JPH086722B2 - Inertial Drive Unit - Google Patents

Description

Description: TECHNICAL FIELD The present invention relates to an inertial body drive device provided in a construction machine such as a hydraulic excavator, and more particularly to an inertial body drive device that prevents the inertial body from rotating in reverse when the inertial body is stopped.

BACKGROUND ART Generally, in construction machines such as hydraulic excavators, an upper rotating body is provided on a lower traveling body so as to be able to rotate, and the upper rotating body is driven to rotate by a hydraulic circuit for rotation, which is an inertial body drive device.

11 and 12 show an example of a hydraulic circuit for swinging a hydraulic excavator as an inertial body drive device according to this type of prior art.

In the figure, reference numeral 1 denotes a hydraulic motor for swinging, and the hydraulic motor 1 is connected to a hydraulic pump 2 as a hydraulic source, a tank 3, a pump-side pipe 4, a tank-side pipe 5, and first and second main pipes.
The hydraulic motor 1 is driven to rotate by the supply and discharge of pressurized oil from the hydraulic pump 2, thereby rotating the upper rotating body, which serves as an inertial body, on the lower traveling body (neither of which is shown).

7 is the pump side pipe 4, the tank side pipe 5 and the main pipes 6A and 6B
and a directional control valve as a switching valve provided between the
The hydraulic pressure source side port of the directional control valve 7 is connected to the hydraulic pump 2 and the tank 3 via the pump side pipe 4 and the tank side pipe 5, and the motor side port is connected to the hydraulic motor 1 via the first and second main pipes 6A and 6B.
By tilting the operation lever 7A, the operator can change the position from the neutral position (A) to the left or right switching position (B),
(C), and the switching position (B) and the switching position (C)
This switches the direction of the pressure oil supplied to and discharged from the hydraulic pump 2 to the hydraulic motor 1. When the directional control valve 7 returns to the neutral position (A), it cuts off the lines between the pump-side line 4, the tank-side line 5 and the main lines 6A, 6B, stopping the supply and discharge of pressure oil to the hydraulic motor 1.

Numerals 8A and 8B indicate a pair of charge chuck valves located between the hydraulic motor 1 and the directional control valve 7 and connected to the main lines 6A and 6B. The check valves 8A and 8B are connected to the tank 3 via an auxiliary line 9 and a tank line 10, and when negative pressure occurs in the main line 6A or 6B during inertial rotation of the hydraulic motor 1, the hydraulic oil in the tank 3 is replenished into the main lines 6A and 6B.

Reference numerals 11A and 11B denote a pair of overload relief valves as pressure control valves located between the hydraulic motor 1 and the directional control valve 7 and connected to the main lines 6A and 6B. The overload relief valves 11A and 11B are connected to the tank 3 via an auxiliary line 9, etc.
The overload relief valve 11A is connected to the inlet side of the check valves 8A and 8B, and is also connected to the inlet side of the check valves 8A and 8B.
The valve opening pressure of overload relief valve 11A (11B) is preset to a first pressure value PC (see FIG. 12) by springs 12A, 12B. When excessive pressure exceeding the pressure value PC occurs in main line 6A (6B) during inertial rotation of hydraulic motor 1, overload relief valve 11A (11B) opens to relieve this excessive pressure to the other main line 6B (6A) etc. via check valve 8B (8A), thereby limiting the maximum pressure in main lines 6A, 6B to this pressure value PC.

13 denotes a drain line of the hydraulic motor 1,
When a part of the pressure oil supplied to the hydraulic motor 1 leaks as a drain, the leaked oil is discharged into the tank 3 by the drain 13 .

In the conventional technology configured as described above, when the directional control valve 7 is switched from the neutral position (A) to the switching position (B), pressure oil from the hydraulic pump 2 is supplied to the hydraulic motor 1 via the main line 6A, and the hydraulic motor 1 uses this pressure oil to drive the upper rotating body, which serves as an inertial body, to rotate, for example, in the right direction. Return oil from the hydraulic motor 1 is then sequentially discharged into the tank 3 via the main line 6B.

In addition, when the direction switching valve 7 is returned from the switching position (B) to the neutral position (A) in order to stop the upper rotating body in this state, the hydraulic motor 1 is turned on via the main line 6A from the hydraulic pump 2.
Since the supply of pressure oil to the upper rotating body is cut off, the driving force of the hydraulic motor 1 to the upper rotating body is released.

However, the upper rotating body, which acts as an inertial body, rotates the hydraulic motor 1 by its inertial force, so that the hydraulic motor 1 performs a pumping action and discharges the pressure oil in the main pipe 6A to the main pipe 6B side.
When the 6A side becomes negative pressure, the hydraulic oil in the tank 3 flows into the tank pipe.
10. Replenish the main line 6A through the check valve 8A.

As a result, a relatively large amount of pressure oil is confined between the hydraulic motor 1 and the directional control valve 7 in the main pipe 6B.
Braking pressure is generated in the main line 6B to stop the inertial rotation of the hydraulic motor 1. This braking pressure is applied to the overload relief valve 11B as shown by the characteristic line 14 in FIG.
When the pressure exceeds the pressure value PC, which is the valve opening pressure of the overload relief valve 11B, the overload relief valve 11B opens against the spring 12B, and the brake pressure in the main line 6B is released to the main line 6B via the auxiliary line 9 and the check valve 8A.
As a result, the inertial rotation of the hydraulic motor 1 is gradually braked, and then at time t1 (see FIG. 12), the overload relief valve 11B is opened by the spring 12B.
Then, the hydraulic motor 1 stops together with the upper rotating body.

In the above-mentioned prior art, braking pressure is generated in the main line 6B due to the inertial rotation of the hydraulic motor 1, causing the over-load relief valve 11B to open, braking the inertial rotation of the hydraulic motor 1. After this inertial rotation is braked, the over-load relief valve 11B opens at time t1 as shown in FIG.
When valve B opens and hydraulic motor 1 is temporarily stopped, the pressure in main line 6B is in a relatively high pressure state as shown by characteristic line 14. On the other hand, the pressure in main line 6A is in a low pressure state until time t1 as shown by characteristic line 15 in Figure 12.

For this reason, in the prior art, even if the hydraulic motor 1 temporarily stops at time t1 shown in Figure 12, a relatively large pressure difference ΔP may occur between the main pipes 6B and 4A as shown by characteristic lines 14 and 15. This pressure difference ΔP causes the hydraulic motor 1 to reverse in the direction opposite to the direction of rotation during inertial rotation, and begins to rotate in the reverse direction. At this time, pressure oil flows from the main pipe 6B side to the main pipe 6A side, causing the pressure difference ΔP
Gradually becomes smaller.

However, even if the differential pressure ΔP gradually decreases,
The reversal of the hydraulic motor 1 generates an inertial force in the reverse direction on the upper rotating body, which causes the hydraulic motor 1 to continue rotating in the reverse direction. As a result, the pressure on the main line 6A side of characteristic line 15 shown in Figure 12 becomes higher than the pressure on the main line 6B side of characteristic line 14 at time t2, and the hydraulic motor 1 once again reverses and rotates in the opposite direction.

In this way, in the prior art, the hydraulic motor 1
At time t1 when the overload relief valve 11B (11A) closes after opening, the differential pressure ΔP between the main lines 4A and 4B increases and the hydraulic motor 1 temporarily stops. However, the hydraulic motor 1 then starts rotating again in the opposite direction, and this reversing movement is repeated together with the upper rotating body, which is an inertial body.

In order to solve the above problem, for example, Japanese Patent Application Laid-Open No. 57-1982
Patent Publication No. 25570 and Japanese Patent Application Laid-Open No. 58-91902 propose providing separate inertia body anti-sway valves in the first and second main lines connected to a hydraulic motor. However, this requires providing separate inertia body anti-sway valves in the first and second main lines, which increases the number of parts, complicates piping work, and increases the size of the entire hydraulic circuit. Furthermore, this inertia body anti-sway valve uses a poppet-type valve element, which is essentially in line contact with the valve seat and only closes. Therefore, even a slight fluctuation in the driving pressure or braking pressure of the hydraulic motor causes the following problems:
There is a problem that the valve body may accidentally open, hindering the smooth driving or stopping of the inertial body.

Furthermore, Japanese Patent Laid-Open Publication No. 57-1803 discloses an inertial body reverse prevention device in which an on-off valve having a spool is provided between first and second main lines connected to a hydraulic motor, a pair of pressure chambers are formed on both ends of the spool in the on-off valve, and a pair of valve springs are provided in each pressure chamber to urge the spool toward a neutral position, and pressure (pilot pressure) from the pair of main lines is guided to each pressure chamber via a respective throttle.

In this case, however, the driving pressure of the hydraulic motor generated in one main line or the braking pressure generated in the other main line continues to act on the pressure chamber of the on/off valve as pilot pressure via the throttle. Therefore, even when the hydraulic motor is not reversed by the inertia body, the spool of the on/off valve may open against the valve spring if the pilot pressure from the main line becomes large.
The on/off valve disclosed in Publication No. 57-1803 has the problem that it interferes with the smooth driving or stopping of the inertial body, and does not necessarily improve the reliability or safety of the device for preventing reverse rotation of the inertial body.

The present invention has been made in consideration of the problems of the prior art described above, and aims to provide an inertial body drive device equipped with an inertial body anti-reversal valve that can effectively prevent the hydraulic motor from repeating reverse operation after the inertial rotation of the hydraulic motor has stopped, thereby reliably improving safety and reliability, preventing an increase in the number of parts, and simplifying the overall structure.

DISCLOSURE OF THE INVENTION In order to solve the above-mentioned problems, the present invention provides a hydraulic pressure source,
an inertial body drive device comprising a hydraulic motor that drives an inertial body by supplying or discharging pressure oil from the hydraulic power source; a switching valve connected to the hydraulic motor via first and second main lines, that supplies or discharges pressure oil from the hydraulic power source to the hydraulic motor when switched from a neutral position and stops the supply or discharge of pressure oil to the hydraulic motor when returned to the neutral position; and a pressure control valve that is located between the switching valve and the hydraulic motor and is provided between the first and second main lines, and that limits the maximum pressure in the first and second main lines to a preset first pressure value; valve means that is located between the first and second main lines and has a spool that slides between a valve open position and a valve closed position to connect or disconnect the first and second main lines, and that has biasing means that biases the spool toward the valve closed position and a hydraulic chamber that slides and displaces the spool from the valve closed position to the valve open position against the biasing means; The hydraulic control valve is configured to have a variable-capacity oil reservoir chamber connected to the hydraulic chamber of the valve means, and when the pressure in the main line, which is the higher-pressure side of the first and second main lines, falls below a preset second pressure value lower than the first pressure value set by the pressure control valve, pressurized oil supply means for pressurizing the oil in the oil reservoir chamber and supplying it to the hydraulic chamber of the valve means; and flow resistance means provided in the oil line connecting at least one of the oil reservoir chamber of the pressurized oil supply means and the hydraulic chamber of the valve means to a tank, for throttling the oil discharged to the tank via the oil line.

With this configuration, brake pressure is generated in one of the first and second main lines during inertial rotation of the hydraulic motor, and this brake pressure acts on the pressure control valve, causing the pressure in one of the main lines to increase to the first
When the pressure is limited to this value, the pressurized oil supply means stores oil in the oil reservoir chamber. In this state, the hydraulic chamber of the valve means is maintained at a low pressure without being supplied with oil, so the spool of the valve means is maintained in the closed position by the biasing means, and the first and second main lines can continue to be isolated.

When the brake pressure at this time drops below the first pressure value and the pressure control valve closes, the inertial rotation of the hydraulic motor is temporarily stopped. When the hydraulic motor starts to rotate in reverse, the brake pressure drops below the second pressure value, and the pressurized oil supply means pressurizes the oil in the oil reservoir chamber and supplies it to the hydraulic chamber of the valve means. As a result, the spool of the valve means slides to the open position against the biasing means.
By connecting the first and second main lines, the differential pressure between the first and second main lines can be reduced, thereby preventing the hydraulic motor from rotating in the reverse direction.

At this time, the spool is biased by the biasing means to the closed position on the hydraulic chamber side, causing the hydraulic fluid in the hydraulic chamber to be discharged to the tank through the oil passage, but this hydraulic fluid is throttled by the flow resistance means, so that the spool remains open for a relatively long time, and the differential pressure between the first and second main lines can be reliably reduced.

Furthermore, the spool of the valve means, which is slidably displaced by pressurizing the oil from the oil reservoir of the pressurized oil supply means and supplying it to the hydraulic chamber of the valve means, has a certain dead band between the closed and open positions. Therefore, even if the driving pressure or brake pressure of the hydraulic motor fluctuates, this pressure fluctuation can be absorbed by the oil reservoir of the pressurized oil supply means, and even if the capacity of the oil reservoir changes slightly, the spool will not slidably displace significantly, preventing accidental communication between the first and second main lines.

In this case, a high pressure selection valve is provided between the first and second main lines, located between the switching valve and the hydraulic motor, for selecting the higher pressure oil of the first and second main lines, and the pressurized oil supply means is provided with a pilot oil chamber to which the pressurized oil selected by the high pressure selection valve is supplied as a pilot pressure, and when the pilot pressure of the pilot oil chamber exceeds the second pressure value, the pressurized oil supply means sucks oil into the oil reservoir chamber to store it, and when the pressure falls below the second pressure value,
It is preferable that the oil stored in the oil reservoir chamber is pressurized and supplied to the hydraulic chamber of the valve means.

Furthermore, a check valve may be provided in the oil passage in parallel with the flow resistance means, and the check valve allows the oil in the tank to flow toward the hydraulic chamber and prevents flow in the opposite direction. This allows the oil in the tank to flow smoothly toward the hydraulic chamber or the oil sump chamber, and reliably prevents the hydraulic chamber or the oil sump chamber from becoming negative pressure.

On the other hand, in the inertial body drive device of the present invention, the valve means comprises a spool slide hole provided in a casing, a pair of ports formed in the casing at intervals in the axial direction of the spool slide hole and communicating with the first and second main lines, a spool slidably inserted into the spool slide hole, a throttle passage formed in the spool for communicating and blocking the pair of ports, a hydraulic chamber provided between one end of the spool and the casing and connected to a tank via the oil passage, and a valve spring as biasing means disposed between the other end of the spool and the casing for biasing the spool towards the hydraulic chamber. The pressurized oil supply means may comprise a piston sliding hole formed in a casing, a piston slidably inserted into the piston sliding hole, an oil sump chamber formed by the piston on one end side of the piston sliding hole and communicating with the hydraulic chamber, a set spring provided on the other end side of the piston so as to constantly urge the piston toward the oil sump chamber and having a spring force corresponding to the second pressure value, and a hydraulic pilot section provided on one end side of the piston so as to slide and displace the piston against the set spring so as to suck oil into the oil sump chamber when the pressure oil on the higher pressure side of the first and second main pipe lines exceeds the second pressure value, and the valve spring of the valve means may be configured to have a spring force weaker than that of the set spring of the pressurized oil supply means.

With this configuration, when the driving pressure or braking pressure of the hydraulic motor rises above the second pressure value, the hydraulic pilot portion of the pressurized oil supply means slidably displaces the piston against the setting spring, increasing the capacity of the oil sump chamber and drawing oil from the hydraulic chamber of the valve means into the oil sump chamber. As a result, the hydraulic chambers become low-pressure, and the spool in the casing can be slidably displaced toward the hydraulic chamber by the valve spring and held in the closed position, thereby reliably isolating the ports of the casing. Then, when the inertial rotation of the hydraulic motor temporarily stops and the braking pressure at this time falls below the second pressure value, the setting spring of the pressurized oil supply means pushes the piston toward the oil sump chamber, pressurizing the oil in the oil sump chamber and supplying it to the hydraulic chamber of the valve means. As a result, the spool of the valve means can be slidably displaced against the valve spring to the open position,
The ports are connected via a throttle passage formed in the spool, thereby preventing the inertial body from moving in reverse.

In this case, the valve means and the pressurized oil supply means are provided in the same casing, and the casing is provided with a spool slide hole and a pair of ports which constitute a part of the valve means, and is spaced from the ports and the spool slide hole,
It is preferable to provide a piston sliding hole that constitutes a part of the pressurized oil supply means.

A check valve may be provided in the oil passage in parallel with the flow resistance means, and the check valve may be configured to allow oil in the tank to flow toward the hydraulic chamber and prevent reverse flow.

Furthermore, in the inertial body drive device of the present invention, the valve means comprises a spool sliding hole provided in a casing, a pair of ports formed in the casing at intervals in the axial direction of the spool sliding hole and communicating with the first and second main lines, a spool formed as a cylindrical valve element inserted into the spool sliding hole and having an inner circumferential side that serves as a piston sliding hole, a throttle passage that connects or disconnects the pair of ports as the spool slides and displaces, and a hydraulic chamber formed between one end of the spool and the casing.
and a valve spring as a biasing means for biasing the spool toward the hydraulic chamber; the pressurized oil supply means may comprise a piston slidably inserted into a piston sliding hole of the spool, an oil sump chamber formed by the piston on one end side of the piston sliding hole and constantly communicating with the hydraulic chamber, a set spring disposed between the other end side of the piston and the casing so as to constantly bias the piston toward the oil sump chamber and having a spring force corresponding to the second pressure value, and a hydraulic pilot section provided on one end side of the piston for slidingly displacing the piston against the set spring so as to suck oil into the oil sump chamber when the pressurized oil on the higher pressure side of the first and second main lines exceeds the second pressure value; and the valve spring of the valve means may be disposed between the other end side of the spool and the piston, and have a spring force weaker than that of the set spring.

In this way, by inserting the spool formed as a cylindrical valve body into the spool sliding hole of the casing and inserting the piston into the piston sliding hole of the spool, the pressure-receiving area on the hydraulic pilot part side can be reduced, and the spring force of the set spring can be correspondingly reduced. Furthermore, the pressurized oil supply means can be housed compactly within the valve means, making it possible to reduce the overall size.

In this case, the oil reservoir chamber is preferably located radially inward of the hydraulic chamber and formed between one end of the piston and the casing.

In addition, it is preferable that the throttle passage be composed of at least a pair of oil holes drilled radially through the spool and spaced apart in the axial direction of the spool, and an oil groove formed on the outer periphery of the piston and extending in the axial direction to connect the oil holes.

In this case, when the spool and piston are displaced relative to each other, the oil holes in the spool are connected to or disconnected from the oil grooves in the piston, thereby increasing the dead zone between the open position, in which the pair of ports are connected via the throttle passage, and the closed position, in which the ports are disconnected, and reliably preventing the ports from inadvertently connecting to each other.

On the other hand, the hydraulic pilot portion is configured by a pilot oil chamber formed in the piston and positioned between the piston and a pusher slidably inserted into one end side of the piston,
The pilot oil chamber may be configured to be supplied with pressure oil from the higher pressure side of the first and second main lines via a high pressure selection valve.

The hydraulic pilot section may also be configured to include a pilot oil chamber formed within the casing and positioned between the pilot oil chamber and a pusher slidably inserted into one end of the casing, and the pilot oil chamber may be supplied with pressurized oil from the higher pressure side of the first and second main lines via a high pressure selection valve.

BRIEF DESCRIPTION OF THE DRAWINGS FIG. 1 is a hydraulic circuit diagram for swing provided with an inertia body anti-reverse valve according to a first embodiment of the present invention.

FIG. 2 is a vertical cross-sectional view of the inertial body anti-reversal valve in FIG. 1 with the spool returned to the valve-closed position.

FIG. 3 is a hydraulic circuit diagram for swinging provided with an inertia body anti-reversal valve according to a second embodiment.

FIG. 4 is a hydraulic circuit diagram for swinging provided with an inertia body anti-reversal valve according to a third embodiment.

FIG. 5 is a vertical cross-sectional view of the inertial body anti-reversal valve in phase 4 with the piston returned to its initial position.

FIG. 6 is a vertical cross-sectional view similar to FIG. 5, showing the piston in a state where it has reached the stroke end.

FIG. 7 is a vertical cross-sectional view similar to FIG. 5, showing the spool slidably displaced to the valve-open position.

FIG. 8 is a vertical cross-sectional view similar to FIG. 5, showing the spool being pushed back from the valve open position toward the hydraulic chamber.

FIG. 9 is a vertical cross-sectional view showing the inertial body anti-reversal valve according to the fourth embodiment with the piston returned to its initial position.

FIG. 10 is a vertical cross-sectional view showing the inertial body anti-reversal valve according to the fifth embodiment with the piston returned to its initial position.

FIG. 11 is a hydraulic circuit diagram for turning according to the prior art.

FIG. 12 is a pressure characteristic diagram of each main line showing a state in which the hydraulic motor repeats reverse operation.

BEST MODE FOR CARRYING OUT THE INVENTION Hereinafter, an embodiment of the present invention will be described with reference to Figures 1 to 10. In the embodiment, the same components as those in the prior art shown in Figure 11 are designated by the same reference numerals, and their description will be omitted.

1 and 2 show a first embodiment of the present invention.

In the figure, reference numeral 20 denotes an inertial body anti-reverse valve used in this embodiment, and the inertial body anti-reverse valve 20 comprises an anti-reverse valve body 21 as a valve means, which will be described later, and an oil suction/supply valve as a pressurized oil supply means.
The supply unit 32 is also included.

Reference numeral 21 denotes a reversal prevention valve body as a valve means disposed between the main pipes 6A and 6B of the hydraulic motor 1.
The valve casing 22 is configured with a two-port, two-position spool-type switching valve, and is provided in the same casing, a valve casing 22, together with a hydraulic fluid suction/supply device 32, which will be described later. As shown in FIG. 2, the valve casing 22 has a spool slide hole 22A extending in the axial direction, and annular oil grooves 22B, 22C formed on the outer periphery of the spool slide hole 22A and spaced apart by a predetermined distance in the axial direction.
and a pair of ports 23A, 23B that are always in communication with the oil grooves 22B, 22C.
The valve casing 22 is provided in a casing (not shown) of the hydraulic motor 1 together with the charge check valves 8A, 8B and the overload relief valves 11A, 11B, etc., and ports 23A, 23B of the valve casing 22 are connected to the main lines 6A, 6B between the hydraulic motor 1 and the directional control valve 7 via bypass lines 24A, 24B.

2, a piston slide hole 22D is formed in the valve casing 22 at a position separated from the spool slide hole 22A as a stepped hole, and the piston slide hole 22D constitutes a hydraulic fluid suction/supply device 32 together with a piston 31 and the like, which will be described later. The piston slide hole 22D is connected to a tank pipe line 2, which will be described later.
7, the spool slide hole 22A is connected to the valve casing 22 through a branch pipe 35. The valve casing 22 also has an annular step 22A1 on the other end side of the spool slide hole 22A.
22A1 and 22D1 are formed on the other end side of the piston slide hole 22D, and the annular steps 22A1 and 22D1 regulate the stroke ends of a spool 25 and a piston 31, which will be described later.

Reference numeral 25 denotes a spool inserted into the spool slide hole 22A, and the spool 25 is formed in a cylindrical shape with an outer diameter corresponding to the hole diameter of the spool slide hole 22A, and a pair of annular grooves 25A and 25B are provided on the outer periphery of the spool 25, spaced apart by a predetermined distance in the axial direction.
An "H"-shaped throttle passage 25C is formed between the spool 25 and the annular grooves 25A and 25B, and the throttle passage 25C communicates the ports 23A and 23B with the oil grooves 22B and 22C when the spool 25 is in the valve-open position. The spool 25 is normally biased in the valve-closing direction by a valve spring 30, which will be described later, and the annular stepped portion 22A1 of the valve casing 22
At the stroke end where it contacts the valve, the valve is in an open state.

Reference numeral 26 denotes a hydraulic chamber formed between one end face of the spool 25 and the valve casing 22, and the hydraulic chamber 26 is connected to a tank pipe as an oil passage in the tank 3 which contains hydraulic fluid.
The hydraulic chamber 26 is connected to the spool 25 via a branch line 35. The hydraulic chamber 26 is connected to the spool 25 via a branch line 35, and is also connected to an oil reservoir chamber 34, which will be described later. The hydraulic chamber 26 is supplied with pressurized oil and discharged, thereby slidingly displacing the spool 25 between the valve open position and the valve closed position.

Reference numeral 28 denotes a spring chamber formed between the other end face of the spool 25 and the valve casing 22. The spring chamber 28, together with a spring chamber 36 of the hydraulic fluid suction/supply device 32 (described later), is connected to the tank 3 via a common pipe 29 and is constantly filled with low-pressure hydraulic fluid.

30 is a spring chamber located between the other end face of the spool 25
The valve spring 30 serves as a biasing means and is disposed within the hydraulic chamber 28. The valve spring 30 biases the spool 25 with a relatively weak spring force.
26, causing the spool 25 to return to the valve-closed position shown in FIG.

Therefore, the anti-reverse valve body 21 as a valve means comprises a spool slide hole 22A provided in the valve casing 22, and a pair of ports 2A formed in the valve casing 22 at a distance in the axial direction of the spool slide hole 22A and communicating with the first and second main lines 6A and 6B.
3A, 23B, a spool 25 slidably inserted into the spool slide hole 22A, and a pair of ports formed in the spool 25.
A throttle passage 25C that connects and disconnects the spool 25A and 23B.
and the valve casing 22, and is connected to the tank 3 via a tank pipe 27 serving as an oil passage.
and a valve spring 30 which is disposed between the other end of the spool 25 and the valve casing 22 and serves as a biasing means for biasing the spool 25 towards the hydraulic chamber 26 side.

Reference numeral 31 denotes a piston that is inserted into the piston sliding hole 22D of the valve casing 22 and constitutes an oil suction/supply device 32 as pressurized oil supply means. The piston 31 is formed in a stepped cylindrical shape with a large diameter portion 31A and a small diameter portion 31B. The tip end surface of the small diameter portion 31B forms a pilot oil chamber 33 as a hydraulic pilot portion connected to a pilot pipe line 40, which will be described later, through the piston sliding hole.
The large diameter portion 31A of the piston 31 is located around the small diameter portion 31B, and the piston slide hole 22D is defined between the large diameter portion 31A and the small diameter portion 31B.
An annular oil reservoir 34 is defined between the piston 31 and the tank 3 through a branch line 35 and a tank line 27 as an oil passage.
The branch pipe 35 is always in communication with the hydraulic chamber 26.
At the branch point 35A between the tank line 27 and the hydraulic chamber 26,
It branches off from

Reference numeral 36 denotes another spring chamber formed between the large diameter portion 31A of the piston 31 and the valve casing 22, and the spring chamber 36 is a common pipe line.
The piston 31 is connected to the tank 3 via the valve 29 and is filled with low-pressure hydraulic oil. 37 denotes a setting spring disposed in the spring chamber 36 in contact with the large diameter portion 31A of the piston 31.
37 constitutes the oil suction/supply device 32 together with the piston 31, etc.
The piston 31 is constantly biased toward the pilot oil chamber 33.

Here, the setting spring 37 is the overload relief valve 11A,
The spring force is set to about 75 to 85% of the pressure value Pc that becomes the valve opening pressure of the valve 11B, and when the pilot pressure in the pilot oil chamber 33 exceeds a second pressure value (hereinafter referred to as the set pressure) of, for example, about 0.85×Pc, the piston 31 moves toward the annular step portion 22D1.
The oil suction and supply device 32 is then allowed to slide toward the side until it reaches the stroke end.
35, oil is sucked into the oil reservoir chamber 34 from the hydraulic chamber 26 and the tank 3 via the tank pipe line 27, and a relatively large amount of oil is filled and stored in the oil reservoir chamber 34.

When the pilot pressure in the pilot oil chamber 33 drops below the set pressure, the setting spring 37 presses the piston 31
The piston 31 is pushed toward the oil reservoir chamber 34, so that the oil in the oil reservoir chamber 34 is pressurized by the piston 31 and flows through the branch pipe 3.
5, and is supplied into the hydraulic chamber 26 via the tank pipe 27. At this time, some of the hydraulic fluid is discharged into the tank 3 via the tank pipe 27, but the flow of this hydraulic fluid is restricted by the throttle 38 provided in the tank pipe 27, so that a large amount of hydraulic fluid is supplied into the hydraulic chamber 26. The pressurized hydraulic fluid supplied into the hydraulic chamber 26 applies pressure to the end face of the spool 25, causing the spool 25 to slide toward the stroke end within the spool sliding hole 22A against the valve spring 30. As a result, the spool 25 communicates between the ports 23A and 23B via the throttle passage 25C, and the anti-reverse valve body 21 is switched from the closed position (a) to the open position (b) shown in FIG. 1.

Therefore, the oil suction/supply device 32 as a pressurized oil supply means
The valve casing 22 has a piston slide hole 22D, and the piston 3 is slidably inserted into the piston slide hole 22D.
1, an oil sump chamber 34 formed by the piston 31 on one end side of the piston sliding hole 22D and communicating with the hydraulic chamber 26, a set spring 37 provided on the other end side of the piston 31 so as to constantly urge the piston 31 toward the oil sump chamber 34 and having a spring force corresponding to a set pressure which is a second pressure value, and a pilot oil chamber provided on one end side of the piston 31 and serving as a hydraulic pilot portion which slides and displaces the piston 31 against the set spring 37 so as to suck oil into the oil sump chamber 34 when the pressure oil on the higher pressure side of the first and second main pipe lines 6A, 6B exceeds the set pressure.
It consists of 33 parts.

Reference numeral 38 denotes a throttle as a flow resistance means provided in the tank pipe line 27 between the branch point 35A of the branch pipe line 35 and the tank 3. When the oil in the hydraulic chamber 26 flows out into the tank 3 through the tank pipe line 27, the throttle 38 exerts a throttling action on the oil to restrict the outflow flow rate, thereby:
This extends the time it takes for the spool 25 to return from the stroke end to the closed position. This causes a predetermined time delay before the anti-reverse valve body 21 returns from the open position (b) shown in Figure 1 to the closed position (a), and the main lines 6A, 6B
B are kept in communication for a relatively long time via bypass pipes 24A, 24B and ports 23A, 23B.

Furthermore, reference numeral 39 denotes a shuttle valve serving as a high-pressure selection valve disposed between the main lines 6A and 6B between the hydraulic motor 1 and the directional control valve 7. The shuttle valve 39 connects the main line 6A or 6B on the high-pressure side of the main lines 6A and 6B to the pilot oil chamber 33 via a pilot line 40, and supplies high-pressure oil into the pilot oil chamber 33 as pilot pressure.

The hydraulic circuit for swinging of the hydraulic excavator according to this embodiment has the above-mentioned configuration, and by switching the direction switching valve 7 from the neutral position (A) to the switching positions (B) and (C), the hydraulic motor 1 drives the upper swing body as an inertial body to swing left and right, or by switching the direction switching valve 7 from the neutral position (A)
There is no particular difference from the prior art in that the rotation of the hydraulic motor 1 is stopped by returning the hydraulic motor 1 to the normal position.

Therefore, the reverse prevention operation of the inertial body reverse prevention valve 20, which is a feature of this embodiment, will be described.

First, when the directional control valve 7 is switched from the neutral position (A) to the switching position (B) and the upper rotating body is rotated by the hydraulic motor 1, high-pressure oil flowing through the main pipe 6A is supplied from the shuttle valve 39 to the pilot oil chamber 33 as a pilot pressure higher than the set pressure. This pilot pressure causes the piston 31 of the oil suction/supply device 32 to slide to the stroke end against the set spring 37, increasing the capacity of the oil sump chamber 34. This causes the oil suction/supply device 32 to suck oil from the oil chamber 26, the tank 3, the tank pipe 27, and the branch pipe 35 into the oil sump chamber 34, filling it with oil and storing it. The spool 25 in the spool sliding hole 22A remains biased toward the oil chamber 26 by the valve spring 30, and the anti-reverse valve body 21 is held in the closed position (a) shown in FIG. 1.

Next, when the directional control valve 7 is returned from the switching position (B) to the neutral position (A) during the rotation of the upper rotating body, the hydraulic motor 1 is rotated by inertia due to the upper rotating body, which serves as an inertial body.
For example, braking pressure occurs in the main line 6B, and the over-load relief valve 11B is opened by the braking pressure at this time. At this time, the pressure in the main line 6B has risen to the pressure value Pc of the over-load relief valve 11B, so
Pressure oil flows into the main line 6B from the shuttle valve 39 as a pilot pressure higher than the set pressure of the setting spring 37.
33. Therefore, the oil suction and supply device 32 is
As in the previous case, the piston 31 is slid to the stroke end to keep a large amount of oil in the oil reservoir chamber 34, and the anti-reverse valve body 21 remains in the closed valve position (a).

In this case, for example, at the moment when the pressure of the pressurized oil selected by the shuttle valve 39 is switched from the driving pressure on the main line 6A side to the brake pressure on the main line 6B side, the pilot pressure supplied to the pilot oil chamber 33 may momentarily drop below the set pressure of the setting spring 37. As a result, the oil reservoir chamber 34 of the oil suction/supply device 32
A small amount of oil is supplied to the hydraulic chamber 26 from the spool sliding hole.
The spool 25 moves slightly upward within 22A.
Since a dead zone is provided in the spool 25 between the oil grooves 22B, 22C and the annular grooves 25A, 25B, the main lines 6A, 6B will not be inadvertently connected to each other.

Then, after the inertial rotation of the hydraulic motor 1 is braked by opening the overload relief valve 11B, the relief valve 11B
When the valve 6B is closed, the inertial rotation of the hydraulic motor 1 is temporarily stopped. At this time, a pressure difference ΔP occurs between the main lines 6B and 6A as shown in Figure 12, and this pressure difference ΔP causes the hydraulic motor 1 to try to rotate in reverse. However, when the hydraulic motor 1 starts to rotate in reverse, the pressure in the main line 6B leaks from the hydraulic motor 1 and is transferred to the tank 3 via the drain line 13.
Therefore, the pressure in the main line 6B is lower than the pressure value Pc of the overload relief valve 11B by, for example, about 75 to 85%.

As a result, the pilot pressure in the pilot oil chamber 33 drops below the set pressure of the setting spring 37, and the oil suction/supply device 32 pushes the piston 31 toward the oil reservoir chamber 34 by the setting spring 37. As a result, the piston 31 pressurizes the oil in the oil reservoir chamber 34 and supplies it into the hydraulic chamber 26 via the branch line 35 and the tank line 27, and at this time the orifice 38 prevents this oil from being discharged into the tank 3.
The oil supplied into 26 slides and displaces the spool 25 in the spool slide hole 22A toward the stroke end against the valve spring 30. At this time, the spool 25 communicates between the ports 23A and 23B via the throttle passage 25C, so that the anti-reverse valve body 21 moves from the valve closed position (a) shown in FIG.
It will be possible to switch to

Then, the valve spring 30 is attached to the spool 25 of the anti-reverse valve body 21.
1 的时间可以延长，由此可以通过将开始反向开位置(b)向开关闭位置(a)，使其电流可能经过程24A、24B及端口(b)23A、23B等来进行进行进行。 When the hydraulic chamber 26 is pushed from the stroke end toward the hydraulic chamber 26 by the force of ...

Thus, according to this embodiment, when the inertial rotation of the hydraulic motor 1 stops, the spool 25 of the anti-reverse valve body 21 can be held at the open position (b) for a relatively long time before it returns from the open position (b) to the closed position (a) by the valve spring 30. As a result, for example, high pressure in the main line 6B can be released to the main line 6A side while providing a throttling effect via the throttling passage 25C of the spool 25, etc., and the differential pressure ΔP (see FIG. 12) generated between the main lines 6A, 6B can be reliably reduced as described in the prior art, and the hydraulic motor 1 can be effectively prevented from repeating reverse rotation.

When the spool 25 is gradually pushed toward the hydraulic chamber 26 by the valve spring 30 and reaches the valve-closed position shown in Figure 2, the space between the ports 23A and 23B is separated from the annular grooves 25A and 25B of the spool 25 and is blocked by the outer peripheral surface of the spool 25.
The main line 6 through the inertial body anti-reverse valve body 21 of the anti-reverse valve 20
Communication between A and 6B is cut off, the hydraulic motor 1 can be reliably held in a stopped state, and the anti-reverse valve body 21 can be reliably prevented from opening the next time the hydraulic motor 1 is driven.

That is, the spool 25 is inserted into the spool sliding hole of the valve casing 22.
The hydraulic motor 1 slides and displaces within the annular grooves 25A and 25B until the annular grooves 25A and 25B communicate with the oil grooves 22B and 22C of the valve casing 22. A dead zone having a certain stroke is set.
When the hydraulic motor 1 is driven or stopped, the driving pressure and braking pressure of the hydraulic motor 1 fluctuate, and the hydraulic fluid suction/supply device 32 flows into the pilot oil chamber 33.
When the pilot pressure acting on the valve body fluctuates, a small amount of oil may be supplied from the oil reservoir chamber 34 to the hydraulic chamber 26 of the valve body 21. However, even in such a case,
The spool 25 only moves slightly upward in the spool sliding hole 22A, and the annular grooves 25A, 25B can be kept shut off from the oil grooves 22B, 22C, and the anti-reverse valve body 21 can be reliably prevented from opening inadvertently.

The above explanation is for the case where there is a large amount of pressure oil leaking from the hydraulic motor 1. However, if the amount of pressure oil leaking from the hydraulic motor 1 is small, when the inertial rotation of the hydraulic motor 1 stops, the hydraulic motor 1 will not be able to move in the main lines 6A, 6B.
However, in this case, even a slight reversal of the hydraulic motor 1 causes the pressure (brake pressure) in the main line 6B (6A), which was in a high state before the reversal, to immediately drop below the set pressure of the set spring 37. As a result, hydraulic fluid is supplied to the hydraulic chamber 26 by the displacement of the piston 31, and the spool 25 establishes communication between the main lines 6A and 6B, immediately preventing the hydraulic motor 1 from reversing any further.

Therefore, according to this embodiment, after the inertial rotation of the hydraulic motor 1 has stopped, pressurized hydraulic fluid is supplied from the oil reservoir chamber 34 of the hydraulic fluid suction/supply device 32 to the hydraulic chamber 26, and the anti-reverse valve body
By sliding the spool 25 of 21 to the open position, the hydraulic motor 1 can be effectively prevented from reversing, and the hydraulic motor 1 can be quickly stopped together with the inertial body such as the upper rotating body.

Furthermore, since only one inertial body check valve 20, which is composed of the check valve main body 21, which is a two-port, two-position spool valve, and the oil suction/supply device 32, needs to be installed between the main lines 6A and 6B via a shuttle valve 39 or the like, the inertial body check valve 20 can be easily incorporated into, for example, the casing of the hydraulic motor 1, preventing the entire hydraulic circuit from becoming complicated. Furthermore, since it is only necessary to arrange the spool 25, piston 31, etc. inside the valve casing 22, the inertial body check valve 20 can be made compact, thereby achieving miniaturization and simplifying the structure of the entire hydraulic circuit.

Furthermore, the reversal prevention valve body 21 is provided with a dead zone with a certain stroke amount by using a spool 25, and the oil reservoir chamber 34 and the hydraulic chamber 26 of the oil suction/supply device 32 are connected to the tank 3 via a tank line 27, a branch line 35, and a throttle 38, thereby forming a separate hydraulic closed circuit separated from the hydraulic motor 1 and the main lines 6A, 6B. Then, by supplying and discharging oil from the oil reservoir chamber 34 of the oil suction/supply device 32 to the hydraulic chamber 26, the reversal prevention valve body 21
The spool 25 is adapted to be slidably displaced.
As a result, even if the driving pressure or braking pressure of the hydraulic motor 1 fluctuates, this pressure fluctuation can be effectively absorbed by the oil reservoir chamber 34 of the oil suction/supply device 32, etc., and the spool 25 is directly affected by this pressure fluctuation, and the anti-reverse valve body
This can reliably prevent valve 21 from opening inadvertently.

Next, FIG. 3 shows a second embodiment of the present invention. In this embodiment, the same components as those in the first embodiment are given the same reference numerals and their explanation will be omitted. The feature of this embodiment is that a bypass line 41 is provided in the tank line 27 to connect the upstream and downstream sides of the throttle 38, and a check valve 42 is provided in the bypass line 41 in parallel with the throttle 38.

Here, the check valve 42 allows the oil in the tank 3 to flow toward the hydraulic chamber 26 and the oil sump chamber 34. For this reason, when, for example, a pilot pressure exceeding the set pressure of the setting spring 37 is supplied to the pilot oil chamber 33 of the oil suction/supply device 32 and the capacity of the oil sump chamber 34 increases, the check valve 42 smoothly sucks the oil in the tank 3 into the oil sump chamber 34, preventing the oil sump chamber 34 from becoming negative pressure. Therefore, when the capacity of the oil sump chamber 34 increases, the oil in the hydraulic chamber 26 is sucked into the oil sump chamber 34, preventing the oil sump chamber 34 from becoming negative pressure.
It is also possible to prevent the inside from becoming negative pressure.

Then, after the inertial rotation of the hydraulic motor 1 has stopped temporarily, when the pilot pressure acting on the pilot oil chamber 33 of the oil suction/supply device 32 drops below the set pressure of the set spring 37, the piston 31 is pushed by the set spring 37, pressurizing the oil in the oil reservoir chamber 34 and sending it out toward the hydraulic chamber 26 and the tank 3 via the branch pipe 35, etc. However,
At this time, the check valve 42 closes and prevents the flow of oil through the check valve 42, so that the flow of oil to the tank 3 side can be restricted by the orifice 38, and oil from the oil reservoir chamber 34 can be reliably supplied into the hydraulic chamber 26.

Furthermore, even when the oil in the hydraulic chamber 26 flows out toward the tank 3 thereafter, the check valve 42 continues to close.
The oil from the hydraulic chamber 26 flows into the tank 3 only through the throttle 38, and the throttle 38 allows the anti-reverse valve body 21
The speed at which the valve returns from the open position (b) to the closed position (a) can be slowed down.

Thus, with this embodiment configured as above, it is possible to obtain substantially the same effects as the first embodiment, but in particular, in this embodiment, since the check valve 42 is provided in parallel with the orifice 38, it is possible to effectively prevent negative pressure from building up in the hydraulic chamber 26 and the oil sump chamber 34, which would otherwise cause cavitation, and it is possible to more reliably compensate for the anti-reverse action of the hydraulic motor 1 by the inertia body anti-reverse valve 20.

In the first and second embodiments, the spool 25 has an "H"-shaped throttle passage 2 located between the annular grooves 25A and 25B.
However, instead of this, the throttle passage 25C may be formed as a "U"-shaped passage that connects the annular grooves 25A and 25B, or a throttle passage of another shape may be formed.
Instead of B, an oil groove extending in the axial direction may be formed, thereby forming a throttle passage that connects and disconnects the oil grooves 22B, 22C of the valve casing 22.

Furthermore, in the first and second embodiments, the hydraulic chamber 26 and the oil reservoir chamber 34 of the oil suction/supply device 32 are described as being connected to each other via the tank line 27 and the branch line 35. However, instead of this, the hydraulic chamber 26 and the oil reservoir chamber 34 may be connected directly to each other without using the tank line 27 or the like. In this case, for example, the branch line 35 may be omitted.

Next, FIGS. 4 to 8 show a third embodiment of the present invention. The feature of this embodiment is that the spool provided in the spool sliding hole of the casing is formed as a cylindrical valve body, a piston sliding hole is formed on the inner peripheral side of the spool, and the piston of the pressurized oil supply means is inserted into the piston sliding hole, thereby enabling the pressurized oil supply means to be housed compactly within the casing of the valve means.

In the figure, 50 indicates an inertia body anti-reverse valve according to this embodiment,
The inertial body anti-reverse valve 50 comprises an anti-reverse valve body 51 as valve means, which will be described later, and an oil suction/supply device 59 as pressurized oil supply means.

Reference numeral 51 denotes a reverse-reversal valve body as a valve means disposed between the main pipes 6A and 6B of the hydraulic motor 1.
As shown in FIG. 5, the valve casing 52 includes a spool slide hole 52A extending in the axial direction, and a spool slide hole 52B formed on the outer periphery of the spool slide hole 52A.
Annular oil grooves 52B spaced apart by a predetermined distance in the axial direction of 2A,
Notches 52D, 52D are formed in the oil grooves 52B, 52C, and the oil grooves 52B, 52C are always in communication with a pair of ports 53A, 53B through which pressure oil flows in and out.
The valve casing 52 has an annular groove 52E formed at the left end of the spool slide hole 52A, and the annular groove 52E is connected to a hydraulic chamber 57 (described later).
The annular groove 52E and the oil groove 52B
Between these, a pilot pipe line 70 (described later) communicating with the spool slide hole 52A is formed in order to supply pilot pressure to a pusher 65 (described later).

Here, the valve casing 52 is integrally formed with the casing (not shown) of the hydraulic motor 1 together with the charge check valves 8A, 8B and the overload relief valves 11A, 11B, etc.
It is compactly housed in the middle of the hydraulic circuit for turning.
Ports 53A and 53B of the valve casing 52 are connected between main lines 6A and 6B via bypass lines 54A and 54B.

Reference numeral 55 denotes a cylindrical spool inserted into the spool slide hole 52A. The spool 55 constitutes the anti-reverse valve body 51 together with a hydraulic chamber 57 and a valve spring 64 (to be described later), and communicates between ports 53A and 53B (main lines 6A and 6B) via a throttle passage 56 (to be described later).
The spool 55 is formed as a cylindrical valve element having an outer dimension corresponding to the diameter of the spool sliding hole 52A, and a piston sliding hole 55A is formed on the inner circumferential side thereof. The spool 55 also has an annular recessed groove 55B on the left side thereof, spaced apart from the annular groove 52E of the valve casing 52 and always communicating with the pilot pipe 70.
A pair of radial oil holes 55C, 55C communicating with the recessed groove 55B are formed in the right end of the spool 55. On the other hand, a spring receiving portion 55D having a stepped shape is formed in the right end of the spool 55.

The spool 55 also has throttle holes 55D, each consisting of two pairs of axial holes, positioned between the recessed groove 55B and the spring receiving portion 55D and spaced apart from each other in the axial direction by a distance corresponding to each of the notch portions 52D.
The throttle holes 5E, 55E are drilled in the radial direction.
5E is a port 53A of the valve casing 52 together with an oil groove 58D described later.
3B in the open position shown in FIG.

Reference numeral 57 denotes an annular hydraulic chamber formed between the left end face of the spool 55 and the valve casing 52. Pressurized hydraulic fluid is supplied to this hydraulic chamber 57 from an oil reservoir chamber 60 of a hydraulic fluid suction/supply device 59 (described later), and the pressure acts on the end face of the spool 55, causing the spool 55 to slide against the valve spring 64 from the valve closed position shown in FIG. 6 to the valve open position shown in FIG. 7.

Therefore, the anti-reverse valve body 51 as a valve means comprises a spool slide hole 52A provided in a valve casing 52, and a pair of ports 5A, 5B formed in the valve casing 52 at intervals in the axial direction of the spool slide hole 52A and communicating with the first and second main lines 6A, 6B.
3A, 53B, a spool 55 formed as a cylindrical valve element inserted into the spool sliding hole 52A and having a piston sliding hole 55A on its inner periphery, a throttle passage 56 which opens and closes communication between the pair of ports 53A, 53B as the spool 55 slides and displaces, a hydraulic chamber 57 formed between one end of the spool 55 and the valve casing 52, and a valve spring 64 which serves as biasing means for biasing the spool 55 towards the hydraulic chamber 57.

Reference numeral 58 denotes a piston inserted into the piston sliding hole 55A of the spool 55, and the piston 58, together with an oil reservoir chamber 60 and a setting spring 63 (described later), constitutes an oil suction/supply device 59 as pressurized oil supply means.

The piston 58 is formed as a stepped rod-like body longer than the spool 55, and the spool is inserted into the valve casing 52.
55. The piston 58 is provided, at its left side, with an annular groove 58A that always communicates with each oil hole 55C in the spool 55, a radial oil hole 58B that always communicates with the groove 58A, and a pusher slide hole 58C that is drilled in the axial direction from the oil hole 58B toward the left end face and slidably supports a pusher 65 (described below). The piston 58 is further formed with an annular oil groove 58D that extends in the axial direction at a position spaced a predetermined distance to the right of the groove 58A, and the oil groove 58D, together with the throttle holes 55E in the spool 55, constitutes a throttle passage 56.

On the other hand, the right side of the piston 58 protrudes into the valve casing 52 through the piston slide hole 55A of the spool 55.
The portion axially facing the spring receiving portion 55D is an annular stepped portion 58E.
When the spool 55 is slidably displaced to the valve-open position shown in FIG.
The piston 58 has a spring receiving portion 55D which contacts the tip of the spring receiving portion 55D of the spool 55 to regulate the stroke end of the spool 55.
A spring receiving portion 58F is formed projecting radially from the right side portion of the piston 58E, and one end of a setting spring 63 is engaged with the spring receiving portion 58F.
A rod-shaped stopper portion 58G extends inward in the axial direction,
The stopper portion 58G restricts the stroke end of the piston 58 by contacting the end face of the spring chamber 61 as shown in FIG.

Reference numeral 60 denotes a variable-capacity oil reservoir chamber located radially inside the hydraulic chamber 57 and formed between the left end face of the piston 58 and the valve casing 52. The oil reservoir chamber 60 is radially connected to the hydraulic chamber 57 and is connected to the tank 3 via a tank pipe line 67, which will be described later. As shown in FIG. 6, the volume of the oil reservoir chamber 60 increases when the piston 58 slides and displaces toward the stroke end, and at this time the oil reservoir chamber 60 draws in and stores oil from the tank 3 via the hydraulic chamber 57.
When the piston 58 slides toward the oil reservoir chamber 60 as shown in FIG. 7, the volume of the oil reservoir chamber 60 decreases, and the oil stored therein is pressurized by the piston 58 and supplied to the hydraulic chamber 57.

Reference numeral 61 denotes a spring chamber formed in the valve casing 52, located to the right of the spool 55 and piston 58, and the spring chamber 61 is connected to the tank 3 via a pipe 62 and is filled with hydraulic oil from the tank 3.

Reference numeral 63 denotes a setting spring disposed in the spring chamber 61 around the stopper portion 58G of the piston 58, and the tip side of the setting spring 63 abuts against the spring receiving portion 58F of the piston 58, and constantly urges the piston 58 toward the hydraulic chamber 57, thereby returning the piston 58 to the initial position shown in FIG.
The setting spring 63 is set to a spring force of about 75 to 85% of the first pressure value PC, which is the valve opening pressure of the overload relief valves 11A and 11B, and when the pilot pressure from the pilot line 70 exceeds a second pressure value (hereinafter referred to as the set pressure) which is preset to, for example, about 0.85 × PC, the piston 58
is allowed to be pushed by the pusher 65 to the stroke end shown in FIG.

64 is located in the spring chamber 61 and is a spring receiving portion 55 of the spool 55.
The valve spring 64 is set to a spring force weaker than that of the set spring 63, and constantly urges the spool 55 toward the hydraulic chamber 57. The valve spring 64 is set to a spring force weaker than that of the set spring 63, and constantly urges the spool 55 toward the piston 58 in the spool sliding hole 52A of the valve casing 52.
When the sprue 55 slides toward the spring chamber 61 and assumes the valve open position as shown in FIG. 7, the ports 53A and 53B of the valve casing 52 communicate with each other via the throttle passage 56.

65 is inserted into the pusher slide hole 58C of the piston 58,
The pusher 65 has a small diameter and its tip end projects into the oil reservoir chamber 60. The pusher 65 defines a pilot oil chamber 66 as a hydraulic pilot section between the oil hole 58B of the piston 58 and the pusher slide hole 58C.
Pressure oil on the higher pressure side of the main lines 6A and 6B is supplied as pilot pressure to the line 66 from a shuttle valve 69, which will be described later, via a pilot line 70 and the like.

When the pilot pressure in the pilot oil chamber 66 exceeds the set pressure of the setting spring 63, the pusher 65 is pressed toward the oil reservoir chamber 60 by the pilot pressure at this time and protrudes, causing the piston 58 to slide against the setting spring 63 to the stroke end shown in FIG.
The capacity of the oil reservoir 60 increases as shown in FIG. 6 according to the stroke of the piston 58, and the oil reservoir 60 is connected to the tank 3 through a tank pipe line.
67, which absorbs and stores hydraulic oil through the hydraulic chamber 57;

Therefore, the oil suction and supply device 59 as a pressurized oil supply means
The piston 58 is slidably inserted into the piston sliding hole 55A of the spool 55, an oil reservoir chamber 60 is formed at one end of the piston sliding hole 55A by the piston 58 and is always in communication with the hydraulic chamber 57, and the other end of the piston 58 is connected to the valve casing 52 so as to always bias the piston 58 toward the oil reservoir chamber 60.
and a setting spring 63 having a spring force corresponding to the second pressure value, which is provided on one end side of the piston 58.
1. A pilot oil chamber 66 as a hydraulic pilot section which slides and displaces the piston 58 against the set spring 63 so as to suck oil into the oil reservoir chamber 60 when the pressure oil on the high-pressure side of the second main lines 6A, 6B exceeds the set pressure.

Reference numeral 67 denotes a tank pipe line serving as an oil passage connecting the hydraulic chamber 57 to the tank 3, and 68 denotes a throttle provided midway along the tank pipe line 67 as flow resistance means, and the throttle 68 acts to throttle the hydraulic fluid (working oil) discharged from the oil sump chamber 60 to the tank 3 via the hydraulic chamber 57 when the piston 58 is pushed back by the set spring 63 from the stroke end shown in Figure 6 toward the oil sump chamber 60 as shown in Figure 7. As a result, pressurized hydraulic fluid is supplied from the oil sump chamber 60 to the hydraulic chamber 57, and the pressure within the hydraulic chamber 57 at this time causes the spool 55 to slide toward the spring chamber 61 against the valve spring 64.

When the piston 58 slides from the initial position toward the stroke end, the tank pipe 67
The hydraulic oil in the oil reservoir chamber 60 is supplied so as to be drawn into the oil reservoir chamber 60.
At this time, the oil reservoir 60 tends to have a negative pressure due to the throttle 68. However, when replenishing the oil, the throttle 68
5. The oil reservoir chamber 60 can be instantly restored from a negative pressure tendency without substantially affecting the operation of the piston 58.

As a result, when the piston 58 is pushed toward the spring chamber 61 by the pusher 65, the tank 3 is filled in the oil reservoir chamber 60.
The hydraulic fluid (hydraulic oil) is smoothly sucked (supplied) from the piston 58, and the piston 58 quickly slides from the initial position shown in Fig. 5 to the stroke end shown in Fig. 6. However, the piston 58 is pushed back to the initial position by the set spring 63, and the pressure in the oil reservoir chamber 60 and the hydraulic chamber 57 rises, and the spool 55 is pushed toward the spring chamber 61 as shown in Fig. 7.
When the valve spring 64 pushes the valve 55 back to the hydraulic chamber 57,
The throttle 68 throttles the oil discharged from the hydraulic chamber 57. Therefore, the throttle 68 keeps the sliding speed (return speed) of the spool 55 low, and the spool 55 maintains the open position shown in Figures 7 and 8 for a relatively long period of time.

Furthermore, 69 denotes a shuttle valve as a high-pressure selection valve disposed between the main lines 6A and 6B between the hydraulic motor 1 and the directional control valve 7. The shuttle valve 69 connects the high-pressure side main line 6A or 6B of the main lines 6A and 6B to a pilot line 70.
High pressure oil is supplied as a pilot pressure from a pilot line 70 to the pilot oil chamber 66 via the groove 55B of the spool 55, the oil holes 55C, and the groove 58A and oil hole 58B of the piston 58.

The inertial body anti-reverse valve 50 used in this embodiment has the above-described structure, and its anti-reverse operation will now be described.

First, when the directional control valve 7 is switched from the neutral position (A) to the switching position (B) and the upper rotating body is rotated by the hydraulic motor 1, the high-pressure oil flowing in the main pipe 6A is supplied from the shuttle valve 69 to the piston 58 as a high-pressure pilot pressure.
The pilot oil pressure is supplied to a pilot oil chamber 66 inside the piston 58 via a pilot line 70. The pusher 65 is pushed toward the oil reservoir chamber 60 by this pilot pressure, causing the piston 58 to protrude into the oil reservoir chamber 60. As a result, a pressure-receiving area difference created by the pusher 65 inside the pilot oil chamber 66 causes the piston 58 to slide from the initial position shown in Figure 5 to the stroke end shown in Figure 6 against the set spring 63.

At this time, the volume of the oil reservoir 60 increases in accordance with the stroke of the piston 58 as shown in FIG.
Hydraulic oil is supplied from the tank 3 through the tank line 67 to fill the oil chamber 60. At this time, the hydraulic chamber 57 and the oil reservoir chamber 60 tend to become negative pressure, so the spool 55 is held in contact with the end face of the valve casing 52 on the hydraulic chamber 57 side, and does not slide against the valve spring 64 toward the spring chamber 61.

Next, when the directional control valve 7 is returned from the switching position (B) to the neutral position (A) during the rotation of the upper rotating body, the hydraulic motor 1 is rotated by inertia due to the upper rotating body, which serves as an inertial body.
Brake pressure is generated in the main line 6B. When the over-load relief valve 11B is opened by this brake pressure, the pressure in the main line 6B rises to a first pressure value PC, which is the valve opening pressure of the over-load relief valve 11B. The high-pressure oil generated at this time is supplied as pilot pressure from the shuttle valve 69 to the pilot oil chamber 66 in the piston 58 via the pilot line 70, etc. Thus, the pusher 65 continues to protrude into the oil reservoir chamber 60, causing the piston 58 to slide and displace to the stroke end as shown in Figure 6, and increasing the capacity of the oil reservoir chamber 60, filling the oil reservoir chamber 60 with oil. In this state, oil flows into the hydraulic chamber 66.
57 side into the oil reservoir chamber 60,
The spool 55 remains in contact with the end face of the valve casing 52 on the hydraulic chamber 57 side.

In this case, for example, the moment the pressure of the pressurized oil selected by the shuttle valve 69 is switched from the drive pressure on the main line 6A side to the brake pressure on the main line 6B side, the pilot pressure in the pilot oil chamber 66 may drop momentarily. However, this momentary drop in pilot pressure causes the piston 58 to move from the position shown in Figure 6 to the oil reservoir chamber 60 side (left side), and a small amount of oil is supplied from the oil reservoir chamber 60 into the hydraulic chamber 57, causing the spool 55 to pressurize the valve spring 64.
However, even in such a case, a relatively large dead zone is provided between the throttle hole 55E of the spool 55 and the oil groove 58D of the piston 58 in the state shown in Figure 6, so that the main lines 6A and 6B will not be inadvertently connected to each other.

Then, after the inertial rotation of the hydraulic motor 1 is braked by opening the overload relief valve 11B, the relief valve 11B
When the valve is closed, the inertial rotation of the hydraulic motor 1 is temporarily stopped. However, at this time, the pressure in the main pipe 6B leaks from the hydraulic motor 1, and this leaked oil flows into the drain pipe.
Since the pressure in the main line 6B is higher than the pressure value PC of the overload relief valve 11B,
For example, the pressure will be about 75 to 85% lower.

As a result, the pilot pressure acting on the pilot oil chamber 66 from the main line 6B via the shuttle valve 69, the pilot line 70, etc., also drops. Therefore, the pusher 65 allows the piston 58 to be pushed by the set spring 63 toward the oil reservoir chamber 60 from the stroke end shown in FIG. 6. Then, when the piston 58 moves from the stroke end to the oil reservoir chamber 60,
When the oil reservoir chamber 60 is pushed to the side, the volume of the oil reservoir chamber 60 decreases, and the oil in the oil reservoir chamber 60 is discharged to the tank 3 via the hydraulic chamber 57 and the tank pipe line 67.

However, when the oil in the oil reservoir chamber 60 is discharged to the tank 3 via the hydraulic chamber 57 and the tank pipe line 67, the restrictor 68
This provides a throttling action, restricting the flow rate of the oil discharged from the hydraulic chamber 57 toward the tank 3.
Pressurized oil is supplied into the hydraulic chamber 57, and the spool
55 is pushed toward the spring chamber 61 against the valve spring 64 by the pressure of the hydraulic chamber 57. Then, the spool 55 is brought to the valve open position shown in FIG. 7, and oil flows through the ports 53A and 53B (bypass line 58) via the throttle passage 56 consisting of the throttles 55E and the oil groove 58D of the piston 58.
At this time, the stroke end of the piston 58 is restricted at the position where the tip of the spring receiving portion 55D of the spool 55 abuts against the annular stepped portion 58E of the piston 58.

When the piston 58 is returned to its initial position as shown in FIG. 7, the spool 55 is pushed toward the hydraulic chamber 57 by the valve spring 64, and the oil in the hydraulic chamber 57 is discharged from the tank line 67 through the orifice 68 to the tank 3.
As a result, the spool 55 is moved by the restrictor 68 as shown in FIG.
5B. The return speed when the spool 55 slides from the valve open position shown in FIG. 5B toward the hydraulic chamber 57 can be slowed down, and the throttle holes 55E of the spool 55 can be adjusted.
The valve open time during which communication between ports 53A, 53B (bypass lines 54A, 54B) can be extended via throttle passage 56 formed of oil groove 58D of piston 58. As shown in Figure 5, spool 55 reaches the valve closed position when each throttle hole 55E passes the position of each notch 52D connected to oil grooves 52B, 52C in valve casing 52, and spool 55 can communicate between bypass lines 54A, 54B from the valve open position shown in Figure 7 to the position shown in Figure 5.

Thus, according to this embodiment, when the inertial rotation of the hydraulic motor 1 is temporarily stopped, the piston 58 of the inertial body anti-reversal valve 50 returns from the stroke end to the initial position, causing the spool 55 to slide and displace to the valve open position.
The spool 55 can be held in an open state for a relatively long time until the valve spring 64 returns the spool 55 to the closed position. This allows, for example, high pressure in the main line 6B to be released reliably to the main line 6A side through the bypass lines 54A and 54B while providing a throttling action through the throttling passage 56 of the anti-reversal valve body 51, and as described in the prior art, the pressure in the main lines 6A and 6B can be reduced.
This reduces the differential pressure ΔP (see FIG. 12) that occurs between the two, effectively preventing the hydraulic motor 1 from repeatedly performing reverse rotation.

Then, when the spool 55 is gradually pushed toward the hydraulic chamber 57 by the valve spring 64 and slides to the position shown in FIG. 8, the spool 55 closes the space between the ports 53A and 53B, cutting off communication between the main lines 6A and 6B by the anti-reverse valve body 51 via the bypass lines 54A and 54B, and the hydraulic motor 1 can be held in a stopped state.

Furthermore, when the inertial rotation of the hydraulic motor 1 stops, if the amount of pressure oil leaking from the hydraulic motor 1 is small, the hydraulic motor 1 may reverse due to the differential pressure ΔP between the main lines 6A and 6B. However, in this case, even if the hydraulic motor 1 reverses only slightly, the pressure in the main line 6B may exceed the set spring.
The pressure drops to the set pressure 63, and the main lines 6A and 6B are communicated with each other by the throttle passage 56 of the spool 55 and piston 58, so that the hydraulic motor 1 is immediately prevented from rotating in reverse any further.

Therefore, according to this embodiment, the hydraulic motor 1 can be effectively prevented from reversing after the inertial rotation of the hydraulic motor 1 has stopped, and the hydraulic motor 1 can be quickly stopped together with the inertial body such as the upper rotating body, thereby achieving effects almost similar to those of the first embodiment.

In particular, in this embodiment, the inertial body anti-reverse valve 50 is provided with a spool 55 and an oil suction valve 56 in a valve casing 52 of an anti-reverse valve body 51.
The inertial body anti-reverse valve 50 can be constructed by providing the spool 55 and the piston 58 of the supply device 59 so that they can be displaced relative to each other, and only one inertial body anti-reverse valve 50 needs to be provided between the main lines 6A, 6B via a shuttle valve 69 or the like. As a result, the inertial body anti-reverse valve 50 can be easily incorporated into the casing of the hydraulic motor 1, for example, simplifying the overall system, and the inertial body anti-reverse valve 50 can be formed compactly, including the spool 55, piston 58, etc., allowing for miniaturization.

In addition, when the spool 55 is in the closed position shown in FIG.
6, one of the throttle holes 55E is blocked from the oil groove 58D of the piston 58 by the piston 58. Then, in the valve casing 52, the spool 55 and the piston
7, even if only one of the spool 55 and the piston 58 slides, the ports 53A and 53B can be reliably shut off from each other.
The spool 55 and piston 58 can provide a dead zone with a relatively large stroke amount.

As the driving pressure or braking pressure of the hydraulic motor 1 fluctuates, the piston 58 moves from the position shown in FIG. 6 to the oil reservoir chamber 60 side (left side), and a small amount of oil is supplied from the oil reservoir chamber 60 to the hydraulic chamber 57, causing the spool 55 to move in the direction of the valve spring 64.
Even in such a case, a relatively large dead zone is provided between the throttle hole 55E of the spool 55 and the oil groove 58D of the piston 58 in the state shown in Figure 6, so that the main lines 6A, 6B do not communicate with each other inadvertently, and the inertial body anti-reversal valve 50 can be reliably prevented from opening inadvertently.

Furthermore, a pusher slide hole 58C formed in the piston 58
is formed to have a smaller diameter than the piston sliding hole 55A of the spool 55. Therefore, the pressure-receiving area of the pusher 65 for the pilot pressure supplied from the pilot pipe 70 to the pilot oil chamber 66 can be significantly reduced. As a result, the spring force of the setting spring 63 can be set small corresponding to the pressure-receiving area of the pusher 65, and the piston 58 can be reliably returned to its initial position even with a small and lightweight setting spring 63. Furthermore, since the volume of the oil sump chamber 60 can be made larger than that of the pusher sliding hole 58C, a relatively large amount of oil can be drawn into and stored in the oil sump chamber 60 when the piston 58 reaches the stroke end. Therefore, when oil is supplied from the oil sump chamber 60 to the hydraulic chamber 57, the valve open time until the spool 55 is returned from the valve open position to the valve closed position by the valve spring 64 can be effectively extended.

Furthermore, by changing the dimensions of spool 55, piston 58, pusher 65 and throttle 68 (or the spring forces of set spring 63 and valve spring 64) in accordance with the size of an inertial body such as the upper rotating body, the opening time of spool 55 can be adjusted appropriately. Meanwhile, because throttle holes 55E, 55E are drilled in spool 55 at right angles to the sliding direction of spool 55, the fluid force when hydraulic oil passes through throttle passage 56 can be minimized, and the sliding displacement of spool 55 can be effectively prevented from being affected and disturbed by fluid force.

Next, Figure 9 shows a fourth embodiment of the present invention. This embodiment is characterized in that a pusher slide hole communicating with the spool slide hole is formed at one end of the valve casing, a pilot oil chamber to which pilot pressure is supplied from outside is formed between one end of a pusher inserted into the pusher slide hole and the pusher slide hole, and the other end of the pusher abuts against one end face of the piston so that the piston is pushed by the pusher. In this embodiment, the same components as those in the third embodiment are designated by the same reference numerals and their description will be omitted.

In the figure, 80 indicates the inertia body anti-reverse valve according to this embodiment,
The inertial body anti-reverse valve 80 comprises an anti-reverse valve body 81 as valve means, which will be described later, and an oil suction/supply device 87 as pressurized oil supply means.

Reference numeral 81 denotes the anti-reverse valve body, and 82 denotes the valve casing that constitutes the casing of the anti-reverse valve body 81. The valve casing 82 has a spool slide hole 82A, annular oil grooves 82B and 82C, notches 82D, and annular groove 82E, similar to the valve casing 52 described in the third embodiment. However, the valve casing 82 has a spool slide hole 82A at one end (left side in the drawing).
A small diameter pusher slide hole 82F is formed in communication with the port 53. The oil grooves 82B and 82C of the valve casing 82 are
The pipes 53A and 53B are connected to the main pipes 6A and 6B via bypass pipes 54A and 54B.

Reference numeral 83 denotes a spool as a cylindrical valve element inserted into the spool slide hole 82A, and the spool 83 is formed in substantially the same manner as the spool 55 described in the third embodiment, except for the annular groove 55B and the oil holes 55C, and the spool 83 is provided with a piston slide hole 83A and a receiving portion 83B. The valve spring 64 is disposed between the spring receiving portion 83B of the spool 83 and a spring receiving portion 80C of a piston 86 (described later), and the spool 83
is constantly biased toward the hydraulic chamber 84 by the valve spring 64.

Here, the spool 83 constitutes a valve means together with the valve spring 64 and the hydraulic chamber 84, and the hydraulic chamber 84 is connected to the tank 3 via the tank pipe line 67 and the throttle 68, similar to the hydraulic chamber 57 described in the third embodiment. Also, throttle holes 83C as oil holes are drilled in the radial direction in the spool 83, and the throttle holes 83C are formed with axial spacing dimensions corresponding to the notch portions 82D. The throttle holes 83C are connected to the oil grooves 82C, which will be described later.
6A constitutes a throttle passage 85 that connects the ports 53A and 53B of the casing 82.

Reference numeral 86 denotes a piston inserted into a piston sliding hole 83A of the spool 83, and the piston 86, together with the setting spring 63 and an oil reservoir chamber 88 (to be described later), functions as a pressurized oil supply means.
The piston 86 is connected to the piston 58 in the third embodiment, the annular groove 58A, and the oil hole 5
Therefore, the piston 86 has the throttle holes 8B and 58C of the spool 83 at positions spaced a predetermined distance to the right from the left end face of the piston 86.
An oil groove 86A is formed between the oil groove 86A and the oil groove 86A, which together with the oil groove 83C constitutes the throttle passage 85, and extends in the axial direction.

The right side of the piston 86 protrudes into the valve casing 82 through the piston slide hole 83A of the spool 83.
An annular step portion 86 is provided at a position axially opposite to the spring receiving portion 83B of the third spring.
The annular step 86B is provided on the spool.
When the spool 83 slides to the right in FIG.
The piston 86 has an annular stepped portion 86B which abuts against the tip of the spring receiving portion 83B and restricts the stroke end of the spool 83.
A spring receiving portion 86C is formed on the right side of the piston 86 so as to protrude radially, and a rod-shaped stopper portion 86D is provided extending in the axial direction to regulate the stroke end of the piston 86. A setting spring 63 is disposed between the spring receiving portion 86C of the piston 86 and the end face of the spring chamber 61, and the piston 86 is constantly biased by the setting spring 63 toward the oil reservoir chamber 88.

Reference numeral 88 denotes an oil sump chamber located radially inside the hydraulic chamber 84 and formed between the left end face of the piston 86 and the valve casing 82, and the oil sump chamber 88 is constructed in substantially the same manner as the oil sump chamber 60 described in the third embodiment.
When the piston 86 slides to the right in FIG. 9, the oil reservoir chamber 88 increases in capacity, sucking in and storing the oil in the tank 3. When the piston 86 slides to the left, the oil stored in the oil reservoir chamber 88 is pressurized and the oil is pumped into the hydraulic chamber 88.
Supply to 84.

Reference numeral 89 denotes a small-diameter cylindrical pusher slidably inserted into the pusher sliding hole 82F of the valve casing 82, and the pusher 89 defines a pilot oil chamber 90 serving as a hydraulic pilot section between its left end face and the pusher sliding hole 82F. The pilot oil chamber 90 applies pressurized oil supplied from the pilot line 70 as pilot pressure to the left end face of the pusher 89. As a result, when the pusher 89 is pushed rightward in Figure 9 by the high pilot pressure, the pusher 89 protrudes from the pusher sliding hole 82F into the oil reservoir chamber 88, sliding the piston 86 to the stroke end against the set spring 63.

Thus, in this embodiment configured as above, the pilot oil chamber 90 becomes highly pressurized during inertial rotation of the hydraulic motor 1, and the piston 86 slides and displaces integrally with the pusher 89 to the stroke end. On the other hand, when the inertial rotation stops, the pressure in the pilot oil chamber 90 becomes lower than the set pressure of the setting spring 63, and the piston 86 returns to its initial position. At this time, oil in the oil reservoir chamber 88 is supplied to the hydraulic chamber 84, and is gradually discharged into the tank 3 via the orifice 68. As a result, the throttling action of the orifice 68 causes the spool
The return speed when the valve 83 slides from the valve open position toward the hydraulic chamber 84 can be slowed down, and the same effects as those of the third embodiment can be obtained.

10 shows a fifth embodiment of the present invention, in which the same components as those in the fourth embodiment are given the same reference numerals and their description will be omitted. However, a feature of this embodiment is that an oil groove 91D that communicates between the throttle holes 91C is formed integrally with the inner periphery of the spool 91, and the oil groove 91D and the throttle holes 91C form a throttle passage 92 that communicates between the ports 53A and 53B in the valve casing 82.

Here, the spool 91 is formed in substantially the same manner as the spool 83 described in the fourth embodiment, except for the oil groove 91D, and is provided with a piston sliding hole 91A, a spring receiving portion 91B, and various throttle holes 91C. The piston 93 inserted into the piston sliding hole 91A of the spool 91 is constructed in the same manner as the piston 86 described in the fourth embodiment, except for the piston 86A and the oil groove 86A, and is provided with an annular step 93A, a spring receiving portion 93B, and a stopper portion 93C.

Thus, in this embodiment configured as above, it is possible to obtain substantially the same effects as in the fourth embodiment. However, in this embodiment, the throttle passage is formed on the outer peripheral surface of the piston 93.
Since it is not necessary to form a part of 92, the shape of the piston 93 can be simplified.

In the third embodiment, the throttle passage 56 is described as being composed of the throttle holes 55E formed in the spool 55 and the oil groove 58D formed on the outer periphery of the piston 58. However,
As with the spool 91 described in the fifth embodiment, an annular oil groove that connects the throttle holes 55E may be formed on the inner periphery of the spool 55, thereby forming a throttle passage.

In the third embodiment, the spool 55 is provided with throttle holes.
However, the present invention is not limited to this. For example, in the valve open position shown in FIG.
The throttle passage that connects between B and C may be formed, for example, by bending in a "U" shape inside the spool 55.
This point is the same as that of the spool 83 (9
The same applies to 1).

Furthermore, in the third, fourth and fifth embodiments, the throttle 68 serving as flow resistance means is provided midway through the tank pipe 67, which serves as the oil passage. However, even in this case, a check valve similar to the check valve 42 used in the second embodiment may be provided midway through the tank pipe 67 in parallel with the throttle 68.

On the other hand, in each of the above-described embodiments, the tank pipe line 27 which serves as the oil passage
However, for example, the tank pipe 27 (67) is provided with a throttle 38 (68) as a flow resistance means.
The flow resistance means according to the present invention may be configured by reducing the flow area of the oil passages such as the restrictor 38 (68) and increasing the pipe resistance.
etc. can be omitted.

Furthermore, in each of the above embodiments, the hydraulic circuit for swinging a hydraulic excavator has been described as an example, but the present invention is not limited to this and can be widely applied to inertial body drive devices used to drive various inertial bodies, such as a hydraulic circuit for swinging a hydraulic crane or a hydraulic circuit for a rope winch.

INDUSTRIAL APPLICABILITY As described above in detail, according to the present invention, when the pressure of the pressurized oil on the high-pressure side of the first and second main lines connected to a hydraulic motor for driving an inertial body exceeds a preset second pressure value, oil is stored in the oil reservoir chamber of the pressurized oil supply means and the spool of the valve means is brought to a closed position by the biasing means, and when the pressure of the pressurized oil falls below the second pressure value, the pressurized oil supply means pressurizes the oil in the oil reservoir chamber and supplies it to the hydraulic chamber of the valve means, and A flow resistance means is provided in the oil passage connecting the hydraulic chamber to the tank, to throttle the oil discharged through the oil passage into the tank. Therefore, when the inertial rotation of the hydraulic motor temporarily stops and the pressure of the pressurized oil falls below the second pressure value, pressurized oil is supplied from the oil reservoir of the pressurized oil supply means to the hydraulic chamber, causing the spool of the valve means to slide to the open position against the biasing means, and the first and second main lines can be inserted to reduce the pressure difference.

In this state, the spool is biased toward the hydraulic chamber by the biasing means, and the oil that is discharged from the hydraulic chamber to the tank through the oil passage is throttled by the flow resistance means.
The open time of the spool can be effectively extended,
The differential pressure between the first and second main lines can be reliably reduced, and the hydraulic motor can be effectively prevented from repeating reverse rotation when the inertial rotation stops.

Furthermore, it is only necessary to provide an inertial body reverse prevention valve consisting of valve means and pressurized oil supply means between the first and second main lines, which simplifies the overall structure and allows it to be formed compactly.

On the other hand, by configuring the valve means as a spool valve, a relatively large dead band can be formed between the closed and open positions, and by supplying and discharging oil from the oil reservoir of the pressurized oil supply means to the hydraulic chamber of the valve means to slide and displace the spool, even if the driving pressure or braking pressure of the hydraulic motor fluctuates, these pressure fluctuations can be effectively absorbed by the oil reservoir of the pressurized oil supply means, etc. This reliably prevents the spool from accidentally opening due to the direct influence of pressure fluctuations, greatly improving safety and reliability.

Furthermore, if a spool formed as a cylindrical valve body is inserted into a spool sliding hole formed in the casing and a piston or the like of a pressurized oil supply means is provided within the spool,
The entire inertia body anti-reverse valve can be made more compact.
Furthermore, by providing the piston of the pressurized oil supply means within the spool, the pressure-receiving area of the hydraulic pilot section that drives the piston can be reduced, and the spring force of the setting spring that sets the second pressure value can be correspondingly reduced, thereby ensuring that the entire inertia body anti-reverse valve can be made smaller and lighter.

Claims (11)

[Claims] [Claim 1] An inertial body drive device comprising: a hydraulic power source; a hydraulic motor that drives an inertial body by supplying and discharging pressurized oil from said hydraulic power source; a switching valve connected to said hydraulic motor via first and second main lines, that supplies and discharges pressurized oil from said hydraulic power source to the hydraulic motor when switched from a neutral position and stops supplying and discharging pressurized oil to said hydraulic motor when returned to the neutral position; and a pressure control valve located between said switching valve and the hydraulic motor and between said first and second main lines, that limits the maximum pressure in said first and second main lines to a preset first pressure value; valve means located between said first and second main lines, that has a spool that slides between a valve open position and a valve closed position to connect and disconnect said first and second main lines, and that has biasing means that biases said spool towards the valve closed position and a hydraulic chamber that slides and displaces said spool from the valve closed position to the valve open position against the biasing means; an inertial body drive system comprising: pressurized oil supply means having a variable-capacity oil reservoir chamber communicating with the hydraulic chamber of said valve means, and pressurizing the oil in said oil reservoir chamber and supplying it to the hydraulic chamber of said valve means when the pressure in the higher-pressure main line of said first and second main lines falls to or below a preset second pressure value lower than the first pressure value set by said pressure control valve; and flow resistance means provided in an oil line connecting at least one of the oil reservoir chamber of said pressurized oil supply means and the hydraulic chamber of said valve means to a tank, and providing a throttling effect on the oil discharged to the tank via said oil line. [Claim 2] A high-pressure selection valve is provided between the first and second main lines, between the switching valve and the hydraulic motor, for selecting the higher-pressure oil of the first and second main lines, and the pressurized oil supply means is provided with a pilot oil chamber to which the pressurized oil selected by the high-pressure selection valve is supplied as a pilot pressure, and when the pilot pressure of the pilot oil chamber exceeds the second pressure value, the pressurized oil supply means sucks oil into the oil reservoir chamber to store it, and when the pressure falls below the second pressure value,
2. The inertial body drive device according to claim 1, wherein the oil stored in the oil reservoir is pressurized and supplied to the hydraulic chamber of the valve means. [Claim 3] An inertial drive device as described in claim 1 or 2, wherein a check valve is provided in parallel with the flow resistance means in the oil passage, and the check valve is configured to allow oil in the tank to flow toward the hydraulic chamber and prevent flow in the reverse direction. [Claim 4] The valve means comprises a spool slide hole provided in a casing, a pair of ports formed in the casing at intervals in the axial direction of the spool slide hole and communicating with the first and second main lines, a spool slidably inserted into the spool slide hole, a throttle passage formed in the spool for communicating and blocking the pair of ports, a hydraulic chamber provided between one end of the spool and the casing and connected to a tank via the oil passage, and a valve spring as biasing means disposed between the other end of the spool and the casing for biasing the spool toward the hydraulic chamber, 2. The inertial body drive device according to claim 1, wherein the pressurized oil supply means comprises: a piston sliding hole formed in a casing; a piston slidably inserted into the piston sliding hole; an oil sump chamber formed by the piston on one end side of the piston sliding hole and communicating with the hydraulic chamber; a set spring provided on the other end side of the piston so as to constantly urge the piston toward the oil sump chamber and having a spring force corresponding to the second pressure value; and a hydraulic pilot section provided on one end side of the piston so as to slidably displace the piston against the set spring so as to suck oil into the oil sump chamber when pressurized oil on the higher pressure side of the first and second main lines exceeds the second pressure value; and 5. The valve means and the pressurized oil supply means are provided in the same casing, and the casing is provided with a spool slide hole and a pair of ports which constitute a part of the valve means, and spaced apart from the ports and the spool slide hole:
5. The inertial body drive device according to claim 4, further comprising a piston slide hole that constitutes a part of said pressurized oil supply means. [Claim 6] An inertial drive device as described in claim 4 or 5, wherein a check valve is provided in the oil passage in parallel with the flow resistance means, and the check valve is configured to allow oil in the tank to flow toward the hydraulic chamber and prevent flow in the reverse direction. [Claim 7] The valve means comprises a spool sliding hole provided in a casing, a pair of ports formed in the casing at intervals in the axial direction of the spool sliding hole and communicating with the first and second main lines, a spool formed as a cylindrical valve body inserted into the spool sliding hole and having an inner circumferential side that serves as a piston sliding hole, a throttle passage that opens and closes communication between the pair of ports as the spool slides and displaces, a hydraulic chamber formed between one end of the spool and the casing, and a valve spring as biasing means that biases the spool toward the hydraulic chamber, 2. The inertial body drive device according to claim 1, wherein the pressurized oil supply means comprises: a piston slidably inserted in the piston sliding hole of the spool; an oil sump chamber formed by the piston on one end side of the piston sliding hole and constantly communicating with the hydraulic chamber; a set spring disposed between the other end side of the piston and the casing so as to constantly urge the piston toward the oil sump chamber and having a spring force corresponding to the second pressure value; and a hydraulic pilot section provided on one end side of the piston for slidingly displacing the piston against the set spring so as to suck oil into the oil sump chamber when the pressure oil on the higher pressure side of the first and second main lines exceeds the second pressure value; and a valve spring of the valve means disposed between the other end side of the spool and the piston and having a spring force weaker than that of the set spring of the pressurized oil supply means. 8. The inertial body drive device according to claim 7, wherein the oil sump chamber is positioned radially inward of the hydraulic chamber and is formed between one end side of the piston and a casing. [Claim 9] An inertial body drive device as described in Claim 7, wherein the throttling passage is composed of at least a pair of oil holes drilled radially in the spool and spaced apart in the axial direction of the spool, and an oil groove formed on the outer periphery of the piston and extending in the axial direction to connect the oil holes. [Claim 10] An inertial body drive device as described in claim 7, 8 or 9, wherein the hydraulic pilot section is composed of a pilot oil chamber formed within the piston, positioned between the piston and a pusher slidably inserted into one end side of the piston, and the pilot oil chamber is supplied with pressurized oil from the higher pressure side of the first or second main pipe line via a high pressure selection valve. [Claim 11] An inertial body drive device as described in claim 7, 8 or 9, wherein the hydraulic pilot section is composed of a pilot oil chamber formed within the casing, positioned between it and a pusher slidably inserted into one end side of the casing, and the pilot oil chamber is supplied with pressurized oil from the higher pressure side of the first and second main lines via a high pressure selection valve.