JP2768491B2 - Hydraulic drive for tracked vehicles - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION <Industrial Application Field> The present invention is provided in a tracked vehicle such as a hydraulic shovel,
The present invention relates to a hydraulic drive device capable of performing a running operation by splitting and supplying pressure oil of a main hydraulic pump to each of a left traveling motor and a right traveling motor via respective flow dividing valves.

<Prior Art> FIG. 19 is a circuit diagram showing a hydraulic drive device of a hydraulic shovel as an example of this type of conventional hydraulic drive device of a tracked vehicle.

The conventional hydraulic drive device shown in FIG.
A variable displacement hydraulic pump driven by the prime mover 1, that is, a main hydraulic pump 2; a left running motor 3 driven by pressure oil discharged from the main hydraulic pump 2 to drive a crawler belt (not shown); A traveling motor 4 is provided.

Also, a flow control valve for controlling the flow of pressure oil supplied from the main hydraulic pump 2 to the left traveling motor 3, that is, a left traveling direction control valve 6, and a differential pressure across the left traveling direction control valve 6 is controlled. A flow control valve for controlling the flow of pressure oil supplied from the main hydraulic pump 2 to the right traveling motor 4, that is, a right traveling direction control valve 8; And a shunt compensation valve 9 for controlling the pressure difference between the front and rear.

One of the driving portion 7a of the flow compensating valve 7, the control force Fa ₁ by the pressure and the load pressure of the upstream side of the flow dividing compensation valve 7 is given by the shunt compensating valve 7 is opened, the other of the drive unit 7b The control force Fa by the pressure on the downstream side of the shunt compensating valve 7 and the maximum load pressure of the circuit led through the shuttle valve 12.
₂ is provided so that the shunt compensating valve 7 is closed, and similarly, the control force Fb ₁ by the pressure on the upstream side of the shunt compensating valve 9 and the load pressure is applied to one drive unit 9 a of the shunt compensating valve 9. given by the shunt compensating valve 9 is opened, the other drive section 9b, the control force Fb ₂ by the maximum load pressure of the pressure and the circuit downstream of the flow dividing compensation valve 9 is the shunt compensation valve 9 closes As given.

The displacement of the main hydraulic pump 2 is supplied to a control actuator 15 driven by a flow regulating valve 14 which is switched according to the magnitude of the differential pressure between the pump pressure of the main hydraulic pump 2 and the maximum load pressure of the circuit. Controlled by

In the combined driving of the left traveling motor 3 and the right traveling motor 4, the pressure on the upstream side of the left traveling direction control valve 6 is set to P
z ₁ , the pressure on the downstream side is P _L1 , the pressure on the upstream side of the right traveling direction control valve 9 is Pz ₂ , the pressure on the downstream side is P _L2 , the pump pressure is Ps, the maximum load pressure of the circuit is Pamax, and the pump pressure is Ps and maximum load pressure Pamax
ΔP _LS , and the pressure receiving area of the drive unit on which the pressure P _L1 of the shunt compensating valve 7 acts.
a _L1 , the pressure receiving area of the drive unit on which the pressure Pz ₁ acts is az ₁ , the pressure receiving area of the drive unit on which the pump pressure Ps acts is as ₁ , the maximum load pressure P
The pressure receiving area of the drive unit on which amax acts is am ₁ ,
The pressure receiving area of the drive unit on which the pressure P _L2 acts is a _L2 , and the pressure Pz ₂
The pressure receiving area of the drive unit on which the pressure acts is az ₂ , the pressure receiving area of the drive unit on which the pump pressure Ps acts as a ₂ , the pressure receiving area of the drive unit on which the maximum load pressure Pamax acts am _2, and a _L1 = A _L2 = az ₁ = az ₂ = am ₁ = am ₂ = as ₁ = as _{2 From} the balance of the forces acting on the drive of the shunt compensating valve 7, P _L1 · a _L1 + Ps · as ₁ = Pz ₁ · az ₁ + Pamax · am ₁ (1) Here, a _L1 = as ₁ = az ₁ = am ₁ and the difference between the pump pressure Ps and the maximum load pressure Pamax is ΔP _LS. The differential pressure P _Z1 -P _L1 across the directional control valve 6 becomes Pz ₁ -P _L1 = ΔP _LS (2).

Similarly, from the balance of the forces acting on the drive unit of the shunt compensating valve 9, P _L2 · a _L2 + Ps · as ₂ = P _Z2 · az ₂ + Pamax · am ₂ (3) where a _L2 = as ₂ = az since it is ₂ = am _2, the differential pressure Pz ₂ -P _L2 of right travel directional control valve 8, the _{P Z2 -P L2 = Ps-Pamax} = ΔP LS (4). As can be seen from the above equations (2) and (4), the differential pressure between the left and right directional control valves 6 and 8 is equal to the load sensing differential pressure ΔP due to the operation of the branch flow compensating valves 7 and 9. _Therefore , even if the load pressure of each of the left traveling motor 3 and the right traveling motor 4 changes individually,
The influence of the change in the load pressure is not exerted on other traveling motors. As a result, the shunt ratio of the pressure oil discharged from the main hydraulic pump 2 to the left traveling motor 3 and the right traveling motor 4 is always kept constant, and, for example, it is not intended to drive an actuator such as a boom cylinder or an arm cylinder (not shown). When only the left traveling motor 3 and the right traveling motor 4 are intended to be driven, a half of the discharge flow rate Q _P of the main hydraulic pump 2 is used.
Thus, the flow can be supplied in accordance with the operation amounts of the left-direction control valve 6 and the right-direction control valve 8, and the vehicle can travel straight, turn, or the like.

<Problems to be Solved by the Invention> In the conventional hydraulic drive device described above, for example, when the running speed is changed from a high speed to a low speed, the rotation speed of the prime mover 1 is reduced, and the discharge from the main hydraulic pump 2 is performed. In this case, the relationship between the flow rate Q passing through the directional control valves 6 and 8 and the opening areas of the directional control valves 6 and 8 is shown in FIG. As is clear from the characteristic diagram, the opening area H until the maximum passing flow rate is obtained at a high speed indicated by the characteristic line A1 is obtained.
And the opening area L at which the maximum flow rate at a low speed indicated by the characteristic line A3 is obtained, that is, a region B1 where the speed at a high speed can be changed, and a speed at a low speed. There is a large difference between the region B3 where the direction can be changed, and the metering, which is the relationship between the operation amount of the lever for driving the directional control valves 6 and 8 and the change in the traveling speed, is bad, and the operation by the operator becomes complicated. , It is easy to give the operator a feeling of fatigue. In addition, there is a problem that it is difficult to operate at a low speed at a low speed because the region B3 in the low speed traveling is narrow.

In FIG. 20, A2 is a characteristic line at an intermediate speed between a high speed and a low speed, that is, a characteristic line at a medium speed, and M is an opening area at which a maximum flow rate at a medium speed is obtained.

SUMMARY OF THE INVENTION The present invention has been made in view of the above circumstances in the prior art, and has as its object to provide a hydraulic drive device for a tracked vehicle capable of improving metering during traveling by using a device having a shunt compensation valve. Is to provide.

<Means for Solving the Problems> In order to achieve this object, the present invention provides a main hydraulic pump, a left traveling motor and a right traveling motor driven by hydraulic oil supplied from the main hydraulic pump, A flow control valve for controlling the flow of the pressure oil supplied to the left traveling motor and the right traveling motor, a flow compensating valve for controlling a differential pressure across these flow control valves, respectively, and discharged from the main hydraulic pump. Flow control means for controlling the flow rate, the hydraulic oil of the main hydraulic pump is supplied to each of the left traveling motor and the right traveling motor via the above-mentioned branching compensation valve and the flow control valve, respectively, to perform traveling. In a hydraulic drive device for a tracked vehicle, a switching unit that outputs a switching signal that changes the speeds of a left traveling motor and a right traveling motor, and a flow control valve in accordance with the switching signal output from the switching unit. A controller that selects a relationship between the maximum passing flow rate and the differential pressure across the flow control valve, and outputs a control force signal corresponding to the selected differential pressure, and a control force signal that is output from the controller. And a driving means for driving the diverting compensation valve.

<Operation> The present invention, as configured above, may operate the switching unit when a change in traveling speed is intended,
The controller selects a differential pressure that corresponds to the maximum flow rate of the flow control valve corresponding to the value of the switching signal output from the switching unit, and drives the flow-division compensating valve by driving means so as to correspond to the selected differential pressure. Is driven. Therefore, by previously setting the relationship between the maximum passing flow rate of the flow control valve and the front-rear differential pressure so that the front-rear differential pressure decreases as the maximum passing flow rate decreases, at high speed, the shunt compensating valve can be used. The differential pressure before and after the flow control valve increases, and a large flow rate can be supplied to the left and right traveling motors to achieve high-speed traveling. In addition, the differential pressure across the flow control valve is reduced, and a small flow rate can be supplied to the left and right traveling motors to achieve low-speed traveling, and at the same time, the ratio between the maximum passing flow rate and the differential pressure across the front and rear can be kept constant. The opening area at which the flow rate of the flow control valve reaches a maximum can be made substantially constant regardless of whether the speed is high or low, and a constant metering that is not affected by a change in speed can be obtained.

<Example> Hereinafter, a hydraulic drive device for a tracked vehicle of the present invention will be described with reference to the drawings.

FIG. 1 is a circuit diagram showing an embodiment of the present invention, in which a hydraulic drive device shown in this embodiment is applied to a hydraulic shovel.

The embodiment shown in FIG. 1 also has a prime mover 1, a main hydraulic pump 2, which is a variable displacement hydraulic pump, a left traveling motor 3, a right traveling motor 4, A flow control valve for controlling the flow of pressure oil supplied to the left traveling motor 3, ie, a left traveling direction control valve 6; and a flow control valve for controlling the flow of pressure oil supplied to the right traveling motor 4, ie, right traveling. Directional control valve 8, a shunt compensation valve 7 for controlling the differential pressure across the left directional control valve 6, a shunt compensation valve 9 for controlling the differential pressure between the right and left directional control valves 8, and the left traveling motor 3. And a shuttle valve 12 for taking out the larger one of the load pressure of the right traveling motor 4 and the load pressure of the right traveling motor 4.

Further, this embodiment has a boom cylinder 5 as another actuator, a flow control valve for controlling the flow of the pressure oil supplied to the boom cylinder 5, that is, a boom directional control valve 10, and a boom directional control valve. A quantity compensating valve 11 for controlling the pressure difference between the front and rear of the control valve 10, the pressure taken out from the aforementioned shuttle valve 12 and the boom cylinder 5.
And a shuttle valve 13 for taking out the larger one of the load pressures as the maximum load pressure Pamax.

Further, a main relief valve 22 for regulating the discharge pressure of the main hydraulic pump 2 and a differential pressure between the pump pressure Ps and the above-mentioned maximum load pressure Pamax, that is, a load sensing differential pressure ΔP _LS is detected and output as a signal. A sensor 27, a pilot pump 21 driven in synchronization with the main hydraulic pump 2, a relief valve 23 for regulating the pilot pressure of the pilot pump 21, a control actuator 20 for controlling a displacement of the main hydraulic pump 2, The pilot side pump 20 is connected to the rod side chamber 20a and the head side chamber 2 of the control actuator 20.
0b, and the electromagnetic switching valve 24
An electromagnetic switching valve 26 which is communicated with the control valve 24 and is interposed between the head side chamber 20b of the control actuator 20 and the tank 25.
And

The above-described pilot pump 21, the relief valve 23, the control actuator 20, and the electromagnetic switching valves 24 and 26 constitute a flow control means for controlling the flow discharged from the main hydraulic pump 2.

The embodiment shown in FIG. 1, the rotational speed of the prime mover 1, for example, an engine rotational speed detector 40 for detecting an engine target speed N _P, displacement of the main hydraulic pump 2 volumes, i.e. tilting position theta _N , A drive detector 42 for detecting whether the actuator other than the left and right traveling motors 3 and 4, that is, the boom cylinder 5 is driven, and a left traveling motor 3 and a right traveling motor 4. A switching unit 28 for outputting a switching signal for changing the speed, and this switching unit
According to the switching signal output from 28, the relationship between the maximum passing flow rate flowing through the left traveling direction control valve 6 and the right traveling direction control valve 8 and the differential pressure across the direction control valves 6, 8 is selected. The controller 30 includes a controller 30 that outputs a control force signal corresponding to the selected differential pressure before and after, and a drive unit that drives the shunt compensation valves 7 and 9 in accordance with the control force signal output from the controller 30.

The above-described switching unit 28 has, for example, three switching positions for switching the left traveling motor 3 and the right traveling motor 4 to high speed, medium speed, and low speed, respectively, and outputs a switching signal W corresponding to each switching position.

Further, the above-mentioned driving means is, for example, a control pressure generating means.
The control pressure generating means 31 includes the pilot pump 21 and the relief valve 23, and the driving units 7b, 9 of the pilot pump 21 and the flow dividing valves 7, 9, 11, respectively.
Electromagnetic valves 32, 33 and 34 are provided between b and 11b.

The controller 30 includes a signal output from the differential pressure sensor 27, a signal output from the switching unit 28, a signal output from the engine speed detector 40, and the tilt sensor 41.
And an input unit for inputting a signal output from the drive detector 42 and a signal input to the input unit, respectively. A calculation unit that performs calculation of the maximum passage flow rate, and a storage unit that stores each function relationship described below,
An output unit is provided for outputting a signal corresponding to the calculation result in the calculation unit to the electromagnetic switching valves 24 and 26 and the electromagnetic sources 32, 33 and 34, respectively.

The storage section of the control 30 stores the maximum displacement position θ _max that can provide the maximum displacement determined by the structure of the main hydraulic pump 2, the set differential pressure ΔPx that is a load sensing differential pressure desired in the circuit, and the left. In addition to the flow rate error ΔQ corresponding to a conceivable maximum value of the manufacturing error between the traveling motor 3 and the right traveling motor 4, the switching signal W of the switching unit 28 and the pump maximum dischargeable flow rate Q _P shown in FIG. And a second relationship between the differential pressure difference DP, which is the difference between the set pressure difference ΔPx and the load sensing pressure difference ΔP _LS shown in FIG. 3, and the tilt movement speed θx of the main hydraulic pump 2. functionally related, the third functional relationship between the differential pressure Pz-P _L of the maximum passing flow Q _V and left and right travel directional control valves 6, 8 of the left and right travel directional control valve 6, 8 shown in FIG. 4 If, difference before and after shown in Fig. 5 pressure Pz-P _L and the solenoid valve 32, 33
And the fourth functional relationship with the control pressure Pc applied to the drive unit.

Incidentally, the first functional relationship described above, as shown in FIG. 2, the relationship switching signal W is slow, medium, has been to be switched to the high speed position connexion maximum pump discharge flow rate Q _P gradually increases The second functional relationship described above is a relationship in which the tilting movement speed θx decreases as the differential pressure deviation DP increases. The third functional relationship described above is that the maximum passing flow rate Q _V is As the pressure becomes smaller, the differential pressure P _Z −P _L
Yes in the smaller relation fourth functional relationship described above was the differential pressure Pz-P _L is small but connexion control pressure Pc
Is larger.

The operation in the embodiment configured as described above is as follows.

If the switching unit 28 is switched to the high-speed position for the purpose of performing only traveling and for high-speed traveling, a switching signal W corresponding to a high speed is output from the switching unit 28 to the controller 30. The controller 30 performs the processing according to the procedure shown in the flowchart of FIG.

First, as shown in step S1, the first functional relationship shown in FIG. 2 stored in the storage unit is transmitted to the arithmetic unit by the switching signal W input to the arithmetic unit via the input unit of the controller 30. It is read out, and the maximum dischargeable flow rate Q of the pump corresponding to the value of the switching signal W is calculated by the arithmetic unit based on the first functional relationship.
_P (= Q _P1 ) is calculated. Next, the procedure proceeds to step S2, where the left / right traveling direction control valve 6, the flow rate error ΔQ stored in the storage unit and the maximum pump dischargeable flow rate Q _P = Q _P1
The following operation is performed to find the maximum passing flow Q _V apparent through the 8.

Q _V = ( _1/2 ) Qp ₁ + ΔQ (5) Then, the process proceeds to step S3, where the third functional relationship shown in FIG. 4 is read out by the arithmetic unit, and the arithmetic unit performs the above-described operation based on the third functional relationship. operation is performed to determine the differential pressure Pz-P _L corresponding to the maximum passing flow Q _V of. Then, proceed to step S4, where the fifth
The fourth functional relationship is read out to the arithmetic unit shown in figure, the control pressure Pc corresponding to the differential pressure Pz-P _L obtained in Step S3 of the above
Is required. Next, the procedure proceeds to step S5, where a control force signal corresponding to the control pressure Pc obtained in step S4 is output to the drive units of the solenoid valves 32 and 33 shown in FIG.

Then, the solenoid valves 32 and 33 are appropriately opened, whereby the pilot pressure Pso of the pilot pump 21 is changed to the control pressure Pc and supplied to the drive units 7b and 9b of the flow dividing valves 7 and 9 respectively. As a result, the diversion compensating valves 7 and 9 are appropriately opened, and the pressure oil discharged from the main hydraulic pump 2 is divided by the diversion compensating valves 7 and 9 respectively, and the left traveling direction control valve 6 and the right traveling direction control valve 8 and is supplied to the left traveling motor 3 and the right traveling motor 4.

In this case, for example, the relationship between the forces acting on the driving portions 7a and 7b of the branch flow compensating valve 7 will be described with reference to FIG. 8, as in the above-mentioned equation (1), P _L1 · a _L1 + Ps ₀ · as ₁ = Pz ₁ · az ₁ + Pc · am ₁ (6) Here, since a _L1 = as ₁ = az ₁ = am ₁ , the pressure difference Pz ₁ −P _L1 between the left and right direction control valve 6 is expressed as Pz ₁ −P _L1 = Pso−Pc (7) Further, the differential pressure Pz ₂ -P
Considering _{L2 in the} same way, Pz ₂ −P _L2 = Pso−Pc (8) As is clear from the above equations (7) and (8),
The front-rear differential pressures Pz ₁ -P _L1 and Pz ₂ -P _L2 of the left traveling direction control valve 6 and the right traveling direction control valve 8 are equal to each other. The flow rates according to the opening degrees of the respective directional control valves 6 and 8 are supplied to the left traveling motor 3 and the right traveling motor 4 without variation in the load pressure of each other affecting the other party.
The vehicle can turn around. In this case, the differential pressure
Pz ₁ −P _L1 and Pz ₂ −P _L2 are determined by the control pressure Pc because the pilot pressure Pso is constant, and the control pressure Pc
Is the maximum passing flow rate Q _V according to the relationship shown in FIGS. 5 and 4.
Therefore, the condition that (1/2) Q _P1 + ΔQ by the above equation (5) can be supplied to each of the left traveling motor 3 and the right traveling motor 4 is maintained.

FIG. 9 is a characteristic diagram showing the relationship between the flow rate Q flowing through each of the left traveling direction control valve 6 and the right traveling direction control valve 8 and the opening area, and the characteristic line 50 indicates the case of high-speed traveling by the above-described control. It shows the characteristic of. Note that the characteristic line 51 shows the characteristic when the flow error ΔQ is not considered, that is, when there is no manufacturing error between the left traveling motor 3 and the right traveling motor 4.

Here, since the maximum flow rate passing through the left traveling direction control valve 6 and the right traveling direction control valve 8 is Q _P1 when the flow rate discharged from the main hydraulic pump 2 is high, it is (1/2) respectively.
In the case of the control according to the procedure of FIG. 6 described above, the diverting compensating valves 7 and 9 are controlled so as to correspond to the flow rate error ΔQ, ie, the directional control valves 6 and 8 are apparently (Q _P1) . Since the flow rate larger than 1/2) Q _P1 is maintained, the flow area is smaller than the opening area Sm that allows the maximum flow rate (1/2) Q _{P1 to} be obtained when there is no flow rate error ΔQ. The maximum flow rate (1/2) Q _P1 can be obtained with the opening area Sn slightly in front. The horizontal line 50a of the characteristic line 50 is a region where a manufacturing error between the left traveling motor 3 and the right traveling motor 4 can be absorbed. Accordingly, in this embodiment in which the procedure of FIG. Even if a production error occurs between the left traveling motor 3 and the right traveling motor 4, the vehicle does not curve due to any production error, and can realize straight traveling.

Then, if medium speed from this state, medium-speed position switching unit 28 with the intention of low speed running, when switched to the low speed position, the maximum pump discharge flow rate Q _P steps S1 shown in FIG. 6 is a second As shown in the figure, Q _P2 and Q _P3 smaller than Qp ₁ respectively.
And the maximum passing flow rate Q _{V of} the left and right traveling direction control valves 6 and 8
As in the case of high-speed running, Q _V = (1/2) Q _P2 + ΔQ (9) Q _V = (1/2) Q _P3 + ΔQ (10), respectively, and the characteristic line 52 in FIG. , 53 are obtained.
The characteristic lines 54 and 55 shown in FIG. 9 show the characteristics in the case of middle-speed running and low-speed running when there is no flow rate error ΔQ, respectively.

An alien control shown in FIG. 6 described above, since the maximum passing flow Q _V as shown in FIG. 4 has a smaller satisfies the connexion differential pressure Pz-P _L decreases relationship, ratio of the maximum passing flow Q _V and the differential pressure Pz-P _L is constant, therefore,
Regardless of whether the vehicle is traveling at high speed, medium speed, or low speed, the maximum flow rate can be obtained when the opening areas of the left and right direction control valves 6 and 8 reach Sn.

Next, another operation in this embodiment, that is, control of the displacement of the main hydraulic pump 2 will be described with reference to FIG.

FIG. 7 is a flowchart showing a procedure of a process performed by the controller 30 for controlling the displacement of the main hydraulic pump 2.

 Suppose now that only high-speed running is intended.

First, as shown in step S10, the switching signal W output from the switching unit 28, the engine target speed N _P output from the engine speed detector 40, the difference The load sensing differential pressure ΔP _LS output from the pressure sensor 27, the tilt position θ _N output from the tilt sensor 41, and the other actuator output from the drive detector 42, that is, the drive signal F of the boom cylinder 5 are read. .

Then proceeds to Step S11, the first functional relationship is read as shown in FIG. 2 to the arithmetic unit from the storage unit of the controller 30, the maximum pump discharge flow rate Q _P in accordance with the value of the switching signal W in the calculation unit The calculation is performed, and Q _P = Q _P1 is obtained.

Then proceeds to step S12, the following calculation seeking maximum tilt theta _P on the basis of the Q _P = Q _P1 and an engine target speed N _P obtained in Step S11 is performed by the calculation unit.

θ _P = Q _P1 / N _P (11) Then, the process proceeds to step S13, where it is determined whether or not the drive signal F of the boom cylinder 5 has been input by the arithmetic unit. now,
Since only high-speed running is intended, the drive signal F is not input, and the process proceeds to step S14.

This procedure S14, sets the tilting control elements theta ₁ to theta _P.
That is, θ _P = θ ₁ (12). Next, proceeding to step S15, the set pressure difference ΔPx stored in the storage unit is read out to the calculation unit, and the calculation unit calculates the pressure difference DP from the load sensing pressure difference ΔP _LS and the set pressure difference ΔPx as follows. It is.

DP = ΔPx−ΔP _LS (13) Then, the process proceeds to step S16, where the second functional relationship shown in FIG. 3 stored in the storage unit is read out by the arithmetic unit, and obtained by the arithmetic unit in step S15 described above. The tilt movement speed θx corresponding to the differential pressure deviation DP is calculated.

Next, the procedure proceeds to step S17, and the following calculation is performed by the calculation unit to obtain the load sensing target displacement θ _LS from the displacement speed θx and the displacement position θ _N obtained in step S16 described above.

θ _LS = θx + θ _N (14) Then, the process proceeds to step S18, where the computing unit compares the displacement position control element θ ₁ obtained in step S14 with the load sensing target displacement θ _LS obtained in step S17, and determines the minimum value. Target value θ ₂
Is performed.

Then proceeds to step S19, tilting control signal corresponding to the tilting target value theta ₂ obtained in Step S18 is outputted to the driving unit of the electromagnetic switching valve 24, 26 shown in FIG. 1 from the output of the controller 30 .

In this case, when the load sensing differential pressure ΔP _LS detected by the differential pressure sensor 27 is larger than the set differential pressure ΔPx, a signal is output from the controller 30 to the drive unit of the electromagnetic switching valve 26, and the electromagnetic switching valve 26 The position is switched to the lower position, and the head side chamber 20b of the control actuator 20 communicates with the tank 25. Then, the pilot pressure of the pilot pump 21 is supplied to the rod side chamber 20a of the control actuator 20,
The pressure in the head side chamber escapes to the tank 25, the piston of the control actuator 20 moves to the right in FIG. 1, the displacement changes so that the flow rate discharged from the main hydraulic pump 2 decreases, and the differential pressure ΔP _LS is controlled so as to approach the set differential pressure ΔPx. When the load sensing differential pressure ΔP _LS detected by the differential pressure sensor 27 is smaller than the set differential pressure ΔPx,
A 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 lower position. As a result, the pilot pressure of the pilot pump 21 is supplied to both the rod side chamber 20a and the head side chamber 20b of the control actuator 20, and the piston of the control actuator 20 is caused by the pressure receiving area difference between the head side and the rod side of the control actuator 20. Moving to the left in the figure, the displacement changes so that the flow rate discharged from the main hydraulic pump increases, and the differential pressure Δ
_PLS is controlled so as to approach the set differential pressure ΔPx. In this way, the displacement of the main hydraulic pump 2 is controlled so that the desired setting error ΔPx is maintained, and high-speed traveling can be realized in conjunction with the processing shown in FIG.

The operation of the boom is intended to be performed from the state where only the high-speed traveling is performed as described above. When the boom cylinder 5 is driven, the drive signal F is input from the drive detector 42. The determination in S13 is satisfied, and the routine goes to Step S20 in FIG. Setting this procedure tilting position control elements theta ₁ step S20 to the maximum tilting position theta _max that is stored and the storage unit. That is, θ _P = θ _max (15), and processing is performed so that a flow rate that does not interfere with the combined operation of the traveling and the boom flows into the circuit. Thereafter, the process proceeds to step S15, and the same processing as steps S15 to S19 described above is performed.

In the embodiment configured as described above, the ninth embodiment described above is used.
As shown by the characteristic lines 50, 52, and 53 in the figure, regardless of the change in the traveling speed, the opening area where the flow rate Q flowing through the left and right traveling direction control valves 6, 8 reaches a maximum can be made substantially equal, Therefore, the metering, which is the relationship between the operation amount of the lever for driving the left traveling direction control valves 6 and 8 and the speed change, can be kept constant irrespective of the change in the traveling speed, and the feeling of fatigue for the operator can be reduced. It is possible to reduce the speed, and particularly to easily perform the low-speed operation in low-speed traveling.

In this embodiment, the maximum flow rate Q _V shown in FIG.
Includes the flow rate error ΔQ, as described above, a stable straight running can be realized regardless of the manufacturing error between the left running motor 3 and the right running motor 4.
Therefore, there is no need for the operator to operate the lever for straight running based on such a manufacturing error, and the fatigue of the operator can also be reduced in this regard.

FIG. 10 is an explanatory view showing another example of the flow control means for controlling the discharge flow rate of the main hydraulic pump 2.

In the flow control means shown in FIG. 10, the rod side chamber 20a of the control actuator 20 for controlling the displacement of the main hydraulic pump 2 is connected to the discharge line of the main hydraulic pump 2 and the rod of the control actuator 20 is controlled. A flow control valve 60 is arranged between the side chamber 20a and the head side chamber 20b. The pump pressure Ps acts on the drive unit on the spring chamber side of the flow regulating valve 60, and the maximum load pressure _Pamax acts on the other drive unit. Further, a cylinder 61 capable of restricting the movement of the piston of the control actuator 20 and this cylinder 61
A solenoid valve 62 for controlling the driving of the solenoid valve, a hydraulic source 63 connected to the solenoid valve 62, and a relief valve 64 for regulating the pressure of the pressure oil supplied from the hydraulic source 63. Controlled by the signal output from 30.

In the apparatus provided with the flow control means shown in FIG. 10, the processing shown in FIG. 11 is performed by the controller 30.

Steps S30, 31, 32, 33, 34, 35, 36, 37 in FIG.
Are the procedures S10, 11, 12, 13, 14, 17, 1 in FIG.
It corresponds almost to each of 9 and 20. That is, this tenth
Differential pressure filed by the flow rate adjusting valve 60 and the pump pressure Ps and the maximum load pressure P _amax to the flow control means shown in FIG., That is controlled by the load sensing differential pressure [Delta] P _LS, to the load sensing differential pressure [Delta] P _LS calculation by the control is not necessary according, as shown in Step S35, the tilting position control elements theta ₁ and tilt sensor
The difference from the tilt position θ _N detected by the controller 41 is obtained as the tilt target value θ _{2 by} the arithmetic unit of the controller 30, and the procedure proceeds to step S 36.
Tilting control signal corresponding to the tilting target value theta ₂ as shown in is output to the driving unit of the solenoid valve 62 shown in Figure 10. As a result, the solenoid valve 62 is switched, and pressure oil is supplied from the hydraulic pressure source 63 to the cylinder 61 via the solenoid valve 62, and the piston rod moves as appropriate to regulate the movement of the control actuator 20,
Tilting such that the target value theta ₂ is 0, i.e. the difference between the tilting position control elements theta ₁ and tilting position theta _N is controlled to be zero. Instead of the flow control means shown in FIG.
The flow control means shown in the figure may be provided.

FIG. 12, FIG. 13, and FIG. 14 are explanatory diagrams each showing another example of the diversion compensating valve.

The shunt compensating valve 66 shown in FIG. 12 receives the downstream pressure P _L of the directional control valve at one drive unit 66a, receives the force of the spring 67, and applies the upstream pressure Pz of the directional control valve to the other drive unit 66b. Yes have a structure for receiving a control pressure Pc generated by the control pressure generator with receiving, shunt compensation valve 68 shown in FIG. 13, receives the downstream pressure P _L and the control pressure Pc of the directional control valve to one of the drive unit 68a The other drive unit 68b receives only the downstream pressure Pz of the direction control valve, and the shunt compensation valve 69 shown in FIG.
It is is provided with a Purisetsuto force variation means 71 to vary the Purisetsuto force of the spring 70 in response to the control pressure Pc with receiving a downstream pressure P _L of the directional control valve on one of the drive unit 69a, the other driving portion 69b It is configured to receive only the upstream pressure Pz of the directional control valve. The shunt compensating valves 7, 9, 11 shown in FIG. 1 can be replaced with the shunt compensating valves shown in FIGS.

In particular, the shunt compensating valve 68 shown in FIG. 13 has a simple structure since it does not require a spring for biasing the drive unit.
Therefore, the manufacturing error can be kept small, and the control accuracy is excellent. Further, the shunt compensation valve 69 shown in FIG.
The pressure receiving area of the preset force varying means 71 can be set irrespective of the size of the pressure receiving area of the drive section 69a of the shunt compensating valve 69, so that the degree of freedom in designing and manufacturing is large.

 FIG. 15 is a circuit diagram showing another embodiment of the present invention.

In the embodiment shown in FIG. 15, the engine speed detector 40 and the tilt sensor 41 are removed from the configuration shown in FIG. 1 described above. The second functional relationship shown in FIGS. 3, 4, and 5, the third functional relationship, the fourth functional relationship, the fifth functional relationship shown in FIG. 16, and the second functional relationship shown in FIG. 6 and the maximum engine speed _Nmax are stored.

The fifth functional relationship described above is a relationship in which the engine speed N gradually increases as the switching signal W is switched to the low speed, the medium speed, and the high speed position, as shown in FIG. functional relationship of the 6, as shown in FIG. 17, although the engine speed N increases are the connexion maximum pump discharge flow rate Q _P gradually increases relationship.

15, 72 is a rotation speed control lever of the prime mover 1, 73 is a cylinder for rotating this rotation speed control lever,
74 is a hydraulic pressure source for supplying pressure oil for driving the cylinder 73,
75 is disposed between the hydraulic pressure source 74 and the cylinder 73,
The solenoid valve controls the flow of the pressure oil supplied from the to the cylinder 73.

FIG. 18 is a flowchart showing a procedure performed by the controller 30 provided in the embodiment shown in FIG. 15,
Among the procedures shown in FIG. 18, procedures S82, 83, 84, and 85 are the same as procedures S2, 3, 4, and 5 in FIG. 6 showing the procedure according to the embodiment shown in FIG. That is, this fifteenth
In the embodiment shown in the figure, the operation of the switching unit 28 causes the arithmetic unit of the controller 30 to store the fifth functional relationship shown in FIG. 16 stored in the storage unit, as shown in step S80 in FIG. Is read out, and the calculation section calculates the engine speed N according to the switching signal. Then, the procedure shifts to step S81 in FIG. 18, where the arithmetic unit of the controller 30 stores the 17th stored in the storage unit.
The sixth functional relationship shown in the figure is read out, and in this arithmetic unit,
Pump maximum dischargeable flow rate Q _P corresponding to engine speed N
Is calculated. Thereafter, as shown in steps S82 to S85, the same processes and operations as those in steps S2 to S5 in FIG. 6 according to the first embodiment described above are performed.

In the embodiment configured as shown in FIG. 15,
The third functional relationship shown in FIG. 4, i.e. the differential pressure Pz-P _L of the maximum passing flow Q _V in the horizontal direction control valve 6 and 8 has a smaller connexion left and right directional control valve 6, 8 gradually Since control is performed to satisfy the relationship of decreasing, a characteristic similar to the characteristic lines 50, 52 and 53 in FIG. 9 can be obtained, whereby a constant metering can be obtained. left traveling motor 3 to the maximum passing flow rate Q _V of the third functional relationship shown in FIG., by including a flow error ΔQ corresponding to manufacturing error between the right traveling motor 4, a left travel motor 3,
Stable straight running can be realized regardless of the manufacturing error between the right running motors 4.

Note that the drive signal F from the drive detector 42 is
When the vehicle is not driven, that is, when only traveling is performed, a control signal corresponding to the engine speed N obtained from the fifth functional relationship shown in FIG. While the boom cylinder 5
Is activated and the drive signal F is sent from the drive detector 42 to the controller.
When input to the controller 30, a control signal corresponding to the maximum engine speed N _max stored in the storage unit of the controller 30 is output from the controller 30 to the drive unit of the solenoid valve 75,
In each case, the solenoid valve 75 is switched, the pressure oil of the hydraulic pressure source 74 is supplied to the cylinder 73, the cylinder 73 is operated, and the engine speed control lever 72 is appropriately rotated, thereby operating the switching unit 28. Is maintained at the engine speed N or the maximum engine speed _Nmax, and the flow rate of the pressure oil flowing through the circuit can be set to an amount corresponding to the type of operation.

<Effect of the Invention> Since the hydraulic drive device for a tracked vehicle of the present invention is configured as described above, the metering during traveling can be improved in comparison with the conventional device having a shunt compensation valve. Therefore, fatigue of the operator can be reduced, and operability at a low speed can be improved.

[Brief description of the drawings]

FIG. 1 is a circuit diagram showing an embodiment of a hydraulic drive device for a tracked vehicle, and FIGS. 2 to 5 are first functional relationships stored in a controller provided in the embodiment shown in FIG. 1, respectively. ~
FIGS. 6 and 7 are diagrams showing a fourth functional relationship, and FIGS. 6 and 7 are flow charts each showing a procedure of processing performed in a controller provided in the embodiment shown in FIG. 1, and FIG. 8 is an embodiment shown in FIG. FIG. 9 is a diagram for explaining the relationship between the forces acting on the drive unit of the branch flow compensating valve provided in the example, FIG. 9 is a diagram showing characteristics obtained in the embodiment shown in FIG. 1, and FIG. FIG. 11 shows an example, FIG. 11 is a flow chart showing a procedure of a process performed in a controller provided with the flow rate control means shown in FIG. 10, and FIGS. FIG. 15 is a circuit diagram showing another embodiment of the hydraulic drive device of the tracked vehicle, FIG. 16, FIG.
FIG. 17 is a diagram showing a fifth functional relationship and a sixth functional relationship stored in a controller provided in another embodiment shown in FIG. 15, and FIG. 18 is provided for the embodiment shown in FIG. 19 is a circuit diagram showing a hydraulic drive device of a conventional tracked vehicle, and FIG. 20 shows characteristics obtained in the hydraulic drive device shown in FIG. 19. FIG. 1 ... prime mover, 2 ... main hydraulic pump, 3 ... left traveling motor, 4 ... right traveling motor, 6 ... left traveling direction control valve,
7: diversion compensating valve, 8: right traveling direction control valve, 9:
Dividing compensation valve, 12, 13 Shuttle valve, 20 Actuator for control, 21 Pilot pump, 23 Relief valve, 24, 26 Electromagnetic switching valve, 27 ... Differential pressure sensor, 28 ...
Switching unit, 30 ... Controller, 31 ... Control pressure generating means (driving means), 32, 33, 34 ... Solenoid valve, 40 ... Engine speed detector, 41 ... Tilt sensor, 42 ... Drive Detector,
60: Flow control valve, 61: Cylinder, 62: Solenoid valve, 63
…… Hydraulic source, 64 …… Relief valve, 66 …… Diversion compensation valve, 67
…… Spring, 68, 69 …… Diversion compensation valve, 70 …… Spring, 71 ……
Preset force variable means, 72: engine speed control lever, 73: cylinder, 74: hydraulic pressure source, 75: solenoid valve.

──────────────────────────────────────────────────続 き Continued on the front page (58) Field surveyed (Int.Cl. ⁶ , DB name) B62D 11/00-11/24 B62D 9/00

Claims (5)

(57) [Claims] 1. A main hydraulic pump, a left traveling motor and a right traveling motor driven by pressure oil supplied from the main hydraulic pump, and a pressure oil supplied to the left traveling motor and the right traveling motor. A flow control valve for controlling the flow, a diversion compensating valve for controlling the differential pressure across the flow control valve, and flow control means for controlling the flow discharged from the main hydraulic pump. Pressurized oil through the diversion compensation valve and the flow control valve through the left traveling motor,
In a hydraulic drive device for a tracked vehicle that can supply to each of the right traveling motors and perform traveling, a switching unit that outputs a switching signal that changes the speeds of the left traveling motor and the right traveling motor, and the switching unit The relationship between the maximum flow rate flowing through the flow control valve and the differential pressure across the flow control valve is selected according to the switching signal output from the control valve, and a control force signal corresponding to the selected differential pressure is output. A hydraulic drive device for a track-mounted vehicle, comprising: a controller that drives the shunt compensation valve according to a control force signal output from the controller. 2. A hydraulic drive system for a tracked vehicle according to claim 1, wherein said drive means is control pressure generation means for generating a control pressure according to a control force signal. 3. A controller, wherein a functional relationship between a preset switching signal and a pump maximum dischargeable flow rate, a functional relationship between a preset maximum flow rate of the flow control valve and a differential pressure across the flow control valve, A storage unit for storing a functional relationship between a pre-set differential pressure before and after the flow control valve and a control pressure for driving the branch flow compensation valve, and a calculation unit for calculating the maximum passage flow rate based on the pump maximum dischargeable flow rate The hydraulic drive system for a tracked vehicle according to claim 2, further comprising: And a controller for setting a functional relationship between a preset switching signal and a prime mover speed, a functional relationship between a preset prime mover speed and a pump maximum dischargeable flow rate, and a preset maximum flow rate control valve. A storage unit that stores a functional relationship between the passing flow rate and the differential pressure across the flow control valve, and a functional relationship between a preset differential pressure across the flow control valve and a control pressure that drives the branching compensation valve, The hydraulic drive device for a tracked vehicle according to claim 2, further comprising a calculation unit that calculates a maximum passage flow rate based on the pump maximum dischargeable flow rate. 5. The controller according to claim 1, wherein the controller determines the discharge flow rate of the main hydraulic pump.
5. The hydraulic drive of a tracked vehicle according to claim 3, further comprising a calculation unit for calculating a maximum flow rate of the flow control valve by adding a predetermined flow rate error to 1/2. apparatus.