JPH02240403A - Hydraulic driving device for construction equipment for civil engineering - Google Patents

Abstract

PURPOSE:To eliminate fluctuation in a speed change caused by temperatures of actuating oil by varying a pilot pressure which serves to open branch flow compensation valves controlling differencial pressures in front and behind of flow rte control valves, according to the discharge flow rate of a main pump. CONSTITUTION:A pilot pressure is increased by a pilot pressure variable means 61 when a discharge flow rate from a main hydraulic pump 22 is high, while branch flow compensation valves 35 to 40 are so controlled as to be relatively opened, so that pressure oil is rapidly passed through rotary left-run and right- rum motors 23 to 25 and flow rate control valves 29 to 34 which drive a boom, arm, and bucket cylinders 26 to 28. When the discharge flow rate is low, the pilot pressure is decreased so that the pressure oil is slowly passed. In the case that a large rate of flow passes through the main hydraulic pump 22 even at the time of starting under a low temperature, the same rate of oil as this under the normal temperature is supplied to an actuator irrespective of viscosity of actuating oil. In the case that the discharge flow rate is constant, the actuator is driven at a constant speed irrespective of the oil temperature.

Description

[Detailed Description of the Invention] <Industrial Application Field> The present invention is provided in civil engineering/construction machines such as hydraulic excavators,
The pressure oil of the main hydraulic pump is divided and supplied to each of the plurality of corresponding actuators via the plurality of branch compensation valves,
The present invention relates to a hydraulic drive device that can perform a desired combined operation by driving these actuators in a combined manner.

<Prior Art> FIG. 12 is a circuit diagram showing a hydraulic drive system for a hydraulic excavator as an example of this type of conventional hydraulic drive system.

The hydraulic drive device shown in FIG. 12 includes a prime mover 1, a variable displacement hydraulic pump, that is, a main hydraulic pump 2, which is driven by the prime mover 1, and a boom (not shown) that is driven by pressure oil discharged from the main hydraulic pump 2. A boom cylinder 3 for rotating a packet, and an actuator including a packet cylinder 4 for rotating a packet (not shown).

In addition, there is also a flow control valve that controls the flow of pressure oil supplied from the main hydraulic pump 2 to the boom cylinder 3, that is, a boom directional control valve 5, and a shunt compensation that controls the differential pressure across the boom directional control valve 5. A valve 6, a flow rate control valve that controls the flow of pressure oil supplied from the main hydraulic pump 2 to the packet cylinder 4, that is, a packet directional control valve 7, and a pressure difference across the packet directional control valve 7. A branch compensating valve 8 is provided.

A control force Fal based on the pressure on the upstream side of the quantity compensation valve 6 and the load pressure is applied to one drive part 6a of the division compensation valve 6 so that the division compensation valve 6 opens, and the other drive part 6b
The pressure on the downstream side of this branch compensation valve 6 and the shuttle valve 9 are
．． A control force Fa2 based on the maximum load pressure of the circuit led through the circuit 10 is applied to close the branch compensating valve 6, and similarly, one drive section 8a of the branch compensating valve 8 Control force Fb due to upstream pressure and load pressure
, is applied to open the branch compensating valve 8, and the other drive unit 8b receives a control force Fb2 based on the pressure on the downstream side of the branch compensating valve 8 and the maximum load pressure of the circuit. is given to close. The displacement of the main hydraulic pump 2 is determined by a control actuator 12 driven by a flow rate regulating valve 11 that is switched according to the magnitude of the differential pressure between the pump pressure Ps of the main hydraulic pump 2 and the maximum load pressure P amax of the circuit. controlled by Further, the maximum pressure of the pressure oil discharged from the main hydraulic pump 2, that is, the relief pressure is regulated by the main relief valve 13.

For example, when the boom cylinder 3 and the packet cylinder 4 having different driving pressures are combinedly driven, the pressure on the upstream side of the boom directional control valve 5 is set as PZl, the pressure on the downstream side as PLl, and the pressure on the upstream side of the boom directional control valve 5 is set as PLl. The upstream pressure is Pa
2. The downstream pressure is PL2, the pump pressure is PS, the maximum load pressure of the circuit is P amax, the pump pressure Ps and the maximum load pressure P
The differential pressure of amax, that is, the load sensing differential pressure, is Δ
PL5, the pressure receiving area of the drive section on which the pressure PL1 of the branch compensation valve 6 acts is at, t, the pressure receiving area of the drive section on which the pressure Pzl acts is aZl, the pressure receiving area of the drive section on which the pump pressure Ps acts is asl, the maximum load The pressure receiving area of the drive section on which the pressure P amax acts is aml, the pressure receiving area of the drive section on which the pressure PL2 of the branch compensating valve 8 acts is aL2, the pressure receiving area of the drive section on which the pressure Pz2 acts is az2, and the pump pressure Ps acts. The pressure-receiving area of the drive section to which the maximum load pressure Panax acts is as2, and the pressure-receiving area of the drive section on which the maximum load pressure Panax acts is am2, and for convenience, a L1 = aZ1 = ai, 2 = aZ 2 = aSl
= aml = aS 2 = am2, then from the balance of forces acting on the drive part of the shunt compensating valve 6, PLl-aLl+P S °a S 1 = Pzl °
aZl + Pamax・aml (1) Here, aL
1=asl=aZl ”aml, pump pressure Ps
Since the differential pressure between the maximum load pressure Pamax and the maximum load pressure Pamax is set to ΔPl, the differential pressure across the boom directional control valve 5 is Pz1-PLI
is Pzl −Pt, t=Ps−Pamax=ΔPLS
(2) becomes.

Similarly, from the balance of forces acting on the drive part of the shunt compensation valve 8, PL2・aL2+Ps Has3 = Pz2− aZ2+Pamax−am2 (3) Here, a L2 = a S 2 = a Z 2 =
Since it is a m 2, the packet directional control valve 7
The differential pressure Pz2 PL2 before and after is P Z2-pL2==ps-Pamax=ΔPLS
(4) becomes.

As can be seen from equations (2) and (4) above, even if the load pressures of the boom cylinder 3 and packet cylinder 4 change individually, due to the action of the shunt compensation valve 6.8, the change in load pressure is No influence is exerted on other actuators, and as a result, the differential pressure across the boom directional control valve 5 and the differential pressure across the packet directional control valve 7 are the same ΔPLS.
is held at the value of

Therefore, the division ratio of the pressure oil discharged from the main hydraulic pump 2 to the boom cylinder 3 and the packet cylinder 4 is kept constant, and the pressure oil of the main hydraulic pump 2 is transferred to the boom cylinder 3.
, a flow rate can be supplied to each of the packet cylinders 4 according to the respective operation amounts of the boom directional control valve 5 and the packet directional control valve 7, and work by combined operation of the boom and packet, such as excavation work of earth and sand, can be performed. I can do it.

<Problems to be Solved by the Invention> By the way, in the conventional hydraulic drive device described above, when starting operation in a low-temperature environment such as in the middle of winter, the temperature of the hydraulic oil flowing through the circuit is low, so the hydraulic oil is The viscous resistance is large, and a relatively small flow rate reaches a predetermined differential pressure across the flow rate control valves for actuators such as the boom cylinder 3 and packet cylinder 4. Also, as the hydraulic oil is circulated appropriately, the oil temperature increases. However, when the viscous resistance of the hydraulic oil decreases and the temperature rises above a certain level, a relatively large flow rate will result in a predetermined differential pressure across the flow rate control valve related to the actuator. Therefore, as shown in FIG. 13, when the actuator speed is low, the oil temperature is t. and the high temperature oil temperature tl, and as a result, the operability of the actuator drive tends to deteriorate.

In addition, in this conventional hydraulic drive device, when changing the boom cylinder 3 from a high-speed drive state to a low-speed drive state, for example, the rotation speed of the prime mover 1 is lowered from a high rotation speed, and the main hydraulic pump is Although it is necessary to reduce the discharged flow rate compared to the case of high-speed drive, in this conventional hydraulic drive system, the difference between the front and back of the boom directional control valve 5 that drives the boom cylinder 3, regardless of whether it is high-speed drive or low-speed drive. Pressure Pz
1 Since the flow compensation valve 6 operates to keep pt and i constant, the directional control for the boom can obtain the maximum flow rate shown by the characteristic line A1 during high-speed operation of the directional control valve 5 for the boom in FIG. 14. The opening area of the valve 5, that is, the lever stroke SMI, and the opening area of the boom directional control valve 5 that provides the maximum flow rate shown by the characteristic line A2 during low-speed operation of the boom directional control valve 5 in FIG. The lever stroke SM2 is greatly different from the lever stroke SM2, and the metering is poor, which gives an uncomfortable operating feeling to the operator, and the operability of driving the boom cylinder 3 is likely to deteriorate.

Particularly in low-speed drive, since the maximum flow rate lever stroke is 3M2, the range for changing the actuator flow rate is small, and therefore, pressure oil is supplied little by little to the boom cylinder 3 to control the boom cylinder 3. It tends to be difficult to perform fine operation to operate slowly.

Further, it is assumed that the required flow rate of the packet directional control valve 7 corresponding to the required flow rate to drive the packet cylinder 4 is smaller than the required flow rate of the boom directional control valve 5 corresponding to the required flow rate to drive the boom cylinder 3. , when changing from high-speed combined boom and packet operation to low-speed combined operation, the boom directional control valve 5 changes from characteristic line A1 to characteristic line A in FIG. 14 as described above.
Regarding the packet directional control valve 7 whose required flow rate is small, although the required flow rate changes to 2, as shown by characteristic line B1 in FIG. A situation occurs in which the maximum flow rate does not change, and in this case as well, the operator feels uncomfortable in the operation, and the operability tends to deteriorate.

The present invention has been made in view of the actual situation in the prior art described above, and its purpose is to be able to drive an actuator at a constant speed without being affected by the high or low temperature of the hydraulic oil, and to An object of the present invention is to provide a hydraulic drive device for civil engineering and construction machinery that can ensure good metering of a flow control valve that drives an actuator regardless of changes in the number of flow control valves that drive actuators.

Means for Solving the Problems To achieve this object, the present invention provides a prime mover, a main hydraulic pump driven by the prime mover, and a plurality of pumps driven by pressure oil supplied from the main hydraulic pump. Actuators, flow control valves that control the flow of pressure oil supplied to these actuators, branch compensation valves that control the differential pressure across these flow control valves, and flow rates discharged from the main hydraulic pump. A flow rate control means that controls the pressure oil of the main hydraulic pump according to the differential pressure between the pressure of the main hydraulic pump and the maximum load pressure of the load pressure of the actuator, and the pressure oil of the main hydraulic pump is controlled by each of the above-mentioned flow compensation valve and flow control valve. In a hydraulic drive system for civil engineering and construction machinery that is capable of combined driving of these actuators, it is supplied to each of the above actuators via A pilot hydraulic pressure source is provided to supply pilot pressure to the driving parts of these branch compensation valves, and a pilot pressure variable means is provided to vary the magnitude of the pilot pressure in accordance with the magnitude of the flow rate discharged from the main hydraulic pump. The configuration is as follows.

<Function> Since the present invention is elliptical as described above, when the flow rate discharged from the main hydraulic pump is large, the pilot pressure is increased by the pilot pressure variable means, and this pilot pressure is applied to the diversion compensation valve. This is applied to the drive section of the main hydraulic pump, which drives this branch compensating valve as close as possible to allow the pressure oil to quickly pass through the flow control valve that drives the actuator.On the other hand, when the flow rate discharged from the main hydraulic pump is small, The pressure is reduced, and this pilot pressure drives the branch compensation valve toward closing, allowing pressure oil to gradually pass through the flow control valve that drives the actuator. Therefore, even when starting operation in a low-temperature environment, When a large flow rate flows from the main hydraulic pump, a pilot pressure equivalent to that at normal temperature is applied to the drive part of the diversion compensation valve to compensate for the diversion, regardless of the viscosity of the hydraulic fluid as the temperature of the hydraulic fluid decreases. When the valve is driven, the flow rate determined simply by the magnitude of the flow rate discharged from the main hydraulic pump without being affected by the high or low temperature of the hydraulic oil is divided by the flow compensation valve and supplied to the actuator. If the flow rate discharged from the pump is constant, the actuator can be driven at a constant speed regardless of the temperature of the hydraulic oil.

In addition, as mentioned above, when the flow rate discharged from the main hydraulic pump is large, the pressure oil passes through the flow control valve quickly, and when the flow rate discharged from the main hydraulic pump is small, the pressure oil passes through the flow control valve. passes gradually, the opening area of the flow control valve that produces the maximum flow rate in each case, that is, the lever stroke, can be made almost the same, allowing flow control to drive the actuator regardless of changes in the rotational speed of the prime mover. Good metering of the valve can be ensured.

Furthermore, during combined operation, hydraulic oil is supplied to each actuator through the drive of a branch compensation valve, which is controlled by the same pilot pressure and also by a pilot pressure that decreases as the flow rate discharged from the main hydraulic pump decreases. is supplied, so regardless of the magnitude of the required flow rate of the flow control valve related to each actuator, the maximum flow rate passing through each of the flow control valves can be changed depending on the magnitude of the flow rate discharged from the main hydraulic pump. It is possible to ensure good metering even in this complex operation.

<Example> Hereinafter, a hydraulic drive system for civil engineering/construction machinery according to the present invention will be explained based on the drawings.

FIG. 1 is a circuit diagram showing one embodiment of the present invention.

This embodiment is applied to a hydraulic excavator, and is driven by a prime mover, that is, an engine 21, a variable displacement hydraulic pump, that is, a main hydraulic pump 22, which is driven by the engine 21, and pressure oil discharged from the main hydraulic pump 22. A plurality of actuators, i.e., swing motors 23
, a left travel motor 24, a right travel motor 25, a boom cylinder 26, an arm cylinder 27, and a packet cylinder 28. Note that the swing motor 23 drives a rotating body (not shown), the left running motor 24 and the right running motor 25 drive crawlers (not shown), and a boom cylinder 26
, arm cylinder 27, and packet cylinder 28 drive a boom, arm, and packet (not shown), respectively.

Also, a flow control valve, that is, a swing direction control valve, controls the flow of pressure oil supplied to each of the swing motor 23, left travel motor 24, right travel motor 25, boom cylinder 26, arm cylinder 27, and packet cylinder 28. 29
, left travel direction control valve 30, right travel direction control valve 31,
Boom directional control valve 32, arm directional control valve 33, bucket directional control valve 34, and branch compensation valves 35.36.37.38.3 provided corresponding to these flow rate control valves.
9.40.

Further, the displacement volume of the main hydraulic pump 22 described above is controlled by a control actuator 41, and the drive of this control actuator 41 is controlled by a flow rate adjustment valve 42. The flow rate adjustment valve 42 controls the pump pressure Ps guided through the pipe line 43 and the maximum load pressure P am guided through the pipe line 44.
The differential pressure with ax, that is, the load sensing differential pressure ΔPL5
Driven by. These control actuators 41
The flow rate adjustment valve 42 constitutes a flow rate control means that controls the flow rate discharged from the main hydraulic pump 22 according to the load sensing differential pressure ΔPLs.

One drive section 35a of the above-described branch compensation valves 35 to 40,
Each of 36a, 37a, 38a, 39a, and 40a is connected to a pilot hydraulic power source, that is, a pilot pump 52, which is driven in synchronization with the main hydraulic pump 22, and receives a pilot pressure Ps generated by this pilot pump 52.
Control force by o is applied to open these branch compensation valves 35 to 40, and the other drive parts 35b, 36b, 37b
, 38b, 39b, and 40b are controlled by the pressure on the downstream side of these branch compensating valves 35, 36, 37, 38, 39, 40, and the control pressure described below led via the pipes 51a to 51f. The control force is applied to these branch compensation valves 35 to 4.
0 is given to close.

In addition, the pump pressure Ps and the maximum load pressure P of the actuator
Differential pressure with amax, that is, load sensing differential pressure Δp
Differential pressure detection device 53 that detects t and s and outputs it as a signal
According to the signal output from the differential pressure detection device 53, the respective drive units 35b to 40 of the branch compensation valves 35 to 40 are activated.
A control force adding means is provided for applying a control force to b.

The control force applying means described above includes a controller 59 and a control pressure generator 60. Among these, the controller 59 has an input section 55 that inputs the load sensing differential pressure ΔPLS outputted from the differential pressure detection device 53, and a load sensing differential pressure ΔP1. and the drive unit 3 of the shunt compensation valves 35 to 40.
5b to 4°b, and a storage unit 57 that stores the functional relationship with the control force applied to each of the valves 5b to 4°b.
A calculation unit 56 that calculates the control force given to each of the drive units 35b to 40b for each branch compensation valve, and a control force signal a corresponding to the control force calculated by the calculation unit 56.
~f to each drive unit 35b of the shunt compensation valves 35 to 40
- 40b. The storage unit 57 stores the functional relationship shown in FIG. 3 as a functional relationship corresponding to the boom cylinder 26 among the functional relationships related to the actuator, that is, the control force F1 increases as the load sensing differential pressure ΔPL5 decreases.
Although the first functional relationship that outputs ΔPLS and the functional relationship shown in FIG. 4 as the functional relationship corresponding to the packet cylinder 28, that is, the characteristic line shown in FIG. Therefore, a second functional relationship that outputs a large control force F2 is stored.

Further, the above-mentioned control pressure generating means 60 is connected to the above-mentioned pilot pump 52 and is connected to this pilot pump 52, and is also provided corresponding to each of the branch compensating valves 35 to 40, and is connected to the above-mentioned pilot pump 52, and is provided corresponding to each of the branch compensating valves 35 to 40. The drive units 3 of the branch compensating valves 35 to 40 operate in accordance with the output control force signals a to f, and change the pilot pressure Pso of the pilot pump 52 to a control pressure of an appropriate magnitude.
5b to 40b are included.

Further, in this embodiment, depending on the magnitude of the flow rate discharged from the main hydraulic pump 22, the pilot pump 52 generates a
The pilot pressure variable means 61 is provided to vary the magnitude of the pilot pressure Pso supplied to each of the pilot pressures Pso.

The pipe pressure variable means 61 defines, for example, the discharge pressure of the pilot pump 52, and includes a relief valve 61b that operates according to the differential pressure between the pressure on the inlet side and the pressure on the outlet side. It consists of a resistance element, such as a throttle valve 61C, provided between the outlet side and the tank.

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

For example, assuming that high-speed combined operation of the boom and packet is intended in a room temperature environment, the main hydraulic pump 22 and pilot pump 52 are driven by the operation of the engine 21, and the boom directional control valve 32 and the packet The direction control valve 34 is now in a state where it can be operated. The load sensing differential pressure ΔPL5 in this state is detected by the differential pressure detection device 53 and input to the input section 55 of the controller 59. In response to this, as shown in step S1 in FIG. 6, the load sensing differential pressure ΔPLs is read into the calculation unit 56 of the controller 59. Next, the process moves to step S2, where the first functional relationship shown in FIG. 3 and the second functional relationship shown in FIG. A control force F corresponding to the load sensing differential pressure ΔPLs is used as a control force for controlling the shunt compensation valve 38.
1 is obtained, and the load sensing differential pressure ΔP is used as the control force for controlling the branch flow compensation valve 40 related to the packet cylinder 28.
A control force F2 corresponding to Ls is determined. Next, the process moves to step S3 in FIG. 6, where a control force signal d corresponding to the control force F1 obtained in step S2 and a control force signal f corresponding to the control force F2 are sent from the output section 58 of the controller 59 to the solenoid valve 62d.
, 62f. This drives the solenoid valves 62d and 62f, and the pilot pump 52
The discharge pressure, that is, the pilot pressure Pso, is changed as appropriate and is used as the control pressure, Fc1, Fc2, to control the flow compensation valve 38.4.
0 to the drive units 38b and 40b. On the other hand, the pilot pressure PsO generated by the pilot pump 52 is applied to each of the other driving parts 38a and 40a of the branch compensation valve 38.40.

At this time, for example, the relationship between the control forces acting on the drive parts 38a and 38b of the branch compensation valve 38 on the boom cylinder 26 side will be explained with reference to FIG. 2.Similar to the above-mentioned equation (1), P, 4-aL1+Ps o -a s 1=Pzl
-aZl +Fcl -am+ (5) Here, since all=asl=az1:aml, the differential pressure across the boom directional control valve 32 PzI pL
l is Pzl -PL1=P! 3O-PCI (
6). Further, considering the differential pressure Pz2-PL2 before and after the packet directional control valve 34 in the same way, Pz2-PL2=PS0-Fe2 (7). Here, Pso-Fcl in equation (6) is obtained from the first functional relationship and can be expressed as a function of load sensing differential pressure ΔPLs, and similarly Pso-Fcl in equation (7)
c2 is obtained from the second functional relationship and can be expressed as a function of the load sensing differential pressure ΔPL5, α,
If β is a proportionality constant, then Pzl-PLI-α・ΔP LS (8) Pz
2 Pt2” β ・ Δ PLS
(9) holds true.

Then, the flow rate passing through the boom directional control valve 32 is determined by Ql.
, the opening area is SA, and the proportionality constant is 1, then there is a relationship between Q, = SA-, and t, t (10).Similarly, the flow rate passing through the packet directional control valve 34 is If Q2 is the opening area, SB is the opening area, and the proportionality constant is 2, then Q2=SB-27 lower 77 PL2 (II)
There is a relationship between

These equations 10) and (11) are expressed as (8) and (
9) Substituting the formula, Q, = SA-K, V7A P ts (12)
Q2 = SB- 2f, P LS (1
3) From these equations (12) and (13), the flow division ratio Ql/Q2 of the boom cylinder 26 and packet cylinder 28 is calculated as Q
l /Q2 =SA, ni1 si'error-/SB
・ Kl”l Takumi “Nari, SA, SB, Kl, Kl
, α, and β are constants of proportionality, so Q 1 / Q
2 remains constant.

That is, even in this embodiment, when the boom and packet are driven at high speed at room temperature, a stable flow rate can be supplied to the boom cylinder 26 and the packet cylinder 28 without being affected by load fluctuations of other actuators, and each Boom directional control valve 32, packet directional control valve 3
Good composite drive can be realized at a speed corresponding to the operation amount of No. 4.

A1 and B1 in FIG. 8 are characteristic lines showing the relationship between the flow rate that passes through the boom directional control valve 32 and the packet directional control valve 34 obtained during such high-speed driving, and their opening area, that is, the lever stroke. These characteristic lines A1 and B1 are equivalent to those shown in FIG. 14 described above.

It should be noted that when the above-mentioned boom and packet are driven at high speed, the flow rate discharged from the main hydraulic pump 22 is large, and therefore the flow rate of the pilot pump 52, which is driven in synchronization with the main hydraulic pump 22, is large. On the outlet side, the flow of the pilot flow rate is regulated by the throttle valve 61C and the pressure is increased, so that the pilot pressure variable means 6
The pressure difference between the pressure on the inlet side and the pressure on the outlet side of the relief valve 61b constituting the relief valve 61b becomes small, and the force of the spring causes the relief valve 6
1b is almost closed, and thereby a relatively large pilot pressure Pso is defined, and this large pilot pressure Ps
o is given to the drive parts 35a-40a of the branch compensation valves 35-40. Pilot pressure P of equations (6) and (7) mentioned above
so is the pilot pressure defined in this way.

When the rotational speed of the engine 21 is lowered in order to shift from high-speed combined boom/packet operation to low-speed combined operation in such a normal temperature environment, the discharge flow rate of the main hydraulic pump 22 becomes smaller, and Accordingly, the discharge flow rate of the pilot pump 52 becomes smaller. As a result, the pressure on the outlet side of the relief valve 61b becomes low because the pilot flow rate easily flows through the throttle valve 61c, and therefore the pressure difference between the pressure on the inlet side and the pressure on the outlet side of the relief valve 61b increases, and the spring The relief valve 61b almost opens against the force of Ps, which is relatively small.
o, and this small pilot pressure Pso is given to the driving parts 35a to 40a of the branch compensation valves 35 to 40.

Therefore, in the case of this low-speed combined operation compared to the case of the high-speed combined operation described above, from the relationship of equations (6) and (7) above, the differential pressure Pzl - P
Ll, differential pressure across the packet directional control valve 34 Pz2-P
L2 is forcibly changed to become smaller, and along with this, the flow rate Ql passing through the boom directional control valve 32 and the packet directional control valve 3 are
The flow rate Q2 passing through 4 is also changed to a small value. As a result, pressure oil gradually passes through the boom directional control valve 32 and the packet directional control valve 34.

In addition, FIG. 5 shows the pilot flow rate Qp and pilot pressure Ps that change as described above by the pilot pressure variable means.
It is a characteristic diagram which shows the relationship with o.

A2 and B2 in FIG. 8 are characteristic lines showing the relationship between the flow rate passing through the boom directional control valve 32 and the packet directional control valve 34 obtained during low-speed driving, respectively, and their opening area, that is, the lever stroke. . As shown in FIG. 8, in this embodiment, pressure oil passes through the boom directional control valve 32 and packet directional control valve 34 relatively quickly at high speeds, and gradually passes through the boom directional control valves 34 at low speeds. The opening area, ie, the lever stroke, at which the maximum flow rate is obtained in each case can be made to be approximately the same lever stroke SM.

Furthermore, unlike the above-mentioned normal temperature environment, in a low temperature environment such as a severe winter, the viscosity of the hydraulic oil flowing in the circuit increases; When the flow rate Qp of the pilot pump 52 is increased, the flow rate Qp of the pilot pump 52 is also increased, and in this case, as described above, the pilot pressure Pso becomes high, so that the diversion compensation valve 38 etc. can be driven almost without being affected by the hydraulic oil temperature. A desired large flow rate, that is, a flow rate equivalent to that at room temperature, can be supplied to actuators such as the boom cylinder 26. After operation, even if the temperature of the hydraulic oil increases due to the circulation of the hydraulic oil and reaches a constant temperature, the pilot pressure Pso is controlled by the flow rate of the main hydraulic pump 22, so the boom directional control valve 32, etc. When the amount of operation of As shown,
A constant actuator speed is obtained regardless of changes in hydraulic fluid temperature.

In the embodiment configured in this manner, as described above, when the flow rate of the main hydraulic pump 22 is large, the pilot pressure Pso is increased to force the branch compensating valves 35 to 40 to open as close as possible, that is, at room temperature. Since the opening amount is controlled to be the same as that at the time of operation, the actuator can be driven at a constant speed without being affected by the high or low temperature of the hydraulic oil, even in low-temperature environments in the depths of winter, and the actuator drive is stable. Provides ease of use.

In addition, as is clear from Figure 8, which shows the characteristics during high-speed drive and low-speed drive, the maximum flow rate of each directional control valve can be obtained with the same lever stroke regardless of whether it is high-speed drive or low-speed drive. It is possible to ensure accurate metering, and even in this case, excellent operability of the actuator drive can be obtained.

In particular, as shown by characteristic lines A2 and B2 in Fig. 8, the lever stroke SM that reaches the maximum flow rate is large even during low-speed operation, so the range in which the flow rate is changed is large, and therefore the fine operation that drives the actuator extremely slowly is easy.

Furthermore, for example, when a boom/packet is operated in combination, the characteristic line B changes at the same time as the characteristic line Al in FIG. 8 changes to A2.
1 changes to 82, and therefore, even if the required flow rate of the packet directional control valve 34 is smaller than the maximum flow rate of the boom directional control valve 32 when the characteristic line is changed to the characteristic line A2, the low speed compound drive Directional control valve 34 for packets
The maximum flow rate of the boom directional control valve 32 can also be reduced in accordance with the maximum flow rate of the boom directional control valve 32, so that the operator does not feel uncomfortable and excellent operability can be obtained even in this complex operation.

Note that although the boom/packet combined operation has been described above for ease of explanation, similar operations can be performed in other types of combinations of combined operations.

FIGS. 9 and 10 are explanatory diagrams showing other examples of flow rate control means for controlling the flow rate of the main hydraulic pump 22, respectively.

The flow rate control means shown in FIG. 9 includes a pilot hydraulic pressure source 70, a control actuator 71 that controls the displacement of the main hydraulic pump 22, and a rod side chamber 71 of the control actuator 71 that is connected to the pilot hydraulic pressure source 70.
an electromagnetic switching valve 7 connected to both a and the head side chamber 71b;
2, and an electromagnetic switching valve 73 connected to the electromagnetic switching valve 72 and interposed between the head side chamber 71b of the control actuator 71 and the tank. The above electromagnetic switching valves 72 and 73 are connected to the output section 58 of the controller 59.
The controllers are driven by control signals outputted from the controllers, and the memory sections 57 of the five controllers store a set differential pressure Px, which is a desirable load sensing differential pressure on the circuit.

Then, the load sensing differential pressure ΔPLS detected by the differential pressure detection device 53 and the set differential pressure ΔPx are determined by the controller 59.
In this case, when the load sensing differential pressure ΔPLS is larger than the set differential pressure ΔPX, a signal is output from the controller 59 to the drive unit of the electromagnetic switching valve 73. is switched to the lower position, and the head side chamber 71b of the control actuator 71 and the tank communicate with each other. Then, the pilot pressure of the pilot oil pressure source 70 is supplied to the rod side chamber 71a of the control actuator 71, the pressure of the head side chamber 71b escapes to the tank, and the piston of the control actuator 71 moves to the right in FIG. The displacement volume changes so that the flow rate discharged from the hydraulic pump 22 decreases, and the differential pressure ΔPLS is controlled so as to approach the set differential pressure ΔPx. In addition, the load sensing differential pressure Δpt detected by the differential pressure detection device 53
When s is smaller than the set differential pressure ΔPx, a signal is output from the controller 59 to the drive section of the electromagnetic switching valve 72, and the electromagnetic switching valve 72 is switched to the lower position. As a result, the pilot pressure of the pilot oil pressure source 70 is applied to the rod side chamber 71a of the control actuator 71 and the head side chamber 71.
Due to the difference in pressure receiving area between the head side and the rod side of the control actuator 71, the piston of the control actuator 71 moves to the left in the figure, and the flow rate discharged from the main hydraulic pump 22 increases. The displacement volume changes as follows, and the differential pressure ΔPLS is controlled so as to approach the set differential pressure ΔPx. In place of the flow rate control means shown in FIG. 1, a flow rate control means shown in FIG. 9 may be provided.

The flow rate control means shown in FIG. 10 is equipped with a fixed capacity expansion hydraulic pump 75, and a flow rate adjustment valve 76 is provided between the discharge pipe and the tank.
is operated by the differential pressure between the pump pressure Ps and the maximum load pressure P amax.

In place of the main hydraulic pump 22 and flow rate control means shown in FIG. 1 described above, a configuration may be provided in which a flow rate control means consisting of a hydraulic pump and a flow rate adjustment valve 76 shown in FIG. 10 is provided.

FIG. 11 is an explanatory diagram showing the main part of another embodiment of the present invention. In this embodiment, the main part of which is shown in FIG. a relief valve 82 that regulates the discharge pressure of the pilot pump 52 according to the pressure of the main hydraulic pump 22 and the spring force of the spring 81; The oil temperature sensor 84 detects the oil temperature, and the preset force variable means 85 changes the preset force of the spring 81 of the relief valve 82 in accordance with the signals output from the flow meter 83 and the oil temperature sensor 84. . The other configurations are the same as those shown in FIG. 1 described above, for example.

In the embodiment configured as shown in FIG. 11, when the rotation speed of the prime mover is high and the discharge flow rate of the main hydraulic pump 22 is large, the flow meter 83 detects that a large flow rate is flowing, and the flow rate signal is set as a preset signal. The force is applied to the force variable means 85. In response, the preset force variable means 85 operates to increase the preset force of the spring 81.
Along with this, the relief valve 82 is operated to be closed. As a result, the pilot pressure P of the pilot pump 52
so is defined as a relatively high pressure. Conversely, when the rotation speed of the prime mover is low and the discharge flow rate of the main hydraulic pump 22 is small, the flow meter 83 detects that a small flow rate is flowing, and the flow rate signal is transmitted to the preset force variable means 8.
5, the preset force variable means 85 operates to reduce the preset force of the spring 81, and accordingly the relief valve 82 operates to open. As a result, the pilot pressure PSO of the pilot pump 52 is set to a relatively low pressure.

Further, when the operating oil temperature is low in a low temperature environment, the oil temperature sensor 84 detects the low temperature and outputs a temperature signal to the preset force variable means 85. The preset force variable means 85 operates to increase the preset force of the spring 81 in accordance with the degree of lowness of the hydraulic oil temperature detected by the oil temperature sensor 84, and accordingly closes the relief valve 82. It operates. As a result, the pilot pump 52
The pilot pressure Pso controls the flow compensating valve so that the same flow rate as at room temperature flows to the actuator.

Even in another embodiment configured in this way, a constant flow rate can be supplied to the actuator regardless of changes in the hydraulic oil temperature, a constant actuator speed unaffected by the hydraulic oil temperature can be obtained, and the main oil pressure When the flow rate discharged from the pump 22 is large, a large flow rate passes through the drive section of the division compensation valve, and conversely, when the flow rate discharged from the main hydraulic pump 22 is small, a small pilot pressure Pso is applied to the drive section of the division compensation valve. This allows a small flow rate to pass through the flow control valve, thereby ensuring good metering of the flow control valve, providing the same effect as the embodiment shown in FIG. 1 described above.

<Effects of the Invention> Since the hydraulic drive device for civil engineering and construction machinery of the present invention is configured as described above, the actuator speed can be driven at a constant speed without being affected by the temperature of the hydraulic oil. Therefore, excellent operability of the actuator drive can be obtained even in a low temperature environment. In addition, it is possible to ensure good metering of the flow control valve that drives the actuator regardless of changes in the rotation speed of the prime mover.
Excellent operability is obtained without giving the operator a sense of discomfort unlike in the past, and fine control is easy, especially since the range in which the flow rate passing through the flow rate control valve can be changed is large even when the prime mover is running at low rotation speeds. Furthermore, during combined operations, the metering of the flow control valve associated with each actuator can be changed regardless of the required flow rate of the flow control valve, and in this respect, there is no discomfort to the operator unlike in the past. , excellent operability can be obtained.

[Brief explanation of drawings]

Fig. 1 is a circuit diagram showing an embodiment of the hydraulic drive system for civil engineering and construction machinery of the present invention, and Fig. 2 explains the balance of forces acting on the shunt compensation valve provided in the hydraulic drive system shown in Fig. 1. Figures 3 and 4 are diagrams showing the functional relationships stored in the controller of the embodiment shown in Figure 1, respectively.
The figure shows the characteristics obtained by the pilot pressure variable means provided in the hydraulic drive device shown in Figure 1, and Figure 6 shows the characteristics obtained by the pilot pressure variable means provided in the hydraulic drive device shown in Figure
FIG. 7 is a flowchart showing the procedure of processing carried out by the controller provided in the hydraulic drive device shown in the figure; FIG. Figure 8 shows the hydraulic drive system shown in Figure 1! Characteristic diagrams showing the relationship between lever stroke and directional control valve flow rate obtained in Figures 9 and 10.
11 is an explanatory diagram showing the main parts of another embodiment of the present invention, and FIG. 12 is a circuit showing a conventional hydraulic drive device for civil engineering and construction machinery. Figure 13 is a characteristic diagram showing the relationship between actuator speed and changes in hydraulic oil temperature in the hydraulic drive system shown in Figure 12, and Figure 14 is a characteristic diagram showing the relationship between lever stroke and directional control valve flow rate in the hydraulic drive system shown in Figure 12. FIG. 21...Engine (prime mover), 22...
・Main hydraulic pump, 23...Swivel motor, 24.
...Left travel motor, 25...Right travel motor, 26...Boom cylinder, 27...
・Arm cylinder, 28...Packet cylinder, 29...Swivel directional control valve, 30...
・Directional control valve for left travel, 31... Directional control valve for right travel, 32... Directional control valve for boom, 33.
Directional control valve for arm, 34... Directional control valve for baguette, 35.36.37.38.39.4o.
- Diversion compensation valve, 41... Control actuator, 42... Flow rate adjustment valve, 43.44...
... Pipeline, 51a, 51b, 51c, 51d, 51
e, 51 f..., pipe line, 52... pilot pump (pilot hydraulic power source), 53...
Differential pressure detection device, 55... Input section, 56...
...Arithmetic section, 57...Storage section, 58.Old...Output section, 5...Controller, 60...
- Control pressure generation means, 61...Pilot pressure variable means, 61B...Relief valve, 61c...
... Throttle valve, 62a, 62b, 62c, 62d, 62
e, 62 f--Solenoid valve, 70... Pilot oil pressure source, 71... Control actuator, 7
1a...Rod side chamber, 71b...Head side chamber, 72.73...Solenoid switching valve, 75...
...Main hydraulic pump, 76...Flow rate adjustment valve. 3 ``1 Mountain 2 Figure 5 ノ\70l-tf-o P Figure 4 Figure 6 Figure 7 4'\ Hydraulic oil Figure 8 V/G゛-nut c-7:; ”' / , Figure 9, Figure 1O, Figure 12FA

Claims (4)

[Claims] (1) A prime mover, a main hydraulic pump driven by the prime mover, a plurality of actuators driven by pressure oil supplied from the main hydraulic pump, and the flow of pressure oil supplied to these actuators. a directional control valve that controls the flow rate control valve, a branch flow compensation valve that controls the differential pressure across these flow rate control valves, and a flow rate discharged from the main hydraulic pump that is the maximum of the pressure of the main hydraulic pump and the load pressure of the actuator. and a flow rate control means for controlling according to the differential pressure with respect to the load pressure, and supplies pressure oil from the main hydraulic pump to each of the actuators via the branch compensation valve and the flow rate control valve, and Civil engineering/
In a hydraulic drive system for construction machinery, a pilot hydraulic source that is connected to the driving parts of the respective shunt compensating valves and applies pilot pressure to the driving parts of the shunt compensating valves to operate the shunt compensating valves in a direction to open them; 1. A hydraulic drive system for civil engineering and construction machinery, characterized in that a pilot pressure variable means is provided for varying the magnitude of the pilot pressure in accordance with the magnitude of the flow rate discharged from the main hydraulic pump. (2) A relief valve in which the pilot pressure variable means regulates the discharge pressure of the pilot pump and operates according to the differential pressure between the inlet side pressure and the outlet side pressure, and the outlet side of this relief valve and the tank. 2. The hydraulic drive system for civil engineering and construction machinery according to claim 1, further comprising a resistance element provided in the hydraulic drive system for civil engineering and construction machinery. (3) The hydraulic drive system for civil engineering and construction machinery according to claim (2), wherein the resistance element is a throttle valve. (4) The pilot pressure variable means detects a relief valve that regulates the discharge pressure of the pilot pump according to the spring force, a flow meter that detects the flow rate discharged from the main hydraulic pump, and the temperature of the hydraulic oil flowing through the circuit. Claim (1) characterized in that it includes an oil temperature sensor and a preset force variable means for changing the preset force of the spring of the relief valve in accordance with the signals output from these flowmeters and the oil temperature sensor. Hydraulic drive system for civil engineering and construction machinery.