JPH0341201A - Hydraulic control valve system - Google Patents

Abstract

PURPOSE:To reduce pump driving power by constituting the pump of a power steering in a variable capacity pump with a negative flow control valve regulator and controlling it with a center by-pass type direction changeover valve downstream side throttle device pressure. CONSTITUTION:A hydraulic control valve system of a power steering drives a bucket cylinder 26 and a boom cylinder 28 through a multiple valve 16 with excess oil from a steering valve 14. Here, a pump 70 driven with an engine 10 is constituted in a variable capacity pump with a negative flow control valve regulator 72 and controls pump discharge flow with control pressure from a throttle device 74 provided on the downstream side of center by-pass type direction changeover valves 30, 32 of the multiple valve 16. The pump discharge flow is thus restricted to the necessary minimum, and it is possible to attempt to reduce pump driving power.

Description

DETAILED DESCRIPTION OF THE INVENTION [Field of Industrial Application] The present invention relates to a hydraulic control valve device for a construction machine such as a shovel loader having a power steering device.

[Conventional technology]

As shown in FIG. 6, a conventional hydraulic control valve device of this type includes a hydraulic pump 12 driven by an engine 10, a main switching valve 22 that drives and controls a steering cylinder 20 using the oil discharged from the pump 12, and a pressure control valve device. A steering valve 14 having a compensation valve 24 and a bucket cylinder 2 to which surplus oil of the steering valve 14 is supplied via a connecting pipe 18.
6. Two center bypass type directional valves 30°3 that drive two working cylinders such as the boom cylinder 28
2, and the pump 12 is of a fixed displacement type.

Next, the operation of the hydraulic control valve device having such a configuration will be explained. First, the figure shows that the engine 10 has stopped and the pump 12 has stopped.
indicates a state in which oil is not being discharged, but when the engine 10 is driven in this state, the discharged oil from the pump 10 flows from the oil inlet 34 of the steering valve 14 to the inlet oil passage 36 and the pressure compensation valve 24. through which it flows into the oil passage 38,
Since this oil passage 38 is blocked by the main switching valve 22, the pressure in this oil passage 38 increases. Here, the pressure compensation valve 24 has its right end connected via the oil passage 38 and the pilot oil passage 40, and its left end connected via the pilot oil passage 42, the main switching valve 22, the return oil passage 44, and the oil outlet 46. Since the pressure in the oil line 38 reaches the predetermined control pressure of the pressure compensation valve 24 (the value obtained by dividing the force of the spring 50 by the spool cross-sectional area of the pressure compensation valve 24), the pressure The compensation valve 24 is actuated and moved to the position 24b, and the oil flowing into the oil passage 38 flows out to the HPCO boat 54 through the variable orifice 52b of the pressure compensation valve 24. Note that the pressure compensation valve 24 is moved to the position 24a. Kara position ff
24b, variable orifice 52a decreases in opening area, while variable orifice 52b increases in opening area. The oil that spilled into the HPCO boat 54 was
enters the oil inlet 56 of the multiple valve 16 through the communication pipe 18;
Tank 48 via center bypass 58, oil outlet 60
Return to

Next, when the main switching valve 22 is operated to the position 22a to drive the steering cylinder 20, the pilot oil passage 42 is cut off from the return oil passage 44 and the variable orifice 62
It is connected to the downstream side of a. In this state, oil does not initially flow into the variable orifice 62a, but if oil does not flow in this way, the oil passage 3 on the upstream side of the variable orifice 62a
The pressures in the pilot oil passage 40 connected to the pilot oil passage 8 and the pilot oil passage 42 become equal, and the pressure compensating valve 24 is moved to this position by the force of the spring 50. eZ24b toward position 24a, thereby constricting variable orifice 52b and opening variable orifice 52a. Therefore,
The pressure in the inlet oil passage 36 and oil passage 38 increases, oil flows into the variable orifice 62a, and this oil flows into the cylinder bow)2.
0a to the steering cylinder 20;
The oil from the steering cylinder 20 is returned from the cylinder boat 20b to the tank 4 via a return oil path 44 and an oil outlet 46.
Return to 8. That is, the steering cylinder 20 is driven. In this case, when oil flows through the variable orifice 62a and a pressure difference is generated in the variable orifice 62a, the pressure compensation valve 24, which is controlled by this pressure difference, determines whether the pressure difference is equal to the predetermined pressure compensation valve control pressure. In the position where
Due to the force of the spring 50 and the pressure difference, the pressure compensating valve 2
The force generated on the spool 4 is balanced and stopped.

In this way, the pressure difference generated in the variable orifice 62a is kept constant (predetermined pressure compensation valve control pressure), so the flow rate that passes through the variable orifice 62a and is supplied to the steering cylinder 20 is 62a
(or 62b) is proportional to the opening area.

Note that the discharge flow rate Q of the pump 10 flowing into the oil inlet 34
Flow rate Qs supplied from P to the steering cylinder 20
The remaining surplus flow i (Q, −Qs) after subtracting H
Since the oil is supplied from the PCO boat 54 to the oil inlet 56 of the multiple valve 16 via the communication pipe Fl@18, by operating the center bypass type directional control valve 30 or 32, the excess oil can be used to control the packet cylinder 26 or Boom cylinder 28 can be driven.

[Problem to be solved by the invention]

However, this type of conventional hydraulic control valve device has had several problems as described below.

Generally, in this type of hydraulic control valve device, the speed of the steering cylinder 20 must be constant regardless of the rotational speed of the engine 10, so the steering cylinder 20 has the variable orifice 62a or 6 of the main switching valve 22.
2b must be supplied with a flow rate proportional to the opening area of the pump 12. For this reason, the pump 12 has a flow rate per revolution so that it can discharge a sufficient flow rate even when the engine 10 is running at a low idle speed. It is set.

Therefore, as the rotational speed of the engine 10 increases, an extra flow rate will be discharged from the pump 12.

However, the extra flow rate, that is, the surplus flow rate, flows through the oil outlet 6 of the multiple valve 16 without doing any work, except when driving the working cylinders 26, 28 in the multiple valve 16.
0 and is returned to the tank 48. For this reason, (pump 12
(discharge pressure) x (excess flow rate) energy is lost, resulting in a worsening of the fuel consumption rate of the engine 10.

SUMMARY OF THE INVENTION Therefore, an object of the present invention is to provide a hydraulic control valve device that can reduce the fuel consumption rate of an engine by appropriately controlling the pump discharge flow rate and the amount of oil supplied to the steering valve and the multiple valve. Our goal is to provide the following.

[Means to solve the problem]

In order to achieve the above object, a hydraulic control valve device according to the present invention includes: a hydraulic pump; a steering valve having a main switching valve and a pressure compensation valve that control driving of a steering cylinder using oil discharged from the hydraulic pump; A hydraulic control valve device includes a multiple valve having a center bypass type directional control valve to which surplus oil from the steering valve is supplied via a communication pipe to control the drive of the work system cylinder, and the hydraulic control valve device has a negative flow rate to the hydraulic pump. A control regulator is provided to configure the pump as a variable displacement pump, and a throttle device is provided on the center bypass downstream of the center bypass type directional valve, and the pressure in the center bypass between the throttle device and the center bypass type directional valve is controlled. The invention is characterized in that a pilot line is arranged to transmit the pressure as a pump control pressure to the negative flow rate control regulator.

In this case, as an alternative, the pump may consist of first and second variable displacement pumps with negative flow control regulators, with the first variable displacement pump connected to the steering valve and the second variable displacement pump connected to the steering valve. The negative flow rate control regulators of the first and second variable displacement pumps, which are connected to the communication pipe connecting between the valves and to which the pump control pressure is applied, are connected to the negative flow rate control regulators of the first and second variable displacement pumps, which are connected to the communication pipe line connecting between the communication valves, and are connected to the negative flow rate control regulators of the first and second variable displacement pumps, which are connected to the communication pipe line connecting between the communication valves and to which the pump control pressure is applied. The discharge amount of the first variable displacement pump may be increased in advance, and then the discharge amount of the variable displacement pump of the cylinder 2 may be increased.

[Effect]

In the hydraulic control valve device of the present invention, the discharge oil fX of the variable displacement pump is controlled by the pump control pressure generated by the throttle device provided downstream of the center bypass type directional control valve, and this pump control pressure Since the discharge flow rate can be suppressed, the amount of oil that passes through this throttle device, that is, the amount of oil that passes through the center bypass and is returned to the tank without doing any work, is reduced, reducing the engine's fuel consumption rate. be able to.

In this case, by configuring two pumps, the amount of oil supplied to the steering valve can also be minimized.

In other words, a portion of the amount of oil required by the multiple valves is supplied directly to the multiple valves without passing through the steering valve, reducing pressure loss when passing through the steering valves and improving engine performance. Fuel consumption rate can be further reduced.

[Actual example]

Next, an embodiment of a hydraulic control valve device according to the present invention will be described in detail below with reference to the accompanying drawings. For convenience of explanation, the same reference numerals are given to the same components as in the structure of the conventional hydraulic control valve device shown in FIG. 6, and detailed explanation will be omitted.

The basic configuration of the hydraulic control valve device according to the present invention is the same as that of the conventional hydraulic control valve device shown in FIG. That is, in FIG. 1, the hydraulic control valve device of the present invention basically includes a pump 70 driven by an engine 10'''C', a main switching valve 22 that drives a steering cylinder 20, and a pressure compensation valve 24. a steering valve 14 having;
Two center bypass directional valves 30.32 driving the packet cylinder 26 and the boom cylinder 28
and a communication pipe line 18 connecting between the steering valve 14 and the multiple valve 16.

However, in the hydraulic control valve device of the present invention, the pump 70 described above is configured as a variable displacement pump having a negative flow rate control regulator 72, and the center downstream of the center bypass type directional control valve 32 of the multiple valve 16 is configured as a variable displacement pump having a negative flow rate control regulator 72. A throttle device 74 having a fixed orifice 74a is provided in the bypass 58, and the pressure in the center bypass 58 between the throttle device 72 and the directional control valve 32 is used as pump control pressure to control the pilot boat 76 and the pilot line 78. It is connected to the negative flow rate control regulator 72 of the variable displacement pump 70 via. Note that the throttle device 74 is provided with a relief valve 74b.

Next, the operation of the hydraulic control valve device of the present invention will be explained, but before that, the characteristics of the variable displacement pump 70 controlled via the throttle device 74 will be briefly explained with reference to FIG.

In Figure 2, the horizontal axis represents the pump control pressure P;
The vertical axis indicates the flow iQ0 passing through the fixed orifice 74a and the discharge flow JiQp of the pump 70, and in the graph, folds a (a), (b), and (c) represent the pump control pressure P relative to the pump control pressure P, respectively. Curve (d) shows the relationship between the flow rate Qp of each pump at rated rotation, low idle rotation, and intermediate rotation. Curve (d) shows the relationship between the fixed orifice passing flow iQ and the pump control pressure P. In this case, the discharge flow rate q per rotation of the pump 70 decreases as the pump control pressure P increases, but does not change when the pressure P is constant. ) Each discharge flow rate Qp of the pump 70 is determined in proportion to the rotation speed of the pump 70. In addition, each broken line (a), (b), (
The intersection points (A) and (B) of C) and song ff1 (D), respectively.
) and (C) are states in which the fixed orifice passing flow rate Q0 and the pump discharge flow rate QP are the same at each rotation of the pump 70, that is, the steering valve 14 and the multiple valve 16 are not operated and the pump discharge flow rate QP shows the operating point of the pump 70 in a state where the entire amount of is passing through the fixed orifice 74a, and in this state, the pump control pressures P are pA, p, respectively.

PC becomes the pump discharge flow rate Qp and the fixed orifice passing flow rate Qo is the flow rate Q PA” Q OAI
It has been shown that Q pI, = Q OIl+ Q pc = Q oc. In other words, for example, the intersection (B
), when the pump control pressure P and pump discharge flow rate QP become pressure PB and flow rate Q p @, respectively, the negative flow rate control regulator 72 balances the pump control pressure P and pump discharge flow rate Q. It shows that you can keep it. Note that the pump 70 rotates at each rotation, that is, as shown by the broken line (A).

In (b) and (c), the intersection (A).

The range indicated by (B), (C) and points (D), (E), and (F), that is, the pump discharge flow rate Q PA to Q PD
, QpH~Qpe + Qp. ~Qpp, and the flow rate QPC, that is, the maximum pump flow rate during low idle rotation, is set to a flow rate sufficient to drive the steering cylinder 20 at a predetermined speed.

The hydraulic control valve device of the present invention having such a configuration is operated as follows.

In FIG. 1, when the engine 10 is stopped,
That is, when the variable displacement pump 70 is stopped, no oil flows into the fixed orifice 74a of the throttle device 74, so the pump control pressure P is zero. Next, when the engine 10 is rotated and pressure oil is discharged from the variable displacement pump 70, the pressure compensating valve 24 of the steering valve 14 is moved to the position 24b, as described in the prior art shown in FIG. Therefore, the discharged oil is
The multiple valve 1 is passed from the HPCO boat 54 through the communication pipe 18.
6 to the oil inlet 56, and then to the center bypass 58.
, the fixed orifice 74a, the return oil passage 64, and the oil outlet 60, and are returned to the tank 48. In this process, when the pump discharge flow rate QP increases, the pump control pressure P generated in the fixed orifice 74a and applied to the negative flow rate control regulator 72 of the variable displacement pump 70 via the pilot boat 76 and the pilot line 78 increases to the second level.
It rises along the curve (d) shown in the figure.

As a result, the variable displacement pump 70 operates at the intersection (B) corresponding to low idle rotation, intermediate rotation, and rated rotation.
), (C), and (A). That is, when both the steering valve 14 and the multiple valve 16 are not operated, the pump control pressure P and the pump discharge flow rate QP follow the curve (d) according to the pump rotation speed.
Maintain equilibrium at a certain point in (B) to (A) above.

Next, when the main switching valve 22 is operated to position g22a to operate the steering valve 14 in the above-mentioned state, the sixth
As explained in the prior art shown in the figure, the variable orifice 52b of the pressure compensating valve 24 is throttled and the variable orifice 5
2a opens and this opens the variable orifice 6 of the main switching valve 22.
A steering flow fitQ proportional to the opening area of the fixed orifice 74a is supplied to the steering cylinder 20, and return oil from the steering cylinder flows out to the tank 48 via the return oil passage 44 and the oil outlet 46.
The flow rate Q0 passing through is reduced by the flow rate Qs. Therefore, the pump control pressure P decreases, and as a result,
Negative flow control regulator 72 increases the discharge flow iQ2 of variable displacement pump 70.

Here, in order to explain the pump operation in the above state, FIG. 3 shows only the pump discharge flow rate relationship curve (c) and the fixed orifice relationship curve (2) taken out from FIG. 2 during intermediate rotation. However, in FIG. 3, when the steering cylinder 20 is not operated, the variable displacement pump 70 is operated at the pump control pressure PC.
, flow rate Qpc=Qoc. However, when the steering valve 14 is operated and the steering cylinder 20 is supplied with the steering flow rate Qs, the fixed orifice passing flow rate Q. C is the flow rate (Q, c
-Q, ), the pump control pressure Pc decreases, and as a result, the pump discharge flow iQ increases by the negative flow rate control regulator 72, and this pump discharge flow becomes Q pc '' Q s + Q oc・
It is shown that the pump discharge flow rate moves on the broken line (c) until point C'' where it becomes.At this time, the pump control pressure P becomes the pressure PC·, and the fixed orifice passing flow iQ becomes the flow rate Q. C. becomes.

Next, in the above-mentioned state, when the center bypass switching valve 30 is operated to throttle the center bypass in order to further operate the packet cylinder 26, for example, the pressure oil from the oil inlet 56 flows into the packet cylinder 26, and The return oil is returned to the tank 48 via the return oil path 64 and the oil outlet 6o. Therefore, the flow rate Q passing through the fixed orifice is the flow rate Q in the above state. Flow i (Q, c, −Q,
) is reduced to As a result, as in the case described above, the variable displacement pump 70 has a pump discharge flow rate Q, or a flow rate (Q
Point C' where s + Q-+ Qo) = Qpc・
° Move further on the broken line (C) until it reaches (see Figure 3).

In addition, at this time, the pump control pressure P is the pressure PC・
Therefore, the flow rate Q0 passing through the fixed orifice is the flow rate Q. It becomes C-.

Note that when operating only the multiple valve 16 without operating the steering valve 14, the variable displacement pump 70
It is operated in such a way that Q, - becomes the flow rate (Q, +Qo), but since this is easily understood, the explanation will be omitted.

As described above, according to the hydraulic control valve device of the present invention, unlike conventional devices, the pump discharge flow rate, which is bypassed to the tank without performing any work in the steering valve and the multiple valve, is suppressed to the necessary and minimum level. Therefore, the pump oil power is reduced, and the fuel consumption rate of the engine can be reduced.

FIG. 4 is a hydraulic circuit diagram showing a second embodiment of the hydraulic control valve device according to the present invention. It consists of two variable displacement pumps.

That is, in FIG. 4, the pumps include a first variable displacement pump 80 connected to the steering valve 14 and a second variable displacement pump 84 connected to the communication pipe 18 between the steering valve 14 and the multiple valve 16. Pump control pressure generated by the fixed orifice 74a of the throttle device 74 is applied to the first and second negative flow rate control regulators 82.86 of these variable displacement pumps 80.84 via the pilot line 78. Ru. The first and second negative flow rate control regulators 82.
86 increases the discharge amount of the first variable displacement pump 80 in advance when the pump control pressure decreases from a predetermined specified pressure, and then increases the discharge amount of the variable displacement pump 84 of the cylinder 2. Set.

Next, the operation of this embodiment will be explained, but before that, the characteristics of the first and second variable displacement pumps 80, 84 controlled via the throttle device 74 will be briefly explained with reference to FIG. Explain.

In Fig. 5, the horizontal axis represents the pump control pressure P;
The vertical axis represents the fixed orifice passing flow rate Q0, the first pump discharge flow rate QIP, and the second pump.

The discharge flow rate Q 2F and the total pump discharge flow rate QP are shown respectively, and in the graph, the broken lines (E) and (E)
, (G) are the pump control pressure P, respectively.
2nd pump discharge flow i at pump intermediate rotation with respect to
Q2. , the first pump discharge flow rate QIp, and the total pump discharge flow rate Qp. Curve (d) shows the relationship between the fixed orifice passing flow rate Qo and the pump control pressure P.

As can be seen from the graph, in the hydraulic control valve device of this embodiment, when the pump control pressure P decreases and reaches the first specified pressure PH, the first variable displacement pump 80
The discharge discharge amount Q rp is increased, and then the pressure is increased to a second
When the specified pressure P1 of the first variable displacement pump 80 is reached, the discharge flow rate Qrp of the first variable displacement pump 80 is held constant, and the discharge flow rate iQ of the second variable displacement pump 84 is maintained constant.

is configured to increase. As shown in the graph, the first pump discharge flow rate Q +p, the second pump discharge flow iQ, . . . the total pump discharge flow rate Qp are respectively Q IPH ~ Q +p! + Q2PI 'Q2PJ
It has been shown to vary between IQpH'Qpr. Incidentally, by controlling the pump as described above, that is, by increasing the pump control pressure PH~ which increases the discharge flow rate Q1p of the first variable displacement pump 8o.
P, the discharge flow rate of the second variable displacement pump 84 is Q 2
The reason for controlling F so as not to increase is to prevent the following inconvenience from occurring. That is, when the steering flow rate Qs flows into the steering cylinder, as described above, the fixed orifice passing flow rate Q0 decreases, the pump control pressure P decreases, and the pump discharge flow rate QP increases. If such control is not provided, the increase in the pump discharge flow rate Qp is achieved by the increase in both the first pump discharge flow rate iQ and the second pump discharge flow rate Q2P. However, in this case, the increase in the second pump discharge amount Q Therefore, the pump discharge flow rate cannot be increased sufficiently, and the steering cylinder cannot be driven at a predetermined speed. On the other hand, in this embodiment in which the pump is controlled as described above, the steering cylinder is reliably driven at the maximum discharge flow rate Q1p of the first variable displacement pump 80, as will be described in further detail later. , the working system cylinder has a flow rate which is the above-mentioned flow rate plus the discharge flow rate of the second variable flow rate pump 84, that is, Q p −Q rp + Q 2
It can be driven by P.

The hydraulic control device of this embodiment having such a configuration is operated as follows. That is, when the steering flow rate Q8 is supplied to the steering cylinder 20 by operating the steering valve 14, the fixed orifice passing flow rate Qo decreases, the pump control pressure P decreases, and the total pump discharge flow rate Qp increases. In this case, only the first pump discharge iQIF is increased when the pump control pressure P is between PH and P, and all of this increased flow rate is supplied to the steering valve 14.

That is, in FIG. 5, if the intersection (L) of the broken line (G) and the pump control pressure line (D) is taken as the pump operating point when both the steering valve 14 and the multiple valve 16 are not operated, When the steering flow rate Qs is supplied to the steering cylinder 20, the total pump discharge flow rate Q, becomes (Qs + Qo) until the pump control pressure P increases.

-Move on the broken line (G) within the range of P1. In this way, the steering cylinder 20 is reliably driven with a flow rate up to the maximum pump discharge flow rate Q1PI of the first variable flow rate pump 80.

Next, in the above state, when the working cylinder flow rate QW is further supplied to the packet cylinder 26 or the boom cylinder 28 by operating the multiple valve 16, the control pressure P
decreases and the total pump discharge flow rate QP is increased, but in this case, the second pump discharge flow iQ, P is increased or stopped when the pump control pressure P reaches Pl and the first pump discharge flow iQ, , is increased. After that, it starts to increase. The increase in the first pump discharge flow rate QIF is supplied to the steering valve 14, and the increase in the second pump discharge flow rate Q2P is supplied to the multiple valve 1.
6. That is, in FIG. 5, when the working cylinder flow rate Qw is further generated in the pond of the steering flow rate Qs, the above-mentioned L') turns on, and the total pump discharge flow rate QP becomes (Q
The steering cylinder 20 is driven at the flow rate Qs, and the steering cylinder 20 is driven at the flow rate Qs, and the working cylinder 26.28
is the maximum discharge flow rate Q of the first pump, and the maximum discharge flow rate Q of the second pump is the flow rate obtained by subtracting the steering flow rate Qs from IP+.
The flow rate plus 2F, i.e. (Q,,.

It is reliably driven at a flow rate of Qs) +Q2F.

In addition, when only the multiple valve 16 is operated without operating the steering valve 14, the variable displacement pumps 80 and 84 are respectively driven at the point where the total pump flow rate Qp becomes (Qll +QO). , since it is easily understood,
The explanation will be omitted.

In this way, also in the hydraulic control valve device of this embodiment,
The same effects as the hydraulic control valve device of the first embodiment shown in FIG. 1 described above are achieved, and the fuel consumption rate of the engine can be reduced compared to the conventional hydraulic control valve device. However, in this case, in this embodiment, the oil supplied to the work system cylinder by the second variable displacement pump is directly supplied to the multiple valve without passing through the steering valve, so it is different from the first embodiment. Compared to the above, the pressure loss due to the flow rate of the work system cylinder passing through the steering valve is avoided, and the engine fuel consumption rate is further reduced.

Although the present invention has been described above with reference to preferred embodiments, many design changes can be made to the present invention without departing from its spirit.

〔Effect of the invention〕

As explained above, the hydraulic control valve device according to the present invention includes:
It consists of a pump driven by an engine, a steering valve having a main switching valve and a pressure compensating valve, and a multiple valve having a center bypass type directional switching valve, and the multiple valve has excess oil from the steering valve. In the hydraulic control valve device configured to be driven, the pump is configured as a variable displacement pump with a negative flow control valve regulator, and
Variable capacity 4.

Since the volume pump is configured to be controlled by the pump control pressure generated by the throttle device installed downstream of the center bypass type directional control valve, the pump is bypassed to the tank without doing any work at the steering valve or multiple valve. Unlike conventional hydraulic control valve devices, the discharge flow rate is
necessary and suppressed to a minimum.

Therefore, the pump driving force can be reduced, and the fuel consumption rate of the engine can be reduced.

[Brief explanation of drawings]

FIG. 1 is a hydraulic circuit diagram showing a first actual travel mode of the hydraulic control valve device according to the present invention, and FIG. 2 is a diagram showing the relationship between pump control pressure, pump discharge flow rate, and fixed orifice passing flow rate in the hydraulic circuit shown in FIG. FIG. 3 is a graph showing a part of FIG. 2, FIG. 4 is a hydraulic circuit diagram showing a second embodiment of the hydraulic control valve device according to the present invention, and FIG. 5 is a graph showing the relationships. Fig. 4 is a graph showing the relationship between the pump control pressure and the first pump discharge flow rate, second pump discharge flow rate, and total pump discharge flow rate in the hydraulic circuit, and Fig. 6 is a hydraulic circuit showing a conventional hydraulic control valve device. It is a diagram. 10... Engine 14... Steering valve 16... Multiple valve 18... Communication pipe 20.
... Steering cylinder 22 ... Main switching valve 24 ... Pressure compensation valve 26
... Packet cylinder 28 ... Boom cylinder 30.32 ... Center bypass type switching valve 34 ...
Oil inlet 36...Inlet oil path 38...Oil path 40.42...Pilot oil path 44...Return oil path 46...Oil outlet 48...
・Tank 50...Spring 52a, 5
2b...Variable orifice 54...HPCO port
56...Oil inlet 58...Center bypass 60...Oil outlet 62a, 62b...Variable orifice 64...Return oil path 70...Variable displacement pump with negative flow rate control regulator 72...Negative Flow rate control regulator 74...
Squeezing device 74a... fixed orifice 74b...
・Relief valve 76 ・Pilot boat 78 ・
...Pilot line 80...Variable displacement pump with first negative flow rate control regulator 82...First negative flow rate control regulator 84...
...Variable displacement pump 86 with second negative flow rate control regulator...Second negative flow rate control regulator P...
・Pump control pressure Q,... Pump discharge flow rate or total pump discharge flow rate QIP... First pump discharge flow rate Q 2F... Second pump discharge flow rate Q0... Fixed orifice passing flow rate 0 days... Steering Flow rate 0w...Working cylinder flow rate

Claims (2)

[Claims] (1) A hydraulic pump, a steering valve having a main switching valve and a pressure compensation valve that control the drive of the steering cylinder by the oil discharged from the hydraulic pump, and surplus oil from the steering valve being supplied through a communication pipe for operation. A hydraulic control valve device includes a multiple valve having a center bypass type directional switching valve for controlling the drive of system cylinders, the hydraulic pump is provided with a negative flow rate control regulator to configure as a variable displacement pump, and the center bypass type A throttle device is provided on the center bypass on the downstream side of the directional control valve, and a pilot line is configured to transmit the pressure in the center bypass between the throttle device and the center bypass type directional control valve to the negative flow rate control regulator as pump control pressure. A hydraulic control valve device comprising: (2) The first hydraulic pump is a variable capacity pump equipped with a negative flow rate control regulator that supplies pressure oil to the steering valve, and a communication pipe line that supplies excess oil from the steering valve to the center bypass type directional control valve. A second hydraulic pump consisting of a variable displacement pump provided with a negative flow rate control regulator is connected and arranged in the pilot line for transmitting the pump control pressure in the center bypass between the center bypass type directional control valve and the throttle device. A branch connection is made to a negative flow rate control regulator of the second hydraulic pump, and the negative flow rate control regulator of the first and second hydraulic pumps is connected to the negative flow rate control regulator of the first and second hydraulic pumps when the pump control pressure falls below a predetermined specified pressure. 2. The hydraulic control valve device according to claim 1, wherein the hydraulic control valve device is configured to increase the discharge amount of the hydraulic pump first, and then increase the discharge amount of the second hydraulic pump.