JP2011196438A - Hydraulic circuit for working vehicle - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a hydraulic circuit for a working vehicle capable of improving energy efficiency and preventing deterioration of mountability, cost increase, and deterioration of work efficiency.SOLUTION: The hydraulic circuit for a working vehicle includes a load sensing system which controls a discharge rate of a first hydraulic pump 21 for supplying a hydraulic oil to a first actuator group 18 in accordance with the maximum load pressure of load pressures applied to the first actuator group 18, and which controls a discharge rate of a second hydraulic pump 22 for supplying the hydraulic oil to a second actuator group 19 in accordance with the maximum load pressure of load pressures applied to the second actuator group 19. The second actuator group 19, which includes a working hydraulic oil actuator (arm cylinder 14) whose required flow rate of the hydraulic oil is the largest in the first actuator group 18 and second actuator group 19, sets the maximum discharge flow rate of the second hydraulic pump 22 larger than that of the first hydraulic pump 21.

Description

  The present invention relates to a technology of a hydraulic circuit of a work vehicle, and more specifically, a work including two hydraulic pumps and a plurality of work hydraulic actuators driven by hydraulic oil respectively supplied from the two hydraulic pumps. The present invention relates to the technology of vehicle hydraulic circuits.

  Conventionally, a first actuator group including a first hydraulic pump and a second hydraulic pump having the same capacity, and a plurality of working hydraulic actuators driven by hydraulic oil supplied from one (first) hydraulic pump A second actuator group including a plurality of working hydraulic actuators driven by hydraulic fluid supplied from the other (second) hydraulic pump, a discharge pressure of the first hydraulic pump, and a first actuator group Controls the amount of hydraulic fluid discharged by the first hydraulic pump based on the maximum load pressure, and sets the hydraulic fluid discharged by the second hydraulic pump based on the discharge pressure of the second hydraulic pump and the maximum load pressure of the second actuator group. A technique relating to a hydraulic circuit of a work vehicle including a load sensing system that controls the discharge amount of the vehicle is known. For example, as described in Patent Document 1.

  In the hydraulic circuit of the work vehicle described in Patent Document 1, the amount of hydraulic oil discharged by the first hydraulic pump and the second hydraulic pump is independently determined based on the maximum load pressure of the first actuator group and the second actuator group. Can be controlled. Therefore, compared to a hydraulic circuit that supplies hydraulic oil to all working hydraulic actuators with one hydraulic pump, the hydraulic oil discharge pressure and discharge amount by the two hydraulic pumps can be controlled to an appropriate value. Efficiency can be improved.

However, in the hydraulic circuit of the work vehicle described in Patent Document 1, the first hydraulic pump and the second hydraulic pump have the same capacity, which is disadvantageous in the following points.
That is, it is assumed that a large-capacity hydraulic pump is selected as the first hydraulic pump and the second hydraulic pump according to the working hydraulic actuator having the maximum required flow rate among the first actuator group and the second actuator group. . In this case, the capacity of the hydraulic pump that supplies hydraulic oil to the working hydraulic actuator with the maximum required flow rate can be selected suitable for the working hydraulic actuator, but the other hydraulic pump has a larger capacity than necessary. This is disadvantageous in that the mounting performance due to the increase in size of the hydraulic pump deteriorates and the cost increases.
In addition, as the first hydraulic pump and the second hydraulic pump, a small-capacity hydraulic pump is selected in accordance with the working hydraulic actuator that requires a relatively small flow rate (not the maximum) among the first actuator group and the second actuator group. Assuming that In this case, when operating the hydraulic actuator for work having a large required flow rate, it is necessary to always join the hydraulic oil from the two hydraulic pumps and supply the hydraulic oil to the hydraulic actuator for work. However, this is disadvantageous in that the energy loss due to the merge occurs.

Japanese Utility Model Publication No. 6-40406

  An object of the present invention is to improve energy efficiency in a hydraulic circuit of a working vehicle that supplies hydraulic oil from two hydraulic pumps to a plurality of working hydraulic actuators, and to reduce the mounting property of the hydraulic circuit on the working vehicle. It is an object of the present invention to provide a hydraulic circuit for a work vehicle that can prevent an increase in cost.

  The problem to be solved by the present invention is as described above. Next, means for solving the problem will be described.

  That is, in claim 1, the discharge amount of the first hydraulic pump that supplies the hydraulic oil to the first actuator group including at least one working hydraulic actuator is set to the maximum load pressure among the load pressures applied to the first actuator group. The discharge amount of the second hydraulic pump that supplies hydraulic oil to the second actuator group including at least one working hydraulic actuator is set to the maximum load pressure among the load pressures applied to the second actuator group. A hydraulic circuit for a work vehicle having a load sensing system that controls the operation according to the second actuator group, wherein the required flow rate of hydraulic oil is the highest among the first actuator group and the second actuator group. A maximum discharge flow rate of the second hydraulic pump, and a maximum discharge flow rate of the first hydraulic pump. It is to set larger than the discharge flow rate.

  According to a second aspect of the present invention, when hydraulic oil is supplied from the first hydraulic pump to a specific working hydraulic actuator in the first actuator group, the hydraulic oil discharged from the second hydraulic pump is used as the first hydraulic pressure. A merging valve for merging the hydraulic oil supplied from the pump to the specific working hydraulic actuator is provided.

  According to a third aspect of the present invention, there are provided a plurality of working direction switching valves provided corresponding to the plurality of working hydraulic actuators, for switching the direction of hydraulic oil supplied to the working hydraulic actuator, The hydraulic oil discharged from the second hydraulic pump when the spool stroke amount of the working direction switching valve for supplying the working oil to the specific working hydraulic actuator becomes equal to or greater than a predetermined value. The hydraulic oil is joined to the hydraulic fluid supplied from the hydraulic pump to the specific working hydraulic actuator.

  According to a fourth aspect of the present invention, the work vehicle includes a work device that includes a boom that is rotatably attached to a vehicle body, an arm that is rotatably attached to the boom, and a bucket that is attached to the arm. The specific working hydraulic actuator is a boom cylinder that rotates the boom with respect to the vehicle body, and one working hydraulic actuator of the second actuator group includes the arm on the boom. And the merging valve is supplied from the first hydraulic pump to the specific working hydraulic actuator when hydraulic oil is supplied from the second hydraulic pump to the arm cylinder. The supply amount of the hydraulic oil from the second hydraulic pump joined to the hydraulic oil is limited.

  As effects of the present invention, the following effects can be obtained.

According to the first aspect of the present invention, the hydraulic oil can be appropriately supplied to each working hydraulic actuator while the first hydraulic pump and the second hydraulic pump have the minimum capacity. Therefore, energy efficiency can be improved as compared with the case of using one hydraulic pump or the case of using hydraulic oil from two hydraulic pumps with small capacities.
In addition, since a hydraulic pump having a larger capacity than necessary is not mounted, it is possible to prevent the first hydraulic pump and the second hydraulic pump from being increased in size, improving the mounting performance of the hydraulic circuit on a work vehicle and reducing the cost. Can be reduced.

  In claim 2, when operating a specific working hydraulic actuator, the working oil discharged by the first hydraulic pump and the second hydraulic pump is merged, so that the specific working hydraulic pressure is obtained only by the first hydraulic pump. The actuator can be operated more quickly than when the actuator is operated. Furthermore, even if only the first hydraulic pump with a small maximum discharge flow rate cannot supply sufficient hydraulic fluid to the required flow rate for the specific working hydraulic actuator, the hydraulic fluid from the second hydraulic pump must be merged. Thus, the operation speed of the specific working hydraulic actuator can be improved, and the working efficiency can be improved.

  According to a third aspect of the present invention, when the switching operation of the working direction switching valve is performed and the spool stroke amount of the working direction switching valve is less than a predetermined value, the hydraulic oil supplied from the first hydraulic pump Only by operating a specific working hydraulic actuator, the specific working hydraulic actuator can be precisely operated. On the other hand, when the spool stroke amount of the work direction switching valve becomes equal to or greater than a predetermined value, the hydraulic oil discharged from the two pumps can be merged to quickly operate the specific work hydraulic actuator. Thereby, the working efficiency of the working hydraulic actuator can be improved.

  In Claim 4, when operating a boom cylinder and an arm cylinder simultaneously, the fall of the operating speed of an arm cylinder can be prevented by restrict | limiting supply of the hydraulic fluid from a 2nd hydraulic pump to a boom cylinder. . This makes it possible to improve the workability of the working hydraulic actuator, particularly during the work of pulling the arm while raising the boom (so-called water averaging work) with the bucket lightly contacting the ground.

The side view which showed the whole structure of the turning working vehicle which comprises the hydraulic circuit which concerns on 1st embodiment of this invention. The figure which shows the whole structure of the hydraulic circuit which concerns on 1st embodiment of this invention. The figure which shows the 1st direction switching valve group, the 1st actuator group, etc. among the hydraulic circuits which concern on 1st embodiment of this invention. The figure which shows the 2nd direction switching valve group, the 2nd actuator group, etc. among the hydraulic circuits which concern on 1st embodiment of this invention. The enlarged view which shows the direction switching valve for boom cylinders which concerns on 1st embodiment of this invention. The enlarged view which shows the boom junction valve which concerns on 1st embodiment of this invention. The figure which shows the relationship between the spool stroke amount and opening area of the direction switching valve for work which concerns on 1st embodiment of this invention. The figure which shows the whole structure of the hydraulic circuit which concerns on 2nd embodiment of this invention.

  First, a turning work vehicle 1 including a hydraulic circuit 201 according to a first embodiment of the present invention will be described with reference to FIG. In this embodiment, the turning work vehicle 1 is described as an embodiment of the work vehicle. However, the work vehicle is not limited to this, and may be other agricultural vehicles, construction vehicles, industrial vehicles, or the like. .

  The turning work vehicle 1 includes a traveling device 2, a turning device 3, and a working device 4.

The traveling device 2 includes a pair of left and right crawlers 5, 5, a left traveling hydraulic motor 5L, and a right traveling hydraulic motor 5R.
The travel device 2 drives the crawler 5 on the left side of the machine body by the hydraulic motor 5L for left travel and the crawler 5 on the right side of the machine body by the hydraulic motor 5R for right travel, thereby moving the turning work vehicle 1 forward and backward. Can do.

The turning device 3 constitutes the vehicle body of the turning work vehicle 1, and includes a turning table 6, a turning motor 7, a control unit 8, and an engine 9.
The swivel base 6 is disposed above the travel device 2 and is supported by the travel device 2 so as to be capable of swiveling. The turning device 3 can turn the turntable 6 with respect to the traveling device 2 by driving the turning motor 7. In addition, on the swivel base 6, a control unit 8 including various operation tools, an engine 9 serving as a power source, and the like are arranged.

The work device 4 includes a boom 10, an arm 11, a bucket 12, a boom cylinder 13, an arm cylinder 14, and a bucket cylinder 15.
One end of the boom 10 is pivotally supported by the front portion of the swivel base 6 and is rotated by a boom cylinder 13 that is extended and retracted. More specifically, when the boom cylinder 13 is extended, the boom 10 is rotated upward, and when the boom cylinder 13 is contracted, the boom 10 is rotated downward.
One end of the arm 11 is pivotally supported by the other end of the boom 10 and is rotated by an arm cylinder 14 that is extended and retracted. More specifically, when the arm cylinder 14 is extended, the arm 11 is rotated downward (the direction in which the other end of the arm 11 is close to the boom 10), and when the arm cylinder 14 is contracted, the arm 11 is upward. (The other end side of the arm 11 is rotated away from the boom 10).
One end of the bucket 12 is supported by the other end of the arm 11 and is rotated by a bucket cylinder 15 that is driven to extend and retract. More specifically, when the bucket cylinder 15 is extended, the bucket 12 is rotated downward (a direction in which the other end side of the bucket 12 is close to the arm 11), and when the bucket cylinder 15 is contracted, the bucket 12 is upward. (The other end side of the bucket 12 is rotated away from the arm 11).
As described above, the working device 4 has a multi-joint structure that excavates earth and sand using the bucket 12.

  In addition, although the working apparatus which the turning working vehicle 1 which concerns on this embodiment comprises is the working apparatus 4 which has the bucket 12 and performs excavation work, it is not limited to this, For example, it has a hydraulic breaker. A working device that performs crushing work may be used.

  Next, the hydraulic circuit 201 provided in the turning work vehicle 1 will be described with reference to FIGS. The hydraulic circuit 201 includes a first hydraulic pump 21, a second hydraulic pump 22, a first pump flow control actuator 23, a second pump flow control actuator 24, a control valve 30, a first actuator group 18, and a second actuator group 19. It has. The first hydraulic pump 21, the second hydraulic pump 22, and the control valve 30 are attached to the turning device 3. The first actuator group 18 includes a boom cylinder 13 and a bucket cylinder 15. The second actuator group 19 includes the turning motor 7 and the arm cylinder 14.

  The hydraulic circuit 201 constitutes a so-called after-orifice type load sensing system in which a pressure compensation valve is connected after a throttle provided in a working direction switching valve that switches the direction of hydraulic oil supplied to the working hydraulic actuator. ing. The load sensing system can control the amount of hydraulic oil discharged by the first hydraulic pump 21 and the second hydraulic pump 22 in accordance with the load pressure applied to the working hydraulic actuator, thereby improving the energy consumption efficiency.

  For convenience of explanation, the boom cylinder direction switching valve 42, the bucket cylinder direction switching valve 43, the arm cylinder direction switching valve 62, and the swing motor direction switching valve 63 are collectively referred to simply as “working direction switching valve”. ". The boom cylinder pressure compensation valve 52, the bucket cylinder pressure compensation valve 53, the arm cylinder pressure compensation valve 72, and the swing motor pressure compensation valve 73 are collectively referred to simply as “pressure compensation valves”. The boom cylinder 13, arm cylinder 14, bucket cylinder 15, and swing motor 7 are collectively referred to as “working hydraulic actuator”.

The first hydraulic pump 21 and the second hydraulic pump 22 shown in FIGS. 2 to 4 are driven by the engine 9 (see FIG. 1) and discharge hydraulic oil. The first hydraulic pump 21 and the second hydraulic pump 22 are variable displacement pumps that can change the discharge amount of hydraulic oil by changing the inclination angles of the movable swash plate 21a and the movable swash plate 22a, respectively.
As will be described in detail later, a hydraulic pump having a capacity (maximum discharge flow rate) smaller than that of the second hydraulic pump 22 is selected as the first hydraulic pump 21.
The hydraulic oil discharged from the first hydraulic pump 21 and the second hydraulic pump 22 is supplied to the control valve 30. More specifically, the hydraulic oil discharged from the first hydraulic pump 21 is supplied to the first direction switching valve group 40 via the oil passage 21b. Further, the hydraulic oil discharged from the second hydraulic pump 22 is supplied to the second direction switching valve group 60 and the boom junction valve 31 via the oil passage 22b.

  The control valve 30 switches the flow of hydraulic oil. The control valve 30 includes a first direction switching valve group 40, a second direction switching valve group 60, and a boom junction valve 31.

  As shown in FIGS. 2 and 3, the first direction switching valve group 40 includes a boom cylinder direction switching valve 42 and a bucket cylinder direction switching valve 43.

As shown in FIG. 3, the boom cylinder direction switching valve 42 is a pilot-type direction switching valve capable of switching the direction of hydraulic oil supplied to the boom cylinder 13.
A boom cylinder pressure compensating valve 52 is connected to the boom cylinder direction switching valve 42. The boom cylinder pressure compensating valve 52 compensates the pressure after the throttle 42c (or 42f) provided in the boom cylinder direction switching valve 42 to a predetermined value.

  Hereinafter, the boom cylinder direction switching valve 42 and the boom cylinder pressure compensating valve 52 will be described in detail with reference to FIG.

  The boom cylinder direction switching valve 42 can be switched to the position 42X (neutral position), the position 42Y, or the position 42Z by sliding the spool. When pilot pressure is not applied to either the pilot port 42a or the pilot port 42b of the boom cylinder direction switching valve 42, the boom cylinder direction switching valve 42 is held at the position 42X by the biasing force of the spring. When pilot pressure is applied to the pilot port 42a of the boom cylinder direction switching valve 42, the boom cylinder direction switching valve 42 is switched to the position 42Y. When pilot pressure is applied to the pilot port 42b of the boom cylinder direction switching valve 42, the boom cylinder direction switching valve 42 is switched to the position 42Z.

  When the boom cylinder direction switching valve 42 is in the position 42X, hydraulic fluid is not supplied to the boom cylinder 13 from the oil passage 21b.

  When the boom cylinder direction switching valve 42 is in the position 42Y, the hydraulic oil compensates for the boom cylinder pressure through the oil path 21b and the throttle 42c provided in the spool of the boom cylinder direction switching valve 42 and the oil path 42d. Supplied to the valve 52.

  The hydraulic oil supplied to the boom cylinder pressure compensation valve 52 is supplied again from the boom cylinder pressure compensation valve 52 to the boom cylinder direction switching valve 42 via the oil passage 52a.

  The hydraulic fluid supplied to the boom cylinder direction switching valve 42 via the oil passage 52a is supplied to the rod chamber of the boom cylinder 13 via the oil passage 13a. The boom cylinder 13 is contracted by the hydraulic oil supplied through the oil passage 13a, and the boom 10 is rotated downward. Further, the hydraulic oil discharged from the bottom chamber of the boom cylinder 13 is returned to the boom cylinder direction switching valve 42 via the oil passage 13b.

  The hydraulic oil returned to the boom cylinder direction switching valve 42 through the oil passage 13b is supplied from the boom cylinder direction switching valve 42 through the oil passage 42e and the return oil passage 17a (see FIG. 2). ).

  When the boom cylinder direction switching valve 42 is in the position 42Z, the hydraulic oil compensates for the boom cylinder pressure through the oil path 21b and the throttle 42f provided in the spool of the boom cylinder direction switching valve 42 and the oil path 42d. Supplied to the valve 52.

  The hydraulic oil supplied to the boom cylinder pressure compensation valve 52 is supplied again from the boom cylinder pressure compensation valve 52 to the boom cylinder direction switching valve 42 via the oil passage 52a.

  The hydraulic fluid supplied to the boom cylinder direction switching valve 42 via the oil passage 52a is supplied to the bottom chamber of the boom cylinder 13 via the oil passage 13b. The boom cylinder 13 is extended by the hydraulic oil supplied through the oil passage 13b, and the boom 10 is rotated upward. The hydraulic oil discharged from the rod chamber of the boom cylinder 13 is returned to the boom cylinder direction switching valve 42 via the oil passage 13a.

  The hydraulic fluid returned to the boom cylinder direction switching valve 42 via the oil passage 13a is returned from the boom cylinder direction switching valve 42 to the hydraulic oil tank 17 via the oil passage 42e and the return oil passage 17a.

When the boom cylinder direction switching valve 42 is in the position 42Y or the position 42Z, the pressure in the oil passage 42d is compensated to a predetermined value by the boom cylinder pressure compensating valve 52.
Specifically, the maximum load pressure (hereinafter simply referred to as “first maximum load pressure”) among the load pressures applied to the boom cylinder 13 and the bucket cylinder 15 is the pressure compensation valve for the boom cylinder via the oil passage 23b. 52. The boom cylinder pressure compensating valve 52 compensates the pressure in the oil passage 42d so as to be higher than the first maximum load pressure by a value set by a spring included in the boom cylinder pressure compensating valve 52. .

As shown in FIG. 3, the bucket cylinder direction switching valve 43 is a pilot-type direction switching valve capable of switching the direction of hydraulic oil supplied to the bucket cylinder 15.
A bucket cylinder pressure compensation valve 53 is connected to the bucket cylinder direction switching valve 43. The bucket cylinder pressure compensation valve 53 compensates the pressure after the throttle provided in the bucket cylinder direction switching valve 43 to a predetermined value.

  The configurations of the bucket cylinder direction switching valve 43 and the bucket cylinder pressure compensation valve 53 are substantially the same as the configurations of the boom cylinder direction switching valve 42 and the boom cylinder pressure compensation valve 52.

  When pilot pressure is applied to the pilot port 43a or the pilot port 43b of the bucket cylinder direction switching valve 43, the bucket cylinder direction switching valve 43 is switched from the neutral position to another position. In this case, the hydraulic oil supplied via the oil passage 21 b is supplied to the bucket cylinder 15. Thereby, the bucket cylinder 15 expands and contracts, and the bucket 12 is rotated upward (a direction in which the other end side of the bucket 12 is separated from the arm 11) or downward (a direction in which the other end side of the bucket 12 is close to the arm 11).

  As shown in FIGS. 2 and 4, the second direction switching valve group 60 includes an arm cylinder direction switching valve 62 and a swing motor direction switching valve 63.

As shown in FIG. 4, the arm cylinder direction switching valve 62 is a pilot-type direction switching valve capable of switching the direction of the hydraulic oil supplied to the arm cylinder 14.
An arm cylinder pressure compensation valve 72 is connected to the arm cylinder direction switching valve 62. The arm cylinder pressure compensating valve 72 compensates the pressure after the restriction provided in the arm cylinder direction switching valve 62 to a predetermined value.

  The configurations of the arm cylinder direction switching valve 62 and the arm cylinder pressure compensation valve 72 are substantially the same as the configurations of the boom cylinder direction switching valve 42 and the boom cylinder pressure compensation valve 52.

  When pilot pressure is applied to the pilot port 62a or pilot port 62b of the arm cylinder direction switching valve 62, the arm cylinder direction switching valve 62 is switched from the neutral position to another position. In this case, the hydraulic oil supplied via the oil passage 22 b is supplied to the arm cylinder 14. As a result, the arm cylinder 14 expands and contracts, and the arm 11 is rotated upward (the direction in which the other end of the arm 11 is separated from the boom 10) or downward (the direction in which the other end of the arm 11 is close to the boom 10).

The turning motor direction switching valve 63 is a pilot-type direction switching valve capable of switching the direction of hydraulic fluid supplied to the turning motor 7.
A swing motor pressure compensation valve 73 is connected to the swing motor direction switching valve 63. The swing motor pressure compensation valve 73 compensates the pressure after the throttle provided in the swing motor direction switching valve 63 to a predetermined value.

  When pilot pressure is applied to the pilot port 63a or the pilot port 63b of the swing motor direction switching valve 63, the swing motor direction switching valve 63 is switched from the neutral position to another position. In this case, the hydraulic oil supplied via the oil passage 22 b is supplied to the turning motor 7. Thereby, the turning motor 7 is rotationally driven.

As shown in FIGS. 2, 4, and 6, the boom junction valve 31 discharges the hydraulic oil discharged from the second hydraulic pump 22 to the bottom chamber of the boom cylinder 13 from the first hydraulic pump 21. This is a pilot-type directional control valve that can be combined with supplied hydraulic oil.
A boom merging pressure compensation valve 32 is connected to the boom merging valve 31. The boom merging pressure compensation valve 32 compensates the pressure after the throttle 31c provided in the boom merging valve 31 to a predetermined value.

  Hereinafter, the boom junction valve 31 and the boom junction pressure compensation valve 32 will be described in detail with reference to FIG.

  The boom junction valve 31 can be switched to the position 31X or the position 31Y by sliding the spool. When the pilot pressure is not applied to the pilot port 31a of the boom junction valve 31, the boom junction valve 31 is held at the position 31X by the biasing force of the spring. When pilot pressure is applied to the pilot port 31a of the boom junction valve 31, the boom junction valve 31 is switched to the position 31Y. Further, even when pilot pressure is applied to the pilot port 31a of the boom junction valve 31 (when the boom junction valve 31 is switched to the position 31Y), a larger pilot is generated by the pilot port 31b of the boom junction valve 31. When pressure is applied, the boom junction valve 31 is switched to the position 31X.

  When the boom junction valve 31 is in the position 31X, hydraulic oil is not supplied to the boom cylinder 13 from the oil passage 22b.

  When the boom merging valve 31 is in the position 31Y, the hydraulic oil is supplied from the oil passage 22b to the boom merging pressure compensation valve 32 via the throttle 31c provided in the spool of the boom merging valve 31 and the oil passage 31d. .

  The hydraulic oil supplied to the boom merging pressure compensation valve 32 is again supplied from the boom merging pressure compensation valve 32 to the boom merging valve 31 via the oil passage 32a. At this time, the differential pressure between the oil passage 31d and the oil passage 32a (the differential pressure across the boom junction valve 31) is compensated to a predetermined value by the boom confluence pressure compensation valve 32.

  The hydraulic fluid supplied to the boom junction valve 31 via the oil passage 32a is supplied to the oil passage 13b via the oil passage 31e and a check valve 31f provided in the middle of the oil passage 31e. The hydraulic oil supplied from the oil passage 31e to the oil passage 13b is merged with the hydraulic oil supplied via the boom cylinder direction switching valve 42 in the oil passage 13b and supplied to the bottom chamber of the boom cylinder 13.

As shown in FIGS. 2 and 3, the first pump flow control actuator 23 is connected to the movable swash plate 21a of the first hydraulic pump 21, and the first hydraulic pressure is changed by changing the inclination angle of the movable swash plate 21a. The discharge amount of hydraulic oil from the pump 21 is controlled.
The first pump flow rate control actuator 23 is connected to the oil passage 21b through the oil passage 23a. The first pump flow control actuator 23 is connected to the boom cylinder pressure compensation valve 52 and the bucket cylinder pressure compensation valve 53 via an oil passage 23b.

As shown in FIGS. 2 and 4, the second pump flow control actuator 24 is connected to the movable swash plate 22a of the second hydraulic pump 22, and the second hydraulic pump is changed by changing the inclination angle of the movable swash plate 22a. The discharge amount of the hydraulic oil 22 is controlled.
The second pump flow rate control actuator 24 is connected to the oil passage 22b via the oil passage 24a. The second pump flow rate control actuator 24 is connected to the boom merging pressure compensation valve 32, the arm cylinder pressure compensation valve 72, and the swing motor pressure compensation valve 73 via the oil passage 24b.

  Below, the operation | movement aspect of the 1st pump flow control actuator 23 and the 2nd pump flow control actuator 24 is demonstrated using FIG.3, FIG4 and FIG.6.

  As shown in FIG. 3, the discharge pressure of the first hydraulic pump 21 is applied to the first pump flow control actuator 23 via the oil passage 21b and the oil passage 23a. The first pump flow control actuator 23 is given the maximum load pressure (first maximum load pressure) among the load pressures applied to the boom cylinder 13 and the bucket cylinder 15 through the oil passage 23b. The first pump flow control actuator 23 holds the differential pressure between the discharge pressure of the first hydraulic pump 21 and the first maximum load pressure at a predetermined value (a value determined by a spring provided in the first pump flow control actuator 23). Thus, the inclination angle of the movable swash plate 21a of the first hydraulic pump 21 is controlled.

  As shown in FIG. 4, when the boom junction valve 31 is in the position 31X, the discharge pressure of the second hydraulic pump 22 is applied to the second pump flow rate control actuator 24 via the oil passage 22b and the oil passage 24a. . The second pump flow control actuator 24 has a maximum load pressure (hereinafter simply referred to as “second maximum load pressure”) among the load pressures applied to the arm cylinder 14 and the swing motor 7 through the oil passage 24b. Is granted. The second pump flow control actuator 24 holds the differential pressure between the discharge pressure of the second hydraulic pump 22 and the second maximum load pressure at a predetermined value (a value determined by a spring provided in the second pump flow control actuator 24). Thus, the angle of the movable swash plate 22a of the second hydraulic pump 22 is controlled.

  When the boom junction valve 31 is in the position 31Y (see FIGS. 4 and 6), the discharge pressure of the second hydraulic pump 22 is applied to the second pump flow rate control actuator 24 via the oil passage 22b and the oil passage 24a. The Further, the second pump flow control actuator 24 has a maximum load pressure (hereinafter simply referred to as “joining maximum load pressure”) applied to the arm cylinder 14, the swing motor 7, and the boom cylinder 13 via the oil passage 24 b. Is written). The second pump flow control actuator 24 holds the differential pressure between the discharge pressure of the second hydraulic pump 22 and the combined maximum load pressure at a predetermined value (a value determined by a spring provided in the second pump flow control actuator 24). In addition, the angle of the movable swash plate 22a of the second hydraulic pump 22 is controlled.

As described above, the first pump flow control actuator 23 can maintain the differential pressure between the first maximum load pressure and the discharge pressure of the first hydraulic pump 21 at a predetermined value. Further, when the boom junction valve 31 is in the position 31X, the second pump flow control actuator 24 can maintain the differential pressure between the second maximum load pressure and the discharge pressure of the second hydraulic pump 22 at a predetermined value. When the boom junction valve 31 is in the position 31Y, the second pump flow control actuator 24 can maintain the differential pressure between the maximum combined load pressure and the discharge pressure of the second hydraulic pump 22 at a predetermined value. As a result, the discharge pressure and discharge amount of the hydraulic fluid by the first hydraulic pump 21 and the second hydraulic pump 22 are controlled to optimum values according to the work state (work load magnitude) of the work device 4.
Further, the differential pressure before and after the throttle provided in each of the work direction switching valves is compensated to a predetermined value by the after orifice type load sensing system.
Accordingly, the flow rate of the hydraulic oil supplied to the working hydraulic actuator depends only on the spool stroke amount of the working direction switching valve (the opening area of the flow path formed in the spool of the working direction switching valve). That is, by controlling the pilot pressure applied to the work direction switching valve, the flow rate of the hydraulic oil supplied to the work hydraulic actuator can be controlled with high accuracy.

  In addition, although the 1st pump flow control actuator 23 and the 2nd pump flow control actuator 24 which concern on this embodiment demonstrated as a control piston provided with the spring, this invention is not limited to this. That is, the first pump flow control actuator 23 and the second pump flow control actuator 24 may be configured by a regulator valve and a control piston. The load pressure of the working hydraulic actuator, the first hydraulic pump 21 and the second hydraulic pressure Any configuration that can hold the differential pressure from the discharge pressure of the pump 22 at a predetermined value may be used.

As shown in FIGS. 2 to 4, the first remote control valve 81 includes a pilot port 42a of the boom cylinder direction switching valve 42 through an oil passage 81a and a boom cylinder direction switching valve 42 through an oil passage 81b. Are connected to the pilot port 42b and the pilot port 31a of the boom junction valve 31, respectively.
The first remote control valve 81 is connected to the pilot port 43a of the bucket cylinder direction switching valve 43 via the oil passage 81c and to the pilot port 43b of the bucket cylinder direction switching valve 43 via the oil passage 81d. The
The first remote control valve 81 uses hydraulic oil supplied from a pilot pump (not shown) as pilot hydraulic oil, a boom cylinder direction switching valve 42 (specifically, a pilot port 42a or a pilot port 42b), and a bucket cylinder direction. The switching valve 43 (specifically, the pilot port 43a or the pilot port 43b) and the boom junction valve 31 (specifically, the pilot port 31a) are distributed.

  The first remote control valve 81 is interlocked and connected to a first operation lever 82 as an operation tool arranged in the control unit 8. By operating the first operating lever 82, the first remote control valve 81 is switched, and the direction of the hydraulic oil supplied to the boom cylinder direction switching valve 42, the bucket cylinder direction switching valve 43, and the boom junction valve 31 is switched. At the same time, the pilot pressure can be adjusted according to the operation amount of the first operation lever 82.

2 and 4, the second remote control valve 91 is connected to the pilot port 62a of the arm cylinder direction switching valve 62 and the pilot port 31b of the boom merging valve 31 via the oil passage 91a and the oil passage 91b. Are connected to the pilot port 62b of the arm cylinder direction switching valve 62, respectively.
The second remote control valve 91 is connected to the pilot port 63a of the turning motor direction switching valve 63 via the oil passage 91c and to the pilot port 63b of the turning motor direction switching valve 63 via the oil passage 91d. The
The second remote control valve 91 uses hydraulic oil supplied from a pilot pump (not shown) as pilot hydraulic oil, a direction switching valve for arm cylinder 62 (specifically, pilot port 62a or pilot port 62b), and direction for the swing motor. The switching valve 63 (specifically, the pilot port 63a or the pilot port 63b) and the boom junction valve 31 (specifically, the pilot port 31b) are distributed.

  The second remote control valve 91 is interlocked and connected to a second operation lever 92 as an operation tool arranged in the control unit 8. By operating the second operation lever 92, the second remote control valve 91 is switched to switch the direction of the hydraulic oil supplied to the arm cylinder direction switching valve 62, the swing motor direction switching valve 63, and the boom junction valve 31. At the same time, the pilot pressure can be adjusted according to the operation amount of the second operation lever 92.

  In the present embodiment, the first remote control valve 81 is the boom cylinder direction switching valve 42 and the bucket cylinder direction switching valve 43, and the second remote control valve 91 is the arm cylinder direction switching valve 62 and the swing motor direction switching. Although the valves 63 are connected to the valves 63, the present invention is not limited thereto. That is, the combination of the first remote control valve 81 and the second remote control valve 91 and the working direction switching valve connected to the first remote control valve 81 and the second remote control valve 91 is not particularly limited.

  As described above, a hydraulic pump having a maximum discharge flow rate smaller than that of the second hydraulic pump 22 is selected as the first hydraulic pump 21.

More specifically, as the first hydraulic pump 21, the operation required when the bucket cylinder 15 is extended, that is, when the bucket 12 is operated (bucket cloud operation) so that the other end side of the bucket 12 is close to the arm 11. A hydraulic pump having a maximum discharge flow rate comparable to the oil flow rate is selected.
As the second hydraulic pump 22, the arm cylinder 14 is extended, that is, approximately equal to the flow rate of hydraulic oil required when the arm 11 is operated (pulling) so that the other end of the arm 11 is close to the boom 10. The hydraulic pump having the maximum discharge flow rate is selected.
In general, the flow rate of hydraulic oil required for the arm cylinder 14 when the arm 11 is pulled is sent to the other working hydraulic actuators (the boom cylinder 13, the bucket cylinder 15, and the swing motor 7). More than the required flow rate of the hydraulic oil (the required flow rate of the hydraulic oil in the arm cylinder 14 is the maximum among the working hydraulic actuators). That is, the flow rate of hydraulic oil required for the arm cylinder 14 when performing the pulling operation of the arm 11 is larger than the flow rate of hydraulic oil required for the bucket cylinder 15 when performing the bucket cloud operation. Therefore, the maximum discharge flow rate of the first hydraulic pump 21 is smaller than the maximum discharge flow rate of the second hydraulic pump 22.

Thus, the maximum discharge flow rate of the first hydraulic pump 21 is based on the working hydraulic actuator of the first actuator group 18, and the maximum discharge flow rate of the second hydraulic pump 22 is based on the working hydraulic actuator of the second actuator group 19. By determining each of these, the hydraulic oil is appropriately supplied from the first hydraulic pump 21 and the second hydraulic pump 22 to the working hydraulic actuator while setting the first hydraulic pump 21 and the second hydraulic pump 22 to the minimum necessary capacity. Can only be supplied. Therefore, energy efficiency can be improved as compared with the case of using one hydraulic pump or the case of using hydraulic oil from two hydraulic pumps with small capacities.
Further, since a hydraulic pump having a discharge flow rate larger than necessary is not mounted, it is possible to prevent the first hydraulic pump 21 and the second hydraulic pump 22 from being enlarged, and the hydraulic circuit 201 can be connected to the turning work vehicle 1. It is possible to improve mountability and reduce costs.

  Below, the operation | movement aspect of the hydraulic circuit 201 comprised as mentioned above is demonstrated using FIG.6 and FIG.7. First, an operation mode of the hydraulic circuit 201 when the boom cylinder 13 is extended to rotate the boom 10 upward will be described.

  As shown in FIG. 6, when the first operating lever 82 is operated, the first operating lever is connected to the pilot port 42b of the boom cylinder direction switching valve 42 and the pilot port 31a of the boom junction valve 31 via the oil passage 81b. A pilot pressure corresponding to the operation amount 82 is applied. When the pilot pressure is applied, the boom cylinder direction switching valve 42 and the boom junction valve 31 are simultaneously slid toward the position 42Z and the position 31Y by the spool stroke amount S corresponding to the pilot pressure. .

  In FIG. 7, the boom cylinder direction switching valve 42 and the boom merging valve 31 spool stroke amount S, and the boom cylinder direction switching valve 42 and the boom merging valve 31 are switched to position 42Z and position 31Y. The relationship with the opening area A of the flow path formed in each spool of the direction switching valve 42 and the boom junction valve 31 is shown.

When the spool stroke amount S of the boom cylinder direction switching valve 42 and the boom junction valve 31 increases, when the spool stroke amount S becomes S1, first, the opening area A of the boom cylinder direction switching valve 42 (the broken line in FIG. 7). B) starts to increase. As the spool stroke amount S of the boom cylinder direction switching valve 42 and the boom junction valve 31 increases, the opening area A of the boom cylinder direction switching valve 42 increases.
Further, when the spool stroke amount S of the boom cylinder direction switching valve 42 and the boom junction valve 31 is increased, the opening area A of the boom junction valve 31 (see the one-dot chain line C in FIG. 7) when the spool stroke amount S becomes S2. ) Begins to increase. Thus, the shapes of the spools are determined so that the spool stroke amount S at which the opening area A starts to increase is different between the boom cylinder direction switching valve 42 and the boom junction valve 31. As the spool stroke amount S of the boom cylinder direction switching valve 42 and the boom junction valve 31 increases, the opening area A of the boom cylinder direction switching valve 42 and the boom junction valve 31 increases.
Further, when the spool stroke amount S of the boom cylinder direction switching valve 42 and the boom junction valve 31 is increased, the opening area A of the boom cylinder direction switching valve 42 and the boom junction valve 31 is reached when the spool stroke amount S becomes S3. Respectively become the maximum (A1 and A2).

When the boom cylinder direction switching valve 42 opens (the spool stroke amount S becomes S1 and the opening area A becomes larger than 0), the hydraulic oil is supplied from the first hydraulic pump 21 to the boom cylinder direction switching valve 42, And is supplied to the bottom chamber of the boom cylinder 13 via the oil passage 13b.
Further, when the boom junction valve 31 is opened (the spool stroke amount S is S2 and the opening area A is larger than 0), the hydraulic oil is supplied from the second hydraulic pump 22 to the boom junction valve 31, the oil passage 31e, And is supplied to the bottom chamber of the boom cylinder 13 via the oil passage 13b. That is, the hydraulic oil from the second hydraulic pump 22 is merged with the hydraulic oil from the first hydraulic pump 21 in the oil passage 13 b and supplied to the bottom chamber of the boom cylinder 13.

  In this case, the differential pressure across the boom cylinder direction switching valve 42 is compensated to a predetermined value by the load sensing system. Therefore, the flow rate of the hydraulic oil flowing through the boom cylinder direction switching valve 42 depends only on the opening area A of the boom cylinder direction switching valve 42. Further, the differential pressure across the boom junction valve 31 is compensated to a predetermined value by the load sensing system. Therefore, the flow rate of the hydraulic oil flowing through the boom junction valve 31 depends only on the opening area A of the boom junction valve 31. That is, an amount of hydraulic oil corresponding to the opening area A of the boom cylinder direction switching valve 42 and the boom junction valve 31 can be supplied to the bottom chamber of the boom cylinder 13.

  As described above, when the boom cylinder 13 is driven (extended), the hydraulic oil from the second hydraulic pump 22 can be supplied to the boom cylinder 13 in addition to the hydraulic oil from the first hydraulic pump 21. Thereby, the amount of hydraulic oil supplied to the boom cylinder 13 can be increased, the boom cylinder 13 can be operated quickly, and the working efficiency can be improved.

Further, the maximum discharge flow rate of the first hydraulic pump 21 is determined based on the flow rate of hydraulic oil required for the bucket cylinder 15 when the bucket 12 is operated in the bucket cloud operation, but the boom 10 is rotated upward. At this time, the flow rate of the hydraulic oil required for the boom cylinder 13 is generally larger than the flow rate of the hydraulic oil required for the bucket cylinder 15. In other words, only the hydraulic oil supplied from the first hydraulic pump 21 is insufficient for the hydraulic oil supplied to the boom cylinder 13, and the operating speed of the boom cylinder 13 may be reduced.
However, as described above, when the boom cylinder 13 is driven (stretched), in addition to the hydraulic oil from the first hydraulic pump 21, the hydraulic oil from the second hydraulic pump 22 can be supplied to the boom cylinder 13. . Thereby, it is possible to prevent the hydraulic oil supplied to the boom cylinder 13 from being insufficient, to operate the boom cylinder 13 at a sufficient speed, and to improve work efficiency.

  The flow rate of the hydraulic fluid supplied from the second hydraulic pump 22 to the boom cylinder 13 can be set to a desired flow rate by arbitrarily setting the opening area A of the boom junction valve 31.

  Further, the flow rate of the hydraulic oil supplied to the bottom chamber of the boom cylinder 13 is determined according to the sum of the opening areas A of the boom cylinder direction switching valve 42 and the boom junction valve 31.

  That is, by operating the first operating lever 82, the opening area A of the boom merging valve 31 is 0 when the spool stroke amount S is from S1 to S2. The boom cylinder 13 is extended only with the hydraulic oil supplied through the switching valve 42. In this case, the boom cylinder 13 is driven (extended) at a speed corresponding to the flow rate of the hydraulic oil supplied to the boom cylinder 13, that is, the opening area A of the boom cylinder direction switching valve 42.

  When the spool stroke amount S is S2 or more, the hydraulic oil supplied from the first hydraulic pump 21 via the boom cylinder direction switching valve 42 and the hydraulic oil supplied from the second hydraulic pump 22 via the boom junction valve 31 Thus, the boom cylinder 13 is extended. In this case, the boom cylinder 13 corresponds to the flow rate of the hydraulic oil supplied to the boom cylinder 13, that is, the sum of the opening areas A of the boom cylinder direction switching valve 42 and the boom junction valve 31 (see the solid line D in FIG. 7). Drive at speed (stretch).

As described above, until the spool stroke amount S reaches S2, that is, as long as the operation amount of the first operation lever 82 is small after the operation starts, the boom cylinder 13 is moved only by the hydraulic oil supplied from the first hydraulic pump 21. Can be driven. That is, since the hydraulic oil is not supplied from the second hydraulic pump 22 to the boom cylinder 13, the boom cylinder 13 can be driven at a low speed. Therefore, after the operation of the first operation lever 82 is started, the boom 10 can be precisely operated without rapidly operating, and the fine operability of the work device 4 can be improved.
Further, when the spool stroke amount S is greater than or equal to S2, that is, when the operation amount of the first operation lever 82 becomes greater than a predetermined operation amount, the hydraulic oil supplied from the first hydraulic pump 21 and the second hydraulic pump 22 is increased. Thus, the boom cylinder 13 can be driven. That is, the boom cylinder 13 can be driven at a high speed. Therefore, the boom cylinder 13 can be operated at a sufficient speed, and the working efficiency can be improved.

  Next, an operation mode of the hydraulic circuit 201 when the arm cylinder 14 is extended and the arm 11 is operated (pulling operation) when the boom cylinder 13 is extended and the boom 10 is rotated upward will be described.

  As described above, when the first operation lever 82 is operated by a predetermined operation amount or more, the boom junction valve 31 is switched to the position 31Y. Thereby, in addition to the hydraulic oil from the first hydraulic pump 21, the hydraulic oil from the second hydraulic pump 22 can be supplied to the boom cylinder 13. The boom cylinder 13 is extended by the supplied hydraulic oil, and the boom 10 can be rotated upward.

  When the second operation lever 92 is operated in a state in which the first operation lever 82 is operated by a predetermined operation amount or more, the pilot port 62a of the arm cylinder direction switching valve 62 and the boom merge are connected via the oil passage 91a. A pilot pressure corresponding to the operation amount of the second operation lever 92 is applied to the pilot port 31 b of the valve 31.

  When a pilot pressure is applied to the pilot port 62a of the arm cylinder direction switching valve 62, the spool of the arm cylinder direction switching valve 62 is moved from the neutral position to another position (arm cylinder by the spool stroke amount corresponding to the pilot pressure). 14 to the position of supplying hydraulic oil to the bottom chamber 14).

  Further, when a pilot pressure is applied to the pilot port 31b of the boom junction valve 31, the pilot pressure counters the pilot pressure applied to the pilot port 31a, and the spool of the boom junction valve 31 slides toward the position 31X. Moved. Accordingly, the flow of the hydraulic oil from the oil passage 22b to the oil passage 31e is restricted by the boom junction valve 31, and the supply amount of the hydraulic oil supplied from the second hydraulic pump 22 to the boom cylinder 13 is restricted. Further, when the spool of the boom junction valve 31 is completely switched to the position 31X, the flow of hydraulic oil from the oil passage 22b to the oil passage 31e is blocked by the boom junction valve 31, and the hydraulic oil from the second hydraulic pump 22 is shut off. Is not supplied to the boom cylinder 13.

  As described above, when the boom cylinder 13 is extended and the boom 10 is rotated upward, when the arm cylinder 14 is extended and the arm 11 is operated (pulling), the second hydraulic pump 22 is moved to the boom cylinder 13. Limit the supply of hydraulic oil. As a result, the flow rate of the hydraulic oil supplied from the second hydraulic pump 22 to the arm cylinder 14 can be secured, and a decrease in the operating speed of the arm cylinder 14 can be prevented.

Further, with such a configuration, it is possible to prevent a decrease in work efficiency, particularly when the arm 11 is operated (pulling operation) while the boom 10 is raised.
As the work using the work device 4, the arm 11 is pulled while the boom 10 is raised, and the bucket 12 is moved in parallel with the ground in a state where the tip of the bucket 12 is in light contact with the ground (so-called horizontal Leveling work). When performing the water leveling operation, it is necessary to raise the boom 10 little by little with respect to the pulling operation of the arm 11.
In the hydraulic circuit 201 according to the present embodiment, when water averaging is performed, the supply of hydraulic oil from the second hydraulic pump 22 to the boom cylinder 13 is restricted as described above. Therefore, the operating speed of the boom 10 does not increase more than necessary, and the amount of operating oil supplied to the arm cylinder 14 does not become insufficient and the operating speed of the arm 11 does not decrease. As described above, the hydraulic circuit 201 according to the present embodiment can average the water and operate the boom 10 and the arm 11 at an appropriate speed, thereby preventing a reduction in work efficiency.

Furthermore, when the boom cylinder 13 is extended and the boom 10 is rotated upward, and the arm cylinder 14 is extended and the arm 11 is operated (pulling operation), the boom merging valve 31 causes the boom cylinder 13 to be moved from the second hydraulic pump 22. It is assumed that hydraulic oil will be supplied to In this case, when there is a large difference between the loads applied to the boom cylinder 13 and the arm cylinder 14, the throttle loss in the boom merging pressure compensation valve 32 increases and the energy loss increases.
However, in the hydraulic circuit 201 according to the present embodiment, when the boom cylinder 13 is extended and the boom 10 is rotated upward, and the arm cylinder 14 is extended and the arm 11 is operated (pulling), the boom junction valve is used. 31 restricts the flow of the hydraulic oil from the oil passage 22b to the oil passage 31e. Therefore, even when a large difference occurs between the loads applied to the boom cylinder 13 and the arm cylinder 14, no throttle loss occurs in the boom merging pressure compensation valve 32, and a reduction in energy efficiency can be prevented.

As described above, the hydraulic circuit 201 of the turning working vehicle 1 according to the present embodiment supplies the first hydraulic oil to the first actuator group 18 including at least one working hydraulic actuator (the boom cylinder 13 and the bucket cylinder 15). The discharge amount of the hydraulic pump 21 is controlled according to the maximum load pressure among the load pressures applied to the first actuator group 18 and includes at least one working hydraulic actuator (the arm cylinder 14 and the swing motor 7). The turning work vehicle 1 having a load sensing system that controls the discharge amount of the second hydraulic pump 22 that supplies hydraulic oil to the actuator group 19 according to the maximum load pressure among the load pressures applied to the second actuator group 19. In the hydraulic circuit 201, the second actuator group 19 is configured so that the required flow rate of hydraulic oil is the first flow rate. It includes the working hydraulic actuator (arm cylinder 14) which is the largest among the tutor group 18 and the second actuator group 19, and the maximum discharge flow rate of the second hydraulic pump 22 is larger than the maximum discharge flow rate of the first hydraulic pump 21. It is to set.
With this configuration, the maximum discharge flow rates of the first hydraulic pump 21 and the second hydraulic pump 22 are selected according to the required flow rates of the working hydraulic actuators included in the first actuator group 18 and the second actuator group 19. Thus, the hydraulic oil can be appropriately supplied to each working hydraulic actuator. Therefore, energy efficiency can be improved as compared with the case of using one hydraulic pump or the case of using hydraulic oil from two hydraulic pumps with small capacities.
In addition, since a hydraulic pump having a discharge flow rate larger than necessary is not mounted, it is possible to prevent the first hydraulic pump 21 and the second hydraulic pump 22 from being enlarged, and to improve mounting performance and cost. be able to.

Further, the hydraulic circuit 201 of the turning work vehicle 1 of the present embodiment is configured so that the hydraulic oil is supplied from the first hydraulic pump 21 to the specific working hydraulic actuator (boom cylinder 13) in the first actuator group 18 when the hydraulic oil is supplied. A merging valve (boom merging valve 31) for merging the hydraulic oil discharged from the two hydraulic pumps 22 with the hydraulic oil supplied from the first hydraulic pump 21 to the boom cylinder 13 is provided.
With this configuration, when the boom cylinder 13 is operated, the hydraulic oil discharged from the first hydraulic pump 21 and the second hydraulic pump 22 is merged, so that the boom cylinder 13 is obtained only by the first hydraulic pump 21. It can be operated more quickly than when operating. Therefore, working efficiency can be improved. Further, even if only the first hydraulic pump 21 having a small maximum discharge flow rate cannot supply sufficient hydraulic oil to the required flow rate of the boom cylinder 13, the boom cylinder can be obtained by merging the hydraulic oil from the second hydraulic pump 22. 13 can be driven at a sufficient speed.

Further, the hydraulic circuit 201 of the turning work vehicle 1 of the present embodiment is provided corresponding to a plurality of work hydraulic actuators, and a plurality of work direction switches for switching the direction of hydraulic oil supplied to the work hydraulic actuators. And the boom junction valve 31 is provided when the spool stroke amount of the working direction switching valve (boom cylinder direction switching valve 42) for supplying hydraulic oil to the boom cylinder 13 becomes equal to or greater than a predetermined value S2. The hydraulic oil discharged from the second hydraulic pump 22 is merged with the hydraulic oil supplied from the first hydraulic pump 21 to the boom cylinder 13.
With this configuration, when the spool stroke amount of the boom cylinder direction switching valve 42 is less than the predetermined value S2, the boom cylinder 13 can be operated only with the hydraulic oil supplied from the first hydraulic pump 21. it can. Thereby, when the spool stroke amount is less than the predetermined value S2, the boom cylinder 13 can be operated precisely.
When the spool stroke amount of the boom cylinder direction switching valve 42 is equal to or greater than the predetermined value S2, the hydraulic oil discharged from the two pumps can be merged, and the boom cylinder 13 can be operated quickly, thereby improving the work efficiency. Can be improved.

In addition, the turning work vehicle 1 of the present embodiment is attached to the arm 10, a boom 10 that is rotatably attached to the vehicle body (the turning device 3), an arm 11 that is rotatably attached to the boom 10, and the arm 11. And the specific working hydraulic actuator is a boom cylinder 13 that rotates the boom 10 with respect to the turning device 3, and is one of the second actuator group 19. The working hydraulic actuator is an arm cylinder 14 that rotates the arm 11 with respect to the boom 10, and the boom junction valve 31 is a first hydraulic pressure when hydraulic oil is supplied from the second hydraulic pump 22 to the arm cylinder 14. The amount of hydraulic oil supplied from the second hydraulic pump 22 joined to the hydraulic oil supplied from the pump 21 to the boom cylinder 13 is limited.
With this configuration, when the boom cylinder 13 and the arm cylinder 14 are operated at the same time, the supply of hydraulic oil from the second hydraulic pump 22 to the boom cylinder 13 is restricted, so that the operating speed of the arm cylinder 14 can be reduced. A decrease can be prevented. Thereby, it is possible to improve the workability at the time of pulling the arm 11 while raising the boom 10 (so-called water averaging work).

  Below, the hydraulic circuit 202 which concerns on 2nd embodiment is demonstrated using FIG.

The hydraulic circuit 202 according to the second embodiment is different from the hydraulic circuit 201 according to the first embodiment (see FIG. 2) in that the first hydraulic pump 121 and the second hydraulic pump 22 are replaced with the first hydraulic pump 121 and the second hydraulic pump 22. The second hydraulic pump 122 is provided with an arm merging valve 131 and an arm merging pressure compensating valve 132 instead of the boom merging valve 31 and the boom merging pressure compensating valve 32, respectively.
Therefore, hereinafter, only differences from the hydraulic circuit 201 according to the first embodiment will be described, and members having substantially the same configuration as the hydraulic circuit 201 will be denoted by the same reference numerals, and description thereof will be omitted.

The first hydraulic pump 121 and the second hydraulic pump 122 are driven by the engine 9 to discharge hydraulic oil. The first hydraulic pump 121 and the second hydraulic pump 122 are variable displacement pumps that can change the discharge amount of hydraulic oil by changing the inclination angle of the movable swash plate.
The hydraulic oil discharged from the first hydraulic pump 121 and the second hydraulic pump 122 is supplied to the control valve 30. More specifically, the hydraulic oil discharged from the first hydraulic pump 121 is supplied to the first direction switching valve group 40 and the arm merging valve 131. Further, the hydraulic oil discharged from the second hydraulic pump 22 is supplied to the second direction switching valve group 60.

  As the first hydraulic pump 121, a hydraulic pump having a maximum discharge flow rate larger than that of the second hydraulic pump 122 is selected.

More specifically, as the first hydraulic pump 121, a hydraulic pump that drives the boom cylinder 13, that is, a hydraulic pump having a maximum discharge flow rate that is approximately the same as the flow rate of hydraulic oil required when operating the boom 10 is selected. .
As the second hydraulic pump 122, a hydraulic pump that drives the turning motor 7, that is, a hydraulic pump having a maximum discharge flow rate that is similar to the flow rate of hydraulic oil required when the swivel base 6 is swung with respect to the traveling device 2 is selected. Is done.
In general, the flow rate of hydraulic oil required for the boom cylinder 13 when operating the boom 10 is larger than the flow rate of hydraulic oil required for the swing motor 7 when operating the swivel base 6. Therefore, the maximum discharge flow rate of the first hydraulic pump 121 is larger than the maximum discharge flow rate of the second hydraulic pump 122. In the present embodiment, the flow rate of hydraulic oil required for the boom cylinder 13 when operating the boom 10 is required for the other working hydraulic actuators (the arm cylinder 14, the bucket cylinder 15, and the turning motor 7). It is assumed that the flow rate is higher than the flow rate of the hydraulic oil (the flow rate required for the hydraulic oil in the boom cylinder 13 is the maximum among the working hydraulic actuators).

Thus, the maximum discharge flow rate of the first hydraulic pump 121 is based on the working hydraulic actuator of the first actuator group 18, and the maximum discharge flow rate of the second hydraulic pump 122 is based on the working hydraulic actuator of the second actuator group 19. As a result, the hydraulic oil can be supplied from the first hydraulic pump 121 and the second hydraulic pump 122 to the working hydraulic actuator by an appropriate amount. Therefore, energy efficiency can be improved as compared with the case where the maximum discharge flow rates of the first hydraulic pump 121 and the second hydraulic pump 122 are determined to be the same.
In addition, since a hydraulic pump having a discharge flow rate larger than necessary is not mounted, it is possible to prevent the first hydraulic pump 121 and the second hydraulic pump 122 from being enlarged, and to improve mounting performance and reduce costs. be able to.

The arm merging valve 131 is a pilot type capable of merging the hydraulic oil discharged from the first hydraulic pump 121 with the hydraulic oil discharged from the second hydraulic pump 122 and supplied to the bottom chamber of the arm cylinder 14. This is a direction switching valve.
An arm merging pressure compensation valve 132 is connected to the arm merging valve 131. The arm merging pressure compensation valve 132 compensates the pressure after the throttling provided in the arm merging valve 131 to a predetermined value.

  When the second operating lever 92 is operated, the pilot port 62a of the arm cylinder directional switching valve 62 and the pilot port 131a of the arm merging valve 131 are set in accordance with the operation amount of the second operating lever 92 through the oil passage 91a. Pilot pressure is applied. By the pilot pressure, the arm cylinder direction switching valve 62 is switched from the neutral position to another position (position for supplying hydraulic oil to the bottom chamber of the arm cylinder 14), and the arm merging valve 131 is switched to the position 131Y. As a result, the hydraulic oil from the first hydraulic pump 121 is supplied to the bottom chamber of the arm cylinder 14 via the arm cylinder direction switching valve 62, and the hydraulic oil from the second hydraulic pump 122 passes through the arm merging valve 131. To the bottom chamber of the arm cylinder 14.

  In this case, when the first operation lever 82 is operated, the first operation lever 82 is operated to the pilot port 42b of the boom cylinder direction switching valve 42 and the pilot port 131b of the arm junction valve 131 via the oil passage 81b. A pilot pressure corresponding to the amount is applied. With the pilot pressure, the boom cylinder direction switching valve 42 is switched to the position 42Z, and the arm junction valve 131 is switched to the position 131X. As a result, when the arm cylinder 14 is extended and the arm 11 is operated (pulling operation), when the boom cylinder 13 is extended and the boom 10 is rotated upward, the first hydraulic pump 121 moves to the arm cylinder 14. Stop supplying hydraulic fluid. Therefore, the flow rate of the hydraulic oil supplied from the first hydraulic pump 121 to the boom cylinder 13 can be ensured, and the operating speed of the boom cylinder 13 can be prevented from decreasing.

7 Swing motor (working hydraulic actuator)
13 Boom cylinder (working hydraulic actuator)
14 Arm cylinder (working hydraulic actuator)
15 Bucket cylinder (working hydraulic actuator)
21 First hydraulic pump 22 Second hydraulic pump 31 Boom merging valve (merging valve)
42 Boom cylinder direction switching valve (working direction switching valve)
43 Bucket cylinder direction switching valve (working direction switching valve)
62 Directional switching valve for arm cylinder (working direction switching valve)
63 Directional switching valve for slewing motor (working direction switching valve)
201 Hydraulic circuit

Claims (4)

Controlling the discharge amount of the first hydraulic pump that supplies the hydraulic oil to the first actuator group including at least one working hydraulic actuator according to the maximum load pressure of the load pressure applied to the first actuator group;
A load sensing system that controls a discharge amount of a second hydraulic pump that supplies hydraulic oil to a second actuator group including at least one working hydraulic actuator according to a maximum load pressure among load pressures applied to the second actuator group A hydraulic circuit for a work vehicle comprising:
The second actuator group includes:
Including a working hydraulic actuator in which the required flow rate of the hydraulic oil is maximum in the first actuator group and the second actuator group;
The maximum discharge flow rate of the second hydraulic pump is set larger than the maximum discharge flow rate of the first hydraulic pump;
Hydraulic circuit of work vehicle.   When hydraulic fluid is supplied from the first hydraulic pump to a specific working hydraulic actuator in the first actuator group, the hydraulic fluid discharged from the second hydraulic pump is transferred from the first hydraulic pump to the specific operation. The hydraulic circuit for a work vehicle according to claim 1, further comprising a merging valve for merging with hydraulic oil supplied to the hydraulic actuator. A plurality of working direction switching valves provided corresponding to the plurality of working hydraulic actuators, for switching the direction of hydraulic oil supplied to the working hydraulic actuator,
The junction valve is
When the spool stroke amount of the working direction switching valve that supplies the working oil to the specific working hydraulic actuator exceeds a predetermined value, the working oil discharged from the second hydraulic pump is supplied to the first hydraulic pump. The hydraulic circuit for the work vehicle according to claim 2, wherein the hydraulic oil is supplied to the hydraulic oil supplied to the specific work hydraulic actuator from the engine. The work vehicle includes a work device having a boom rotatably attached to a vehicle body, an arm rotatably attached to the boom, and a bucket attached to the arm.
The specific working hydraulic actuator is a boom cylinder that rotates the boom with respect to the vehicle body,
One working hydraulic actuator of the second actuator group is an arm cylinder that rotates the arm with respect to the boom;
The junction valve is
When the hydraulic oil is supplied from the second hydraulic pump to the arm cylinder, the hydraulic oil from the second hydraulic pump merges with the hydraulic oil supplied from the first hydraulic pump to the specific working hydraulic actuator. The hydraulic circuit for a work vehicle according to claim 3, wherein the supply amount of the work vehicle is limited.

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