JP2001029834A - Hydraulic driving device for self-travelling crusher - Google Patents

Abstract

PROBLEM TO BE SOLVED: To sufficiently enhance the quality of a crushed product and to improve the productivity by sufficiently stabilizing the revolution speed of a hydraulic motor for crushing regardless of increase/decrease of the load imposed on the hydraulic motor. SOLUTION: This device is provided with first and second hydraulic pumps 19, 20, hydraulic motors for crushing and a feeder, control valves for crushing and the feeder, a crusher speed setting dial 36c for setting the target discharge flow rate of the first hydraulic pump 19, a pressure sensor 105 for detecting the discharge pressure P1 of the pump 19, a first servo valve 95 for controlling the discharge flow rate Q1 of the pump 19 according to the target discharge flow rate and a second servo valve 97 for limiting input horsepower to the pump 19 based on P1 blow 1/2 of the effective output horsepower of an engine 21. The control valve for the feeder is controlled so that the output horsepower of the pump 19 based on P1 does not exceed 1/2 of the effective output horsepower of the engine 21.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a self-propelled crusher having a crushing device for crushing a crushed raw material, such as a jaw crusher, a roll crusher, a shredder, etc. Regardless, the present invention relates to a hydraulic drive device for a self-propelled crusher, which can sufficiently stabilize its rotation speed, sufficiently improve the quality of a crushed product, and improve productivity.

[0002]

2. Description of the Related Art Crushers transport various kinds of large and small rocks, construction waste materials, industrial wastes, etc. generated at construction sites such as concrete lumps carried out when dismantling buildings and asphalt lumps discharged during road repairs. By crushing to a predetermined size at the work site before the work is performed, waste materials can be reused, work can be smoothly performed, and costs can be reduced.

[0003] Among such crushers, for example, a self-propelled crusher is a crushing device that crushes a crushed material input from a hopper into a predetermined size, with a traveling body having left and right crawler tracks. And an auxiliary machine for performing operations related to the crushing operation by the crushing device, for example, a feeder for guiding the crushed raw material supplied from the hopper to the crushing device, and a conveyor for transporting the crushed material crushed and reduced by the crushing device. And a magnetic separator provided above the conveyor and magnetically attracting and removing magnetic substances contained in the crushed material being conveyed on the conveyor.

[0004] In such a configuration, the crushed raw material charged into the hopper above the crusher is guided to the crusher by a feeder below the hopper, and is crushed to a predetermined size by the crusher. The crushed crushed material falls from a space below the crushing device onto a conveyor below the crushing device, and is conveyed by this conveyor. In the middle of this transportation, the magnetic separator placed above the conveyor adsorbs and removes, for example, rebar pieces mixed in the concrete lump, and is almost uniform in size and finally from the front or rear of the crusher. It is carried out.

At this time, the endless track, the crushing device, the feeder, the conveyor, and the magnetic separator are driven by corresponding hydraulically driven actuators. That is, the hydraulic drive device of the self-propelled crusher including these hydraulic actuators includes, for example, a plurality of (for example, two) variable displacement hydraulic pumps driven by one prime mover and the pressure discharged from these hydraulic pumps. A hydraulic motor for crushing and a hydraulic motor for auxiliary machines (for example, a hydraulic motor for a feeder, a hydraulic motor for a conveyor, and a hydraulic motor for a magnetic separator) driven by oil, respectively, for driving the crushing device and the auxiliary machine, and the hydraulic pump And a plurality of control valves for controlling the flow of hydraulic oil supplied from the hydraulic pump to the hydraulic motor, and pump control means for controlling the discharge flow rate of the hydraulic pump. Is supplied to each hydraulic motor via each control valve.

By the way, generally, in a crushing device,
The particle size of the crushed material is the rotation speed (number of rotations) of the crushing hydraulic motor
It is known that there is a characteristic that the particle size decreases as the rotation speed increases, and the particle size increases as the rotation speed decreases. Therefore, in order to obtain a high-quality crushed product having a uniform particle size, it is preferable to keep the rotation speed of the crushing hydraulic motor as constant as possible.

From such a viewpoint, Japanese Patent Application Laid-Open No.
As described in Japanese Patent Publication No. 25, a hydraulic source, a crushing hydraulic motor for driving a crushing device, a feeder hydraulic motor for driving a feeder, and the crushing hydraulic motor and the feeder hydraulic motor from the hydraulic source In a hydraulic drive of a self-propelled crusher having a crushing control valve and a feeder control valve for controlling the flow of pressurized oil to the crushing device, a rotation speed setting means for setting a rotation speed of the crushing device, and a crushing hydraulic pressure. A rotation speed detection unit for detecting a rotation speed of the motor; and a control unit for switching a feeder control valve based on the rotation speed set by the rotation speed setting unit and the rotation speed detected by the rotation speed detection unit. A configuration has been proposed.

[0008]

In the above prior art,
When the load on the crushing device increases and the rotation speed of the crushing hydraulic motor falls below the first predetermined value, the control means stops the feeder by setting the feeder control valve to the neutral position and stops supplying the crushing material to the crushing device. Then, when the rotation speed of the crushing hydraulic motor becomes equal to or higher than the second predetermined value, the control valve for the feeder is switched again to restart supply of the crushing raw material to the crushing device. As a result, the amount of the crushed material in the crushing device is always set to an appropriate amount, the load of the crushing device is maintained in an appropriate range, and the fluctuation of the rotation speed of the crushing hydraulic motor is reduced, thereby improving the particle size distribution of the crushed material. Maintain and improve product quality.

[0009] However, the above prior art has the following problems. That is, although not specifically described in the above-described conventional technology, in this case, as in a normal hydraulic drive of this type, a variable displacement hydraulic pump driven by a prime mover (eg, an engine) is used as a hydraulic source. Generally, a so-called horsepower control for controlling the discharge flow rate of the hydraulic pump by the pump control means so that the input horsepower of the hydraulic pump does not exceed the output horsepower of the prime mover is performed. In this horsepower control, when the discharge pressure P of the hydraulic pump increases based on the so-called PQ characteristic line, the maximum value Qmax of the discharge flow rate Q is limited to a small value, whereby the input horsepower of the hydraulic pump is limited to a predetermined value or less. You.

The PQ characteristic line defines the maximum value Qmax of the discharge flow rate Q of the hydraulic pump, and particularly on the high pressure side, represents the maximum discharge flow rate when the maximum horsepower of the prime mover is used. (That is, it is an isopower line). Here, in performing the crushing operation, the power of the prime mover is given a margin, and the prime mover is used with a horsepower slightly smaller than the maximum horsepower of the prime mover. Normally, the pump is controlled so that the pump discharge pressure and the pump discharge flow rate correspond to the region (side).

Here, when the load of the crushing hydraulic motor increases due to an increase in the supply of the crushing raw material to the crushing device or the like, the pump discharge pressure P increases accordingly, so that the operating point becomes high pressure. Side, but initially, the consumed horsepower is smaller than the maximum horsepower of the prime mover.
Since it is the area inside the horsepower line portion such as the -Q characteristic line, only the discharge pressure P increases while the discharge flow rate Q is constant. Then, when the load of the crushing hydraulic motor further increases and the operating point reaches the horsepower line portion such as the PQ characteristic line, the operating point thereafter moves downward to the right on the horsepower line as the pump discharge pressure P increases. As a result, the pump discharge flow rate Q decreases, so that the rotation speed NC of the crushing hydraulic motor decreases.

In this prior art, when the rotational speed NC of the crushing hydraulic motor is reduced to be equal to or less than the first predetermined value, an increase in the load of the crushing hydraulic motor is detected, and the crushing by the feeder is performed. Stop feeding raw materials. Then, the rotation speed NC of the crushing hydraulic motor is increased again, and when the rotation speed NC becomes equal to or higher than the second predetermined value, the load reduction of the crushing hydraulic motor is similarly detected, and the supply of the crushed raw material by the feeder is restarted. By repeating this, the rotation speed NC of the crushing hydraulic motor is maintained at a substantially constant value.

Thus, the rotation speed N of the crushing hydraulic motor
Since the control for correcting the increase and decrease of C is performed, the rotation speed NC of the crushing hydraulic motor actually fluctuates between a first predetermined value and a second predetermined value.
Therefore, the rotation speed NC cannot be sufficiently stabilized, and the particle size distribution of the crushed material varies to some extent, making it difficult to sufficiently improve the quality of the product. Therefore, in order to obtain a high quality product, it is necessary to sieve the crushed material using a sieving means or the like, and the productivity is reduced.

SUMMARY OF THE INVENTION An object of the present invention is to provide a self-propelled crusher capable of sufficiently stabilizing the rotational speed of a crushing hydraulic motor irrespective of an increase or decrease in load and sufficiently improving the quality of a crushed product and improving productivity. Another object of the present invention is to provide a hydraulic drive device.

[0015]

(1) In order to achieve the above object, the present invention provides a crushing device for crushing a crushed material input from a hopper, and a crushing device for crushing the crushed material input to the hopper. A first and a second driven by a prime mover provided in a self-propelled crusher having
A hydraulic pump, a crushing hydraulic motor and a feeder hydraulic motor that respectively drive the crushing device and the feeder with pressurized oil discharged from the first and second hydraulic pumps, and the first and second hydraulic pumps Crushing control valve means and feeder control valve means for controlling the flow of hydraulic oil supplied to the crushing hydraulic motor and the feeder hydraulic motor, respectively, and at least one of the first and second hydraulic pumps A first hydraulic pump is a variable displacement type hydraulic pump. In a hydraulic drive device of a self-propelled crusher, a flow rate setting means for setting a target discharge flow rate of the first hydraulic pump, and a discharge pressure of the first hydraulic pump. A first discharge pressure detecting means for detecting, a pump flow control means for controlling a discharge flow rate of the first hydraulic pump according to the set target discharge flow rate, The according to discharge pressure of the first hydraulic pump, the first to limit the input horsepower of the first hydraulic pump to below the standard horsepower associated with the output horsepower of the prime mover
A first pump horsepower limit control unit that controls a discharge flow rate of the hydraulic pump, and based on the detected discharge pressure of the first hydraulic pump, the output horsepower of the first hydraulic pump does not exceed the reference horsepower. Feeder control means for controlling the feeder control valve means.

During the crushing operation, the hydraulic oil discharged from the second hydraulic pump is supplied to the hydraulic motor for the feeder via the control valve means for the feeder, while the hydraulic oil discharged from the first hydraulic pump is controlled for the crushing. The crushing hydraulic motor is supplied through a valve means. As a result, the feeder and the crushing device operate, the crushed raw material supplied to the hopper is transported to the crushing device by the feeder, and the crushing device crushes the transported crushed raw material. Here, when the target discharge flow rate of the first hydraulic pump is set by the flow rate setting means, the discharge flow rate of the first hydraulic pump is controlled by the pump flow rate control means to a value corresponding to the set value, for example, the same value as the set value. Therefore, the crushing hydraulic motor rotates at a constant speed corresponding to the discharge flow rate of the first hydraulic pump, and the particle size of the crushed product is adjusted to a size corresponding to the constant speed.

At this time, the discharge flow rate of the first hydraulic pump is also controlled by the first pump horsepower restriction control means so that the input horsepower of the first hydraulic pump is equal to a reference horsepower related to the output horsepower of the prime mover, for example, one of the effective output horsepower of the prime mover. / 2 or less (so-called horsepower control).

Here, when the load of the crushing hydraulic motor increases due to, for example, a large supply of the crushing raw material from the feeder to the crushing device or a high compressive strength, the discharge pressure of the first hydraulic pump increases accordingly. Therefore, the operating point shifts to the high pressure side on the PQ characteristic diagram of the horsepower control. Usually, at this time, the operating point is controlled by the above-described flow rate control means so that the operating point is equal to the PQ characteristic line. The area is inside the horsepower line portion, and only the discharge pressure increases while the discharge flow rate of the first hydraulic pump remains constant under the control of the flow rate control means. That is, the operating point moves horizontally to the right in the area inside the equihorse power line of the PQ characteristic line.

Thereafter, when the load of the crushing hydraulic motor further increases, the operating point further moves to the high pressure side and reaches the equihorse power line portion of the PQ characteristic line. At this time, the first discharge pressure detecting means Based on the detected discharge pressure of the first hydraulic pump, the feeder control means ensures that the actual output horsepower of the first hydraulic pump does not exceed the reference horsepower, that is, the aforementioned horsepower line portion such as the PQ characteristic line (the reference horsepower). The control valve means for the feeder is controlled so that the operating point that has reached the upper position does not further move to the lower right in the equihorse power line portion. For example, the control valve means for the feeder is returned to the neutral position or the opening is reduced, whereby the supply of the pressure oil from the second hydraulic pump to the hydraulic motor for the feeder is stopped or decelerated. As a result, the feeder operation is stopped or delayed, and the input of the crushed raw material to the crushing device is stopped or reduced. Therefore, the load on the crushing hydraulic motor is reduced, the discharge pressure of the first hydraulic pump is reduced, and the operating point is set to the above-mentioned point. It separates from the horsepower line portion such as the PQ characteristic line and shifts to the low pressure side again to return.

(2) In the above (1), preferably,
The feeder control means switches the feeder control valve means to a neutral position or reduces the opening so that the output horsepower of the first hydraulic pump does not exceed the reference horsepower.

(3) In the above (2), more preferably, the feeder control means includes a reference for calculating a reference discharge pressure of the first hydraulic pump corresponding to the set value of the flow rate setting means and the reference horsepower. A signal for switching the feeder control valve means to a neutral position or outputting a signal for decreasing the opening in accordance with the discharge pressure calculation means and the reference discharge pressure and the detected discharge pressure of the first hydraulic pump. Output means.

(4) In the above (3), more preferably, the signal output means is configured to determine that a difference between the reference discharge pressure and the detected discharge pressure of the first hydraulic pump is equal to or less than a first predetermined value. The control valve means for the feeder is switched to the neutral position or the opening is reduced, and when the control valve returns to the second predetermined value or more, the opening of the control valve means for the feeder is returned to the set value.

(5) In the above (1), preferably, the flow rate setting means includes a rotation speed setting means for setting a rotation speed of the crushing hydraulic motor.

(6) In the above (1), preferably, the first pump horsepower limit control means uses approximately 1/2 of the effective output horsepower of the prime mover as the reference horsepower, and The discharge flow rate of the first hydraulic pump is controlled so as to limit the input horsepower to approximately 1/2 or less of the effective output horsepower of the prime mover.

(7) In the above (1), preferably, there is further provided a second discharge pressure detecting means for detecting a discharge pressure of the second hydraulic pump, and the first pump horsepower limit control means comprises: As the reference horsepower, a first distribution reference horsepower obtained by distributing an output horsepower of the prime mover to a first hydraulic pump side at a ratio according to the detected discharge pressures of the first and second hydraulic pumps is used. The discharge flow rate of the first hydraulic pump is controlled so that the input horsepower of the hydraulic pump is limited to the first distribution reference horsepower or less.

The first hydraulic pump for the crushing hydraulic motor having a relatively high load by performing full horsepower control for distributing the output horsepower of the prime mover at a ratio corresponding to the discharge pressures of the first and second hydraulic pumps. The power of the prime mover can be effectively distributed to the second hydraulic pump related to the feeder hydraulic motor, which has a relatively low load, in accordance with the load difference. That is, the surplus of the horsepower to the second hydraulic pump, which is surplus due to the small required horsepower of the feeder hydraulic motor, can be supplied to the first hydraulic pump for the crushing hydraulic motor having a large required horsepower. As described above, the horsepower of the prime mover can be effectively used, so that energy can be saved.

(8) In order to achieve the above object, the present invention also provides a crushing device for crushing crushed raw material supplied from a hopper, and a feeder for conveying the crushed raw material supplied to the hopper to the crushing device. A self-propelled type having a conveyor for carrying out the crushed material crushed by the crushing device, a magnetic separator for magnetically sucking and removing a magnetic material contained in the crushed material being conveyed on the conveyor, and running means. Installed in the crusher,
Variable capacity first and second hydraulic pumps driven by a prime mover, and the crushing device, the feeder, the conveyor, the magnetic separator, and the traveling by pressure oil discharged from the first and second hydraulic pumps A hydraulic motor for crushing, a hydraulic motor for a feeder, a hydraulic motor for a conveyor, a hydraulic motor for a magnetic separator, a hydraulic motor for left and right running, and a hydraulic motor for crushing from the first and second hydraulic pumps. Crushing control valve means for controlling the flow of pressure oil supplied to the feeder hydraulic motor, the conveyor hydraulic motor, the magnetic separator hydraulic motor, and the left / right traveling hydraulic motor, respectively, and feeder control. Valve means,
Conveyor control valve means, magnetic separator control valve means, and
A hydraulic drive device for a self-propelled crusher having control valve means for right running; a flow rate setting means for setting a target discharge flow rate of the first hydraulic pump; Pump flow rate control means for controlling the discharge flow rate of the hydraulic pump, first and second discharge pressure detection means for detecting discharge pressures of the first and second hydraulic pumps, respectively, and the output horsepower of the prime mover is detected. The input horsepower of the first and second hydraulic pumps is determined using the first and second distribution reference horsepower distributed to the first and second hydraulic pumps at a ratio corresponding to the discharge pressures of the first and second hydraulic pumps. The first
And below the second distribution reference horsepower, respectively.
First and second pump horsepower limit control means for controlling discharge flow rates of the first and second hydraulic pumps, respectively, and the first hydraulic pump corresponding to the set target discharge flow rate and the first distribution reference horsepower A reference discharge pressure calculating means for calculating a reference discharge pressure of the first hydraulic pump; and a control valve means for the feeder being neutralized when a difference between the reference discharge pressure and the detected discharge pressure of the first hydraulic pump becomes equal to or less than a first predetermined value. Signal output means for outputting a signal for returning the opening degree of the feeder control valve means to a set value when the position is switched to a position or the opening degree is reduced and returned to a second predetermined value or more.

[0028]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS An embodiment in which the present invention is applied to a self-propelled crusher will be described below with reference to FIGS.

A first embodiment of the hydraulic drive system for a self-propelled crusher according to the present invention will be described with reference to FIGS.

FIG. 1 is a side view showing the entire structure of a self-propelled crusher to which the present embodiment is applied. In FIG. 1, a self-propelled crusher 1 receives crushed raw material by a working tool such as a bucket of a hydraulic shovel and receives the crushed raw material, and receives the crushed raw material into a hopper 2 having a substantially V-shaped cross section. Crusher body 5 having a crusher for crushing the crushed raw material to a predetermined size, for example, a jaw crusher 3, and a feeder 4 for transporting and guiding the crushed raw material received in the hopper 2 to the jaw crusher 3, and a jaw crusher 3. A conveyor 6 for transporting and carrying out the crushed material crushed and reduced to the rear side (the right side in FIG. 1) of the crusher 1 and the crushed material provided above the conveyor 6 and being transported on the conveyor 6 are included. Magnetic separator 7 for magnetically attracting and removing magnetic substances
And a traveling body 8 provided below the crusher body 5 and provided with left and right crawler tracks 8a and track frames 8b.

The jaw crusher 3 has moving teeth (not shown).
And driving teeth generated by the crusher hydraulic motor 9 are converted into oscillating motion of moving teeth by a known conversion mechanism, and the moving teeth are swung back and forth with respect to the fixed teeth. By moving, the crushed raw material supplied from the feeder 4 is crushed to a predetermined size.

The feeder 4 is a so-called grizzly feeder, and a bottom plate including a plurality of saw-tooth plates 4a on which the crushed material from the hopper 2 is placed by a driving force generated by a hydraulic motor 10 for the feeder. Excite the part. As a result, the crushed raw material supplied to the hopper 2 is sequentially conveyed and supplied to the jaw crusher 3, and fine earth and sand attached to the crushed raw material fall down from the gap between the saw teeth of the saw tooth plate 4a during the conveyance. ing.

The conveyor 6 includes a hydraulic motor 11 for the conveyor.
This drives the belt 6a, thereby transporting the crushed material that has fallen onto the belt 6a from the jaw crusher 3.

The magnetic separator 7 is mounted on a power unit 18 to be described later via a support member 12, and a magnetic separator belt 7a disposed above the conveyor belt 6a so as to be substantially orthogonal to the conveyor belt 6a. By driving the magnetic force from the magnetic force generating means (not shown) around the belt 7a by driving the magnetic force from the magnetic force generating means (not shown) by the hydraulic motor 13 for the magnetic separator, the magnetic material is attracted to the belt 7a, and then the conveyor belt is driven. The conveyor belt 6a is conveyed in a direction substantially perpendicular to the conveyor belt 6a and dropped to the side of the conveyor belt 6a.

Each of the endless track crawler tracks 8a is bridged between a drive wheel 14 provided on the track frame 8b and an idler 15, and is provided on the left and right sides provided on the drive wheel 14 side.
The crusher 1 is caused to travel by being provided with a driving force by right traveling hydraulic motors 16 and 17 (see FIG. 5 described later for 17).

The track frame 8b is provided with a power unit 18 at the upper part of its rear end in the longitudinal direction (right side in FIG. 1).
It is equipped with. Then, hydraulic oil is discharged to hydraulic actuators such as the crusher hydraulic motor 9, the feeder hydraulic motor 10, the conveyor hydraulic motor 11, the magnetic separator hydraulic motor 13, the left and right traveling hydraulic motors 16 and 17, and the like. Hydraulic pumps 19 and 20 (see FIG. 7 to be described later), an engine 21 (the same) as a prime mover for driving the hydraulic pumps 19 and 20, and hydraulic oil supplied from these hydraulic pumps 19 and 20 to the hydraulic actuator First and second valve groups 22 and 23 for controlling the direction and flow rate (see FIG.
And control valve unit 9 equipped with
1 and 92 (see FIGS. 5 and 6 described later).

The front side of the power unit 18 (FIG. 1)
A driver's seat 24 on which the operator boards is provided on the left side (middle left).

Here, the jaw crusher 3, the feeder 4, the conveyor 6, the magnetic separator 7, and the traveling body 8 constitute driven members driven by a hydraulic driving device provided in the self-propelled crusher 1. ing. 5, 6, and 7
FIG. 1 is a hydraulic circuit diagram illustrating a first embodiment of a hydraulic drive device for a self-propelled crusher of the present invention.

In FIGS. 5 to 7, the hydraulic drive unit is the same as the engine 21 and the variable displacement type first hydraulic pump 19 and the second hydraulic pump 20 driven by the engine 21. The hydraulic motors 9, 10, 11, 13, which are supplied with hydraulic oil discharged from the first and second hydraulic pumps 19, 20, respectively.
16, 17 and the flow (direction and flow rate or only flow rate) of the pressure oil supplied from the first and second hydraulic pumps 19, 20 to the hydraulic motors 9, 10, 11, 13, 16, 16 and 17 are controlled. Six control valves 26, 27,
28, 29, 30, 31 and a control valve 2 for left and right running provided in the driver's seat 24 (see FIG. 1).
Left and right traveling operation levers 32a and 33a for switching operations of the first and second 7, 28 (described later);
Pump control means for adjusting the discharge flow rates of the hydraulic pumps 19 and 20, for example, regulator devices 34 and 35, and the crusher main body 5 (for example, in the above-mentioned operator's seat 24) are provided, and the jaw crusher 3, feeder 4, And magnetic separator 7
And an operation panel 36 for an operator to input instructions to start and stop the operation.

The six hydraulic motors 9, 10, 11, 13,
The crushing hydraulic motor 9 for generating the driving force for operating the jaw crusher 3, the feeder hydraulic motor 10 for generating the driving force for operating the feeder 4, and the conveyor 6 for operating the conveyor 6, as described above. The conveyor hydraulic motor 11 for generating a driving force, the magnetic separator hydraulic motor 13 for generating a driving force for operating the magnetic separator 7, and the left and right for generating a driving force to the left / right trackless track 8a. The traveling hydraulic motors 16 and 17 are provided.

The control valves 26 to 31 are two-position switching valves or three-position switching valves, and include a crushing control valve 26 connected to the crushing hydraulic motor 9 and a left traveling A control valve 27, a right traveling control valve 28 connected to the right traveling hydraulic motor 17, a feeder control valve 29 connected to the feeder hydraulic motor 10, and a conveyor control connected to the conveyor hydraulic motor 11. It comprises a valve 30 and a magnetic separator control valve 31 connected to the magnetic separator hydraulic motor 13.

At this time, the first and second hydraulic pumps 19,
20, the first hydraulic pump 19 discharges hydraulic oil to be supplied to the left traveling hydraulic motor 16 and the crushing hydraulic motor 9 via the left traveling control valve 27 and the crushing control valve 26. Has become. Each of these control valves 27 and 26 is a three-position switching valve capable of controlling the direction and flow rate of hydraulic oil to the corresponding hydraulic motors 16 and 9, and is connected to the discharge line 37 of the first hydraulic pump 19. Center bypass line 22a
In the first valve group 22 provided with the above, the control valve 27 for left running and the control valve 26 for crushing are arranged in this order from the upstream side. In addition, a pump control valve 38 (details will be described later) is provided at the most downstream side of the center bypass line 22a.

On the other hand, the second hydraulic pump 20 is provided with a feeder hydraulic motor 10, a conveyor hydraulic pressure, via a right traveling control valve 28, a feeder control valve 29, a conveyor control valve 30, and a magnetic separator control valve 31. Pressure oil to be supplied to the motor 11 and the hydraulic motor for magnetic separator 13 is discharged. Of these, the right travel control valve 28
Is a three-position switching valve capable of controlling the flow of pressure oil to the corresponding right-travel hydraulic motor 17, and the other control valves 28, 29, 30, 31 are connected to the corresponding hydraulic motors 10, 11, 13. A two-position switching valve capable of controlling the flow rate of the pressure oil of the second hydraulic pump 20.
In the second valve group 23 having a center bypass line 23a connected to the control line 9 and a center line 23b further connected downstream of the center bypass line 23a, the right running control valve 28, the magnetic separator control valve 31 , A conveyer control valve 30 and a feeder control valve 29 in this order. The center line 23b is closed on the downstream side of the feeder control valve 29 on the most downstream side.

Of the control valves 26 to 31, the left and right traveling control valves 27 and 28 are center bypass type pilot operated valves which are operated using pilot pressure generated by the pilot pump 25. These left and right running control valves 2
7 and 28 are operating lever devices 32 generated by the pilot pump 25 and provided with the operating levers 32a and 33a described above.
It is operated by the pilot pressure reduced to a predetermined pressure at 33.

That is, the operation lever devices 32 and 33 are
Operation levers 32a and 33a and a pair of pressure reducing valves 32b, 32b and 3 for outputting a pilot pressure corresponding to the operation amount thereof
3b and 33b. When the operation lever 32a of the operation lever device 32 is operated in the direction a in FIG. 5 (or in the opposite direction, the same applies hereinafter), the pilot pressure increases via the pilot line 40 (or 41). It is led to the driving unit 27a (or 27b),
Thereby, the left traveling control valve 27 is switched to the upper switching position 27A (or the lower switching position 27B) in FIG. The hydraulic pressure is supplied to the left traveling hydraulic motor 16 via the switching position 27A (or the lower switching position 27B) of the left traveling control valve 27, and the left traveling hydraulic motor 16 is driven in the forward (or reverse) direction. You.

When the operating lever 32a is set to the neutral position shown in FIG. 5, the left traveling control valve 27 returns to the neutral position shown in FIG. Stop.

Similarly, when the operating lever 33a of the operating lever device 33 is operated in the direction b (or the opposite direction) in FIG. 5, the pilot pressure is increased via the pilot line 42 (or 43). Drive unit 28a
(Or 28b), and the upper switching position 28A in FIG.
(Or the lower switching position 28B), and the right traveling hydraulic motor 17 is driven in the forward direction (or the reverse direction). When the operating lever 33a is set to the neutral position, the right traveling control valve 28 returns to the neutral position by the urging force of the springs 28c and 28d, and the right traveling hydraulic motor 17
Stops.

Here, a solenoid control valve 46 which is switched by a drive signal St (described later) from a controller 45 is provided in the pilot introduction pipes 44a and 44b for guiding the pilot pressure from the pilot pump 25 to the operation lever devices 32 and 33. Is provided. This solenoid control valve 46
Is a drive signal St input to the solenoid drive unit 46a.
7 is switched to the communication position 46A on the left side in FIG. 7, the pilot pressure from the pilot pump 25 is guided to the operation lever devices 32 and 33 via the introduction conduits 44a and 44b, and left and right by the operation levers 32a and 33a. The above operation of the traveling control valves 27 and 28 is enabled.

On the other hand, when the drive signal St is turned OFF, the solenoid control valve 46 returns to the shut-off position 46B on the right side in FIG. The introduction pipe 44b is connected to the tank 47
To the tank line 47a.
b is the tank pressure, and the operating levers 32a, 33a
The above-mentioned operation of the left and right traveling control valves 27 and 28 is not possible.

The crushing control valve 26 is a center bypass type electromagnetic proportional valve having solenoid driving parts 26a and 26b at both ends. The solenoid drive unit 26a,
The solenoid 26b is provided with a solenoid driven by a drive signal Scr from the controller 45, and the crushing control valve 26 is switched according to the input of the drive signal Scr.

That is, the drive signal Scr is a signal corresponding to the forward rotation (or reverse rotation, hereinafter the same) of the jaw crusher 3, for example, the drive signal Scr to the solenoid drive units 26a and 26b is ON and OFF (or the solenoid, respectively). When the drive signals Scr to the drive units 26a and 26b are turned OFF and ON, respectively, the crushing control valve 26 is switched to the upper switching position 26A (or the lower switching position 26B) in FIG. As a result, the pressure oil from the first hydraulic pump 19 is supplied to the crushing hydraulic motor 9 via the discharge pipe 37, the center bypass line 22a, and the switching position 26A (or the lower switching position 26B) of the crushing control valve 26. And the crushing hydraulic motor 9 is driven in the forward (or reverse) direction.

The driving signal Scr is a signal corresponding to the stop of the jaw crusher 3, for example, the solenoid driving units 26a and 26a.
When both the drive signals Scr to 6b are turned off, the control valve 26 is actuated by the urging forces of the springs 26c and 26d as shown in FIG.
And the crushing hydraulic motor 9 stops.

The pump control valve 38 has a function of converting a flow rate into a pressure, and a piston 38a capable of connecting and disconnecting the center bypass line 22a and the tank line 47b through a throttle portion 38aa.
And a spring 38 for biasing both ends of the piston 38a.
b, 38c and a line 80 connected to the discharge line 79 of the pilot pump 25 to guide the pilot pressure, the upstream side is connected, the downstream side is connected to the tank line 47c, and the relief pressure is set by the spring 38b. And a variable relief valve 38d variably set.

With such a configuration, the pump control valve 38 functions as follows. That is, as described above, the left traveling control valve 27 and the crushing control valve 26 are center bypass type valves, and the flow rate flowing through the center bypass line 22a depends on the operation amount of each of the control valves 27 and 26 (that is, the spool). Changeover stroke amount). When the control valves 27 and 26 are in a neutral state, that is, when the required flow rate to the first hydraulic pump 19 is small, most of the hydraulic oil discharged from the first hydraulic pump 19 is pump-controlled through the center bypass line 22a. A relatively large flow of pressure oil is introduced into the valve 38 and a tank line 47 flows through the throttle portion 38aa of the piston 38a.
b. As a result, the piston 38a moves to the right in FIG.
The set relief pressure d becomes low, and the pipelines 81a and 81 branch from the pipeline 80 and reach a first servo valve 95 for negative tilt control (negative control) described later.
b, a relatively low control pressure (negative control pressure) Pc1 is generated.

Conversely, when each of the control valves 27 and 26 is operated to be in the open state, that is, when the required flow rate to the first hydraulic pump 19 is large, the flow rate flowing through the center bypass line 22a is reduced by the hydraulic motor Since it is reduced by the flow rate flowing to the 16 and 9 sides, the piston throttle
5a, the flow rate of the pressure oil led to the tank line 47b becomes relatively small, and the piston 38a moves to the left in FIG. 5 to increase the set relief pressure of the relief valve 38d. Pc1 increases.

In this embodiment, as will be described later, the first servo valve 95 controls the tilt angle of the swash plate 19A of the first hydraulic pump 19 based on the fluctuation of the control pressure (negative control pressure) Pc1. (Details will be described later).

The first and second hydraulic pumps 19, 20
A relief valve 89 and a relief valve 90 are provided in pipes 87 and 88 branched from the discharge pipes 37 and 39 of the first and second hydraulic pumps 19 and 20, respectively, so that the maximum discharge pressures P1 and P2 of the first and second hydraulic pumps 19 and 20 are provided. The value of the relief pressure for limiting the value is set by the biasing force of the springs 89a and 90a provided respectively. Also, the first hydraulic pump 1
9 is provided with a pressure sensor 105 for detecting the discharge pressure P1 in the discharge pipe 37. The pressure sensor 105 outputs a corresponding detection signal to the controller 45.

The feeder control valve 29 is an electromagnetic switching valve having a solenoid drive section 29a. The solenoid drive section 29a is provided with a solenoid driven by a drive signal Sf from the controller 45, and the feeder control valve 29 is switched according to the input of the drive signal Sf. That is, when the drive signal Sf becomes an ON signal for operating the feeder 4, the feeder control valve 29 is switched to the upper switching position 29A in FIG.

As a result, the pressure oil from the second hydraulic pump 20 guided through the discharge pipeline 39, the center bypass line 23a, and the center line 23b is transferred to the switching position 29.
A from the throttle means 29Aa provided in A, a pipe 50 connected thereto, a pressure control valve 51 provided in this pipe 50
(Details will be described later), port 2 provided at switching position 29A
9Ab and supply line 5 connected to this port 29Ab
2, is supplied to the feeder hydraulic motor 10, and the hydraulic motor 10 is driven. When the drive signal Sf becomes an OFF signal corresponding to the stop of the feeder 4, the feeder control valve 29 returns to the shut-off position shown in FIG. 6 by the urging force of the spring 29b, and the feeder hydraulic motor 10 stops.

The control valve 30 for the conveyor, like the control valve 29 for the feeder, sends a drive signal S from the controller 45 to its solenoid drive section 30a.
A solenoid driven by com is provided. Drive signal Sc
When om becomes the ON signal for operating the conveyor 6, the control valve 30 for the conveyor moves to the upper communication position 3 in FIG.
0A, and the pressure oil from the center line 23b flows from the throttle means 30Aa at the switching position 30A to the pipeline 5A.
3. The pressure control valve 54 (details will be described later), the port 30Ab of the switching position 30A, and the supply pipe 55 connected to the port 30Ab are supplied to the conveyor hydraulic motor 11 to be driven. When the drive signal Scom becomes an OFF signal corresponding to the stop of the conveyor 6, the conveyor control valve 30 returns to the shut-off position shown in FIG. 6 by the urging force of the spring 30b, and the conveyor hydraulic motor 11 stops.

As with the control valve 29 for the feeder and the control valve 30 for the conveyor, the solenoid of the solenoid selector 31a is driven by the drive signal Sm from the controller 45. When the drive signal Sm becomes an ON signal, the control valve 31 for the magnetic separator is switched to the communication position 31A on the upper side in FIG. 31Ab → supplied to the hydraulic motor 13 for magnetic separator via the supply line 58 to be driven. When the drive signal Sm becomes the OFF signal, the control valve 31 for the magnetic separator returns to the shut-off position by the urging force of the spring 31b.

The feeder hydraulic motor 1 described above
0, supply of pressurized oil to the conveyor hydraulic motor 11 and the magnetic separator hydraulic motor 13 from the viewpoint of circuit protection, etc.
Relief valves 62, 63, 64 are provided in pipes 59, 60, 61 connecting the supply pipes 52, 55, 58 and the tank line 47b, respectively.

Here, the functions related to the pressure control valves 51, 54, 57 provided in the pipes 50, 53, 56 will be described.

The port 29Ab at the switching position 29A of the control valve 29 for the feeder and the port 30A at the switching position 30A of the control valve 30 for the conveyor.
b, and the switching position 3 of the control valve 31 for the magnetic separator
The 1A port 31Ab has a load detection port 29Ac, a load detection port 30Ac for detecting the load pressure of the corresponding feeder hydraulic motor 10, conveyor hydraulic motor 11, and magnetic separator hydraulic motor 13, respectively.
The load detection port 31Ac is in communication. At this time,
The load detection port 29Ac is connected to the load detection line 65, the load detection port 30Ac is connected to the load detection line 66, and the load detection port 31Ac is connected to the load detection line 6
7 is connected.

Here, the load detection line 65 to which the load pressure of the feeder hydraulic motor 10 is led and the load detection line 6 to which the load pressure of the conveyor hydraulic motor 11 is led.
6 further includes a load detection line 6 via a shuttle valve 68.
9, and the load pressure on the high pressure side selected via the shuttle valve 68 is guided to the load detection line 69. The load detection line 69 and the load detection line 6 to which the load pressure of the magnetic separator hydraulic motor 13 is guided.
7 is the maximum load detection line 71 via the shuttle valve 70.
The load pressure on the high pressure side selected by the shuttle valve 70 is guided to the maximum load detection pipe 71 as the maximum load pressure.

The maximum load pressure guided to the maximum load detection line 71 is transmitted to the corresponding pressure control valve 51 via lines 72, 73, 74, and 75 connected to the maximum load detection line 71. , 54, 57, respectively. At this time, the pressure in the pipes 50, 53, 56, that is, the downstream pressure of the throttle means 29Aa, 30Aa, 31Aa is guided to the other side of the pressure control valves 51, 54, 57.

As described above, the pressure control valves 51, 54, 57
Is the throttle means 2 of the control valves 29, 30, 31
Each hydraulic motor operates in response to the differential pressure between the downstream pressure of 9Aa, 30Aa, and 31Aa and the maximum load pressure of the feeder hydraulic motor 10, the conveyor hydraulic motor 11, and the magnetic separator hydraulic motor 13. The differential pressure is maintained at a constant value irrespective of changes in the load pressures 10, 11, and 13. That is, the throttle means 29Aa, 3
The downstream pressures of 0Aa and 31Aa are set higher than the maximum load pressure by the set pressures of the springs 51a, 54a and 57a.

On the other hand, the discharge line 39 of the second hydraulic pump 20
Bleed-off line 76 branched from center bypass line 23a and center line 23b connected to
A relief valve (unload valve) 77 having a spring 77a is provided. On one side of the relief valve 77, a maximum load pressure is guided through a maximum load detection pipe 71 and a pipe 78 connected thereto, and the other side of the relief valve 77 is bleed off through a port 77b. The pressure in line 76 is led. Thereby, the relief valve 77 is configured to increase the pressure in the pipeline 76 and the center line 23b by the pressure set by the spring 77a from the maximum load pressure. That is, when the pressure in the pipe 76 and the pressure in the center line 23b becomes a pressure obtained by adding the spring force of the spring 77a to the pressure in the pipe 78 to which the maximum load pressure is introduced, the relief valve 77 The pressure oil in the passage 76 is led to the tank 47 via the pump control valve 82, whereby the feeder control valve 29,
Control is performed so that the flow rates from the control valve 30 for the conveyor and the control valve 31 for the magnetic separator become constant.

At this time, the relief pressure set by the spring 77a corresponds to the above-described relief valve 89 and relief valve 90.
Is set to a value smaller than the set relief pressure.

A pump control valve 82 having a flow-pressure conversion function similar to that of the pump control valve 38 is provided downstream of the relief valve 77 of the bleed-off line 76. A piston 82a capable of connecting / disconnecting the tank line 47d to the tank line 47d via a throttle portion 82aa;
The upstream side is connected to the springs 82b and 82c for biasing both ends of the 2a and the line 83 connected to the discharge line 79 of the pilot pump 25 to guide the pilot pressure, and the downstream side is connected to the tank line 47d. And the spring 8
A variable relief valve 82d whose relief pressure is variably set by 2b.

With such a configuration, during the crushing operation, the pump control valve 82 functions as follows. That is, as described above, the center line 23b
Is closed at the most downstream end, and the right-running control valve 28 is not operated during the crushing operation, as described later. Valve 30, control valve 3 for magnetic separator
It changes according to the amount of operation (i.e., the amount of spool switching stroke). Each control valve 29, 30, 31
Is neutral, that is, when the required flow rate to the second hydraulic pump 20 is small, the pressure oil discharged from the second hydraulic pump 20 is hardly introduced into the supply pipes 52, 55, and 58. It is led to the downstream side and introduced into the pump control valve 82. As a result, a relatively large flow rate of the pressurized oil flows into the throttle portion 82a of the piston 82a.
6, the piston 82a moves to the right in FIG. 6 and the set relief pressure of the relief valve 82d by the spring 82b decreases, and the pipe 83
, A relatively low control pressure (load sensing pressure) Pc2 is generated in a pipeline 84 which is provided to branch to a first servo valve 96 for later-described load sensing tilt control.

Conversely, when each control valve is operated to be opened, that is, when the required flow rate to the second hydraulic pump 20 is large, the flow rate flowing through the bleed-off line 76 is reduced by the hydraulic motors 10 and 11. , 13 is reduced by the flow rate flowing to the tank line 47d through the piston throttle portion 82aa, and the piston 82a moves to the left in FIG. 5 to set the relief valve 82d. Since the relief pressure increases, the pipe 84
, The load sensing pressure Pc2 increases. FIG. 8 shows an example of such a relationship between the flow rate of the piston 82a of the pump control valve 82 and the control pressure Pc2. In the present embodiment, as described later, this load sensing pressure P
The tilt angle of the swash plate 20A of the second hydraulic pump 20 is controlled based on the variation of c2 (details will be described later).

The pressure control valves 51, 54, 5 described above
7, control between the downstream pressure of the throttle means 29Aa, 30Aa, 31Aa and the maximum load pressure, and the relief valve 7
7, the control between the pressure in the bleed-off line 76 and the maximum load pressure allows the throttling means 29Aa, 30Aa, 3
The pressure compensating function for keeping the differential pressure of 1 Aa constant is performed. Thereby, each hydraulic motor 10, 11, 13
Control valve 2 regardless of load pressure change
The pressure oil corresponding to the opening degree of 9, 30, 31 can be supplied to the corresponding hydraulic motor. The pressure compensation function and the tilt angle control of the swash plate 20A of the hydraulic pump 20 described later based on the output of the load sensing pressure Pc2 from the pump control valve 82 result in the discharge pressure of the second hydraulic pump 20 And aperture means 29Aa,
The difference from the downstream pressure of 30Aa, 31Aa is kept constant (details will be described later).

Further, a relief valve 85 is provided between the pipe line 78 from which the maximum load pressure is introduced and the tank line 47d, and the maximum pressure in the pipe line 78 is limited to a set pressure of the spring 85a or less. It is designed for protection. That is,
The relief valve 85 and the relief valve 75 constitute a system relief valve. When the pressure in the pipe 78 becomes larger than the pressure set by the spring 85a, the action of the relief valve 85 causes the pressure in the pipe 78 to be reduced. The pressure is reduced to the tank pressure, whereby the above-described relief valve 77 is operated to be in a relief state.

In the above arrangement, the crushing control valve 26 and the left traveling control valve 27 of the first valve group 22, the right traveling control valve 28 of the second valve group, and the pump control valve 38 , The relief valves 89 and 90 are combined as a high-pressure side system, and are integrated into the main valve unit 91. On the other hand, the control valve 29 for the feeder, the control valve 30 for the conveyor, and the control valve 31 for the magnetic separator of the second valve group 23,
A relief valve 77, a pump control valve 82,
The relief valve 85 is integrated as a low-pressure side system, and is integrated into the sub-valve unit 92. Center bypass line 23 of main valve unit 91
The carry-over port 91a on the downstream side of a is connected to the pump port 92a of the sub-valve unit 92 communicating with the center line 23b.

At this time, although the detailed structure is not shown,
Control valve 29 for feeder, control valve 30 for conveyor, and control valve 3 for magnetic separator
The diameter of each of the spools 1 is smaller than the diameter of the spool of the crushing control valve 26, the left traveling control valve 27, and the right traveling control valve 28.

The regulator devices 34 and 35 include tilt actuators 93 and 94, first servo valves 95 and 96, and second servo valves 97 and 98. These servo valves 95
To 98, the pressure of the hydraulic oil acting on the tilt actuators 93, 94 from the pilot pump 25 or the first and second hydraulic pumps 19, 20 is controlled, and the swash plates 19A, 19A, The tilt (ie, displacement) of the 20A is controlled. The tilt actuators 93 and 94 have large-diameter pressure receiving portions 93a and 94 at both ends.
a and operating pistons 93c and 94c having small-diameter pressure receiving portions 93b and 94b, and pressure receiving portions 93a, 93b and 94, respectively.
a and 94b respectively have pressure receiving chambers 93d and 93e and 94d and 94e. And both pressure receiving chambers 93
When the pressures d, 93e and 94d, 94e are equal to each other, the working pistons 93c, 94c move rightward in FIG.
The tilt of 20A increases, and the pump discharge flow rate increases. When the pressure in the large-diameter pressure receiving chambers 93d, 94d decreases, the working pistons 93c, 94c move to the left in FIG. It is supposed to. The large-diameter-side pressure receiving chambers 93d and 94d are connected to a pipe 99 communicating with the discharge pipe 79 of the pilot pump 25 via first and second servo valves 95 to 98. The chambers 93e and 94e are directly connected to the pipe 99.

Of the first servo valves 95 and 96, the first servo valve 95 of the regulator 34 is controlled by the pump control valve 38 via the pipes 81a and 81b and the solenoid control valve 102 (described later) as described above. A negative tilt control servo valve driven by the pressure (negative control pressure) Pc1;
The servo valve 96 is a load sensing control servo valve driven by the control pressure (load sensing pressure) Pc2 guided from the pump control valve 82 via the pipe line 84, as described above. Has become.

That is, when the control pressures PC1 and PC2 are high, the valve bodies 95a and 96a move rightward in FIG. 7, and the pilot pressure PP from the pilot pump 25 is not reduced and the tilting actuators 93 and 94 receive the pressure. This is transmitted to the chambers 93d and 94d, whereby the tilt of the swash plates 19A and 20A is increased, and the discharge flow rates of the first and second hydraulic pumps 19 and 20 are increased. As the control pressures PC1 and PC2 decrease, the valves 95a and 96a move leftward in FIG. 7 by the force of the springs 95b and 96b, and reduce the pilot pressure PP from the pilot pump 25 to reduce the pressure in the pressure receiving chambers 93d and 94d. To reduce the discharge flow rate of the first and second hydraulic pumps 19 and 20. These servo valves 95,
FIG. 9 shows an example of control characteristics of the discharge flow rates of the first and second hydraulic pumps 19 and 20 with respect to the control pressures PC1 and PC2 executed by the operation 96.

As described above, the first of the regulator device 34
In the servo valve 95, the pump control valve 38 is supplied from the center bypass line 22a so that a discharge flow rate corresponding to the required flow rate of the control valves 26 and 27 is obtained in addition to the function of the pump control valve 38 described above. The first is to minimize the flow rate
The above-described negative tilt control (negative control) for controlling the tilt of the swash plate 19A of the hydraulic pump 19 is realized.

At this time, a solenoid control valve 102 that is switched by a drive signal Su from the controller 45 is provided in the pipe lines 81a and 81b. When the drive signal Su input to the solenoid drive unit 102a is turned on, the solenoid control valve 102 is switched to the switching position 102A on the left side in FIG.
b. This allows negative control as described above.

On the other hand, when the drive signal Su is turned off, the solenoid control valve 102 returns to the right switching position 102B in FIG.
1b, and a line 81b that is branched from a line 80 communicating with the discharge line 79 of the pilot pump 25 and a line 81b. At this time, the drive signal S from the controller 45
An electromagnetic proportional valve 104 that is switched according to p is provided. The solenoid proportional valve 104 is connected to the pilot pump 2
5 based on the pilot pressure introduced through the discharge pipeline 79, the pipeline 80, and the pipeline 103.
The control pressure (positive control pressure) Pc3 is generated such that the larger the drive current value of the drive signal Sp inputted to the drive signal 4a becomes, the smaller the drive current value becomes. Thus, the first servo valve 95 of the regulator device 34 controls the swash plate 19A of the first hydraulic pump 19 so that the pump discharge flow rate becomes in accordance with the operation amount of the crusher speed setting dial 36c of the operation panel 36 as described later. The so-called positive tilt control (positive control) for controlling the tilt is realized.

In the first servo valve 96 of the regulator device 35, in addition to the function of the pump control valve 82 described above, the discharge pressure P2 of the second hydraulic pump 20 and the downstream pressure of the throttle means 29Aa, 30Aa, 31Aa are controlled. So-called load sensing control in which the discharge flow rate of the second hydraulic pump 20 is controlled so that the difference is kept constant.

On the other hand, the second servo valves 97 and 98 are both servo valves for horsepower control (input torque limiting control), and have the same structure. That is, the second servo valves 97 and 98 are valves which are operated by the discharge pressures P1 and P2 of the first and second hydraulic pumps 19 and 20, respectively.
1, P2 is a discharge pressure detecting pipe 1 branched from discharge pipes 37, 39 of the first and second hydraulic pumps 19, 20.
The pressure receiving chamber 97b of the operation drive unit 97a is provided via
, And are guided to the pressure receiving chamber 98b of the operation drive unit 98a.

That is, the first and second hydraulic pumps 19,
The operation drive units 97a, 97 are controlled by the 20 discharge pressures P1, P2.
When the force acting on 8a is smaller than the force acting on the valve bodies 97d, 98d by the spring force set by the springs 97c, 98c, the valve bodies 97d, 98d move rightward in FIG. First servo valve 95, 96
Is transmitted to the pressure receiving chambers 93d and 94d of the tilt actuators 93 and 94 without reducing the pressure, thereby tilting the swash plates 19A and 20A of the first and second hydraulic pumps 19 and 20. To increase the discharge flow rate. Then, as the force due to the discharge pressures P1, P2 of the first and second hydraulic pumps 19, 20 becomes larger than the force according to the set spring force of the springs 97c, 98c, the valve element 9 is increased.
7d and 98d move to the left in FIG. 7, reduce the pilot pressure PP guided from the pilot pump 25 via the first servo valves 95 and 96, and transmit the reduced pressure to the pressure receiving chambers 93d and 94d. The discharge flow rate of the second hydraulic pumps 19 and 20 is reduced.

As described above, in the regulator device 34, as the discharge pressure P1 of the first hydraulic pump 19 increases, the maximum value Q1ma of the discharge flow rate Q1 of the first hydraulic pump 19 increases.
x is limited to a small value, and the input horsepower (input torque) of the first hydraulic pump 19 is limited to １ ／ or less of the output horsepower (output torque, more specifically, the effective output horsepower described later, the same applies hereinafter) of the engine 21. The tilt of the swash plate 19A of the first hydraulic pump 19 is controlled (known horsepower control). Also,
In the regulator device 35, the second hydraulic pump 20
As the discharge pressure P2 of the second hydraulic pump 20 increases.
The maximum value Q2max of the discharge flow rate Q2 of the second
The input horsepower (input torque) of the hydraulic pump 20 is
The tilt of the swash plate 20A of the second hydraulic pump 20 is controlled so as to limit the output horsepower (output torque) to 1/2 or less (known horsepower control). Thus, the first hydraulic pump 1
By halving the horsepower of the engine 21 on the 9th side and the second hydraulic pump 20 side by 1/2, the total of the input horsepower (input torque) of the first hydraulic pump 19 and the second hydraulic pump is reduced by the output horsepower ( Output torque). That is, the output horsepower of the engine 21 is equally divided into １ ／ on the first hydraulic pump 19 side and １ ／ on the second hydraulic pump 20 side.

The first and second hydraulic pumps 19, which show the horsepower control realized by the second servo valves 97, 98,
An example of the 20 P (pump discharge pressure) -Q (pump discharge flow rate) characteristic line is shown in FIG. In the present embodiment, both the first hydraulic pump 19 and the second hydraulic pump 20 are controlled to approximately the characteristics represented by this characteristic line.

Returning to FIGS. 5 to 7, the operation panel 36 has a crusher start / stop switch 36a for starting / stopping the jaw crusher 3, and the operation direction of the jaw crusher 3 is either forward or reverse. , A crusher forward / reverse selection dial 36b for setting the operation speed of the jaw crusher 3, a crusher speed setting dial 36c for setting the operation speed of the jaw crusher 3, and a feeder start / stop switch 36d for starting / stopping the feeder 4. Conveyor 6
Start / stop switch 36e for starting / stopping the machine, a magnetic separator start / stop switch 36f for starting / stopping the magnetic separator 7, one of a traveling mode for performing a traveling operation and a crushing mode for performing a crushing operation A mode selection switch 36g for selecting one is provided.

When an operator operates various switches and dials of the operation panel 36, operation signals are input to the controller 45. The controller 45 controls the crushing control valve 26 and the feeder control valve 2 based on an operation signal from the operation panel 36.
9, the control valve 30 for the conveyor, the control valve 31 for the magnetic separator, and the solenoid drive units 26a and 26b and the solenoid drive unit 29 of the solenoid control valve 46
a, solenoid drive section 30a, solenoid drive section 31
a, solenoid driving section 46a, solenoid driving section 102
a, and the aforementioned drive signals Scr, Sf, Scom, Sm, St, Su, Sp to the solenoid drive section 104a are generated and output to the corresponding solenoid drive sections.

That is, when the "running mode" is selected by the mode selection switch 36g of the operation panel 36, the drive signal St of the solenoid control valve 46 is turned on to set the solenoid control valve 46 to the communication position on the left side in FIG. To enable the operation of the traveling control valves 27 and 28 by the operation levers 32a and 33a, and turn on the drive signal Su of the solenoid control valve 102 to switch the solenoid control valve 102 to the switching position on the left side in FIG. The aforementioned negative control using the pump control valve 38 is performed. When the "crushing mode" is selected by the mode selection switch 36g of the operation panel 36, the drive signal St of the solenoid control valve 46 is turned off and FIG.
The operation levers 32a and 33 are returned to the middle right cutoff position.
a, the driving signal Su of the solenoid control valve 102 is turned off, and the solenoid control valve 102 is turned off.
After returning to the middle right switching position, the above-described positive control using the electromagnetic proportional valve 104 is performed.

Further, "forward rotation" (or "reverse rotation", hereinafter referred to as "forward rotation" or "reverse rotation") using the crusher forward / reverse selection dial 36b of the operation panel 36.
When the crusher start / stop switch 36a is pressed to the "start" side in a state where the correspondence is the same, the drive signal Scr to the solenoid drive unit 26a (or the solenoid drive unit 26b) of the crushing control valve 26 is transmitted. At the same time, the drive signal Scr to the solenoid drive unit 26b (or the solenoid drive unit 26a) is turned off, and the crushing control valve 26 is set to the upper switching position 26A in FIG.
(Or the lower switching position 26B), and the hydraulic oil from the first hydraulic pump 19 is supplied to and driven by the crushing hydraulic motor 9, so that the jaw crusher 3 is rotated in the normal direction (or in the reverse direction).
To start. At this time, crusher speed setting dial 3
6c to generate a drive signal Sp having a drive current value corresponding to the operation amount signal to generate a solenoid drive portion 104a of the solenoid proportional valve 104.
(See also step 120 in the flow of FIG. 11 to be described later), and the electromagnetic proportional valve 104 is switched. Is supplied to the crushing hydraulic motor 9, and the jaw crusher 3 is operated at a speed corresponding to the operation amount. Thereafter, when the crusher start / stop switch 36a is pushed to the "stop" side, the drive signals Scr of both the solenoid drive unit 26a and the solenoid drive unit 26b of the crushing control valve 26 are turned off to return to the neutral position shown in FIG. The crushing hydraulic motor 9 is stopped, and the jaw crusher 3 is stopped.

When the feeder start / stop switch 36d of the operation panel 36 is pushed to the "start" side, the drive signal Sf to the solenoid drive section 29a of the feeder control valve 29 is turned on to turn on the upper side in FIG. Switching position 29A
And feeds the pressure oil from the second hydraulic pump 20 to the feeder hydraulic motor 10 and drives it to start the feeder 4. Thereafter, when the feeder start / stop switch 36d of the operation panel 36 is pushed to the "stop" side, the drive signal Sf to the solenoid drive portion 29a of the control valve 29 for the feeder is turned off to return to the neutral position shown in FIG. ,
The feeder hydraulic motor 10 is stopped, and the feeder 4 is stopped.

Similarly, the conveyor start / stop switch 36
When e is pushed to the "start" side, the control valve 30 for the conveyor is switched to the upper switching position 30A in FIG. 6, the hydraulic motor 11 for the conveyor is driven to start the conveyor 6, and the conveyor start / stop switch 36e Is "stop"
When pushed to the side, the control valve 30 for the conveyor is returned to the neutral position, and the conveyor 6 is stopped. Also,
When the magnetic separator start / stop switch 36f is pushed to the "start" side, the magnetic separator control valve 31 is switched to the upper switching position 31A in FIG.
Is driven to start the magnetic separator 7, and when the magnetic separator start / stop switch 36f is pushed to the "stop" side, the magnetic separator control valve 31 is returned to the neutral position, and the magnetic separator 7 is stopped.

Here, the main part of the present embodiment is that when the load of the crushing hydraulic motor 9 increases during the crushing operation (that is, during the operation of the jaw crusher 3 and the feeder 4), it is earlier than in the prior art. In other words, the operation of the feeder 4 is stopped or delayed (details will be described later) to prevent a decrease in the rotation speed of the crushing hydraulic motor 9 beforehand. The control contents of the controller 45 relating to this function will be described with reference to FIG.

In FIG. 11, the controller 45 first sets the crusher speed setting dial 36 in step 100.
Input an operation signal corresponding to the operation amount of c.
At 10, a target discharge flow rate Q1 of the first hydraulic pump 19 corresponding to the operation signal (for example, based on a predetermined table stored in advance and substantially in direct proportion to the operation signal).
Is calculated and set.

Thereafter, the process proceeds to step 120, and as described above, a drive signal Sp having a drive current value of a magnitude corresponding to the above Q1 is generated and output to the electromagnetic proportional valve 104, 104 is driven.

Then, in step 130, the PQ characteristic line of the first hydraulic pump 19 previously shown in FIG. 10 (the lower right horsepower line represents the reference horsepower of the first hydraulic pump 19).
, A reference discharge pressure P10 (see FIG. 10) of the first hydraulic pump 19 corresponding to the above Q1 is obtained. Specifically, this is calculated by performing the following calculation.

That is, the maximum (for example, practical) maximum output horsepower uniquely determined by the specifications of the engine 21 is represented by W
[KW], assuming that the efficiency coefficient taking into account all losses including the mechanical loss of the engine 21 is η, the first hydraulic pump 19 and the second hydraulic pump 20 of the total output horsepower of the engine 21 are actually used. The part that can be done (effective output horsepower) is W
× η [kW]. As described above, in the present embodiment, the effective output horsepower is compared with the first hydraulic pump 19 side and the second hydraulic pump power.
Since the hydraulic pump 20 and the hydraulic pump 20 are equally divided by 1/2, the horsepower W1 that can be allocated to the first hydraulic pump 19 is W1 = (W × η) / 2 [kW]. P- when axis is P1
This corresponds to the equi-horsepower line portion (downward curved portion) of the Q characteristic diagram. Thereby, the reference discharge pressure P10 corresponding to the above Q1 is obtained.
(See FIG. 10) is: P10 = (W1 / Q1) × 60 [MPa]

After step 130 as described above, the process proceeds to step 140, in which a feeder stop pressure P11 and a feeder operation restart pressure P12 corresponding to the reference discharge pressure P10 are set (see FIG. 10). As a method of setting P11 and P12, for example, a differential pressure (P10-P11, P10-P
12) may be set to a first predetermined value and a second predetermined value, respectively (however, the first predetermined value <the second predetermined value). These predetermined values may be fixed values, respectively, or P10
(For example, P10−P11 = 0.1 × P10, P10−
(P12 = 0.2 × P10).

Then, in step 150, a detection signal from the pressure sensor 105 for detecting the discharge pressure P1 of the first hydraulic pump 19 is input, and in step 160, it is determined whether or not P1 is larger than P11. If P1 ≦ P11, this determination is not satisfied, and the aforementioned step 150
Return to and repeat the same procedure. If P1> P11, it is determined that the load on the crushing hydraulic motor 9 is too large, and the routine proceeds to step 170. The determination in step 160 may be based on whether or not the state of P1> P11 has continued for a predetermined time. In this case,
There is a merit that more accurate control can be performed except for a transient or instantaneous increase in P1.

In step 170, the drive signal Sf to the feeder control valve 29 is turned off to return the feeder control valve 29 to the neutral position.
Supply of pressure oil to the feeder 4 is stopped, and the feeder 4 is stopped. The feeder control valve 29 is simply turned on.
Instead of using an OFF valve, the amount of pressure oil supplied to the supply line 52 can be controlled in accordance with the drive current value of the drive signal Sf to the solenoid drive unit 29a, and in step 170, the drive signal to the feeder control valve 29 is set. The feeder 4 may be decelerated by reducing the drive current value of Sf to reduce the supply of pressure oil to the feeder hydraulic motor 10.

After step 170 as described above, the process proceeds to step 180, where the detection signal P1 from the pressure sensor 105 is input again, and in step 190, it is determined whether P1 is smaller than P12. If P1 ≧ P12, this determination is not satisfied, and the routine returns to the step 180 and repeats the same procedure. If P1 <P12, it is determined that the excessive load on the crushing hydraulic motor 9 has been eliminated, and the routine proceeds to step 200.

As in step 160, it goes without saying that the determination in step 190 may be based on whether the condition of P1 <P12 has continued for a predetermined time.

In step 200, the drive signal Sf to the feeder control valve 29 is turned ON and switched again (for example, to the full stroke position), whereby the supply of the pressure oil to the feeder hydraulic motor 10 is resumed, and Operation is restarted. The feeder 4 described above
Is decelerated without stopping, this step 200
The drive signal Sf to the feeder control valve 29
Is increased to the initial value (the value before deceleration of the feeder 4), the switching position of the control valve 29 for the feeder is returned to the initial position (the position before deceleration), and the operation speed of the feeder 4 is returned to the original value. Return to Thereafter, the process returns to step 150, and the same procedure is repeated.

In the configuration described above, the crawler track 8a constitutes the traveling means described in the claims, and the left traveling control valve 27 and the right traveling control valve 28 constitute the left / right traveling control valve means. The crushing control valve 26 forms crushing control valve means. Further, the feeder control valve 29 constitutes feeder control valve means, the conveyor control valve 30 constitutes conveyor control valve means, and the magnetic separator control valve 31 constitutes magnetic separator machine control valve means.

The crusher speed setting dial 36c of the operation panel 36 constitutes rotation speed setting means for setting the rotation speed of the crushing hydraulic motor.
Steps 100 and 11 shown in FIG.
0 constitutes a flow rate setting means for setting a target discharge flow rate of the first hydraulic pump. Further, the discharge pressure detection pipe 100 and the pressure sensor 105 constitute a first discharge pressure detecting means for detecting a discharge pressure of the first hydraulic pump, and a step 120 shown in FIG. Proportional valve 104 and first servo valve 9 of regulator device 34
5 constitutes a pump flow control means for controlling the discharge flow rate of the first hydraulic pump in accordance with the set target discharge flow rate, and the second servo valve 97 of the regulator device 34 detects the discharge of the first hydraulic pump which is detected. In accordance with the pressure, the input horsepower of the first hydraulic pump is changed to a reference horsepower related to the output horsepower of the prime mover (in the present embodiment, 1 / の of the effective output horsepower of the prime mover as described above).
2) A first pump horsepower limit control unit that controls the discharge flow rate of the first hydraulic pump so as to limit the following is set.

Further, among the control functions of the controller 45, the step 130 shown in FIG. 11 is a step of calculating a reference discharge pressure calculating means for calculating a reference discharge pressure of the first hydraulic pump corresponding to the set value of the flow rate setting means and the reference horsepower. And step 1
40 to 200 constitute signal output means for switching the feeder control valve means to the neutral position or outputting a signal for decreasing the opening degree in accordance with the reference discharge pressure and the detected discharge pressure of the first hydraulic pump. I do.

Next, the operation and effects of the self-propelled crusher according to the first embodiment of the present invention having the above configuration will be described below.

In the self-propelled crusher 1 having the above-described configuration, when the self-propelled crusher 1 is allowed to travel to a location where crushing work is performed, the operator operates the mode selection switch 3 of the operation panel 36.
At 6g, "running mode" is selected, and the driver gets on the driver's seat 24 and operates the operating levers 32a and 33a forward. As a result, the left and right traveling control valves 27 and 28 are connected as shown in FIG.
It is switched to the middle upper switching position 27A, 28A, and the first
Hydraulic oil guided from the hydraulic pump 19 via the center bypass line 22a is supplied to the left and right traveling hydraulic motors 16 and 17, which are driven in the forward direction, and the crawler tracks 8a on both sides of the crusher 1 are sequentially moved. The driving body 8 is driven in the direction to travel forward.

At the time of the crushing operation, the operator selects the "crushing mode" with the mode selection switch 36g of the operation panel 36 to disable the traveling operation, and then sets the "forward" with the crusher forward / reverse selection dial 36b. , And while rotating the crusher speed setting dial 36c to a position where the desired setting speed is obtained, the magnetic separator start / stop switch 36f, the conveyor start / stop switch 36e, the crusher start / stop switch 36a, and the feeder start / stop. Switch 3
6d is sequentially pushed to the "start" side.

By the above operation, the controller 45 controls the solenoid driving unit 3 of the magnetic separator control valve 31.
The drive signal Sm to 1a is turned ON, the control valve 31 for the magnetic separator is switched to the upper switching position 31A in FIG. 6, and the drive signal Scom from the controller 45 to the solenoid drive unit 30a of the control valve 30 for the conveyor is transmitted. Turned on and control valve 3 for conveyor
0 is switched to the upper switching position 30A in FIG. Further, the drive signal Scr from the controller 45 to the solenoid drive section 26a of the crushing control valve 26 is turned on.
And the drive signal S to the solenoid drive unit 26b.
The cr is turned OFF, the crushing control valve 26 is switched to the upper switching position 26A in FIG. 5, and the solenoid drive section 29a of the feeder control valve 29 is turned on.
A drive signal Sf to ON is turned on, and the feeder control valve 29 is switched to the upper switching position 29A in FIG.

Thus, the pressure oil from the second hydraulic pump 20 is introduced into the pump port 92a and the center line 23b of the sub-valve unit 92 via the center bypass line 23a and the carry-over port 91a of the main valve unit 91. Hydraulic motor for magnetic separator 13,
The magnetic separator 7, the conveyor 6, and the feeder 4 are supplied to the conveyor hydraulic motor 11 and the feeder hydraulic motor 10, and are started. On the other hand, the pressure oil from the first hydraulic pump 19 is supplied to the crushing hydraulic motor 9 and the jaw crusher 3
Is activated in the forward direction.

When the crushed raw material is charged into the hopper 2 by, for example, a bucket of a hydraulic shovel, the input crushed raw material is guided to the jaw crusher 3 while sorting only those having a predetermined particle size or more in the feeder 4, and Crushed by the crusher 3 to a predetermined size. The crushed material is transferred from the space below the jaw crusher 3 to the conveyor 6.
The magnetic material (for example, rebar pieces mixed in the waste construction material) mixed in the crushed material is removed by the magnetic separator 7 in the middle of the transportation, and the size is almost uniform. Finally, the crusher 1 is carried out from the rear (the right end in FIG. 1).

At this time, the discharge flow rate Q 1 of the first hydraulic pump 19 is limited by the second servo valve 97 of the regulator device 34 so that the input horsepower of the first hydraulic pump 19 is 以下 or less of the effective output horsepower of the engine 21. (I.e., to operate in a region including the equihorse power line portion of the PQ characteristic line shown in FIG. 10 and inside the same).
Further, under this restriction condition, the discharge flow rate Q1 of the first hydraulic pump 19 is such that the electromagnetic proportional valve 104 connects the conduit 103 to the conduit 81b with an opening corresponding to the operation amount of the crusher speed setting dial 36c. Therefore, the flow rate is controlled according to the operation amount of the crusher speed setting dial 36c. Thus, the crushing hydraulic motor 9 rotates at a constant speed corresponding to the discharge flow rate Q1.

Here, it is generally known that in a crushing device such as a jaw crusher, the particle size of the crushed material changes according to its operating speed (ie, the rotation speed of the crushing hydraulic motor). When the rotation speed increases, the particle size decreases, and when the rotation speed decreases, the particle size increases. This will be described with reference to FIG.

FIG. 12 shows a crushed material 106 crushed in the jaw crusher 3 having the moving teeth 3a and the fixed teeth 3b.
FIG. 4 is a view schematically showing a state in which is discharged downward in the jaw crusher 3. In FIG.
The rotational speed of the crushing hydraulic motor 9 for swinging the moving teeth 3a is set to N
Assuming that C [rpm], the time (period) t required for the moving tooth 3a to make one reciprocation (open once and then close again to the original position) is t = 60 / NC [s] (Equation 1) It is represented by During this time t [s], the crushed material 1
The distance h at which 06 falls downward by its own weight is given by h = gt2 / 2 (Equation 2), where g is the gravitational acceleration.

That is, according to equations (1) and (2), h is NC
Therefore, as the rotation speed NC increases, the distance h traveled by the crushed material 106 during one cycle t decreases. For this reason, the moving teeth 3a are set between the time when the jaw crusher 3 is inserted from above and the time when the jaw crusher 3 is discharged from below.
And the number of times of pressing and crushing by the fixed teeth 3b increases,
The finer the particles, the smaller they become.
Conversely, as the rotation speed NC decreases, the distance h that the crushed material 106 travels during one cycle t increases. The number of times of pressing by the teeth 3b is reduced, so that the particles are not crushed too finely, and the particle size increases.

In the other types of crushers, that is, roll crushers, shredders, etc., the same principle that the distance that the crushed material travels in one cycle becomes smaller as the rotation speed of the crushing hydraulic motor increases, Characteristics similar to the above are established.

In the present embodiment, as described above,
The crushing hydraulic motor 9 is driven by a crusher speed setting dial 36.
Since it rotates at a constant speed according to the operation amount of c, the particle size of the crushed product carried out from the conveyor 6 is adjusted to a size according to the constant speed.

Here, when the load of the crushing hydraulic motor 9 increases due to, for example, a large supply of the crushing raw material from the feeder 4 to the jaw crusher 3 or a high compressive strength (that is, a high hardness), the crushing hydraulic motor 9 responds accordingly. The discharge pressure P1 of one hydraulic pump 19 increases. For this reason, the first hydraulic pump 19
Will shift to the high pressure side. For example,
In FIG. 13, the first operating point of the first hydraulic pump 19 is a point (0) in an area inside the equihorse power line portion (curved portion) of the PQ characteristic line by operating the crusher speed setting dial 36c. If the load on the crushing hydraulic motor 9 increases, only the discharge pressure P1 increases while the discharge flow rate Q1 of the first hydraulic pump 19 remains constant. In other words, the operating point moves horizontally to the right inside the equal horsepower line portion of the PQ characteristic line (see the rightward arrow in FIG. 13A).

Thereafter, when the load of the crushing hydraulic motor 9 further increases, the discharge pressure P1 further increases, and the operating point further moves to the high pressure side and approaches the horsepower line portion such as the PQ characteristic line. At t1, the discharge pressure P1 becomes larger than the feeder stop pressure P11 described above (FIG. 13).
(See FIG. 13B, corresponding to a point in FIG. 13A). This satisfies the determination at step 160 in the flow of FIG. 11 described above, and the feeder 4 is stopped or decelerated at step 170 (see FIG. 13C). As a result, the supply of the crushing raw material to the jaw crusher 3 is stopped or reduced, so that the load on the crushing hydraulic motor 9 is reduced, and the discharge pressure P1 of the first hydraulic pump 19 is reduced again (FIG. 13).
(See (b)) The operating point of the first hydraulic pump 19 shifts and returns to the direction away from the horsepower line portion such as the PQ characteristic line, that is, to the low pressure side (see the leftward arrow in FIG. 13A).

As is apparent from the control contents for returning to the low pressure side before reaching such an equal horsepower line portion of the PQ characteristic line, the control function of the controller 45 in the step 130 shown in FIG. And steps 140-200
In other words, it corresponds to feeder control means for controlling the feeder control valve means based on the detected discharge pressure of the first hydraulic pump so that the output horsepower of the first hydraulic pump does not exceed the reference horsepower.

Thereafter, when the load on the crushing hydraulic motor 9 further decreases, the discharge pressure P1 of the first hydraulic pump further decreases, and the operating point further moves to the low pressure side opposite to the horsepower line portion such as the PQ characteristic line. (See FIG. 13 (a)). At a certain time t2, the discharge pressure P1 becomes smaller than the above-described feeder operation restarting pressure P12 (see FIG. 13 (b).
FIG. 13 (a) corresponds to a point). This satisfies the determination of step 190 in the flow of FIG.
In step 200, the feeder 4 returns to the normal operation speed (see FIG. 13C). Thus, the supply of the crushed raw material to the jaw crusher 3 is restarted.

Thereafter, by repeating the above control,
As shown in FIG. 13 (b), the discharge pressure P1 of the first hydraulic pump 19 fluctuates only within a range in which P11 and P12 are substantially upper and lower limits, and the operating point of the first hydraulic pump 19 is as shown in FIG.
As shown in (a), even if the vehicle does not reach the horsepower line portion such as the PQ characteristic line (or reaches the horsepower line portion such as the PQ characteristic line), it does not slide further on the horsepower line. Good), and reciprocate between point (1) and point (2).

A conventional technique described in Japanese Patent Application Laid-Open No. 8-257425 will be described as a comparative example with respect to the above-described embodiment (the same reference numerals as in the embodiment are used for clarifying the comparison). ). In this case, as described above, control is performed to correct the rotational speed NC of the crushing hydraulic motor 9 after the rotational speed NC has decreased.

That is, as shown in FIG. 14A, in this case, similarly to the first embodiment of the present invention, the operating point of the first hydraulic pump 19 is first set to the equihorse power line of the PQ characteristic line. When the load on the crushing hydraulic motor 9 increases, the discharge pressure P1 of the first hydraulic pump 19 increases and moves horizontally to the right (FIG. 14A). See arrow pointing right). When the load of the crushing hydraulic motor 9 further increases, the operating point reaches the equihorse power line portion of the PQ characteristic line, but at this time, the rotation speed NC of the hydraulic motor 9 has not yet decreased (see FIG. 14 (d)) Since the feeder 4 continues the normal operation (see FIG. 14 (c)), the load of the crushing hydraulic motor 9 further increases, and the discharge pressure P1 of the first hydraulic pump 19 further increases. (See FIG. 14B). Therefore, the operating point of the first hydraulic pump 19 slides downward to the right on the isohorse power line (FIG. 1).
4 (a), the discharge flow rate Q1 decreases, so that the rotational speed NC of the crushing hydraulic motor 9 also gradually decreases (see FIG. 14 (d)).

At a certain time t1 ', the rotational speed NC becomes smaller than a first predetermined value NC1 (see FIG. 14 (d), which corresponds to point' in FIG. 14 (a)). As a result, the supply of the crushed material from the feeder 4 to the jaw crusher 3 is stopped, and the crushing hydraulic motor 9 is stopped.
The load of the first hydraulic pump 19 is reduced.
Starts to decrease again (see FIG. 14B). for that reason,
The operating point of the first hydraulic pump 19 slides on the horsepower line portion such as the PQ characteristic line so as to return to the upper left again (FIG. 1).
4 (a), the discharge flow rate Q1 of the first hydraulic pump 19 increases, so that the rotation speed NC of the crushing hydraulic motor 9 gradually increases again (FIG. 14 (d)). reference).

Thereafter, at a certain time t2 ', the rotational speed NC becomes larger than a second predetermined value NC2 (see FIG. 14 (d), corresponding to' in FIG. 14 (a)).
Therefore, the supply of the crushed raw material from the feeder 4 to the jaw crusher 3 is restarted.

In such a conventional control, FIG.
As can be seen from FIG. 4D, since the rotational speed NC of the crushing hydraulic motor 9 is corrected after the increase or decrease has occurred, the rotational speed NC of the crushing hydraulic motor 9 is set to the first predetermined value NC1. And a second predetermined value NC2. Therefore, the rotation speed NC cannot be sufficiently stabilized, and the particle size distribution of the crushed material varies to some extent, making it difficult to sufficiently improve the quality of the product. Therefore, in order to obtain a high-quality product, it is necessary to sieve the crushed material using a sieving means or the like, and there is a problem that productivity is reduced.

On the other hand, FIG.
In the above-described present embodiment described with reference to (d), the feeder 4 is stopped when the operating point has reached at least the equi-horsepower line of the PQ characteristic line, thereby rotating the crushing hydraulic motor 9. Since the operating point of the first hydraulic pump 19 is returned to the low pressure side before the speed NC decreases, the rotation speed NC of the crushing hydraulic motor 9 can be prevented from lowering. Therefore, the crushing hydraulic motor 9 can be stably rotated at a constant speed according to the setting of the crusher speed setting dial 36c of the operation panel 36. Thereby, the particle size of the crushed product can be adjusted to a size corresponding to the constant speed, so that productivity can be improved.

In the prior art described in Japanese Patent Application Laid-Open No. Hei 8-257425, the crushing device is switched to a crushing hydraulic motor by switching the crushing control valve so that the rotation speed is set to the rotation speed set by the rotation speed setting means. Because the amount of pressurized oil supplied is controlled, for example, when a relatively low speed is set by the rotational speed setting means, the pressurized oil from the hydraulic pump is supplied to the crushing hydraulic motor by squeezing with the crushing control valve Flow rate to be reduced. As a result, a large pressure loss occurs in the crushing control valve, and accordingly, the horsepower of the prime mover is wasted and energy efficiency is reduced. On the other hand, in the above embodiment, the crusher speed setting dial 36c is
Since the flow rate Q1 corresponding to the operation amount (set value) is discharged, the pressure loss generated in the crushing control valve 26 is significantly reduced. Thereby, energy efficiency can be improved.

Further, by controlling the operating speed of the crushing hydraulic motor 9 in accordance with the setting of the crusher speed setting dial 36c of the operation panel 36, the following effects can be obtained.

For example, conventionally, as disclosed in Japanese Patent Application Laid-Open No. H8-25704, a hydraulic drive device of a self-propelled crusher is used to freely control the rotation speed of a crushing hydraulic motor. A crushing device for crushing the crushed raw material, a variable displacement hydraulic pump driven by a prime mover, and a variable displacement crushing hydraulic motor driving the crushing device by pressure oil discharged from the hydraulic pump; Crushing control valve means for controlling the flow of pressurized oil supplied to the crushing hydraulic motor from the hydraulic pump, means for detecting the magnitude of the load acting on the crushing device, and There has been proposed a configuration in which means for changing the capacity of the crushing hydraulic motor is provided.

In the prior art described above, when the load of the crushing device increases, the capacity of the crushing hydraulic motor is increased so that the crushing device is rotated at a low speed, and the output torque is increased to secure a large crushing force. When it decreases, the capacity of the crushing hydraulic motor is reduced to achieve high speed rotation,
This improves work efficiency.

However, in the prior art, the discharge flow from the hydraulic pump to the crushing hydraulic motor is basically constant (the capacity of the variable displacement hydraulic pump is controlled only by the so-called horsepower control), and the load is reduced. The hydraulic oil supply from the hydraulic pump to the hydraulic motor for crushing via the control valve means for crushing at the time of low-speed rotation and high-speed rotation of the hydraulic motor for crushing, because the capacity of the hydraulic motor for crushing is changed according to the size of The flow through the path does not change.
Therefore, even when the crushing hydraulic motor rotates at a low speed, the same large flow rate as during the high-speed rotation flows through the pressure oil supply path, so that a pressure loss (for example, a pressure loss in a pipe or a control valve) in the pressure oil supply path increases. Accordingly, at the time of low-speed rotation, the horsepower that can be supplied to the crushing device requiring the largest horsepower among the devices of the self-propelled crusher decreases, and the energy efficiency decreases. In addition, the power of the prime mover tends to be insufficient, and the operating speed of the crushing device or the like may be reduced due to the insufficient flow rate, and the productivity may be reduced. In order to prevent this, a larger prime mover is required, which leads to an increase in cost.

On the other hand, in the above embodiment,
The controller 45 performs the processing in step 100 shown in FIG.
Then, an operation signal corresponding to the operation amount of the crusher speed setting dial 36c is input. In step 110, the target discharge flow rate Q1 of the first hydraulic pump 19 corresponding to the operation signal is set. A drive signal Sp having a drive current value of a magnitude corresponding to
4 to drive the electromagnetic proportional valve 104. As a result, the electromagnetic proportional valve 104 connects the conduit 103 with the conduit 81 at an opening corresponding to the operation amount of the crusher speed setting dial 36c.
b, the discharge flow rate Q1 of the first hydraulic pump 19
Is controlled to a flow rate Q1 (= constant) corresponding to the above-mentioned operation amount, and the crushing hydraulic motor 9 rotates at a constant speed corresponding to the discharge flow rate Q1.

That is, when the crushing hydraulic motor 9 is rotated at a desired speed at a high speed, the operation amount of the crusher speed setting dial 36c is increased, so that the discharge flow rate Q1 from the first hydraulic pump 19 is set to a higher value. When the crushing hydraulic motor 9 is rotated at a desired speed at a low speed, the operation amount of the crusher speed setting dial 36c is made small, so that the discharge flow rate Q1 from the first hydraulic pump 19 is set to a small value. Decreases correspondingly.

As described above, the hydraulic oil supply path from the first hydraulic pump 19 to the crushing hydraulic motor 9 via the crushing control valve 26 depends on whether the crushing hydraulic motor 9 is rotated at low speed or at high speed. By changing the flow rate, the flow rate of the crushing hydraulic motor 9 at the low speed rotation can be reduced as compared with the high speed rotation at the high speed rotation, so that the pressure loss in the pressure oil supply path is made smaller than at the high speed rotation. be able to. Therefore, the energy efficiency can be improved correspondingly as compared with the above-described conventional technology. In addition, this can prevent the horsepower of the engine 21 from becoming insufficient and prevent the productivity from lowering. Therefore, it is not necessary to increase the size of the engine 21 and the cost can be prevented from rising.

When the above embodiment is compared with the above prior art based on the above effects, the above embodiment is driven by a crushing device for crushing crushed raw materials supplied from a hopper and a motor. The variable displacement type hydraulic pump, the crushing hydraulic motor that drives the crushing device by the pressure oil discharged from the hydraulic pump, and the flow of the pressure oil supplied from the hydraulic pump to the crushing hydraulic motor. A hydraulic drive device for a self-propelled crusher having crushing control valve means for controlling, a flow rate setting means for setting a target discharge flow rate of the hydraulic pump, and a hydraulic pump for the hydraulic pump according to the set target discharge flow rate. The invention can also be considered as an invention of a hydraulic drive device for a self-propelled crusher, which has a pump flow rate control means for controlling a discharge flow rate.

At this time, in the above invention, as described above, the crushing control valve 26 constitutes a crushing control valve means, and the crusher speed setting dial 36 of the operation panel 36 and the control function of the controller 45 are provided. Among them, steps 100 and 110 shown in FIG. 11 constitute flow rate setting means for setting the target discharge flow rate of the first hydraulic pump, and step 120 of the control function of the controller 45 shown in FIG. 104 and the first servo valve 95 of the regulator device 34 constitute a pump flow control means for controlling the discharge flow rate of the first hydraulic pump according to the set target discharge flow rate.

FIGS. 15 to 17 show a second embodiment of the present invention.
This will be described below. This embodiment is an embodiment in the case where so-called total horsepower control is performed for the first hydraulic pump and the second hydraulic pump. In FIGS. 15 to 17, parts that are the same as in the first embodiment are given the same reference numerals.

FIG. 15 is a hydraulic circuit diagram showing a main structure of a hydraulic drive device of a self-propelled crusher according to the present embodiment.
FIG. 8 is a diagram corresponding to FIG. 7 of the first embodiment. This FIG.
The present embodiment is different from the first embodiment in that the pressure sensor 10 detects the discharge pressure P2 of the second hydraulic pump 20 and outputs a corresponding detection signal to the controller 45.
7, the discharge pressures P1 and P2 of the first and second hydraulic pumps 19 and 20 are introduced into the second servo valves 97 and 98 of the regulator devices 34 and 35, respectively. The second servo valves 97 and 98 perform so-called total horsepower control on the first and second hydraulic pumps 19 and 20.

That is, the second servo valves 97 and 98 of the regulator devices 34 and 35 are connected to the first and second hydraulic pumps 1 and 2, respectively.
These valves are operated by the discharge pressures P1 and P2 of the first and second hydraulic pumps 19 and 20, respectively.
The pressure receiving chambers 97ba and 97bb of the operation driving unit 97a and the pressure receiving chambers 98ba and 98ba of the operation driving unit 98a are connected via the discharge pressure detection pipes 100a to 100c and 101a to 100c branched from the 20 discharge pipes 37 and 39, respectively. 98bb.

Then, the first and second hydraulic pumps 19, 2
The operation driving unit 97a,
When the force acting on the valve body 98a is smaller than the force acting on the valve bodies 97d and 98d by the spring force set by the springs 97c and 98c, the valve bodies 97d and 98d move rightward in FIG. First servo valve 95, 9
6 is transmitted to the pressure receiving chambers 93d and 94d of the tilt actuators 93 and 94 without reducing the pressure, whereby the swash plates 19A and 20A of the first and second hydraulic pumps 19 and 20 are tilted. The rotation is increased to increase the discharge flow rate. Then, as the force based on the sum P1 + P2 of the discharge pressures of the first and second hydraulic pumps 19 and 20 becomes larger than the force based on the set spring force of the springs 97c and 98c, the valve bodies 97d and 98d move leftward in FIG. It moves and reduces the pilot pressure PP guided from the pilot pump 25 via the first servo valves 95 and 96, and transmits the reduced pressure to the pressure receiving chambers 93d and 94d, whereby the first and second hydraulic pumps 19 and 2 are moved.
The discharge flow rate of 0 is reduced.

As described above, the first and second hydraulic pumps 1
As the discharge pressures P1 and P2 of the pumps 9 and 20 increase, the maximum values Q1max and Q2max of the discharge flow rates Q1 and Q2 of the first and second hydraulic pumps 19 and 20 are limited to a small value. Of the input horsepower of the engine 21 (specifically, the effective output horsepower as in the first embodiment,
The tilting of the swash plates 19A and 20A of the first and second hydraulic pumps 19 and 20 is controlled so as to limit the following.
More specifically, according to the sum of the discharge pressure P1 of the first hydraulic pump 19 and the discharge pressure P2 of the second hydraulic pump 20, the total input horsepower of the first and second hydraulic pumps 19 and 20 is calculated by Full horsepower control that limits the output horsepower or less is realized.

The first and second hydraulic pumps 1 showing the full horsepower control realized by the second servo valves 97 and 98
9, 20 P (pump discharge pressure)-Q (pump discharge flow rate)
FIG. 16 shows an example of the characteristic line. In the present embodiment, the first
Both the hydraulic pump 19 and the second hydraulic pump 20 are controlled to approximately the characteristics represented by this characteristic line. That is, the sum of the discharge pressures P1 + P2 of the first and second hydraulic pumps 19 and 20 and the discharge flow rate Q1 of the first hydraulic pump 19 when the first hydraulic pump 19 is controlled by the second servo valve 97 of the regulator device 34. And the sum P1 + P2 of the discharge pressures of the first and second hydraulic pumps 19 and 20 when the second hydraulic pump 20 is controlled by the second servo valve 98 of the regulator device 35, and the second hydraulic pump 2
0 and the maximum value Q2max of the discharge flow rate Q2 are substantially the same as each other (for example, with a width of about 10%), and the discharge flow rates Q1 of the first and second hydraulic pumps 19 and 20 are different. , Q2 are limited to substantially the same value (same) as Q1max and Q2max.

The other hydraulic circuits are the same as in the first embodiment.

FIG. 17 is a diagram showing a control flow of the control contents of the controller 45 in the present embodiment.
FIG. 12 is a diagram corresponding to FIG. 11 of the embodiment. 17 differs from FIG. 11 in that step 125 is provided between step 120 and step 130 in cooperation with the above-described total horsepower control.

That is, the controller 45 outputs the drive signal Sp to the electromagnetic proportional valve 104 in step 120 to drive the same, and then proceeds to step 125, where the pressure sensor 1
07, the discharge pressure P2 of the second hydraulic pump 20 is input.

Thereafter, in step 130, based on the PQ characteristic lines of the first and second hydraulic pumps 19 and 20 previously shown in FIG. 16, the reference discharge pressure P10t of the first hydraulic pump 19 corresponding to the above Q1. (See FIG. 16). This is calculated by performing the following calculation, which is slightly different from step 130 in the first embodiment.

That is, as described above in the first embodiment, the first hydraulic pump 1 out of the total output horsepower of the engine 21
9 (the effective output horsepower) of the second hydraulic pump 20 is the maximum value of the output horsepower of the engine 21 as W
[KW] and the efficiency coefficient as η, W × η [kW].

During the crushing operation, the hydraulic oil from the second hydraulic pump 20 is supplied to the feeder hydraulic motor 10, the conveyor hydraulic motor 11, and the magnetic separator hydraulic motor 13. Due to the specifications of the motors 10, 11, and 13, the flow rates required during operation are uniquely determined. Further, as described above in the first embodiment, the load pressure P2 of the second hydraulic pump 20 is controlled by the hydraulic motors 10 and 10 during the crushing operation by the load sensing control using the first servo valve 96 of the regulator device 35. 11,
The pressure is always maintained so as to be higher than the maximum load pressure of 13 by a certain value, and is controlled to be the minimum pressure and flow rate necessary for driving each of the hydraulic motors 10, 11, and 13. Therefore, during the crushing operation, the second hydraulic pump 2
The discharge flow Q2 of 0 is controlled by the hydraulic motors 10, 11,
Thirteen required flow rates.

Using Q2 thus obtained and P2 detected by the pressure sensor 107, the engine 21
The horsepower W2 consumed by the second hydraulic pump 20 among the effective output horsepower W × η of the above is represented by W2 = P2 × max (Q1, Q2) / 60 [kW]. Here, the reason why the value is set to max (Q1, Q2) is that P1 and P2 are cross-sensed by full horsepower control. This W2 is an equal horsepower line portion β obtained by sliding the equal horsepower line portion α of the PQ characteristic diagram of FIG. 16 so as to indicate the horsepower distributed to the second hydraulic pump side (second distribution reference horsepower). Is equivalent to

Then, since W2 is obtained in this manner, the first output power W × η of the effective output horsepower of engine 21 is obtained.
The horsepower W1 that can be consumed by the hydraulic pump 19 is W1 = W × η−W2 = W × η−P2 × max (Q1, Q2) / 60. This W1 is obtained by dividing the equal horsepower line portion α in the PQ characteristic diagram of FIG. 16 by the horsepower (first horsepower) distributed to the first hydraulic pump side.
(The reference horsepower distribution) corresponds to the isohorsepower line portion γ that has been slid. As a result, the reference discharge pressure P10t of the first hydraulic pump 19 corresponding to the above Q1 can be obtained by P10t = (W1 / Q1) × 60 [MPa]. Incidentally, the discharge pressure P20t of the second hydraulic pump 20 corresponding to the above Q1 has a value shown in FIG. 16, and P10t-P10 = P10-P20t as shown.

After step 130 as described above, the process proceeds to step 140, in which the feeder stop pressure P11t and the feeder operation restart pressure P12t corresponding to the reference discharge pressure P10t.
Is set (see FIG. 16). The method of setting P11t and P12t is the same as the method of setting P11 and P12 in the first embodiment.

In the subsequent steps 150 to 200, P11 and P12 of the first embodiment are used instead of P11 and P12.
A control procedure similar to that of the first embodiment is performed except that t and P12t are used. As can be seen by comparing FIG. 16 with FIG. 10, in the present embodiment, as a result of performing the full horsepower control as described above, the feeder stop pressure changes from P11 to P11t.
The pressure for restarting the feeder operation increases from P12 to P12t.

In the above, the discharge pressure detecting pipe 10
0a to c and the pressure sensor 105 constitute first discharge pressure detecting means for detecting the discharge pressure of the first hydraulic pump, and the discharge pressure detection lines 101a to 101c and the pressure sensor 107 detect the discharge pressure of the second hydraulic pump. Step 120, the electromagnetic proportional valve 104, and the regulator device 3 shown in FIG. 17 among the control functions of the controller 45.
The fourth first servo valve 95 constitutes a pump flow rate control means for controlling the discharge flow rate of the first hydraulic pump according to the set target discharge flow rate.

Further, the second of the regulator devices 34 and 35
Servo valves 97 and 98 control the output horsepower of the prime mover at a ratio corresponding to the detected discharge pressures of the first and second hydraulic pumps.
And the second and third hydraulic pumps respectively distributed to the hydraulic pump side.
The input horsepower of the first and second hydraulic pumps is reduced below the first and second distribution reference horsepower using the distribution reference horsepower (in this embodiment, the horsepower corresponding to the equal horsepower line portions γ and β in FIG. 16 respectively). First and second pump horsepower limit control means for controlling the discharge flow rates of the first and second hydraulic pumps, respectively, so as to limit them.

Further, among the control functions of the controller 45, steps 125 and 130 shown in FIG. 17 calculate the reference discharge pressure of the first hydraulic pump corresponding to the set value of the flow rate setting means and the first distribution reference horsepower. It constitutes a reference discharge pressure calculating means.

The present embodiment configured as described above has the following effects.

In general, in a self-propelled crusher, during the crushing operation, the load pressure of the crushing hydraulic motor 9 is compared with the load pressure of the feeder hydraulic motor 10, the conveyor hydraulic motor 11, and the magnetic separator hydraulic motor 13. And the horsepower of the engine 21 required for the hydraulic motors 10, 11, 13 is smaller than the horsepower required for the crushing hydraulic motor 9.

Therefore, in the second embodiment, the discharge pressures P of the first and second hydraulic pumps 19 and 20 are controlled by the second servo valves 97 and 98 of the regulator devices 34 and 35.
First and second hydraulic pumps 1 according to the sum of P1, P2
The input horsepower of the engine 9 and 20 is limited to the output horsepower of the engine 21 or less, and the horsepower of the engine 21 is distributed at a ratio corresponding to the discharge pressures P1 and P2 of the first and second hydraulic pumps 19 and 20. Full horsepower control for controlling the discharge flow rates of the first and second hydraulic pumps 19 and 20 is performed, thereby reducing the load relatively with the first hydraulic pump 19 related to the crushing hydraulic motor 9 having a relatively high load. Certain hydraulic motors 10, 11,
The horsepower of the engine 21 can be effectively distributed to the second hydraulic pump 20 according to the thirteenth aspect in accordance with the difference in the load. That is, the surplus of the horsepower to the second hydraulic pump 20 that is surplus due to the small required horsepower of the hydraulic motors 10, 11, and 13 is supplied to the first hydraulic pump 19 for the crushing hydraulic motor 9 having a large required horsepower. can do. As described above, by allocating more horsepower to the crushing hydraulic motor 9 having a large load and a large required horsepower, the horsepower of the engine 21 can be effectively used, and the energy efficiency can be greatly improved. Thereby, energy saving can be achieved.

In the above two embodiments, the first and second hydraulic pumps 19, 20 are provided by the hydraulic regulator devices 34, 35 having the first servo valves 95, 96 and the second servo valves 97, 98. However, the present invention is not limited to this as long as the basic effect of the present invention that the crushing hydraulic motor 9 is stably rotated at a constant speed is obtained. For example, a displacement control controller that inputs detection signals from pressure sensors as first and second discharge pressure detection means and outputs a drive signal in response to the detection signals, and responds to drive signals from the displacement control controller. An electromagnetic proportional pressure reducing valve for reducing the pilot pressure from the pilot pump 25 through the first and second hydraulic pumps 19 and 20
And a hydraulic control piston connected to the swash plates 19A and 20A and operated by pilot pressure via the electromagnetic proportional pressure reducing valve. in this case,
In the second embodiment of the present invention, the second servo valves 97 and 9 are used.
The whole horsepower control performed hydraulically in step 8 is sufficient if a predetermined table is provided in the tilt control controller so that the whole horsepower control can be performed.

In the above two embodiments,
By separately providing the pressure P11 or P11t for stopping the feeder and the pressure P12 or P12t for restarting the feeder operation, a so-called control hysteresis is set, thereby performing hunting in which the stop and restart of the feeder are repeated in a short time. It does not happen, but it is not necessarily limited to this. That is, when there is little possibility that control instability will occur, or when there is little problem, the pressure for stopping the feeder and the pressure for restarting the feeder operation may be set to the same value. At this time, the predetermined continuation time used as the criterion of step 160 and step 190 in the flow of FIG. 11 and FIG. 17 is different between when the feeder is stopped and when the feeder operation is restarted. A configuration that avoids instability is also conceivable.

Further, in the above two embodiments, the maximum load pressure of the hydraulic motors 10, 11, 13 is detected by the lines 65, 66, 67, 69, 71, 78 and the shuttle valves 68, 70, while Pressure control valves 51, 54, 57
Thus, the differential pressure between the downstream pressures of the throttle means 29Aa, 30Aa, 31Aa and the maximum load pressure is kept constant, and the difference between the discharge pressure P2 of the second hydraulic pump 20 and the maximum load pressure is controlled by the unload valve 77. Although the pressure is kept constant and the pressure difference before and after the throttling means 29Aa, 30Aa, 31Aa is kept constant, a reliable distribution function is obtained, but the present invention is not limited to this. For example, a pressure compensating valve may be provided simply to guide the differential pressure across the throttling means 29Aa, 30Aa, 31Aa to both ends, and the differential pressure may be kept constant by the set pressure of the spring. Further, in the above, the load sensing control is performed by using the unload valve 77 to keep the pressure difference between the discharge pressure P2 of the second hydraulic pump 20 and the maximum load pressure constant, but the present invention is not limited to this. Absent. That is, the discharge pressure P2 of the second hydraulic pump 20 is detected by, for example, a pressure sensor or the like, and the maximum load pressure is also detected by a pressure sensor or the like, and both detection results are input to the load sensing controller. The pressure difference may be calculated, and the tilt angle of the second hydraulic pump 20 may be controlled by the pump control means so that the pressure difference is kept constant according to the calculation result. In addition, as long as the above-described basic effects of the present invention can be obtained, the above-described reliable distribution function by the load sensing and the pressure compensation is not always necessary, and the operating point of the first hydraulic pump 19 must be at least the PQ characteristic line. Needless to say, it is sufficient if the rotation speed NC of the crushing hydraulic motor 9 is prevented from lowering by returning to the low pressure side when reaching the constant horsepower line.

Further, in the above two embodiments, of the left and right traveling control valves 27 and 28, the left traveling control valve 27 is in the first valve group 22, and the right traveling control valve 28 is in the second valve group. Although arranged in the valve group 23, such an arrangement is not always necessary as long as the above-described basic effects of the present invention are obtained.
8 may also be arranged in the first valve group 22. However, in this case, from the viewpoint of ensuring straightness, some means for balancing the pressure oil supplied to the left and right traveling hydraulic motors 16 and 17 via the left and right traveling control valves 27 and 28 during traveling (for example, for example) Control valve 2 for left / right running
7 and 28).

In the above two embodiments,
Although the self-propelled crusher 1 provided with the jaw crusher 3 for crushing the moving teeth and the fixed teeth as the crushing device has been described as an example, the present invention is not limited to this. A rotary crusher (a six-axis crusher including a so-called roll crusher, etc.) that crushes the raw material by rotating the pair in the opposite direction to each other with a pair of crushing blades and rotating the pair in the opposite direction to each other to crush the raw material. ) Or a crusher provided with a cutter on a shaft arranged in parallel, and a crushing device (a so-called two-axis shearing machine including a shredder) for rotating the crushed raw material by rotating them in opposite directions to each other. In these cases, a similar effect is obtained.

Further, in the above two embodiments, the feeder 4 uses the driving force of a hydraulic motor to
Although the self-propelled crusher 1 provided with the grizzly feeder for oscillating the bottom plate portion including the plurality of sawtooth plates 4a on which the crushed material is placed has been described as an example, the invention is not limited to this. That is, another type of feeder, for example, a crushed raw material supplied from a hopper is placed on a substantially flat bottom plate provided below the hopper, and the bottom plate is substantially driven by a base drive mechanism based on a driving force generated by a hydraulic motor. By reciprocating in the horizontal direction, the subsequent crushed material is input, the preceding crushed material is sequentially extruded on the bottom plate, and the crusher equipped with a so-called plate feeder that sequentially supplies the crushed material to the crushing device from the front end of the bottom plate. Is also applicable.

In the above two embodiments,
Although the case where the present invention is applied to the self-propelled crusher 1 having the feeder 4, the conveyor 6, and the magnetic separator 7 as an auxiliary machine for performing the work related to the crushing operation by the crusher has been described as an example, the invention is not limited to this. That is, the present invention may be applied to a self-propelled crusher in which at least one of the conveyor 6 and the magnetic separator 7 is appropriately omitted, for example, one in which the magnetic separator 7 is omitted according to work circumstances. Conversely, in addition to the feeder 4, the conveyor 6, and the magnetic separator 7, additional auxiliary machines, for example, an auxiliary conveyor (2) located downstream (or upstream) of the conveyor 6 in order to lengthen the path of the conveyor 6 Jaw crusher 3 for further sorting according to the particle size of the crushed material
May be applied to a self-propelled crusher provided with a vibrating screen positioned downstream of the crusher. When adding auxiliary equipment,
The corresponding control valve is assigned to the second valve group 2
3 so that the pressure oil from the second hydraulic pump 20 is supplied. In these cases, a similar effect is obtained.

[0170]

According to the present invention, when the load on the crushing hydraulic motor increases, the operating point slides on the equihorse power line portion of the PQ characteristic line to the high pressure side to rotate the crushing hydraulic motor. Unlike the prior art, which performs control to correct the speed after the speed has decreased, the operating point is returned to the low pressure side at least when the operating point reaches the equihorse power line portion of the PQ characteristic line. The rotation speed of the hydraulic motor can be prevented from decreasing. Therefore, the crushing hydraulic motor can be rotated sufficiently stably at a constant speed according to the set value of the flow rate setting means. Thus, the particle size of the crushed product can be adjusted to a size corresponding to the constant speed, so that the quality of the crushed product can be sufficiently improved, and the productivity can be improved.

[Brief description of the drawings]

FIG. 1 is a side view showing the entire structure of a self-propelled crusher to which a first embodiment of the present invention is applied.

FIG. 2 is a top view illustrating the entire structure of a self-propelled crusher to which an embodiment of the present invention is applied.

FIG. 3 is a front view as viewed from a direction A in FIG. 1;

FIG. 4 is a rear view as viewed from a direction B in FIG. 1;

FIG. 5 is a hydraulic circuit diagram illustrating an embodiment of a hydraulic drive device for a self-propelled crusher of the present invention.

FIG. 5 is a hydraulic circuit diagram showing a first embodiment of a hydraulic drive device for a self-propelled crusher of the present invention.

FIG. 6 is a hydraulic circuit diagram showing a first embodiment of a hydraulic drive device for a self-propelled crusher according to the present invention.

FIG. 7 is a hydraulic circuit diagram showing a first embodiment of a hydraulic drive device for a self-propelled crusher of the present invention.

FIG. 8 is a diagram showing an example of a relationship between a flow rate through a piston of the pump control valve shown in FIG. 6 and a control pressure.

FIG. 9 is a diagram showing an example of control characteristics of the discharge flow rates of the first and second hydraulic pumps with respect to a control pressure executed by the operation of the first servo valve shown in FIG.

FIG. 10 is a diagram illustrating an example of a PQ characteristic line of the first and second hydraulic pumps, illustrating horsepower control realized by the second servo valve illustrated in FIG. 7;

FIG. 11 is a control flow showing the control contents of the controller shown in FIG. 7;

12 is a schematic diagram for explaining the relationship between the rotation speed of the crushing hydraulic motor and the particle size of the crushed material in the jaw crusher shown in FIG.

FIG. 13 is an explanatory diagram for explaining a behavior when the load of the crushing hydraulic motor is increased in the first embodiment of the hydraulic drive device of the present invention.

FIG. 14 is an explanatory diagram for explaining a behavior when the load of the crushing hydraulic motor is increased in the conventional hydraulic drive device.

FIG. 15 is a hydraulic circuit diagram illustrating a main structure of a hydraulic drive device of a self-propelled crusher according to a second embodiment of the present invention.

FIG. 16 shows first and second hydraulic pumps PQ showing the total horsepower control realized by the second servo valve shown in FIG. 15;
It is a figure showing an example of a characteristic line.

17 is a control flow showing the control contents of the controller shown in FIG.

[Explanation of symbols]

DESCRIPTION OF SYMBOLS 1 Self-propelled crusher 2 Hopper 3 Jaw crusher (crusher) 4 Feeder 6 Conveyor 7 Magnetic separator 8a Endless track crawler (running means) 9 Hydraulic motor for crushing 10 Hydraulic motor for feeder 11 Hydraulic motor for conveyor 13 Hydraulic for magnetic separator Motor 16 Left traveling hydraulic motor 17 Right traveling hydraulic motor 19 First hydraulic pump 20 Second hydraulic pump 21 Engine (motor) 26 Crushing control valve (Crushing control valve means) 27 Left traveling control valve (Left traveling Control valve means) 28 Control valve for right running (Control valve means for right running) 29 Control valve for feeder (Control valve means for feeder) 30 Control valve for conveyor (Control valve means for conveyor) 31 Control valve for magnetic separator (Magnetic separation) Control valve means) 34 Regulator device 3 5 Regulator device 36 Operation panel 36c Crusher speed setting dial (rotation speed setting means, flow rate setting means) 45 Controller 95 1st servo valve (pump flow rate control means) 97 2nd servo valve (1st pump horsepower limit control means) 98th 2 servo valve (second pump horsepower limit control means) 100 discharge pressure detection pipe (first discharge pressure detection means) 100a-c discharge pressure detection pipe (first discharge pressure detection means) 101a-c discharge pressure detection pipe (Second discharge pressure detecting means) 104 Electromagnetic proportional valve (pump flow control means) 105 Pressure sensor (first discharge pressure detecting means) 107 Pressure sensor (second discharge pressure detecting means)

────────────────────────────────────────────────── ───

[Procedure amendment]

[Submission date] August 4, 1999 (1999.8.4)

[Procedure amendment 1]

[Document name to be amended] Statement

[Correction target item name] Brief description of drawings

[Correction method] Change

[Correction contents]

[Brief description of the drawings]

FIG. 1 is a side view showing the entire structure of a self-propelled crusher to which a first embodiment of the present invention is applied.

FIG. 2 is a top view showing the entire structure of the self-propelled crusher to which the first embodiment of the present invention is applied.

FIG. 3 is a front view as viewed from a direction A in FIG. 1;

FIG. 4 is a rear view as viewed from a direction B in FIG. 1 ;

5 is a hydraulic circuit diagram showing a first embodiment of the hydraulic drive system of the self-propelled crushing machine of the present invention.

FIG. 6 is a hydraulic circuit diagram showing a first embodiment of a hydraulic drive device for a self-propelled crusher according to the present invention.

FIG. 7 is a hydraulic circuit diagram showing a first embodiment of a hydraulic drive device for a self-propelled crusher of the present invention.

FIG. 8 is a diagram showing an example of a relationship between a flow rate through a piston of the pump control valve shown in FIG. 6 and a control pressure.

FIG. 9 is a diagram showing an example of control characteristics of the discharge flow rates of the first and second hydraulic pumps with respect to a control pressure executed by the operation of the first servo valve shown in FIG.

FIG. 10 is a diagram illustrating an example of a PQ characteristic line of the first and second hydraulic pumps, illustrating horsepower control realized by the second servo valve illustrated in FIG. 7;

FIG. 11 is a control flow showing the control contents of the controller shown in FIG. 7;

12 is a schematic diagram for explaining the relationship between the rotation speed of the crushing hydraulic motor and the particle size of the crushed material in the jaw crusher shown in FIG.

FIG. 13 is an explanatory diagram for explaining a behavior when the load of the crushing hydraulic motor is increased in the first embodiment of the hydraulic drive device of the present invention.

FIG. 14 is an explanatory diagram for explaining a behavior when the load of the crushing hydraulic motor is increased in the conventional hydraulic drive device.

FIG. 15 is a hydraulic circuit diagram illustrating a main structure of a hydraulic drive device of a self-propelled crusher according to a second embodiment of the present invention.

FIG. 16 shows first and second hydraulic pumps PQ showing the total horsepower control realized by the second servo valve shown in FIG. 15;
It is a figure showing an example of a characteristic line.

17 is a control flow showing the control contents of the controller shown in FIG.

[Description of Signs] 1 Self-propelled crusher 2 Hopper 3 Jaw crusher (crushing device) 4 Feeder 6 Conveyor 7 Magnetic separator 8a Endless track crawler (running means) 9 Hydraulic motor for crushing 10 Hydraulic motor for feeder 11 Hydraulic motor for conveyor 13 Hydraulic motor for magnetic separator 16 Left traveling hydraulic motor 17 Right traveling hydraulic motor 19 First hydraulic pump 20 Second hydraulic pump 21 Engine (motor) 26 Crushing control valve (Crushing control valve means) 27 Left traveling control Valve (control valve means for left running) 28 Control valve for right running (control valve means for right running) 29 Control valve for feeder (control valve means for feeder) 30 Control valve for conveyor (control valve means for conveyor) 31 Magnetic separator Control valve (control valve means for magnetic separator) 34 legi Cutter device 35 Regulator device 36 Operation panel 36c Crusher speed setting dial (rotation speed setting means, flow rate setting means) 45 Controller 95 1st servo valve (pump flow rate control means) 97 2nd servo valve (1st pump horsepower limit control means) 98 Second servo valve (second pump horsepower limiting control means) 100 Discharge pressure detection pipe (first discharge pressure detection means) 100a-c Discharge pressure detection pipe (first discharge pressure detection means) 101a-c Discharge pressure detection Pipe line (second discharge pressure detection means) 104 Electromagnetic proportional valve (pump flow control means) 105 Pressure sensor (first discharge pressure detection means) 107 Pressure sensor (second discharge pressure detection means)

 ──────────────────────────────────────────────────の Continued on the front page (72) Inventor Tadashi Shiohata 650, Kandamachi, Tsuchiura-shi, Ibaraki Pref.Hitachi Construction Machinery Co., Ltd. F-term in Tsuchiura Works (reference) 4D063 AA08 GA07 GA10 GC01 GC08 GC16 GC21 4D067 DD04 DD06 FF01 FF15 GA02 GA06 GA20 GB05

Claims (8)

[Claims] 1. A self-propelled crusher having a crushing device for crushing crushed raw material supplied from a hopper and a feeder for conveying the crushed raw material supplied to the hopper to the crushing device, is driven by a motor. First and second hydraulic pumps, a crushing hydraulic motor and a feeder hydraulic motor that respectively drive the crushing device and the feeder with pressurized oil discharged from the first and second hydraulic pumps, And a crushing control valve means and a feeder control valve means for controlling the flow of hydraulic oil supplied from the second hydraulic pump to the crushing hydraulic motor and the feeder hydraulic motor, respectively, and At least a first hydraulic pump of the second hydraulic pump is a variable displacement type hydraulic pump. Flow rate setting means for setting an output flow rate; first discharge pressure detecting means for detecting a discharge pressure of the first hydraulic pump; and controlling a discharge flow rate of the first hydraulic pump according to the set target discharge flow rate. Pump flow rate control means, and the first hydraulic pressure so as to limit an input horsepower of the first hydraulic pump to a reference horsepower related to an output horsepower of the prime mover according to the detected discharge pressure of the first hydraulic pump. First pump horsepower limit control means for controlling a discharge flow rate of a pump, and the feeder so that an output horsepower of the first hydraulic pump does not exceed the reference horsepower based on the detected discharge pressure of the first hydraulic pump. A hydraulic drive device for a self-propelled crusher, further comprising a feeder control means for controlling a control valve means for use in a crusher. 2. The hydraulic drive system for a self-propelled crusher according to claim 1, wherein said feeder control means controls said feeder control valve so that an output horsepower of said first hydraulic pump does not exceed said reference horsepower. A hydraulic drive for a self-propelled crusher, wherein the means is switched to a neutral position or the opening is reduced. 3. The hydraulic drive system for a self-propelled crusher according to claim 2, wherein said feeder control means includes a reference discharge of said first hydraulic pump corresponding to a set value of said flow rate setting means and said reference horsepower. Switching the feeder control valve means to a neutral position or decreasing the opening in accordance with the reference discharge pressure calculating means for calculating the pressure and the detected discharge pressure of the first hydraulic pump. A hydraulic drive device for a self-propelled crusher, comprising: signal output means for outputting a signal. 4. The hydraulic drive device for a self-propelled crusher according to claim 3, wherein said signal output means determines that a difference between the reference discharge pressure and the detected discharge pressure of the first hydraulic pump is a first pressure. Switching the feeder control valve means to the neutral position or decreasing the opening when the value is equal to or less than a predetermined value, and returning the opening of the feeder control valve means to the set value when returning to the second predetermined value or more. Hydraulic drive for self-propelled crusher. 5. A hydraulic drive system for a self-propelled crusher according to claim 1, wherein said flow rate setting means includes a rotation speed setting means for setting a rotation speed of said crushing hydraulic motor. Hydraulic drive for self-propelled crusher. 6. The hydraulic drive system for a self-propelled crusher according to claim 1, wherein said first pump horsepower limit control means uses substantially half of the effective output horsepower of said motor as said reference horsepower. A hydraulic drive system for a self-propelled crusher, wherein a discharge flow rate of the first hydraulic pump is controlled so as to limit an input horsepower of the first hydraulic pump to approximately 1/2 or less of an effective output horsepower of the prime mover. . 7. The hydraulic drive device for a self-propelled crusher according to claim 1, further comprising a second discharge pressure detecting means for detecting a discharge pressure of the second hydraulic pump, and wherein the first pump horsepower is provided. Limiting control means, as the reference horsepower, the output horsepower of the motor;
And using the first distribution reference horsepower distributed to the first hydraulic pump at a ratio corresponding to the discharge pressure of the second hydraulic pump,
A hydraulic drive device for a self-propelled crusher, wherein a discharge flow rate of the first hydraulic pump is controlled so as to limit an input horsepower of the hydraulic pump to be equal to or less than the first distribution reference horsepower. 8. A crusher for crushing the crushed raw material supplied from the hopper, a feeder for conveying the crushed raw material supplied to the hopper to the crusher, and a conveyor for carrying out the crushed material crushed by the crusher. A variable capacity type provided in a self-propelled crusher having a magnetic separator for magnetically attracting and removing a magnetic substance contained in the crushed material being conveyed on the conveyor, and a traveling means, and driven by a motor First and second
A hydraulic pump, a crushing hydraulic motor, and a feeder hydraulic motor that respectively drive the crushing device, the feeder, the conveyor, the magnetic separator, and the traveling unit by hydraulic oil discharged from the first and second hydraulic pumps. A hydraulic motor for a conveyor, a hydraulic motor for a magnetic separator, and a hydraulic motor for left and right running; and a hydraulic motor for the crushing, a hydraulic motor for the feeder, a hydraulic motor for the conveyor from the first and second hydraulic pumps, Crushing control valve means, feeder control valve means, conveyor control valve means, control valve means for magnetic separator, control valve for magnetic separator, and hydraulic motor for magnetic separator, control valve means for controlling flow of pressure oil supplied to the hydraulic motors for left and right running respectively And a hydraulic drive device for a self-propelled crusher having left and right traveling control valve means, wherein a flow rate setting means sets a target discharge flow rate of the first hydraulic pump. Setting means; pump flow control means for controlling the discharge flow rate of the first hydraulic pump in accordance with the set target discharge flow rate; and first and second detecting means for detecting discharge pressures of the first and second hydraulic pumps, respectively. (2) discharge pressure detecting means, and first and second distributions in which the output horsepower of the prime mover is distributed to the first and second hydraulic pumps at a ratio corresponding to the detected discharge pressures of the first and second hydraulic pumps, respectively. Using the reference horsepower, the discharge flow rates of the first and second hydraulic pumps are controlled so as to limit the input horsepower of the first and second hydraulic pumps to the first and second distribution reference horsepowers, respectively. First and second pump horsepower limit control means; reference discharge pressure calculation means for calculating a reference discharge pressure of the first hydraulic pump corresponding to the set target discharge flow rate and the first distribution reference horsepower; Reference spitting When the difference between the pressure and the detected discharge pressure of the first hydraulic pump is equal to or less than a first predetermined value, the control valve means for the feeder is switched to the neutral position or the opening is decreased, and the pressure is increased to the second predetermined value or more. And a signal output means for outputting a signal for returning the opening of the feeder control valve means to a set value when returning.