JP2000257601A - Hydraulic circuit device - Google Patents

Abstract

PROBLEM TO BE SOLVED: To obtain a smooth starting characteristic regardless of size of a driven inertia body, and simplify a constitution so as to be easily applicable for even a control valve of spool type, in a hydraulic circuit device provided with a load sensing system. SOLUTION: A flow divider valve 5-1 and a hold check valve 6-1 are constituted by a valve assembly 56-1, an intermediate chamber 56c between a flow divider valve element 50 and a hold check valve element 60 is connected to a signal detection oil path 9 via an oil path slit 20, lap part 32 having a check valve function, control chamber 70 in the valve element 60, and a spring chamber 72, and the signal detection oil path 9 is connected to a tank T via a throttle 14 (area at). A fixed plug 75 is extended in an axial direction into a control chamber 79 in the inside of the valve element 60, and a circular annular throttle 11 (area ac) positioned between the control chamber 70 and the spring chamber 72 is formed by an internal periphery of the valve element 60 and an external periphery of the fixed plug 75.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention is mounted on a construction machine which can simultaneously operate a plurality of hydraulic actuators, for example, a hydraulic shovel, and can obtain a smooth start-up characteristic regardless of the size of an inertial body to be driven. The present invention relates to a hydraulic circuit device.

[0002]

2. Description of the Related Art A hydraulic circuit device mounted on a construction machine such as a hydraulic shovel uses a center bypass type control valve having a bleed-off circuit or a closed center type control valve having no bleed-off circuit. The latter employs a load sensing system that controls the discharge oil amount of the hydraulic pump so that the flow rate required by the control valve can be basically supplied. When the purpose is to simplify the hydraulic equipment, the latter is advantageous because there is no bleed-off circuit. However, because it does not have a bleed-off circuit, when driving a hydraulic actuator having a large inertia, the pressure rises suddenly and suddenly accelerates, or the vibration of the pressure (pressure pulsation) does not readily attenuate. There is a problem that smooth startup characteristics cannot be obtained.

That is, in the load sensing system, since the discharge oil amount of the hydraulic pump is controlled so as to supply the flow rate required by the control valve, the load driven by the actuator is an inertial body such as a swivel, and the hydraulic pump discharges. When the actuator cannot consume the oil amount, the discharge pressure of the hydraulic pump rises sharply, and the energy discharged by the hydraulic pump is stored in the piping system. As the pressure decreases, the energy stored in the piping system is released, the actuator overshoots, the driving pressure further decreases, and then the driving pressure increases again when the actuator speed decreases. Change, a sudden rise in pressure occurs transiently, and the pressure pulsation slows down. Kunar.

[0004] Therefore, as a method of reducing the supply flow rate to the actuator with an increase in the driving pressure and suppressing a steep rise of the pressure, Japanese Patent Application Laid-Open Nos. 4-191501, 5-263804, and 10-89304 disclose a method.
A method described in Japanese Patent Application Laid-Open No. H10-264, has been proposed.

The methods described in JP-A-4-191501 and JP-A-5-263804 have the same purpose.
In a control valve that controls the displacement of the proportional seat valve having a slit by the valve opening of the pilot valve, the valve displacement of the pilot valve is controlled in accordance with the driving pressure of the actuator,
It controls the valve displacement of the proportional seat valve. That is, the pressure induced from the inlet of the hydraulic motor via the throttle is induced by the pilot valve against the operating force of the pilot valve. The pressure induced through the throttle from the inlet of the hydraulic motor is a pressure that increases in proportion to the driving pressure of the hydraulic motor, and therefore, the valve opening of the pilot valve decreases in proportion to the driving pressure. Accordingly, the valve opening of the proportional seat valve also decreases. As a result, the oil discharged from the hydraulic pump is controlled so as to be reduced, which contributes to alleviation of a sudden rise in pressure and attenuation of pressure pulsation.

[0006] In Japanese Patent Application Laid-Open No. Hei 10-89304, a pressure compensating valve provided to enable complex operation in a load sensing system is provided with a load dependency in which the compensation differential pressure is reduced as the load pressure increases. Thus, when the load pressure increases, the supply flow rate to the actuator decreases, and the discharge amount of the hydraulic pump is controlled to decrease. The load dependency of the pressure compensating valve is determined by the pressure receiving area of the pressure-incoming valve, in which the inlet pressure of the meter-in variable restrictor acts in the closing direction, and the pressure-receiving area in which the outlet pressure of the meter-in variable restrictor acts in the opening direction. It is obtained by making it larger than the area. When the pressure receiving area difference is provided in this manner, the differential pressure receiving area generates a hydraulic pressure in the closing direction that increases as the load pressure increases, so that the differential pressure across the meter-in variable throttle is proportional to the load pressure. It is controlled to be small, and the supply flow rate to the actuator decreases. Due to the decrease in the supply flow rate to the actuator, the hydraulic pump controlled by the load sensing reduces the discharge flow rate, avoids a sudden rise in pressure, and attenuates pressure pulsation early.

On the other hand, in a hydraulic circuit device provided with a load sensing system, only the drive speed of a specific hydraulic actuator is reduced without changing the target differential pressure of the load sensing control set in the pump control means, thereby performing a very low speed operation. As a possible method, there is a proposal described in JP-A-2-296002. This proposal modulates the load pressure by setting the spring force of the check valve for detecting the load pressure to a certain strength and applying a pressure loss at the check valve portion, and the signal pressure is more than the load pressure. By decreasing by the pressure loss, the differential pressure between the discharge pressure and the load pressure of the hydraulic pump that is load-sensing controlled also decreases from the normal differential pressure by the pressure loss, and the control flow rate is reduced accordingly.

In a hydraulic circuit device provided with a load sensing system, a control valve described in International Patent Application Publication No. WO 98/31940 is one in which the structure is simplified as a valve assembly by combining a flow dividing valve and a hold check valve. . In this device, the valve body of the flow dividing valve is partially incorporated in the hollow valve body of the hold check valve, and the load pressure detection oil passage of the control valve is formed as an internal passage (oil passage slit) of the flow dividing valve. And by giving a check valve function using its internal passage,
The check valve as a valve element is not required, and the configuration of the entire control valve is simplified.

[0009]

Problems to be Solved by the Invention
According to the proposals described in Japanese Patent Application Laid-Open No. H05-263804 and Japanese Patent Application Laid-Open No. 5-263804,
Since the supply flow rate to the hydraulic actuator decreases in proportion to the load pressure, and the discharge flow rate of the hydraulic pump decreases, a sudden rise in pressure when the hydraulic actuator is driven can be avoided.
In addition, the hydraulic pulsation is also attenuated at an early stage, and a smooth start-up characteristic can be obtained regardless of the size of the driven inertial body.
However, these conventional techniques have the following problems.

The proposals described in JP-A-4-191501 and JP-A-5-263804 use a control valve that controls the displacement of a proportional seat valve by the valve opening of a pilot valve. It is structurally difficult to implement the proposal using a spool-type control valve. In particular, recent control valves use the inside of the spool as an oil passage for assembling a regeneration circuit.
Difficulty is compounded.

Japanese Patent Application Laid-Open No. Hei 10-89304 proposes
The valve structure of the pressure compensating valve when a spool type control valve is used is shown.However, in order to make the pressure compensating valve have a difference in pressure receiving area, the structure is too complicated in consideration of assemblability, and Area management is also difficult.

Japanese Patent Application Laid-Open No. 2-296002 proposes
It is an object of the present invention to reduce only the drive speed of a specific hydraulic actuator to enable a very low speed operation. However, since the discharge flow rate of the hydraulic pump is reduced, it is possible to prevent a sudden rise in pressure when the hydraulic actuator is driven and to attenuate hydraulic pulsation early. In addition, there is an advantage that the structure is simple because only a pressure loss is applied at the check valve portion for detecting the load pressure. However, since the pressure loss given by the check valve portion is set by the spring force, the pressure loss is a constant value irrespective of the load pressure, and the control characteristics according to the size of the inertial body,
That is, load dependency cannot be obtained. For this reason, depending on the size of the inertial body to be driven, there is a problem that a sudden rise of the pressure occurs and the pressure pulsation does not attenuate early when the actuator is driven.

The control valve described in International Publication No. WO 98/31940 has a valve assembly in which a flow dividing valve and a hold check valve are combined, and various functions are incorporated therein, so that the configuration of the entire control valve is simplified. There are advantages. However, no measures have been taken for a sudden rise in pressure or hydraulic pulsation when driving an actuator having a large inertia, and when the inertial body to be driven is large, a sudden rise in pressure occurs when driving the actuator, This causes a problem that the pressure pulsation does not attenuate early.

An object of the present invention is to provide a hydraulic circuit device equipped with a load sensing system, which can obtain a smooth start-up characteristic regardless of the size of an inertia body to be driven, has a simple structure, and is a spool type control valve. An object of the present invention is to provide a hydraulic circuit device which can be easily applied even if there is.

[0015]

(1) In order to achieve the above object, the present invention provides a hydraulic pump, a plurality of hydraulic actuators driven by hydraulic oil discharged from the hydraulic pump, and the hydraulic pump. And a plurality of control valves disposed between the plurality of actuators, a signal pressure based on the maximum load pressure of the plurality of hydraulic actuators is guided, and a signal detection oil passage connected to the tank via a first throttle. Pump control means for controlling a discharge pressure of the hydraulic pump so as to be higher than the signal pressure by a predetermined value, wherein each of the plurality of control valves controls a flow rate of hydraulic oil supplied to the hydraulic actuator. A main valve having a meter-in variable throttle, and disposed between the meter-in variable throttle of the main valve and the hydraulic actuator,
A valve assembly comprising a combination of a flow dividing valve and a hold check valve, wherein the hold check valve of the valve assembly has a hollow spool-shaped valve body in which a control chamber connected to the signal detection oil passage is formed. The flow dividing valve of the valve assembly has a valve body slidably mounted at least partially in the hollow spool-shaped valve body, and has one end located in the inlet passage and the other end located in the control chamber. A hydraulic circuit device that controls the differential pressure across the metered throttle by controlling the pressure of the inlet passage by controlling the pressure of the inlet passage by balancing the pressure of the control chamber and the pressure of the inlet passage. The valve assembly of each of the plurality of control valves includes an oil passage slit formed on an outer periphery of a valve body of the flow dividing valve, and a valve body of the flow dividing valve having the hole. A wrap that opens the oil passage slit to the control chamber when it moves a predetermined distance or more with respect to the hollow spool-shaped valve element of the check valve, and connects an intermediate chamber between the flow dividing valve and the hold check valve to the control chamber. And a valve assembly of a specific control valve of the plurality of control valves, a fixed plug extending in an axial direction inside a hollow spool-shaped valve body of the hold check valve, and a fixed plug of the hold check valve. It has an annular second throttle formed by the inner periphery of the hollow spool-shaped valve body and the outer periphery of the fixed plug.

By providing the annular second throttle in the valve assembly of a specific control valve among the plurality of control valves, the load pressure of the hydraulic actuator related to the specific control valve is the maximum load pressure. At this time, the oil passage slit in the valve assembly of the specific control valve is opened to the control chamber, the load pressure is guided to the control chamber through the intermediate chamber and the oil passage slit, and the annular second throttle is connected to the first throttle. It cooperates and modulates the load pressure induced in the control room and guides it to the signal detection oil passage. For this reason, as the load pressure (maximum load pressure) of the hydraulic actuator related to the specific control valve increases, the differential pressure across the annular second throttle increases, reducing the signal pressure guided to the signal detection oil passage. Becomes stronger. Since the pump control means controls the discharge pressure of the hydraulic pump so as to be higher than the signal pressure by a predetermined value, the differential pressure before and after the meter-in variable throttle of a specific control valve becomes smaller as the load pressure becomes higher. The function of reducing the flow rate comes out. For this reason, when the hydraulic actuator related to a specific control valve is started, the supply flow rate to the hydraulic actuator decreases in accordance with the load pressure, and the discharge flow rate of the hydraulic pump decreases. Can be avoided, and hydraulic pulsation can be damped early,
Smooth starting characteristics can be obtained regardless of the size of the driven inertial body.

Further, since only the annular second restrictor is additionally formed by using the valve body and the fixing plug of the hold check valve in the valve assembly of the flow dividing valve and the hold check valve, the structure is extremely simple. Even if the main valve of the control valve is a spool type, it can be easily applied.
In addition, there is no risk of malfunction since only the second annular stop is added.

(2) In the above (1), preferably,
The fixed plug is provided on all of the valve assemblies of the plurality of control valves, and is the same spring supporting member having a spring receiving portion that holds the flow dividing valve in a closed position. The second throttle of the above, the inner peripheral shape of the hollow spool-shaped valve element of the hold check valve is different from the inner peripheral shape of the hollow spool-shaped valve element of the hold check valve in the valve assembly of another control valve, It is formed by an annular gap between the inner periphery of the valve body and the outer periphery of the spring receiving portion.

This makes it possible to easily form the annular second throttle only on the valve assembly side of the specific control valve.

(3) In the above (1), preferably, the valve bodies of the respective flow dividing valves in the valve assembly of the plurality of control valves have a pressure receiving area on the inlet side larger than a pressure receiving area on the control chamber side. large.

As a result, the characteristics of the control valve on the low load pressure side are improved, for example, by removing the influence of the flow force acting on the flow dividing valve in the control valve on the low load pressure side during the composite operation, and a good composite operation can be performed. The means for improving the characteristics of the control valve on the high load pressure side (installation of the annular second throttle) described in the above (1) and the means for improving the characteristics of the control valve on the low load pressure side (changing the pressure receiving area) are described. Are independent of each other, the specific improvement of the high load pressure side and the improvement of the characteristics of the low load pressure side can be achieved by independent means, and the degree of freedom of equipment selection is greatly increased.

[0022]

Embodiments of the present invention will be described below with reference to the drawings.

First, a hydraulic circuit device according to a first embodiment of the present invention will be described with reference to FIGS.

In FIG. 1, the hydraulic circuit device of the present embodiment includes a fixed displacement hydraulic pump 1 and a bleed valve 2 capable of bleeding the entire discharge oil amount of the hydraulic pump 1 with a small override. A fixed pump load sensing system is configured in combination with the bleed valve 2.

The hydraulic oil discharged from the hydraulic pump 1 is supplied to a plurality of hydraulic actuators 3-1 and 3-2, and a space between the hydraulic pump 1 and the hydraulic actuators 3-1 and 3-2 is shown in FIG. Control valves 4-1 and 4 having spool-type main valves 4a-1 and 4a-2 provided with such meter-in variable throttle M / I and meter-out variable throttle M / O.
-2 is installed, and the flow direction and flow rate of the pressure oil supplied to the hydraulic actuators 3-1 and 3-2 are controlled by switching the main valves 4a-1 and 4a-2. In this embodiment, only two actuators are shown, but the number of actuators that can be used is not limited to this. Also, in FIG. 1, for convenience of illustration, the main valves 4a-1, 4a
-2 variable stop M of meter-in at one side switching position
Only the / I and the meter-out variable aperture M / O are shown separately for the meter-in side and the meter-out side.

The control valves 4-1 and 4-2 are combined with the main valves 4a-1 and 4a-2 having the meter-in variable throttle M / I and the meter-out variable throttle M / O, respectively. The valve assemblies 56-1 and 56-2 each including a combination of the flow dividing valves 5-1 and 5-2 and the hold check valves 6-1 and 6-2 which enable the operation are incorporated.

The valve assemblies 56-1 and 56-2 include:
The main valves 4a-1 and 4a-2 are respectively installed between the meter-in variable throttle M / I and the hydraulic actuators 3-1 and 3-2. The hold check valves 6-1 and 6-2 are located on the inlet passage 56a side of the valve assembly 56 and the valve assembly 56.
-1, 56-2 on the outlet passage 56b side.

In the control valve 4-1, the hold check valve 6-1 of the valve assembly 56-1 has a hollow spool-shaped valve element 60, and the branch valve 5-1 is at least partially a hollow spool-shaped valve. A spool-shaped valve body 50 slidably provided inside the body 60 and having one end located in the inlet passage 56a and the other end located in the control chamber 70 (described later). An intermediate chamber 56c is formed therebetween.

The valve body 60 of the hold check valve 6-1 is
As shown in FIG. 3, a large diameter portion 61 having an outer diameter D2 and an inner diameter d2 and a small diameter portion 62 having an outer diameter D3 (<D2) and an inner diameter d3 (<d2) are provided. Is provided with a seat portion 63. The large diameter portion 61 is slidably fitted to the casing 1, and the small diameter portion 62 is slidably fitted to the inner diameter portion of the sleeve 64 inserted into the casing 1. The boundary between the large diameter portion 61 and the small diameter portion 62 and the sleeve 6
A load pressure chamber 65 is formed between the end surfaces of the large diameter portion 4, and a plurality of slits 66 for guiding the load pressure from the outlet passage 56 b to the load pressure chamber 65 are formed on the outer periphery of the large diameter portion 61.

The valve body 60 is biased by a spring 79 whose base end is supported by an inner step portion of the sleeve 64 so that the seat portion 63 comes into contact with an edge portion between the intermediate chamber 56c and the outlet passage 56b. ing.

Further, the outer diameter D3 of the small diameter portion 62 of the valve body 60 and the inner diameter d2 of the large diameter portion 61 (= the stem portion 53 of the flow dividing valve 5-1).
(Described later) are manufactured to have the same dimensions, thereby eliminating the influence of the force acting on the valve body 60 due to the hydraulic pressure in the control chamber 70.

The valve body 50 of the flow dividing valve 5-1 has a land 52 in which a metering notch 51 is formed and a stem 53, and the stem 53 is provided with a valve body 6 of the hold check valve 6-1.
The control chamber 70 is slidably fitted to the large-diameter portion 61 of the control valve 70 in which the back of the stem portion 53 of the flow dividing valve 5-1 is located in the valve body 60 of the hold check valve 6-1. Split valve 5-
In operation 1, the control chamber 70 has the stem 5 of the flow dividing valve 5-1.
The hydraulic pressure of the intermediate chamber 56c is guided through the oil passage slit 20 provided on the outer periphery of the third 3 (described later).

The control chamber 70 communicates with a spring chamber 72 which is located behind the small diameter portion 62 of the hold check valve 6-1 and also serves as a second control chamber. The spring chamber 72 is provided with a small hole 73 provided in a sleeve 64. And is connected to the signal detection oil passage 9 via the. The signal detection oil passage 9 is further connected to the tank T via a throttle 14 (area at).

A fixed plug 75 extends in the axial direction from the spring chamber 72 through the inside of the small diameter portion 62 of the valve body 60, and a base end of the fixed plug 75 is held by a sleeve 64. The fixing plug 75 has a flange-like spring receiving portion 76 near the tip.
Is provided, and the spring receiving portion 76 has
A spring 78 for urging the valve element 50 of the flow dividing valve 5-1 to abut against the inner wall 77 of the inlet passage 56a to hold the valve element 50 in the closed position.
The base end is supported. In addition, hold check valve 6
The inner circumference of the small-diameter portion 62 of the valve body 60 of -1 has such an inner diameter as to narrow the annular gap between the outer periphery of the spring receiving portion 76 of the fixed plug 75, and the present invention uses the annular gap. An annular aperture 11 (area ac> at), which is a feature of (1), is formed.

One end of the above-described oil passage slit 20 formed on the outer periphery of the stem portion 53 of the flow dividing valve 5-1 is opened to the intermediate chamber 56c, the other end is cut in the middle of the stem portion 53, and the valve body 50 is closed. When the valve body 50 is in the position, a wrap portion 32 having a wrap amount X for blocking communication between the oil passage slit 20 and the control chamber 70 is formed. 20 opens into the control room 70. That is, the lap portion 32 opens only when the load pressure of the hydraulic actuator 3-1 (see FIG. 1) to which the lap portion 32 is related is the maximum load pressure, and the wrap portion 32 detects the pressure (load pressure) of the intermediate chamber 56c. Performs a valve stop function.

As described above, the wrap portion 32 performs the check valve function, so that the hydraulic actuator 3 to which the wrap portion 32 is related can be used.
Dividing valve 5-1 only when the load pressure of No. 1 is the maximum load pressure
The load pressure is detected from the intermediate chamber 56c between the control chamber 70 and the hold check valve 6-1, and the load pressure is guided to the control chamber 70. Further, between the control chamber 70 and the spring chamber 72, an annular throttle 1 is provided.
1, the annular throttle 11 and the throttle 14 provided in the signal detection oil passage 9 cooperate to form the control chamber 7.
The load pressure induced to 0 is modulated (described later), and the modulated load pressure is guided to the signal detection oil passage 9.

The wrap amount X of the wrap portion 32 is different from the valve body 5.
It functions as a dead zone at the time of the operation of 0, and a dead zone Y for the metering notch 20 is formed on the land 52 of the flow dividing valve 5-1 and has a relationship of Y <X. The dead zone X is constant with respect to the valve element 60 of the hold check valve 6-1, but when the valve element 60 moves upward in the figure, the dead zone X changes according to the position of the valve element 60. Functions as a variable dead zone.

Further, the land 5 of the valve body 50 of the flow dividing valve 5-1 is provided.
Numeral 2 has an outer diameter d1 larger than the outer diameter d2 of the stem portion 53 for reducing the fluid force, and the pressure receiving area Ai on the inlet passage 56a side of the valve body 50 of the flow dividing valve 5-1 and the pressure receiving area Ac on the control chamber 70 side. And Ai> Ac.

The intermediate chamber 56c is larger than the outer diameter d1 of the land 52 and the outer diameter D of the large diameter portion 61 of the hold check valve 6-1.
Having an inner diameter D1 smaller than 2 and a hold check valve 6-
1 can surely fulfill the hold check function.

FIGS. 4 and 5 show equivalent circuits of the control valve 4-1. In the figure, parts that are the same as the parts or functions shown in FIG. 3 are given the same reference numerals.

In the control valve 4-1 of the present embodiment, the valve assembly 56-1 of the flow dividing valve 5-1 and the hold check valve 6-1 is the same as the separated flow dividing valve 5-1 and the hold check valve 6--6 shown in FIG. 1, and the annular stop 11 corresponds to the stop 11 in FIG. That is, the valve assembly 56-1 has a configuration in which the three valve elements of the flow dividing valve 5-1 and the hold check valve 6-1 and the throttle 11 of FIG. In the control valve 4-1 of the present embodiment, the oil passage slit 20, the control chamber 70, and the spring chamber 72 correspond to the load pressure detection oil passage 7-1 in FIG. , Check valve 8 of FIG.
-1. That is, the valve assembly 56-1 includes
The load pressure detection oil passage 7-1 of the control valve shown in FIG. 5 is configured as an internal passage, and the function of the check valve 8-1 is given by using an oil passage slit 20 which is a part of the internal passage. . As described above, various functions are integrated in the valve assembly 56-1.
The overall configuration can be made compact and simple.

The valve assembly 56-2 of the control valve 4-2 is configured similarly to the valve assembly 56-1 of the control valve 4-1 except that the control valve 4-1 does not have the annular throttle 11.

That is, in the valve assembly 56-1 of the control valve 4-1, the inner circumference of the small-diameter portion 62 of the valve body 60 of the hold check valve 6-1 is connected to the spring receiving portion 76 of the fixed plug 75 as described above. The inner diameter dimension is such that an annular gap between the outer circumference and the outer circumference is narrowed, and the annular gap 11 provides an annular aperture 11 (area at). In contrast, the valve assembly 56-2 of the control valve 4-2
Is a small diameter portion 62 of the valve body 60 of the hold check valve 6-1.
Is enlarged on the control chamber 70 side, and an annular gap between the inner periphery of the small diameter portion 62 and the outer periphery of the spring receiving portion 76 of the fixed plug 75 is widened to form a passage structure without a throttle. Thus, by using the same fixed plug (spring support member) and changing the inner peripheral shape of the small diameter portion 62 of the valve body 60, the annular throttle 11 can be easily formed only on the valve assembly 56-1 side.

The bleed valve 2 is provided with a valve body 2a, a spring chamber 2b in which the working end of the valve body 2a in the valve closing direction is located, and is disposed in the spring chamber 2b to urge the valve body 2a in the valve closing direction. Spring 2
c, the spring chamber 2b is connected to the signal detection oil passage 9 via the throttle 15, and the signal pressure detected in the signal detection oil passage 9 is guided to the spring chamber 2b. Set the discharge pressure of the hydraulic pump 1 to P
1. Assuming that the signal pressure in the signal detection oil passage 9 is Pc, the bleed valve 2 recirculates the excess flow from the hydraulic pump 1 to the tank T when the difference between P1 and Pc is equal to or greater than the differential pressure ΔPL set by the spring 2c. Work. This means that the control valve 4-
The inlet pressure of the meter-in variable throttle M / I (= P1), which is the differential pressure created by the amount of oil flowing in 1,4-2.
When the pressure difference between the pressure and the signal pressure Pc of the signal detection oil passage 9 exceeds ΔPL, it means that the excess flow is returned to the tank T.

Reference numeral 21 denotes a main relief valve for protecting the main circuit, and reference numeral 22 denotes an auxiliary relief valve for protecting the signal circuit.

In the above, the valve assembly 56-1
The structure other than the annular throttle 11 or the structure of the valve assembly 56-2 without the annular throttle is disclosed in International Publication WO98 / 9898.
/ 31940, which is basically the same as the valve structure of the “directional control valve with a flow dividing valve” disclosed in US Pat.
Other structural features of 56-2 and operational features other than those described below are described in WO 98/31940.

The operation of the hydraulic circuit device according to the present embodiment will be described. In the following description, the discharge pressure of the hydraulic pump 1 and the signal pressure of the signal detection oil passage 9 are respectively set to P1,
Let Pc be the pressure of the inlet passage 56a of the flow dividing valve 5-1 (hereinafter, appropriately referred to as an inlet pressure) as P2, and the pressure of the outlet passage (intermediate chamber 56c) of the flow dividing valve 5-1 (hereinafter, appropriately referred to as an outlet pressure) as P3. The pressure in the control chamber 70 (hereinafter referred to as control pressure as appropriate) is P4. Further, the valve element 6 of the hold check valve 6-1 is used.
0, the pressure loss is very small, and the outlet pressure P of the branch valve 5-1 is small.
3 is substantially equal to the load pressure of the hydraulic actuator 3-1.

First, the detected load pressure modulation function of the annular throttle 11 will be described.

If the area of the annular throttle 11 is ac, the area of the throttle 14 is at, and the flow rate passing through the throttles 11 and 14 is q, the relationship between the control pressure P4 and the signal pressure Pc is as follows. However, ac> at. Further, it is assumed that the pressure loss at the wrap portion 32 having the check valve function can be ignored.

Q = C · ac · √ (2 g / γ) √ (P 4 −Pc) = C · at√ (2 g / γ) √Pc C: Flow coefficient g: Gravity γ: Viscosity coefficient The detected signal pressure Pc is Pc = {ac ² / (ac ² + at ² )} · P 4, and P 4 -Pc = {at ² / ac ² + at ² } · P 4 From the relationship of (1), P 4 -Pc , Ie, the differential pressure across the annular throttle 11 is determined. From this equation, the hydraulic actuator 3-1
As the control pressure P4 increases, the differential pressure P4-Pc across the annular throttle 11 increases, and the effect of reducing the signal pressure Pc by the annular throttle 11 increases. I understand. In other words, the annular throttle 11 applies the load pressure (the outlet pressure P
It has a modulating function of increasing the differential pressure P4-Pc depending on 3) and reducing the signal pressure.

When the hydraulic actuator 3-1 is operated alone,
Alternatively, the operation of the control valve 4-1 during the combined operation in which the load pressure of the hydraulic actuator 3-1 is the maximum load pressure will be described.

The difference between the inlet pressure P2 of the flow dividing valve 5-1 and the control pressure P4 of the control chamber 70 is defined as ΔPb1. The pressure difference ΔPb1 is a pressure loss in the oil passage from the inlet passage 56a to the control chamber 70 and is a function of the control flow rate. The influence of the flow rate is reduced by reducing the pressure loss as much as possible. In this case, ΔPb1
Is small, and the control pressure P4 is the outlet pressure P3 of the branch valve 5-1.
That is, it is almost equal to the load pressure.

If there is no annular stop 11, P4
= Pc, meter-in variable throttle M of main valve 4a-1
The differential pressure before and after / I is as follows: P1−P2 = (Pc + ΔPL) − (P4 + Pb1) = ΔPL−ΔPb1 (2) On the other hand, when the annular aperture 11 is inserted,
The signal pressure Pc becomes lower than the control pressure P4 due to the detected load pressure modulation function of the annular throttle 11, and the differential pressure across the meter-in variable throttle M / I of the main valve 4a-1 is P1-P2 = (Pc + ΔPL) − (P4 + Pb1) =. DELTA.PL-.DELTA.Pb1- (P4-Pc) (3) The pressure is reduced by the differential pressure between P4 and P4-Pc.

Here, the differential pressure P4-Pc represented by the equation (3) increases as the load pressure (outlet pressure P3) increases due to the modulating function of the annular throttle 11 represented by the equation (1). Therefore, the function of reducing the control flow rate as the load pressure increases becomes apparent. That is, the control valve 4-1
Has a load-dependent characteristic in which the control flow rate Q decreases as the load pressure (outlet pressure P3) increases as shown in FIG.

FIG. 7 shows the result of a simulation performed for examining the effect of the annular stop 11. In FIG.
The moment of inertia of the hydraulic actuator 3-1 is different between (a) and (b), and (b) has three times the moment of inertia with respect to (a). The upper part of (a) and (b) shows the relationship between the discharge oil amount Qp of the hydraulic pump 1, the flow rate Q1 flowing to the load side, and the flow rate Qc bleed by the bleed valve 2. The control valve 4-1 is fully operated in 0.5 seconds. The middle stage of (a) and (b) is the pump discharge pressure P1,
The lower part shows the angular velocity ω of the hydraulic actuator 3-1. In order to see the effect of the annular stop 11, the ratio of the opening area ac of the annular stop 11 to the opening area at of the stop 14, k = ac / at, was selected as a parameter.

1) At k = 25 where there is no effect of the aperture 11,
This is remarkable when the hydraulic pulsation is large, especially when the moment of inertia is large. In this simulation, it was assumed that the main relief valve 21 did not operate.
The discharge pressure (drive pressure) P1 is considerably high due to the large moment of inertia.

2) When k = 5.76, the bleed flow rate by the bleed valve 2 transiently increases, the rotation of the hydraulic actuator 3-1 is smooth, and the pressure pulsation is attenuated immediately. (When k = 5.76 is compared with the diameter of the aperture, there is a relation of dc = 1.2 with respect to dt = 0.5.) When the rotation speed becomes a constant value, the driving pressure decreases, and P4−Pc The value also becomes smaller, and the same rotational speed as when the detected pressure is not modulated is obtained.

In this embodiment, the annular aperture 1
1 is provided between the control chamber 70 and the spring chamber 72,
A pressure difference is generated between the control chamber 70 and the spring chamber 72, and a pushing force upward acts on the valve element 60 of the hold check valve 6-1 to increase the opening degree of the valve element 60 of the hold check valve 6-1. Therefore, pressure loss at the high load pressure side hold check valve 6-1 can be reliably reduced.

Since the lap amount X of the lap portion 32 is in the variable dead zone as described above, the hold check valve 6
When the opening degree of the valve body 60 of -1 increases, the valve body 50 of the flow dividing valve 5-1 moves following the valve body 60 and moves, and the displacement of the valve body 50 of the flow dividing valve 5-1 increases by that much. The opening area increases. For this reason, the pressure loss generated in the flow dividing valve 5-1 is also reduced.

Next, the operation of the control valve 4-2 on the low load pressure side during the combined operation in which the load pressure of the hydraulic actuator 3-1 is the maximum load pressure will be described.

There are some differences in the flow characteristics required on the high load pressure side and the low load pressure side during the combined operation. As one of the flow characteristics required on the low load pressure side during the combined operation, there are cases where it is better that a large amount of pressure oil flows on the low load pressure side. For example, in a combined operation of a boom and a swing of a hydraulic shovel, there is a demand to drive the swing with the extension drive pressure of the boom, and in this case, a characteristic in which the function of the flow dividing valve is considerably relaxed is required. The second is to eliminate the influence of the flow force acting on the flow dividing valve on the low load pressure side. The flow force acting on the flow dividing valve is as follows: FL = 2 · C · A (x) · (Pin−Pout) · cos θ C: Flow coefficient A (x): Opening area determined by the stroke x of the valve body Pin: Inlet pressure Pout : Outlet pressure θ: Flow angle, and the flow force FL increases in accordance with the differential pressure Pin-Pout before and after the restriction of the flow dividing valve. The differential pressure Pin-Pout before and after the restriction of the flow dividing valve increases at the flow dividing valve on the low load side. For this reason, the influence of the flow force acting on the flow dividing valve becomes large on the low load pressure side. As described above, since the annular throttle 11 is installed in the control valve 4-1 on the high load pressure side, when the load pressure (the outlet pressure P3) increases, the control flow rate Q decreases as shown in FIG. Has characteristics. In the branch valve 5-2 of the control valve 4-1 on the low load pressure side, the signal pressure Pc of the signal detection oil passage 9 is guided to the control chamber 70. The valve body 50 of the high load pressure side flow dividing valve 5-1 is balanced with the pressure P2 and the pressure P4, whereas the valve body 50 of the low pressure side flow dividing valve 5-2 is guided to the control chamber 70. The signal pressure Pc is in proportion to the signal pressure Pc. Since the detected signal pressure Pc is a value obtained by reducing the detected load pressure (outlet pressure P3) (= P4) by the annular throttle 11, the flow dividing valve 5-5 on the low load pressure side is used. The second valve body 50 should be balanced at an inlet pressure Pin lower than P2. However, the restrictor (metering notch 5) of the valve body 50 is provided in the valve body 50 of the branch valve 5-2 on the low load pressure side.
The flow force corresponding to the differential pressure Pin-P5 of 1) acts in the valve closing direction. In order to balance the flow force with the signal pressure Pc of the control chamber 50, the inlet pressure Pin of the flow dividing valve 5-2 needs to be increased. A pressure higher than P2 is required. In other words,
On the low load pressure side, due to the influence of the flow force, the differential pressure ΔPb2 between the inlet pressure Pin of the flow dividing valve 5-2 and the control pressure Pc of the control chamber 70 described with reference to the expression (2) in the first embodiment can be ignored. Disappears. As a result, as shown by the dotted line in FIG.
The control flow rate Q may decrease as the differential pressure between P3 and P5 increases. In this case, the control valve 4-1 on the high load pressure side is controlled to decrease the flow rate when the load pressure increases, whereas the control valve 4-2 on the low load pressure side increases as the differential pressure between P3 and P5 increases. The control flow rate decreases, and the operation on the high load pressure side is canceled. Further, this is contrary to the fact that when the pressure on the high load pressure side is constant and the pressure on the low load pressure side decreases, the flow rate consumed on the low load pressure side decreases.

In this embodiment, in order to cancel the influence of the flow force on the flow dividing valve 5-2 of the control valve 4-2 on the low load pressure side, as described above, the pressure receiving area Ai on the inlet passage 56a side and the control chamber are controlled. The pressure receiving area Ac on the 70 side has a relationship of Ai> Ac so that the differential pressure between the inlet pressure and the outlet pressure of the flow dividing valve 5-2 acts on the area of Ai-Ac. As a result, the flow force increases in proportion to the differential pressure P3-P5, and while the valve element 50 acts on the closing side, the force acting on the area Ai-Ac and acting on the opening side also acts on the differential pressure P3-P5. Since it increases in proportion, the effect of the flow force is cancelled, and the characteristic that the control flow rate Q increases as the differential pressure P3-P5 increases as shown by the solid line in FIG. 8 is obtained.

As a result, the control valve 4-2 on the low pressure load side during the combined operation can also remove the influence of the flow force, improve the characteristics, and perform a favorable combined operation. Also,
The means for improving the characteristics of the control valve 4-1 on the high load pressure side is only to install the annular throttle 11, and the means for improving the characteristics of the control valve 4-2 on the low load pressure side reduces the pressure receiving area of the flow dividing valve 5-2. The only difference is that both improvement means are completely independent of each other, so that the required performance on the high load pressure side and the required performance on the low load pressure side can be achieved by independent means, greatly increasing the freedom of equipment selection. I do.

As described above, according to the present embodiment, when the hydraulic actuator 3-1 is operated alone or in a combined operation in which the load pressure of the hydraulic actuator 3-1 is the maximum load pressure, the activation of the hydraulic actuator 3-1 is performed. At times, the supply flow rate to the hydraulic actuator 3-1 decreases according to the load pressure,
Since the discharge flow rate of the hydraulic pump 1 is reduced, a sudden rise in pressure when the hydraulic actuator is driven can be avoided, and the hydraulic pulsation can be attenuated at an early stage, so that a smooth startup characteristic can be obtained regardless of the size of the driven inertial body. Can be

Further, since the annular throttle 11 is additionally formed between the control chamber 70 and the spring chamber 72 of the valve assembly 56-1 using the valve body 60 and the fixed plug 75,
The configuration is extremely simple, and it can be easily applied even if the main valve 4a-1 of the control valve 4-1 is a spool type. In addition, there is no possibility of malfunction since only the annular stop 11 is added.

Further, the valve assembly 56-1 of the control valve 4-1 includes three valve elements, namely, a flow dividing valve 5-1 and a hold check valve 6-1 and a throttle 11 in FIG. The functions of the path 7-1 and the check valve 8-1 are integrated, and the valve assembly 56-2 of the control valve 4-2 is also the same, thereby making the entire configuration of the control valve 4-1 compact. Can be simplified and simplified.

Further, the characteristics of the control valve 4-1 on the high load pressure side in the single operation and in the combined operation are made load-dependent so that the characteristics of the control valve 4-1 are improved. The characteristics of the control valve 4-2 on the side are also improved, for example, by removing the influence of the flow force, and a favorable composite operation can be performed. The only means for improving the characteristics of the control valve 4-1 on the high load pressure side is to install the annular throttle 11, and the means for improving the characteristics of the control valve 4-2 on the low load pressure side is the pressure receiving of the flow dividing valve 5-2. Only the area is different, and both improvement means are completely independent of each other. Therefore, the required performance on the high load pressure side and the required performance on the low load pressure side can be achieved by means independent of each other, and the degree of freedom in selecting equipment is greatly increased.

While the representative embodiment of the present invention has been described above, various modifications can be made to this embodiment within the spirit of the present invention. For example, in the above embodiment, a fixed displacement hydraulic pump is used as the hydraulic pump, and the bleed 2 is used as the pump control means of the load sensing system. However, as shown in FIG. 9, a variable displacement hydraulic pump is used as the hydraulic pump. 1A, as the pump control means of the load sensing system, the discharge pressure P1 of the hydraulic pump 1A is higher than the signal pressure Pc of the signal detection oil passage 9 by the set value ΔPL of the spring 2d.
Alternatively, a tilt controller 2A that performs tilt control of the hydraulic pump 1A so as to be higher may be used. Similar effects can be obtained by applying the present invention to a hydraulic circuit device having such a load sensing system.

[0069]

According to the present invention, when the hydraulic actuator is started, the supply flow rate to the hydraulic actuator is reduced according to the load pressure, and the discharge flow rate of the hydraulic pump is reduced. As a result, the hydraulic pulsation can be attenuated at an early stage, and a smooth start-up characteristic can be obtained regardless of the size of the driven inertial body.

Further, since only the annular second restrictor is additionally formed by using the valve element and the fixed plug of the hold check valve in the valve assembly of the flow dividing valve and the hold check valve, the configuration is extremely simple. Even if the main valve of the control valve is a spool type, it can be easily applied.
In addition, there is no risk of malfunction since only the second annular stop is added.

Furthermore, the three valve elements of the flow dividing valve, the hold check valve, and the annular second throttle are rationally integrated into one valve assembly, and the load pressure detecting oil passage of the control valve is connected to the internal passage of the valve assembly. And the function of the check valve is provided by utilizing the oil passage slit which is a part of the internal passage, so that the configuration of the entire control valve can be made compact and simple.

Further, the characteristics of the control valve on the low load pressure side are improved by removing the influence of the flow force acting on the flow dividing valve in the control valve on the low load pressure side during the composite operation, so that a good composite operation can be performed and the high load can be achieved. The specific improvement of the control valve on the pressure side and the improvement of the characteristics of the control valve on the low load pressure side can be achieved by independent means, and the degree of freedom of equipment selection is greatly increased.

[Brief description of the drawings]

FIG. 1 is a diagram showing a hydraulic circuit device according to a first embodiment of the present invention.

FIG. 2 is a diagram showing functions of a main valve portion of a control valve by hydraulic symbols.

FIG. 3 is an enlarged view showing a valve assembly of a flow dividing valve and a hold check valve of a control valve.

FIG. 4 is a diagram showing an equivalent circuit of a control valve.

FIG. 5 is a diagram showing another equivalent circuit of the control valve.

FIG. 6 is a diagram showing the load dependency of the control valve on the high load pressure side at the time of a single operation or a combined operation obtained by installing a throttle.

FIG. 7 is a diagram showing a result of a simulation performed for investigating an effect due to a load dependency of an aperture.

FIG. 8 is a diagram showing characteristics of a control valve on the low load pressure side during a combined operation.

FIG. 9 is a diagram showing pump control means of a load sensing system when a variable displacement hydraulic pump is used.

[Explanation of symbols]

 DESCRIPTION OF SYMBOLS 1 Fixed displacement type hydraulic pump 1A Variable displacement type hydraulic pump 2 Bleed valve 2A Tilting controller 3-1 and 3-2 Hydraulic actuator 4-1 and 4-2 Control valve 4a-1, 4a-2 Main valve 5 -1,5-2 Split valve 6-1, 6-2 Hold check valve 9 Signal detection oil passage 11 Ring throttle (second throttle) 14 throttle (first throttle) 20 oil passage slit 32 lap portion 50 valve element 51 Metering notch 52 Land 53 Stem section 56-1, 56-2 Valve assembly 56a Inlet passageway 56b Outlet passageway 56c Intermediate chamber 60 Valve element 61 Large diameter section 62 Small diameter section 63 Seat section 64 Sleeve 65 Load pressure chamber 66 Slit 70 Control chamber 72 Spring chamber 73 Small hole 75 Fixed plug 76 Spring receiving portion 77 Inner wall 78 Spring 79 Spring M / I Variable throttle of meter-in of main valve M / O Variable throttle of meter-out of main valve X lap amount T tank

 ──────────────────────────────────────────────────の Continuing on the front page (72) Inventor Yusaku Nozawa 650 Kandamachi, Tsuchiura-shi, Ibaraki Hitachi Construction Machinery Co., Ltd. F-term in Tsuchiura Plant (reference) 2D003 AA01 AB05 BA01 BB02 BB03 CA08 CA09 DA02 DB02 DC05 3H089 AA27 CC01 CC08 CC11 DA02 DB03 DB12 DB75 DB76 DB78 EE04 EE07 HH05 JJ02

Claims (3)

[Claims] A hydraulic pump driven by hydraulic oil discharged from the hydraulic pump; a plurality of control valves disposed between the hydraulic pump and the plurality of actuators; Signal pressure based on the maximum load pressure of the hydraulic actuator of
A signal detection oil passage connected to the tank via a throttle, and pump control means for controlling a discharge pressure of the hydraulic pump so as to be higher than the signal pressure by a predetermined value, wherein each of the plurality of control valves is A main valve having a meter-in variable throttle that controls the flow rate of pressure oil supplied to the hydraulic actuator, and a flow dividing valve and a hold disposed between the meter-in variable throttle of the main valve and the hydraulic actuator. A valve assembly comprising a combination with a check valve, wherein the hold check valve of the valve assembly has a hollow spool-shaped valve body in which a control chamber connected to the signal detection oil passage is formed. Is slidably mounted at least partially within the hollow spool-shaped valve body, and has one end located in the inlet passage and the other end. Having a valve body located in the control chamber, the valve body strokes with a balance between the pressure of the control chamber and the pressure of the inlet passage to control the pressure of the inlet passage, and before and after the variable throttle of the meter-in. In a hydraulic circuit device for controlling a differential pressure, each valve assembly of the plurality of control valves includes an oil passage slit formed on an outer periphery of a valve body of the flow dividing valve, and a valve body of the flow dividing valve being the hold check valve. A wrap portion that opens the oil passage slit to the control chamber when moved over a predetermined distance with respect to the hollow spool-shaped valve element, and communicates an intermediate chamber between the flow dividing valve and the hold check valve with the control chamber. And a valve assembly of a specific control valve of the plurality of control valves is configured so that an inside of a hollow spool-shaped valve body of the hold check valve extends in an axial direction. And a second annular restriction formed by an inner periphery of a hollow spool-shaped valve element of the hold check valve and an outer periphery of the fixed plug. 2. The hydraulic circuit device according to claim 1, wherein the fixed plug is provided on all of the valve assemblies of the plurality of control valves, and has the same spring support having a spring receiving portion for holding the flow dividing valve in a closed position. An annular second member of the valve assembly of the particular control valve.
The throttle is configured such that the inner peripheral shape of the hollow spool-shaped valve element of the hold check valve is different from the inner peripheral shape of the hollow spool-shaped valve element of the hold check valve in the valve assembly of another control valve. A hydraulic circuit device formed by an annular gap between an inner periphery and an outer periphery of the spring receiving portion. 3. The hydraulic circuit device according to claim 1, wherein the valve body of each of the branch valves in the valve assembly of the plurality of control valves has a pressure receiving area on the inlet side larger than a pressure receiving area on the control chamber side. A hydraulic circuit device characterized by the above-mentioned.