branch flow compensating valves 7, 9, 11 are connected to the load pressures (the outlets of the corresponding flow control valves 6, 8, 10) extracted through the detection lines 6a, 8a, 10a via the lines 7a, 9a, 11a, respectively. Pressures PL1, PL2, PL3 are guided, and drive units 7x, 9x, 11x for urging the branch flow compensating valve in the valve opening direction.
And corresponding flow control valves 6, 8, 10 via lines 7b, 9b, 11b
The pressures Pz1, Pz2, and Pz3 on the inlet side are guided, and the same constant pilots, which will be described later, are driven through driving units 7y, 9y, and 11y that urge the branch flow compensating valves in the valve closing direction, and pipes 7c, 9c, and 11c. The pressure Ps is guided, and control pressures Fa, Fa, and Fb, which will be described later, are guided through the drive units 7d, 9d, and 11d that urge the branch flow compensation valve in the valve opening direction, and the pipelines 7e, 9e, and 11e. Drive units 7f, 9f, 11f for urging the compensating valve in the valve opening direction
And The drive units 7x, 9x, 11x and 7y, 9y, 11y feed back the differential pressures Pz1-PL1, Pz2-PL2, and Pz3-PL3 of the flow control valves 6, 8, and 10, and drive units 7d, 9d,
11d and 7f, 9f, 11f are for setting the target value of the differential pressure before and after the pilot pressure and the control pressure described above to these driving units, thereby driving the driving units 7d, 9d, 11d.
The target value of the differential pressure before and after is set in accordance with the difference between the control force generated in the control unit and the control force generated in the drive units 13d, 15d, and 17d.
The control is performed so that the differential pressure between 6, 8, and 10 is maintained at the target value.

In addition, the hydraulic drive device of the present embodiment includes a variable displacement mechanism for the main pump 2, that is, a pump regulator 20 for controlling the amount of displacement (displacement volume) of the swash plate 2a, and a rotational drive synchronized with the main pump 2. Connected to a pilot pump 21 and a discharge pipe line 2b connected to a discharge port of the main pump 2,
A main relief valve 22 that regulates the maximum pressure of the pressure oil discharged from the main pump 2, and a pilot pressure that is connected to a pilot line 21 a connected to a discharge port of the pilot pump 21 and regulates a pilot pressure discharged from the pilot pump 21. And a relief valve 23 to be provided. The pilot pipe 21a is connected to the pipes 7c, 9c, and 11c related to the above-described branch flow compensating valves 7, 9, and 11,
The pilot pressure is guided to the driving units 7d, 9d, 11d.

The pump regulator 20 is connected to the swash plate 2a of the main pump 2 and drives an actuator 19 for driving the swash plate 2a, and two electromagnetic switching valves 24 and 25 for controlling the drive of the actuator 19.
And The actuator 19 is a double-acting cylinder device having a piston 19a having different pressure receiving areas on both end faces, and a small-diameter cylinder chamber 19b and a large-diameter cylinder chamber 19c located on both end faces of the piston 19a.
19b is always connected to the pilot line 21a
The tank 26 is communicated to the tank 26 via 24, 25 and the large-diameter cylinder chamber 19c
Is connected to a pilot line 21a via an electromagnetic switching valve 24 and to a tank 26 via an electromagnetic switching valve 25.

In the illustrated position where the two electromagnetic switching valves 24 and 25 are in the closed position, the piston 19a is stopped, and therefore, the tilt amount of the swash plate 2a does not change. When the solenoid valve 24 is switched to the open position, the piston
The piston 19a is moved leftward in the figure due to the pressure receiving area difference between both end surfaces of the 19a, and the tilt amount of the swash plate 2a is increased. That is, the discharge flow rate of the pump 2 increases. When the electromagnetic switching valve 25 is switched to the open position with the electromagnetic valve 24 in the closed position shown in the figure, the large-diameter cylinder chamber 19c communicates with the tank 26, the piston 19a is moved rightward in the figure, and the amount of tilt of the swash plate 2a is Decrease. That is, the main pump 2
Discharge flow rate decreases. Thus, the electromagnetic switching valve 2
The discharge flow rate of the main pump 2a is controlled by turning on and off the 4, 25.

The hydraulic drive device of the present embodiment is further connected to the detection line 13a and the discharge line 2b, and is a load which is a differential pressure between the discharge pressure Ps of the main pump 2 and the maximum load pressure Pamax of the load pressure of the actuator. A differential pressure sensor 27 for detecting the sensing differential pressure ΔPLS, an operating speed of the left running motor 3 and the right running motor 4, that is, a selecting device 28 for outputting a switching signal W for changing the running speed; A controller 30 for inputting a detection signal and a switching signal W from the selection device 28 to perform a predetermined calculation and to output a control signal of the main pump 2 and control signals of the shunt compensation valves 7, 9, 11; Control pressure generating means 31 for generating control pressures Fa, Fa, Fb based on the control signals 7, 9, 11 is provided. The control signal of the main pump 2 from the controller 30 is output to the electromagnetic switching valves 24 and 25. The controller 30 and the control pressure generating means 31 are driven by a pair of left and right traveling motors 3 and 4 in accordance with the switching signal W output from the selecting device 28.
, The drive units 7f and 9f of the flow compensating valves 7 and 9 according to the first embodiment are controlled to change the target value of the differential pressure across the flow control valves 6 and 8.

The selection device 28 includes a left traveling motor 3 and a right traveling motor 4
Has four switching positions, for example, an OFF position and three ON positions of high speed, medium speed, and low speed, and a switching signal W corresponding to each switching position in the three ON positions. Is output.

The control pressure generating means 31 includes a drive unit 7 for the branch flow compensating valves 7, 9, and 11.
f, 9f, 11f electromagnetic proportional valves 32, which are respectively arranged between the pipelines 7e, 9e, 11e and the pilot pipeline 21a,
The electromagnetic proportional valves 32, 33, and 34 have their opening amounts changed according to a control signal from the controller 30, and output control pressures Fa, Fa, and Fb according to the level of the control signal.

The controller 30 includes an input unit 30 as shown in FIG.
a, a storage unit 30b, a calculation unit 30c, and an output unit 30d. The input 30a of the controller 30 is connected to the differential pressure sensor 27 and the selector 28 described above, and the output 30d is connected to the electromagnetic switching valves 24 and 25 and the electromagnetic proportional valves 32, 33 and 34 described above. Further, the storage unit 30b of the controller 30 includes:
As shown in FIG. 5, the first functional relationship between the switching signal W output from the selection device 28 and the target value of the differential pressure across the flow control valves 6 and 8, that is, the target differential pressure ΔPc, and the sixth functional relationship As shown in the figure,
As shown in FIG. 7, a second functional relationship between the target differential pressure ΔPc and a pressure correction amount PE corresponding to a control force to be generated by the drive units 7f, 9f of the branch flow compensating valves 7, 9, and a differential pressure sensor as shown in FIG. Load sensing differential pressure ΔPLS detected by 27 and drive unit of shunt compensation valve 11
The third functional relationship with the pressure compensation amount PELS corresponding to the control force to be generated in 11f is stored. Also, the storage unit 30b
Stores a target load sensing differential pressure to be held by the circuit of the hydraulic drive device shown in FIG. 3, that is, a target value ΔPx of a differential pressure ΔPLS between the pump discharge pressure and the maximum load pressure.

The first functional relationship shown in FIG. 5 is that the target differential pressure .DELTA.Pc becomes .DELTA.Pc as the switching signal W switches from low speed to medium speed and high speed.
The relationship is set to gradually increase from Pc3 to ΔPc2, ΔPc1, and the second functional relationship shown in FIG.
Since the drive units 7f, 9f of 9 apply a control force in the valve closing direction, as the target differential pressure ΔPc increases from ΔPc3 to ΔPc2, ΔPc1, the control force, that is, the pressure compensation amount PE becomes PF3.
From PF2,0 to PF2,0
The third functional relationship shown in FIG.
Imparts a control force in the valve closing direction. Therefore, when the load sensing differential pressure ΔPLS is near the target value ΔPx, the control force, that is, the pressure compensation amount PELS is 0, and the load sensing differential pressure ΔPLS is smaller than the target value ΔPx. The relationship is set so that the pressure compensation amount PFLS increases as the pressure increases.

Operation The operation of the embodiment configured as described above will be described below. First, the control of the branch flow compensating valves 7, 9, 11 of this embodiment will be described with reference to the flowchart shown in FIG.

First, as shown in step S1, the load sensing differential pressure ΔPLS detected by the differential pressure sensor 27 and the switching signal W output from the selection device 28 are read into the arithmetic unit 30c via the input unit 30a of the controller 30. . Next, the procedure proceeds to step S2, where it is determined whether or not the switching signal W is input by the operation unit 30c. Here, if the switching signal W is not input, it means that it is not specifically intended to change the traveling speed, and the process proceeds to step S3. In step S3, the storage unit 30b
The third functional relationship shown in FIG.
30c, the control force according to the load sensing differential pressure ΔPLS, that is, the pressure compensation amount PELS is calculated.

Next, the procedure proceeds to step S4, where a control signal corresponding to the same corresponding pressure compensation amount PFLS is output to each of the drive units of the electromagnetic proportional valves 32, 33, and. As a result, the same control pressure Fa, Fb (Fa (Fa)
= Fb = PFLS), and the differential pressure across the flow control valves 6, 8, and 10 is controlled to be maintained at the same target differential pressure corresponding to the pressure compensation amount PLS.

That is, the drive units 7x, 7y and 7
As shown in FIG. 9, the balance of the forces acting on d and 7f is such that the pressure receiving area of the driving unit 7x is aL1, the pressure receiving area of the driving unit 7y is az1,
The pressure receiving area of the driving section 7d is as1 and the pressure receiving area of the driving section 7f is am1.
Then, PL1 · aL1 + Ps · as1 = Pz1 · az1 + Fa · am1 (1) Here, for convenience, if aL1 = as1 = az1 = am1, then the differential pressure Pz1−PL1 of the left and right directional control valve 6 becomes Pz1 −PL1 = Ps−Fa (2) Similarly, the differential pressures Pz2-PL2, Pz3-PL3 before and after the right direction control valve 8 and the boom direction control valve 10 are also considered, and Pz2-PL2 = Ps-Fa (3) Pz3-PL3 = Ps-Fb (4) As is apparent from the equations (2) to (4), the differential pressures before and after the left-running directional control valve 6, the right-running directional control valve 8, and the boom directional control valve 10 are all equal. The ratio of the flow rate discharged from each actuator to each actuator is constant, and the flow rate according to the opening of each directional control valve is supplied to each actuator regardless of the load fluctuation of each actuator. Operations can be performed.

If the selection device 28 is switched to, for example, a low-speed position in order to reduce only the traveling speed during the combined operation of the traveling and the boom, the procedure of FIG.
When the determination in S2 is satisfied, the process proceeds to step S5. In this step S5, the first functional relationship shown in FIG. 5 and the second functional relationship shown in FIG. 6 are read out, and the first functional relationship shown in FIG. The target differential pressure ΔPc3 is determined. Next, the process proceeds to step S6, where a large pressure compensation amount PF3 corresponding to the target differential pressure ΔPc3 is obtained from the second functional relationship shown in FIG. 6 as the control force for the branch flow compensating valves 7, 9, that is, the pressure compensation amount PF. . Next, in step S7, the third functional relationship shown in FIG. 7 is read out, and a value corresponding to the load sensing differential pressure ΔPLS is obtained as the pressure compensation amount PFLS related to the shunt compensation valve 11.

Next, the procedure proceeds to step S8, where a control signal corresponding to the pressure compensation amount PE = PF3 is output to the drive units of the electromagnetic proportional valves 32 and 33, and a control signal corresponding to the pressure compensation amount PFLS is output to the electromagnetic proportional valve 34. . This allows the pilot pump 21 to
The same control pressure Fa (Fa = PF = PF3) is given to each of the drive units 7f and 9f of the branch flow compensating valves 7 and 9 via 32 and 33, and the differential pressure across the flow control valves 6 and 8 is Control is performed so as to be maintained at the target differential pressure ΔPc3. In addition, a control pressure Fb (Fb = PFLS) is given from the pilot pump 21 to the drive unit 11f of the shunt compensation valve 11 via the electromagnetic proportional valve 34, and the differential pressure across the flow control valve 10 corresponds to the pressure compensation amount PFLS. Control is performed so as to be maintained at the target differential pressure.

That is, at this time, the control pressure Fa in the above-described equations (2) and (3) corresponds to the pressure compensation amount PF3 shown in FIG.
If the load sensing differential pressure ΔPLS is near the target value ΔPx, the control pressure in the equation (4) corresponding to the pressure compensation amount PFLS
Since the value is larger than the value of Fb, the urging force in the valve closing direction acting on the shunt compensating valves 7 and 9 becomes larger than the urging force in the valve closing direction acting on the shunt compensating valve 11, and the left traveling direction control valve 6 And the differential pressure Pz1−PL1 and P of the right and left direction control valve 8
z2-PL2 are equal to each other, and the boom directional control valve 10
Is controlled to be smaller than the differential pressure Pz3−PL3.

In general, if the flow rate passing through the directional control valve is Q, the opening area of the directional control valve is A, the differential pressure across the directional control valve is ΔP, and the proportionality constant is K, From this, it is known that the flow rate Q flowing through the left-handing directional control valve 6 and the right-handing directional control valve 8 having a small front-back differential pressure flows through the boom directional control valve 10. As a result, the operating speed of the left traveling motor 3 and the right traveling motor 4 is reduced while maintaining the operating speed of the boom cylinder 5 as it is, so that only the traveling speed can be reduced.

FIG. 10 and FIG. 11 show the flow rate characteristics obtained at this time. FIG. 10 is a characteristic diagram showing the relationship between the opening areas of the left-handing direction control valve 6 and the right-handing direction control valve 8 and the flow rate Q. The characteristic line 35 indicates that the selection device 28 is in the low-speed position as described above. The characteristics at the time of switching are shown.
7 shows the characteristics when the selection device 28 is switched to the middle speed position and the high speed position, respectively. FIG. 11 shows the relationship between the opening area of the boom directional control valve 10 and the flow rate Q,
FIG. 9 is a characteristic diagram showing the relationship between the opening areas of the left and right traveling direction control valves 6 and 8 and the flow rate Q when the selection device 28 is at the OFF position.

As described above, the relationship between the opening area of each directional control valve and the passing flow rate Q when the selection device 28 is in the OFF position is unique as shown in FIG. ON
When the selector device 28 is operated to the position, the relationship between the opening areas of the left traveling direction control 6 and the right traveling direction control valve 8 and the passing flow rate Q, as shown in FIG. Characteristic line 3 depending on whether the position is switched to high speed or high speed
5, 36, 37, and the relationship between the opening area of the boom directional control valve 10 and the passing flow rate Q remains the characteristic shown in FIG. Therefore, by switching the selection device 28, only the traveling speed can be changed without changing the boom speed.

Next, control of the main pump 2 in this embodiment will be described with reference to a flowchart shown in FIG.

First, as shown in step S10, the load sensing differential pressure ΔPLS detected by the differential pressure sensor 27 is read into the arithmetic unit 30c via the input unit 30 of the controller 30. Then step S
11, the target value ΔPx of the load sensing differential pressure stored in the storage unit 30c is read, and the target value ΔPx is read.
Px is compared with the load sensing differential pressure ΔPLS.

Here, if the differential pressure ΔPLS detected by the differential pressure sensor 27 is larger than the target value ΔPx, the process proceeds to step S12, where a control signal is output from the controller 30 to the drive unit of the electromagnetic switching valve 25, and the electromagnetic switching valve 25 is switched to the open position, and the large-diameter cylinder chamber 19c of the actuator 19 communicates with the tank 26. Thereby, the small-diameter cylinder chamber 19 of the actuator 19
The piston 19a of the actuator 19 moves rightward in FIG. 3 due to the pilot pressure of the pilot pump 21 being loaded on the swash plate 2a so that the flow rate discharged from the main pump 2 decreases. That is, the displacement is changed so that the pressure difference ΔPLS is controlled to approach the target value ΔPx.

Also, the load sensing Δ detected by the differential pressure sensor 27
When PLS is smaller than the target value ΔPx, the process proceeds to step S13, where a control signal is output from the controller 30 to the drive unit of the electromagnetic switching valve 24, and the electromagnetic switching valve 24 is switched to the open position. As a result, the pilot pressure of the pilot pump 21 is applied to both the small-diameter cylinder chamber 19b and the large-diameter cylinder chamber 19c of the actuator 19, and the piston 19a moves to the left in FIG. The tilt amount of the swash plate 2a changes so that the flow rate discharged from the main pump 2 increases, and the differential pressure ΔPLS is controlled so as to approach the target value ΔPx.

When the load sensing ΔPLS detected by the differential pressure sensor 27 is equal to the target value ΔPx, the process proceeds to step S14, where the controller 30 does not output a control signal to the drive units of the electromagnetic switching valves 24 and 25, and outputs the control signals to the electromagnetic switching valves 24 and 25. Are both held in the closed position.
Thereby, the movement of the piston 19a of the actuator 19 is stopped, and the amount of tilt of the swash plate 2a is held so that the flow rate discharged from the main pump 2 becomes constant.

In the embodiment configured as described above, the selection device 28
Controls the drive units 7f, 9f of the shunt compensating valves 7, 9 related to the left and right traveling motors 3, 4 by the switching operation of the left and right traveling direction control valves 6, and the target value of the differential pressure across the left and right traveling direction control valves 8. The running speed can be changed by changing. For this reason, since it is not necessary to reduce the rotation speed of the prime mover 1, only the traveling speed can be changed independently,
Accordingly, the speed of the boom or the like can be maintained at a desired speed regardless of a change in the traveling speed during the combined operation of the traveling and the boom, and the working efficiency and the safety can be improved.

Second Embodiment A second embodiment of the present invention will be described with reference to FIGS. This embodiment is different from the first embodiment in how to obtain the pressure compensation amount for the shunt compensating valve related to the left and right traveling motor. The hardware configuration is substantially the same as that of the first embodiment shown in FIG. Therefore, in the following description, reference numbers shown in FIG. 3 are cited.

First, in the present embodiment, the selection device 28 does not have an OFF position, and has three ON positions of a high speed, a medium speed, and a low speed as a plurality of switching positions regarding the operating speeds of the left traveling motor 3 and the right traveling motor 4. And outputs a switching signal X corresponding to each switching position.

As shown in FIG. 13, the storage unit 30b of the controller 30 stores the load sensing differential pressure ΔPLS detected by the differential pressure sensor 27 corresponding to the switching signal X when the selector 28 is switched to the high-speed position. A first functional relationship with a pressure compensation amount PFt corresponding to a control force to be generated by the branch flow compensating valves 7 and 9;
As shown in FIG. 14, the load sensing differential pressure .DELTA.PLS and the pressure corresponding to the control force to be generated by the shunt compensating valves 7, 9 corresponding to the switching signal X when the selecting device 28 is switched to the middle speed position. In response to the second functional relationship with the compensation amount PFt and the switching signal X when the selector 28 is switched to the high-speed position, as shown in FIG. Pressure compensation amount P corresponding to the control force to be generated in 9
The third functional relationship with Ft is stored. Also storage unit
Similarly to the first embodiment, a fourth value 30b of the load sensing differential pressure ΔPLS shown in FIG. 7 and the pressure compensation amount PFLS corresponding to the control force to be generated by the drive unit 11f of the shunt compensation valve 11 is provided in 30b. The functional relationship and the target value ΔPx of the load sensing differential pressure are stored.

The first functional relationship shown in FIG. 13 is similar to the fourth functional relationship shown in FIG. 7, and when the load sensing differential pressure ΔPLS is near the target value ΔPx, the pressure compensation amount PFt is 0,
The relationship is set such that the pressure compensation amount PFt increases as the load sensing differential pressure ΔPLS becomes smaller than the target value ΔPx. The second functional relationship shown in FIG. 14 indicates that when the load sensing differential pressure ΔPLS is near the target value ΔPx, the pressure compensation amount PFt becomes a constant value of the pressure compensation amount PFt2 that is larger than 0, and the load sensing differential pressure ΔPLS becomes a predetermined value. Value ΔPLS2
The relationship is set so that the pressure compensation amount PFt increases as the value decreases. The third functional relationship shown in FIG. 15 is that when the load sensing differential pressure ΔPLS is near the target value ΔPx, the pressure compensation amount PFt becomes a constant value of the pressure compensation amount PFt3 which is larger than the pressure compensation amount PFt2, and the load sensing differential pressure The relationship is set such that the pressure compensation amount PFt increases as ΔPLS becomes smaller than the predetermined value ΔPLS3.

The control of the diversion compensating valves 7, 9, 11 of the embodiment configured as described above will be described based on the flowchart shown in FIG.

First, as shown in step S20, the load sensing differential pressure ΔPLS detected by the differential pressure sensor 27 and the switching signal X output from the selection device 28 are read into the arithmetic unit 30c via the input unit 30a of the controller 30. . Next, proceed to step S21.
It is determined whether the switching signal X corresponds to any of the high speed, medium speed, and low speed positions of the selection device 28. At this time, if it is assumed that the selecting device 28 has been switched to the high-speed position for the purpose of high-speed traveling in the combined device of traveling and boom, it is determined in step S21 that the switching signal X is a switching signal for the high-speed position. Move to S22. In step S22, the first functional relationship shown in FIG. 13 is read out, and at this time, if the load sensing differential pressure ΔPLS is near the target value ΔPx, a relatively small value corresponding to the differential pressure ΔPLS is calculated by the shunt compensation. It is obtained as the pressure compensation amount PFt for the valves 7, 9.

Next, the procedure proceeds to step S23, where a similarly small value corresponding to the load sensing differential pressure ΔPLS is obtained as the pressure compensation amount PFLS related to the shunt compensation valve 11 from the fourth functional relationship shown in FIG.

Next, in step S24, a control signal corresponding to the pressure compensation amount PFt is output to the drive units of the electromagnetic proportional valves 32 and 33, and a control signal corresponding to the pressure compensation amount PFLS is output to the electromagnetic proportional valve. As a result, the solenoid proportional valve 3
The control pressure Fa (Fa = PFt) of the same value is given to each of the driving parts 7f, 9f of the branch flow compensating valves 7, 9 via 2, 33, and the differential pressure across the flow control valves 6, 8 is pressure compensated. Control is performed so as to be maintained at the target differential pressure corresponding to the amount PFt. Also, the drive unit of the shunt compensation valve 11 from the pilot pump 21 via the electromagnetic proportional valve 34
The control pressure Fb (Fb = PFLS) is given to 11f, and the flow control valve 1
The differential pressure of 0 is controlled so as to be maintained at the target differential pressure corresponding to the pressure compensation amount PFLS. Here, as described above,
The functional relationship shown in FIG. 13 is similar to the functional relationship shown in FIG. Accordingly, the differential pressure between the left and right traveling direction control valves 6 and 8 becomes substantially equal to the differential pressure between the front and rear boom direction control valves 10, and the combined operation of the traveling and the boom can be performed as in the first embodiment. .

If the selector 28 is switched to the low-speed position in order to lower only the traveling speed in the combined operation of the traveling and the boom, it is determined in step S21 that the switching signal X is the switching signal for the low-speed position. Move to S25. In step S25, the third functional relationship shown in FIG.
At this time, the load sensing differential pressure ΔPLS is equal to the target value ΔPx.
If it is near, a relatively large pressure compensation amount PFt3 corresponding to the differential pressure ΔPLS is obtained as the pressure compensation amount PFt relating to the branching compensation valves 7, 9. Then, step S23 and step S2
Moving to 4, the same processing as described above is executed. As a result, a relatively high control pressure Fa (Fa = PFt3) of the same value is given to each of the driving portions 7f and 9f of the flow dividing valves 7 and 9, and the differential pressure across the flow control valves 6 and 8 is pressure compensated. Control is performed so as to be maintained at a relatively small target differential pressure corresponding to the amount PFt3.
In addition, the control pressure Fb (Fb = PFL) is applied to the drive unit 11f of the branch flow compensating valve 11.
S), the differential pressure across the flow control valve 10 is
Is controlled to be maintained at the same target differential pressure as that at the time of the high-speed position.

Therefore, similarly to the first embodiment, the operating speed of the left traveling motor 3 and the right traveling motor 4 is reduced while the operating speed of the boom cylinder 5 is maintained, and only the traveling speed can be reduced.

Further, if the selection device 28 is switched to the medium speed position with the intention of slightly lowering the traveling speed only from the high speed in the combined operation of the traveling and the boom, in step S21, it is determined that the switching signal X is the switching signal for the medium speed position. It is determined, and it moves to step S26. In step S26, the second functional relationship shown in FIG. 14 is read out, and if the load sensing differential pressure ΔPLS is near the target value ΔPx at this time, the above-described pressure compensation amount PF3 corresponding to the differential pressure ΔPLS Pressure compensation amount smaller than PFt
2 is obtained as the pressure compensation amount PFt related to the branch flow compensating valves 7 and 9. Therefore, in this case, only the differential pressure before and after the flow control valves 6 and 8 is controlled to be maintained at the target differential pressure that is smaller than that at the high-speed position but larger than that at the low-speed position. it can.

As described above, according to this embodiment, substantially the same effects as those of the first embodiment can be obtained. Further, according to the present embodiment, the pressure compensation amounts of the branch flow compensating valves 7, 9 relating to the left and right traveling motors 3, 4 are also determined from the load sensing differential pressure ΔPLS. When the so-called saturation of the main pump 2 in which the flow rate is insufficient occurs, the differential pressure ΔPLS becomes a predetermined value, for example, when the selecting device 28 is switched to the low-speed position, it becomes ΔPLS3 or less.
As in the case of the shunt compensating valve 11 related to the boom cylinder 5 shown in FIG. 7, the pressure compensation PFt is increased, so that the shunt compensation control is performed, and the combined operation of the traveling and the boom according to the required flow rate ratio can be performed. .

Other Embodiments Still other embodiments of the present invention will be described with reference to FIGS. 17 to 22.

First, a modified example of the discharge flow rate control means for controlling the discharge flow rate of the main pump 2 will be described with reference to FIG. 17 and FIG.

FIG. 17 shows an example in which the discharge flow control means is constituted only by a hydraulic control type pump regulator 39. The pump regulator 39 is connected to the swash plate 2a of the main pump 2,
An actuator 40 that drives the actuator 2a, and a switching valve 41 that controls the communication between the bottom chamber of the actuator 40 and the rod 26 of the tank 26 and the actuator 40 to control the driving of the actuator 40 are provided. The switching valve 41 has two opposing driving units 41a and 41b, and the discharge pressure Ps of the main pump 2 is guided to one driving unit 41a via a pipe line 42, and the other driving unit
The maximum load pressure Pamax extracted by the above-mentioned detection line 13a (see FIG. 3) is led to a line 41b via a line 43.
Further, a spring 41c for setting the load sensing differential pressure is arranged on the side to which the maximum load pressure is led.

In the pump regulator 39 configured as described above,
When the maximum load pressure guided to the drive unit 41b increases, the switching valve 4
1 is driven to the left in the figure, the pressure oil in the bottom chamber of the actuator 40 is discharged to the tank 26, and the piston is moved to the left in the figure to increase the tilt angle of the swash plate 30a. Conversely, when the maximum load pressure decreases, the switching valve 41 is driven rightward in the figure,
The bottom chamber of the actuator 40 communicates with the rod chamber,
Due to the pressure receiving area difference between the bottom chamber and the rod chamber, the piston is moved rightward in the figure to reduce the tilt angle of the swash plate 2a.
Thus, the pump discharge flow rate is controlled so that the differential pressure between the pump discharge pressure and the maximum load pressure is maintained at a value set by the spring 73.

FIG. 18 shows an embodiment in which the control of the pump discharge flow rate is performed without depending on the load sensing differential pressure. That is, the discharge flow rate control means shown in FIG.
A pump regulator 20 having the same configuration as that of the embodiment shown in FIG.
The tilt angle sensor 44 detects the tilt angle of the main pump 2 and outputs a detection signal to an input section of the controller 30A.
A command device 45 for outputting the command signal to the input section of A. The command device 45 obtains the total required flow rate from the total of the operation amounts of the operation levers related to the plurality of flow control valves,
The target tilt angle can be determined using the total required flow rate as the target discharge flow rate.

In the discharge flow rate control means, the value of the command signal from the command device 45 and the value detected by the tilt angle sensor 44 are compared in the calculation unit of the controller 30A, and a drive signal corresponding to the difference is output from the controller 30A. The main pump 2 discharges a command signal value, that is, a flow rate corresponding to a target discharge flow rate, selectively to a drive section of the electromagnetic switching valves 24 and 25 from the section.

Even when the discharge flow rate control means of the main pump 2 is configured as described above, the configuration of the first or second embodiment described above is employed for controlling the branch flow compensation valves 7, 9, 11 (see FIG. 3). Thereby, the same effects as those of the embodiments can be obtained.

Next, modified examples of the drive means of the branch flow compensation valve are shown in FIGS.
This will be described with reference to the drawings. Each of the shunt compensating valves will be described as an alternative to the shunt compensating valve 7 related to the left running motor 3 shown in FIG.

The shunt compensating valve 46 shown in FIG. 19 is provided with a spring 47 instead of the driving unit 7d (see FIG. 3) to which the pilot pressure of the shunt compensating valve 7 is guided. This shunt compensation valve 46 is
It will be apparent that it works similarly to.

The diversion compensating valve 48 shown in FIG. 20 is provided with a single driving unit 49 for applying a control force in the valve opening direction instead of the two driving units 7d and 7f of the diversion compensating valve 7. The control pressure Fa from the electromagnetic proportional valve 32 (see FIG. 3) is led to the section 49. In this case, the switching signal W stored in the storage unit of the controller
The first functional relationship between the target differential pressure ΔPc and the pressure compensation amount PF is the same as that shown in FIG.
6 differs from the characteristic in FIG. 6 in that the drive unit 49 applies a control force in the valve opening direction, and the target differential pressure ΔPc increases from ΔPc3 to ΔPc2, ΔPc1 as shown in FIG. Therefore, the pressure compensation amount PF is set to have a relationship that increases similarly from PF3 to PF2 and PF1. FIG. 7 also shows a third functional relationship between the load sensing differential pressure ΔPLS and the pressure compensation amount PFLS when this type of flow compensating valve is used in place of the flow compensating valve 11 related to the boom cylinder 5. Unlike the characteristic, as shown in FIG. 22, the differential pressure ΔPLS is equal to the target value ΔPx
The relationship is set such that the pressure compensation amount PFLS becomes smaller as it becomes smaller.

The shunt compensating valve 48 includes two driving units 7
The control force in the valve-opening direction related to the target differential pressure is obtained by d and 7f, but the control force in the same valve-opening direction is obtained by one drive unit 49. By giving the obtained control pressure, the shunt compensation valve 7 shown in FIG.
Or substantially the same action as 11 is obtained. Further, the shunt compensating valve 48 has a simple structure because only one drive unit is required, and therefore, a manufacturing error can be suppressed to be small, and the control accuracy is excellent.

The shunt compensating valve 49 shown in FIG. 23 includes a spring 50 and a spring 50 in accordance with a control pressure Fa on the side that applies a control force in the valve opening direction instead of the driving units 7d and 7f that are arranged opposite to the shunt compensating valve 7. A preset force varying means 51 for varying the preset force is provided. In this case also, a shunt compensation valve is set by setting a functional relationship as shown in FIGS. 21 and 22 in the storage unit of the controller. The same operation as that of 7 or 11 is obtained. In addition, the branch flow compensating valve 49 can set the pressure receiving area of the preset force variable means 51 irrespective of the size of the pressure receiving area of the drive unit 50, and thus has an advantage that the degree of freedom in design and manufacture is large.

INDUSTRIAL APPLICABILITY As described above, according to the hydraulic drive system for a tracked vehicle of the present invention, the traveling speed can be changed without depending on the discharge flow rate of the main pump, and therefore, the traveling and other In the combined operation with the above operation, the traveling speed can be changed without affecting the other operation speed, and there is an effect that the work efficiency and safety in other operations can be improved as compared with the conventional operation.

Claims (9)

(57) [Claims] A plurality including a hydraulic pump (2), a pair of left and right traveling motors (3, 4) driven by pressure oil discharged from the hydraulic pump, and at least one other actuator (5). Actuators, a plurality of flow control valves (6, 8, 10) for controlling the flow of pressure oil supplied to these actuators, and a plurality of branch flow compensating valves for controlling the differential pressure across the flow control valves respectively (7,8,
11), wherein each of the plurality of flow compensating valves has a driving means (7d, 7f, 9d, 9f, 11d, 11f) for setting a target value of the differential pressure across the corresponding flow control valve. In a hydraulic drive system for an articulated vehicle, first means (28) for outputting a switching signal for changing the operating speed of the pair of traveling motors (3, 4); and a switching signal output from the first means. According to said 1
Second means for controlling the driving means (7f, 9f) of the flow dividing valves (7, 9) related to the pair of traveling motors and changing the target value of the differential pressure across the corresponding flow control valves (6, 8). (30, 31) A hydraulic drive device for a tracked vehicle, comprising: 2. A hydraulic drive system for a tracked vehicle according to claim 1, wherein said first means sets a plurality of switching positions for operating speeds of said pair of traveling motors (3, 4). And a means (28) for outputting a switching signal corresponding to each switching position. 3. The hydraulic drive system for a tracked vehicle according to claim 1, wherein said second means controls a control force in accordance with a switching signal output from said first means. The third means (30) to be determined, and the driving means (7f, 9f) are controlled so that the control force is applied to the shunt compensation valves (7, 9) related to the pair of traveling motors (3, 4). And a fourth means (31) for performing a hydraulic drive for a tracked vehicle. 4. A hydraulic drive system for a tracked vehicle according to claim 3, wherein said third means (30) is provided with:
The functional relationship between the switching signal (W) output from the means (28) and the control force to be applied to the shunt compensation valves (7, 9) relating to the pair of traveling motors (3, 4) is stored. Means (30b)
And a means (30c) for obtaining a control force according to the switching signal from the switching signal output from the first means (28) and the functional relationship. Drive. 5. The hydraulic drive system for a tracked vehicle according to claim 3, wherein said third means (30) is configured to control a discharge pressure of said hydraulic pump (2) and said plurality of actuators (3, 3). Means for detecting the differential pressure from the maximum load pressure of (4,5) (2
Means (30b) storing a plurality of functional relationships between the differential pressure and the control force to be applied to the shunt compensation valves (7, 9) relating to the pair of traveling motors (3, 4); Selecting one of the plurality of functional relationships in accordance with the switching signal (X) output from the first means (28), and determining the difference from the detected differential pressure and the selected functional relationship. Means (30c) for obtaining a control force according to the pressure. 6. The hydraulic drive system for a tracked vehicle according to claim 1, wherein said second means calculates a control force to be applied to said flow compensating valve (7, 9, 11). And a controller (30) for outputting a corresponding signal; and a control pressure generating means (31) for generating a control pressure corresponding to the calculated control force based on the signal. Vehicle hydraulic drive. 7. A hydraulic drive system for a tracked vehicle according to claim 6, wherein said control force generating means (31) includes a pilot hydraulic source (21) and said control pressure based on said hydraulic source. A hydraulic drive device for a tracked vehicle, comprising: a generated proportional solenoid valve (32-34). 8. The hydraulic drive system for a tracked vehicle according to claim 1, wherein the drive means of the shunt compensating valve (7, 9, 11) opens the shunt compensating valve with a constant force. Direction driving means (7d, 9d, 11d), and a driving unit (7f, 9f, 11f) for generating a control force for driving the branching compensation valve in the valve closing direction,
The second means (30, 31) is adapted to switch the switching signal output from the first means (28) to the pair of traveling motors (3, 4).
A hydraulic drive device for a tracked vehicle, wherein the drive unit is controlled so that the control force is increased when the switching signal is a switching signal for reducing the operation speed of the vehicle. 9. A hydraulic drive system for a tracked vehicle according to claim 1, wherein said drive means for said diverting compensation valve (48) generates a control force for driving said diverting compensation valve in a valve opening direction. The second means (30, 3).
1) The control means may reduce the control force when the switching signal output from the first means (28) is a switching signal for decreasing the operating speed of the pair of traveling motors (3, 4). A hydraulic drive system for a tracked vehicle, wherein the drive unit is controlled.