JP5382863B2 - Power steering mechanism - Google Patents

Description

  The present invention relates to an improvement of a power steering mechanism that is a hydraulic steering mechanism and has a variable steering force.

FIG. 8 shows a block diagram of a so-called linkage type power steering system used in a three-axis vehicle with two front wheels in a large commercial vehicle, and FIG. 9 is a perspective view of the main part of the power steering system of the vehicle. It is shown in three dimensions.
In FIG. 9, the second axis of the front wheels and the like are not shown in order to simplify the description.

The power steering system of FIG. 8 includes a hydraulic pump 2 driven by the engine 1, a power cylinder 3, a gear box 4, an oil tank 5, a hydraulic supply line L, and a hydraulic return line Lr.
A flow rate control valve 6 and a pressure control valve 7 are interposed in the hydraulic pressure supply line L, and if necessary, a part of the hydraulic oil in the hydraulic pressure supply line L is supplied to the suction side of the oil pump 2 or an oil tank. Return to 5.

A valve unit 8 is incorporated in the power cylinder 3. The valve unit 8 is, for example, a 4-port 3-position flow path conversion valve, and FIG. 8 shows a neutral state, for example, a state during straight traveling.
The valve unit 8 has a first position (a block having a left-turning flow path) 81 and a second position (neutral block: a straight-travel flow path) in a housing 80 that is a fixed portion integrated with the power cylinder 3. ) 82 and a third position (block having a right-turning flow path) 83 are configured to be switchable.

Return springs 8sl and 8sr are provided at the left and right ends of the valve unit 8 in FIG. One end of the valve unit 8 (the end on the first position 81 side) is connected to the tip of the pitman arm 9 engaged with the gear box 4 via a connection link 89.
Further, in the cylinder body 30 of the power cylinder 3, the left end 31 in FIG. 8 is connected to the knuckle arm 11 of the front wheel Wf via the first link 10, and the piston 32 that slides in the power cylinder 3 is the second It is connected to the frame 13 of the vehicle via the link 12. In other words, the piston 32 is fixed, and the cylinder body 30 is configured to be movable in the left-right direction in FIG.

The gear box 4 is connected to the steering wheel 15 via the steering column 14, and when the driver steers the steering wheel 15, the pitman arm 9 engaged with the gear box 4 rotates as indicated by an arrow R.
When the driver steers the steering wheel 15 to the right (the operation of the arrow Rh), the pitman arm 9 rotates in the counterclockwise direction of the arrow R in FIG.
When the pitman arm 9 rotates in the counterclockwise direction of the arrow R, the valve unit 8 switches to the flow path at position 83, the hydraulic oil flows into the right chamber 34 of the power cylinder 3, and the pressure in the right chamber 34 increases. The pressure in the left chamber 33 is reduced.

As described above, the piston 32 in the power cylinder 3 is fixed to the frame 13 via the second link 12. Therefore, when the right chamber 34 is pressurized, the power cylinder body 30 moves to the right side.
When the power cylinder body 30 moves to the right side, the knuckle arm 11 connected to the first link 10 is rotated clockwise (arrow RKR) with respect to the central axis of the king pin K of the front wheel Wf, and the vehicle turns right. To do.
At this time, part of the oil in the left chamber 33 is returned to the oil tank 5 via the hydraulic pressure return line Lr.

On the other hand, when the driver steers the steering wheel 15 to the left, the pitman arm 9 rotates the arrow R in the clockwise direction, the flow path of the valve unit 8 becomes the flow path of the position 81, and the left chamber 33 is boosted. Then, the power cylinder main body 30 is moved to the left, and the knuckle arm 11 is rotated counterclockwise (arrow RKL direction).
That is, the direction of movement of the pitman arm 9, the power cylinder body 30, and the knuckle arm 11 is opposite to that in the case of right steering.

When returning the vehicle to the straight traveling state, the driver steers the steering wheel 15 to return to the neutral state.
When returning to the neutral state, the valve unit 8 is switched to the flow path (neutral) of the position 82, and the hydraulic pressure from the hydraulic pump 2 and the hydraulic pressure in the left and right oil chambers 33, 34 of the power cylinder 3 are antagonized. The hydraulic oil fed to the oil tank 5 is returned to the oil tank 5 through the hydraulic pressure return line Lr.

In the conventional system shown in FIG. 8, even if the system is neutral, hydraulic pressure acts on the power cylinder 3, so the engine 1 must consume a part of the output in order to drive the oil pump 4.
In this way, applying hydraulic pressure to the power cylinder 3 even in the neutral state prevents the steering wheel 15 from wobbling during straight travel, copes with road surface input (kickback), and excessively moves the arm when moving from straight travel to steering. It is essential in the prior art to prevent unnecessary input.

FIG. 10 shows the relationship between the hydraulic pressure supply amount and the rotational speed of the oil pump 4 in the conventional system shown in FIG. As described above, the flow rate control valve 6 is interposed in the hydraulic pressure supply circuit L in the conventional system of FIG.
The flow control valve 6 uses the inlet pressure of the power cylinder 3 as a pilot pressure for valve operation via the line Lp in FIG. That is, the higher the inlet pressure of the power cylinder 3, the greater the relief amount in the flow control valve 6.
The flow rate control valve 6 operates when the pump rotation speed reaches a predetermined value or more, and returns a certain ratio of pressure oil to the suction side of the pump 2 or the tank 5. In the example of FIG. 10, when the flow control valve 6 is opened, an amount (relief amount) obtained by subtracting the solid line (characteristic Q _r ) from the broken line (characteristic Q ₀ ) is returned to the suction side of the pump 2 or the tank 5. .

FIG. 11 shows the hydraulic pressure supply amount (characteristic Q _r ) and the pump driving force (characteristic W) with respect to the rotation speed of the oil pump 2 in the system of FIG.
In the supply amount characteristic Q _r, the characteristic indicated by the solid line is when the pressure in the power cylinder 3 is low, the characteristic indicated by the broken line is when the pressure in the power cylinder 3 is high.
On the other hand, the driving force characteristic Q _w, characteristics indicated by the solid line is when the pressure in the power cylinder 3 is high, the characteristic indicated by broken line a when the pressure in the power cylinder 3 is low.

As described above, in the systems shown in FIGS. 8 to 11, a constant load is applied to the oil pump 2 even at the neutral time (straight forward). For this reason, the engine 1 consumes output in the steering system even when the vehicle is traveling straight without operating the steering.
Such a situation is unacceptable from the viewpoint of energy conservation.

Here, there is also a technology in which a power cylinder or a hydraulic pump is electrified.
If it is a motorized power cylinder or hydraulic pump, there is no problem of wasted power unlike a hydraulic pump directly connected to the engine.
However, in a large truck, the weight of the power cylinder increases, and it is difficult to assist the steering force with the electric motor.
Further, when a hydraulic pump or a power cylinder is driven by an electric motor, various parts must be newly replaced, and existing parts cannot be used. For this reason, there is a problem that the cost at the time of introduction increases. In other words, it is not possible to meet the demand for providing a system that can be retrofitted using existing parts as much as possible.

As another conventional technique, for example, a switching spool in a switching cylinder is provided with a pressure receiving chamber, and the switching spool is slid by the pressure acting on the pressure receiving chamber to switch between communication between the first port and the second port. A control device for a positive displacement pump provided with a valve has been proposed (Patent Document 1).
Although this technology (Patent Document 1) may reduce the driving power of the hydraulic pump, it is necessary to develop a new valve system and is difficult to retrofit to an existing vehicle.

Further, a power steering device is proposed in which the hydraulic pump is driven only when the handle is operated, and the flow rate of hydraulic oil discharged from the hydraulic pump is increased in response to an increase in the steering load pressure and supplied to the actuator. (Patent Document 2).
However, when this technology (Patent Document 2) is applied to a commercial vehicle such as a truck, particularly a large commercial vehicle, the electric hydraulic pump is larger and has a larger output than a small vehicle such as a passenger car. It is required and becomes a dedicated device, and compatibility with conventional equipment cannot be expected.

In addition, a technology has been proposed to change the steering force pattern by providing a bypass line from the discharge port to the suction port of the hydraulic pump, interposing a flow rate adjusting valve in the bypass line, and controlling the opening degree of the flow rate adjusting valve. (See Non-Patent Document 1).
However, the technology (Non-Patent Document 1) aims to change the steering feeling by controlling the characteristics of the weight of the steering wheel with respect to the vehicle speed, and as described above, reduces unnecessary power in the hydraulic pump. In addition, it does not meet the request in the prior art that enables retrofitting using existing parts as much as possible.

Japanese Examined Patent Publication No. 6-71889 Japanese Patent Publication No. 64-6984

Edited by Kayaba Kogyo Co., Ltd., “Automotive steering system and maneuverability”, Sankaido Co., Ltd., September 10, 1996, p. 90-93

  The present invention has been proposed in view of the above-described problems of the prior art, can reduce wasteful power in a hydraulic pump directly connected to an engine, and can be retrofitted using existing parts of a vehicle. The purpose is to provide a power steering mechanism.

  The power steering mechanism of the present invention powers the hydraulic pump (2) directly connected to the engine (1), the power cylinder unit (3), the reservoir tank (5), and the discharge port (2o) of the hydraulic pump (2). In the power steering mechanism having a pressure oil line (L) communicating with the valve unit (8) of the cylinder (3), the discharge port (2o) and the suction port (2i) of the hydraulic pump (2) are communicated. The bypass line (Lb) has a flow rate adjusting valve (16) interposed in the bypass line (Lb), and the flow rate adjusting valve (16) has a piston (32) of the power cylinder unit (3). The pressure in the cylinder sections (33, 34) partitioned by the cylinder is input as a signal pressure, and closes when the signal pressure is high, and opens when the signal pressure is low. ) Are separated from each other (33, 34) of the cylinder (3) partitioned by the signal pressure input side of the flow rate adjusting valve (16) via signal pressure lines (Ls1 to Ls4). The lines (Ls1 to Ls4) are connected to the cylinder sections (33, 34) partitioned by the piston (32) (signal pressure lines: Ls1, Ls2) and the signal pressure of the flow rate adjusting valve (16). Lines (Ls3, Ls4) communicating with the input side, lines (Ls5) communicating with the upstream side of the valve unit of the hydraulic oil line (L) (discharge side 2o of the hydraulic pump 2), and communication with the signal pressure input side Each of the sections (33, 34) of the branch point (B5) and the cylinder (3) of the line (Ls5) communicating with the region of the line (Ls4) and the valve unit (8) upstream (discharge side 2o of the hydraulic pump 2). ) It is characterized in that the confluence of the passage to the line (Ls1, Ls2) and (G3) has a line (Ls3) communicating.

  The power steering mechanism of the present invention includes a hydraulic pump (2) directly connected to the engine (1), a power cylinder unit (3), a reservoir tank (5), and a discharge port (2o) of the hydraulic pump (2). In the power steering mechanism having a pressure oil line (L) communicating with the valve unit (8) of the power cylinder (3), the discharge port (2o) and the suction port (2i) of the hydraulic pump (2) It has a bypass line (Lb) that communicates, and a flow rate adjustment valve (16) is interposed in the bypass line (Lb). The flow rate adjustment valve (16) has a piston ( 32) The pressure in the cylinder section (33, 34) partitioned by 32) is input as a signal pressure, and closes when the signal pressure is high, and opens when the signal pressure is low. Two flow rate regulating valves (16A, 16B) are interposed in the line (Lb), and the two flow rate regulating valves (16A, 16B) are connected to the piston (32) of the power cylinder unit (3). The pressure in each of the compartments (33, 34) of the partitioned cylinder (3) is input as a signal pressure.

  Here, each of the lines (signal pressure lines: Ls1, Ls2) communicating with each of the cylinder sections (33, 34) partitioned by the piston (32) is closer to the cylinder (3) than the junction (G3). The check valve (181, 182) is interposed in the region of (1), the line (Ls3) connecting the junction (G3) and the branch point (B5) is connected to the junction (G3) rather than the branch point (B5). The check valve (191) is communicated with the region on the cylinder side (cylinder side), and the check valve (191) is connected to the line (Ls5) communicated with the region on the upstream side (discharge side 2o of the hydraulic pump) of the valve unit (8). 192) and a pressure loss mechanism (eg, orifice 193) are preferably interposed (FIG. 1).

  In this case, the signal pressure line (Ls6, Ls7) is connected to the signal pressure input side of each of the sections (33, 34) of the cylinder (3) partitioned by the piston (32) and the flow rate adjusting valves (16A, 16B). The signal pressure lines (Ls6, Ls7) are connected to the flow rate adjusting valves (16A, 16B) when the pressure in the compartments (33, 34) of the cylinder (3) rises. ) When the pressure in the first path (Ls6, Ls7) that transmits to the signal pressure input side and the cylinder compartments (33, 34) drops, the high pressure on the flow rate adjustment valve (16A, 16B) side It is preferable to have the 2nd path | route (Lb6, Lb7) which transmits to a division (33, 34) with a time difference (FIG. 7).

  Here, the first path (Ls6, Ls7) is provided with check valves (201, 211) for transmitting the signal pressure only toward the flow rate adjustment valves (16A, 16B). (Lb6, Lb7) are provided with check valves (202, 212) and pressure loss mechanisms (for example, orifices 203, 213) that transmit signal pressure only toward the cylinder sections (33, 34). Preferred (FIG. 6).

  Further, it is preferable that a pressure loss mechanism (for example, an orifice 17) is interposed in the bypass line (Lb) (FIGS. 1 and 7).

Furthermore, in addition to the bypass line (Lb), there is a second bypass line (L6, L67) that communicates the discharge port (2o) and the suction port (2i) of the hydraulic pump (2). It is preferable that a flow rate control valve (61) is interposed in the bypass line (L6).
Then, in addition to the bypass line (Lb) and the second bypass line (L6, L67), a third outlet communicating the discharge port (2o) and the suction port (2i) of the discharge port (2) of the hydraulic pump. It is preferable to have a bypass line (L7, L67), and a relief valve (pressure regulating valve 71) is interposed in the third bypass line (L7, L67).

According to the present invention having the above-described configuration, the bypass line (Lb) that connects the discharge port (2o) and the suction port (2i) of the hydraulic pump (2) is provided, and a flow rate is provided in the bypass line (Lb). The regulating valve (16, 16A, 16B) is interposed, and the flow regulating valve (16, 16A, 16B) is a compartment of the cylinder (3) partitioned by the piston (32) of the power cylinder unit (3). When the pressure at (33, 34) is low, it opens, and when it is high, it closes. When the load of the power cylinder (3) is low, such as when the vehicle is traveling straight, any of the cylinders (3) The pressure also decreases in the compartments (33, 34). Therefore, the flow rate adjusting valves (16, 16A, 16B) are opened, and the discharge port (2o) and the suction port (2i) of the hydraulic pump (2) are communicated.
As a result, the discharge port (2o) and the suction port (2i) of the hydraulic pump (2) are communicated with each other, so that power consumption for circulating the pressure oil is reduced, and wasteful power consumption is reduced. And the auxiliary machine drive power of an engine (1) is reduced and a fuel consumption is also improved.

On the other hand, when the load of the power cylinder (3) is high as during steering, the pressure in one of the compartments (33 or 34) of the cylinder (3) increases, so that the compartment (33 or 34) The flow regulating valve (16) that communicates is closed, and the bypass line (Lb) is also closed. As a result, the pressure oil discharged from the hydraulic pump (2) is reliably supplied to the valve unit (8) of the power cylinder unit (3), and a sufficient steering assist force can be generated during steering.
For example, in the present invention, the signal pressure line has a line (signal pressure line: Ls1, Ls2) communicating with each of the sections (33, 34) of the cylinder (2) partitioned by the piston (32). The lines (Ls1, Ls2) merge at the junction (G3), and if a check valve (181, 182) is interposed on the cylinder (3) side from the junction (G3), the power cylinder (3) When the load on the cylinder is high, the higher pressure in the compartment (33, 34) of the cylinder (3) partitioned by the piston (32) is applied to the signal of the flow regulating valve (16) via the junction (G3). Apply to the pressure input side. As a result, the flow regulating valve (16) is closed and the bypass line (Lb) is also closed (FIG. 1).

Alternatively, in the present invention, two flow rate adjustment valves (16A, 16B) are interposed in the bypass line (Lb), and the two flow rate adjustment valves (16A, 16B) have a power cylinder unit (3). If the pressure in each of the cylinder sections (33, 34) partitioned by the piston (32) is input as a signal pressure, when the load on the power cylinder (3) is high, one section in the cylinder (3) (33 or 34) is increased in pressure, and when the other section (34 or 33) is depressurized, the high pressure in the increased pressure section (33 or 34) is one of the flow rate regulating valves (16A, 16B). The flow rate regulating valve (16A or 16B) is closed (FIG. 7).
At this time (in the case of FIG. 7), the flow regulating valve (16A or 16B) communicating with the opposite compartment (decompressed compartment 34 or 33) is opened. However, if either one of the flow control valves (16B or 16A) is closed, the bypass line (Lb) is closed as a whole, so that the discharge port (2o) and the suction port (2i) of the hydraulic pump (2) Does not communicate.

Further, according to the present invention (for example, FIG. 1), each of the compartments of the cylinder partitioned by the bypass line (Lb) interposing the flow rate adjusting valve (16) and the piston (32) of the power cylinder unit (3). (33, 34) and a line (Ls5) communicating with the flow rate adjusting valve (16) and a line (Ls5) communicating with the valve unit (8) upstream side (discharge side 2o of the hydraulic pump 2). Since it may be added to the existing power steering mechanism, it can be easily applied to the existing vehicle by so-called “retrofitting” by utilizing the existing parts of the vehicle already having the power steering mechanism.
Therefore, the introduction cost can be kept low.
Alternatively, according to the present invention (for example, FIG. 7), the bypass line (Lb) including two flow rate adjusting valves (16A, 16B), the two flow rate adjusting valves (16A, 16B), and the power cylinder Two lines (signal pressure lines Ls6, Ls7) communicating with each of the sections (33, 34) of the cylinder (2) partitioned by the piston (32) of the unit (3) are used as an existing power steering mechanism. Since it only has to be added, the existing parts of the vehicle already having the power steering mechanism can be utilized, and it can be easily applied to the existing vehicle by so-called “retrofitting”, so that the introduction cost can be reduced.

In the present invention, the signal pressure line is a branch point between a line (Ls4) communicating with the signal pressure input side and a line (Ls5) communicating with the valve unit (8) upstream side (discharge side 2o of the hydraulic pump 2) ( B5) and a line (Ls3) that communicates the junction (G3) of the lines (Ls1, Ls2) that communicate with each of the compartments (33, 34) of the cylinder (3) (FIG. 1), When the power cylinder (3) shifts from a high load state to a low load state, as in the case of transition from steering to straight travel, the pressure in the compartments (33, 34) of the cylinder (3) decreases. The high pressure on the flow regulating valve side (16) is transmitted from the branch point (B5) to the region upstream of the valve unit (8) (discharge side 2o of the hydraulic pump 2) via the line (Ls5).
Here, since the pressure loss mechanism (for example, the orifice 193) is interposed in the line (Ls5) communicating with the region upstream of the valve unit (8) (the discharge side 2o of the hydraulic pump 2), the flow adjustment valve side ( There is a slight delay before the high pressure of Lb) is transmitted to the region upstream of the valve unit (8) (the discharge side 2o of the hydraulic pump 2).

  On the other hand, in the present invention, two flow rate adjusting valves (16A, 16B) are interposed, and each of the signal pressure lines (Ls6, Ls7) has a pressure in the section (34, 33) of the cylinder (3). If it has a second path (Lb6, Lb7) for transmitting a high pressure on the flow rate regulating valve (16A, 16B) side to the section (34, 33) with a time difference when descending (FIG. 7). ) As in the case of transition from steering to straight travel, the power cylinder (3) shifts from a high load state to a low load state, and the pressure in the compartment (34, 33) of the cylinder (3) is low. If it becomes, the high pressure by the side of the flow regulating valve (16A, 16B) will be transmitted to the said division (34, 33) with a time difference via the 2nd path | route (Lb6, Lb7). Here, since the pressure loss mechanism (for example, the orifices 203 and 213) is interposed in the second path (Lb6, Lb7), a high pressure on the flow rate adjusting valve (16A, 16B) side causes a partition of the cylinder (3). There is a slight delay before being transmitted to (34, 33).

That is, according to the present invention (FIGS. 1 and 7), there is a slight delay in the transition from the closed state to the opened state of the flow rate regulating valves (16, 16A, 16B) of the bypass line (Lb). .
For example, when changing lanes, the vehicle is steered and instantaneously goes straight, and then steers in the opposite direction again. When going straight ahead momentarily (when the load on the power cylinder 3 becomes low), the flow control valve (16, 16A, 16B) immediately shifts to the open state and the pressure of the power cylinder (3) is reduced. When the pressure is reduced, the steering assist force is not generated when the steering is performed in the opposite direction, and the steering is very heavy.
In contrast, when the load of the power cylinder (3) is shifted from a high load state to a low load state, the flow rate adjusting valve (16, 16A, 16B) is slightly shifted from the closed state to the opened state. If this delay occurs, the pressure of the power cylinder (3) is not immediately reduced during such instantaneous straight travel, and a very heavy steering state is prevented.

  Similarly, in the present invention, if a pressure loss mechanism (mechanism for generating pressure loss) such as the orifice (17) is interposed in the bypass line (Lb), the flow rate adjustment valve (16) is opened ( In the case of FIG. 1), or the two flow rate adjusting valves (16A, 16B) are both opened (in the case of FIG. 7), the discharge port (2o) and the suction port (2i) of the hydraulic pump (2). And the amount of pressure oil supplied to the power cylinder (3) through the pressure oil line (first pressure oil line L) by the amount that pressure loss occurs in the pressure loss mechanism (17). Can be secured. Therefore, the pressure in the power cylinder (3) is lost, and it is possible to prevent the steering from becoming too heavy when steering is performed thereafter.

1 is a block diagram showing a system configuration according to a first embodiment of the present invention. It is the characteristic figure which showed the time change of the operation angle of the steering wheel in a lane change. FIG. 4 is a characteristic diagram showing a change in valve unit inlet pressure when the bypass line is opened corresponding to FIG. 2. FIG. 3 is a characteristic diagram showing a change in valve unit inlet pressure when the bypass line is closed in correspondence with FIG. 2. FIG. 3 is a characteristic diagram showing changes in the opening degree of the flow rate adjustment valve corresponding to FIG. 2. FIG. 3 is a characteristic diagram illustrating a change in pump driving power corresponding to FIG. 2 between the prior art and an embodiment of the present invention. It is the block diagram which showed the system configuration | structure which concerns on 2nd Embodiment of this invention. It is the block diagram which showed the system configuration | structure in a prior art. It is the perspective view which showed the system principal part of the vehicle carrying the system of FIG. FIG. 5 is a characteristic diagram showing a relationship between a pump rotation speed and an oil supply amount in a linkage type power steering system. It is the characteristic view which showed the relationship between pump rotation speed and oil supply amount in a linkage type power steering system, and pump rotation speed and pump driving force.

Embodiments of the present invention will be described below with reference to the accompanying drawings.
First, a first embodiment of the present invention will be described with reference to FIGS.
In FIG. 1, a steering system denoted as a whole by reference numeral 101 includes a hydraulic pump 2 driven by an engine 1, a power cylinder 3, a gear box 4, an oil tank 5, a hydraulic supply line L, and a hydraulic return line Lr. I have.

The hydraulic pressure supply line L includes lines L1 to L5.
The line L1 connects the discharge side 2o of the hydraulic pump (hereinafter referred to as “pump”) 2 and the first branch point B1.
The line L2 connects the first branch point B1 and the second branch point B2 in the flow control valve 6.
The line L3 connects the second branch point B2 and the third branch point B3 in the pressure control valve 7.
The line L4 connects the third branch point B3 and the fourth branch point B4 that is not included in the pressure control valve 7.
The line L5 connects the fourth branch point B4 and the port P1 of the housing 80 of the valve unit 8.

In FIG. 1, the flow control valve 6 is shown surrounded by a broken-line rectangle denoted by reference numeral 6. The flow control valve 6 includes the line L2, a part of the line L3, the second branch point B2, the flow control valve main body 61, the orifice 62, the line L6, the pilot pressure line Lp, the first junction G1, and the line L7. A line L67.
The line L67 connects the first merge point G1 and the second merge point G2.
The line L6 connects the second branch point B2 and the first junction point G1, and interposes the flow rate control valve main body 61.
The orifice 62 is interposed in the line L3.
Here, the flow control valve main body 61 is an external pilot type sequence valve.

The pressure control valve 7 is shown in FIG. 1 surrounded by a broken-line rectangle denoted by reference numeral 7. The pressure control valve 7 includes the line L3, a part of the line L4, a third branch point B3, a pressure control valve main body 71, an orifice 72, and a line L7.
The line L7 connects the third branch point B3 and the first junction point G1, and interposes the pressure control valve main body 71.
In the line L7, an orifice 72 is interposed in a region between the third branch point B3 and the pressure control valve main body 71.
The pressure control valve main body 71 is an internal pilot type sequence valve.

A valve unit 8 is incorporated in the power cylinder 3. The valve unit 8 is, for example, a 4-port 3-position flow path conversion valve.
The four ports are ports P1 to P4 formed in the housing 80, the port P1 is connected to the line L5, the port P2 communicates with the right chamber 34 of the power cylinder 3, and the port P3 is the left chamber 33 of the power cylinder 3. The port P4 is connected to the line Lr.
FIG. 1 shows a neutral state (a state where P1, P2, P3 and P4 are all in communication, a straight traveling state when traveling).

One end of the line Lr is connected to the oil tank 5.
The oil tank 5 and the suction side 2i of the pump 2 are connected by a line La.
A second merging point G2 is formed in the line La.

The valve unit 8 includes a housing (fixed portion) 80 integrated with the power cylinder 3 in a first position (a block having a left-turning flow path) 81 and a second position (a block having a straight-travel flow path). ) 82, a third position (block having a right-turning flow path) 83 is configured to be switchable.
Return springs 8sl and 8sr are provided at the left and right ends of the valve unit 8 in FIG. One end of the valve unit 8 (the end on the first position 81 side) is connected to the tip of the pitman arm 9 engaged with the gear box 4 via a connection link 89.
In the cylinder body 30 of the power cylinder 3, the left end 31 in FIG. 1 is connected to the knuckle arm 11 of the front wheel Wf via the first link 10.
The piston 32 sliding in the power cylinder 3 is connected to the vehicle frame 13 via the second link 12. In other words, the piston 32 is fixed, and the cylinder body 30 is configured to be movable in the left-right direction in FIG.

The gear box 4 is connected to the steering wheel 15 via the steering column 14, and when the driver steers the steering wheel 15, the pitman arm 9 engaged with the gear box 4 rotates as indicated by an arrow R.
When the driver steers the steering wheel 15 to the right (the operation of the arrow Rh), the pitman arm 9 rotates in the counterclockwise direction of the arrow R in FIG.
When the pitman arm 9 rotates in the counterclockwise direction of the arrow R, the valve unit 8 switches to the flow path at position 83, the hydraulic oil flows into the right chamber 34 of the power cylinder 3, and the pressure in the right chamber 34 increases. The pressure in the left chamber 33 is reduced.

Since the piston 32 in the power cylinder 3 is fixed to the frame 13 via the second link 12, when the right chamber 34 is pressurized, the power cylinder body 30 moves to the right. When the power cylinder body 30 moves to the right side, the knuckle arm 11 connected to the first link 10 is rotated clockwise (arrow RKR direction) with respect to the central axis of the king pin K of the front wheel Wf, and the vehicle is moved to the right. Turn.
At this time, part of the oil in the left chamber 33 is returned to the oil tank 5 via the hydraulic pressure return line Lr.

On the other hand, when the driver steers the steering wheel 15 to the left, the pitman arm 9 rotates clockwise as indicated by the arrow R, the flow path of the valve unit 8 becomes the flow path of the position 81, and the left chamber 33 of the power cylinder 3. Pressure increases, and the power cylinder body 30 moves to the left. Then, the knuckle arm 11 is rotated counterclockwise (arrow RKL direction) with respect to the central axis of the kingpin K, and the vehicle turns to the left.
That is, the pitman arm 9, the power cylinder main body 30, and the knuckle arm 11 move or rotate in the opposite direction to the case of right steering.

In the illustrated first embodiment, the configuration described below is added to the above configuration.
The first branch point B1 on the line L side and the second junction point G2 formed on the line La are connected by a bypass line Lb.
A flow rate adjusting valve 16 is interposed in the bypass line Lb, and an orifice 17 is interposed in a region between the flow rate adjusting valve 16 and the second junction G2.
The flow regulating valve main body 16 is an external pilot type sequence valve, and the pilot pressure is given by a signal pressure line Ls, and the signal pressure line Ls is constituted by lines Ls1 to Ls4.

A bypass flow rate adjustment valve drive circuit 18 and a bypass operation adjustment valve 19 are interposed in the signal pressure line Ls.
The bypass flow rate adjusting valve drive circuit 18 includes two check valves 181 and 182, and a third junction G3 located between the two check valves 181 and 182, and is configured as a so-called double check valve. Yes.
The check valve 181 is interposed in the line Ls1, and the line Ls1 communicates the junction G3 and the left chamber 33 of the power cylinder 3. By inserting the check valve 181 in the line Ls1, only the flow of hydraulic oil in the direction from the left chamber 33 of the power cylinder 3 toward the third junction G3 is passed through the line Ls1.
The check valve 182 is interposed in the line Ls2, and the line Ls2 communicates the junction G3 and the right chamber 34 of the power cylinder 3. By interposing the check valve 182 in the line Ls2, in the line Ls2, only the flow of hydraulic oil in the direction from the right chamber 34 of the power cylinder 3 toward the third junction G3 is passed.

When the hydraulic oil of the hydraulic pump 2 flows into the left chamber 33 of the power cylinder 3 and the cylinder body 30 moves left, the bypass flow rate adjusting valve drive circuit 18 changes the pressure of the hydraulic oil in the left chamber 33 to the line Ls1, It is supplied to the line Ls3 via the check valve 181 and the third junction G3. The pressure of the hydraulic oil in the left chamber 33 is supplied as a pilot pressure (signal pressure) to the flow rate adjustment valve 16 of the bypass line Lb via the line Ls4.
On the other hand, when the working flow flows into the right chamber 34 of the power cylinder 3 and the cylinder body 30 moves to the right, the bypass flow rate adjustment valve drive circuit 18 indicates that the pressure of the hydraulic oil in the right chamber 34 is reversed to the line Ls2. It is supplied to the line Ls3 via the stop valve 182 and the third junction point G3. Then, the pressure of the hydraulic oil in the right chamber 34 is given to the flow rate adjustment valve 16 of the bypass line Lb as a pilot pressure (signal pressure) via the line Ls4.

A portion surrounded by a dotted line denoted by reference numeral 19 indicates a bypass operation regulating valve 19 where the line Ls3 and the line Ls4 are connected.
The bypass operation adjusting valve 19 includes a check valve 191, a check valve 192, a line Ls3, a line Ls4, a line Ls5, a fifth branch point B5, and an orifice 193.
The line Ls3 connects the third junction point G3 and the fifth branch point B5, and interposes a check valve 191.
The line Ls4 connects the fifth branch point B5 and the flow rate adjusting valve 16 (to the signal pressure input side). The line Ls5 is the fifth branch point B5 and the fourth junction formed in the line L5. G4 (the valve unit 8 or the inlet side of the power cylinder 3) is connected, and a check valve 192 is interposed. An orifice 193 is interposed in a region between the check valve 192 and the fourth junction point G4.

When a pilot pressure (signal pressure) higher than a predetermined pressure is applied, the flow rate adjustment valve 16 is closed. On the other hand, when the pilot pressure (signal pressure) is lower than the predetermined pressure, the flow regulating valve body 16 is opened.
When the vehicle turns, the load acting on the power cylinder 3 increases, and either the left chamber 33 or the right chamber 34 of the power cylinder 3 increases in pressure. The hydraulic fluid (or its pressure) on the boosted side of the left chamber 33 and the right chamber 34 flows through one of the lines Ls1 and Ls2, flows through one of the check valves 181 and 182, and the third junction G3, Via the line Ls3, the check valve 191, the fifth branch point B5, and the line Ls4, it is added as the pilot pressure (signal pressure) of the flow rate adjusting valve 16. When the pilot pressure (signal pressure) increases and exceeds a predetermined value, the flow rate adjustment valve 16 is closed.
If the flow regulating valve 16 is closed, the hydraulic oil discharged from the hydraulic pump 2 is reliably sent to the power cylinder 3 without flowing through the bypass line Lb, and the power cylinder 3 exhibits a sufficient steering assist force. To do.
For the one that communicates with the non-pressurized side (the decompressed side) of the left chamber 33 and the right chamber 34 of the lines Ls1, Ls2, the hydraulic oil flowing through the opposite line flows in by the check valves 181 and 182. And the pressure is never transmitted.

When the load acting on the power cylinder 3 is small as when the vehicle is traveling straight, either the left chamber 33 or the right chamber 34 of the power cylinder 3 is not boosted. Further, the pressure on the inlet side of the power cylinder 3 is also lowered, and the pressure at the fourth junction G4 is also lowered.
As a result, the hydraulic oil of the pilot pressure (signal pressure) of the flow rate adjusting valve 16 in the line Ls4 passes through the line Ls4, the fifth branch point B5, and the line Ls5, and the fourth merging on the inlet side of the power cylinder 3 Flowing to point G4, the pilot pressure (signal pressure) of the flow rate adjustment valve 16 becomes lower than a predetermined value, the flow rate adjustment valve 16 is opened, and the bypass line Lb is also opened.
Since the bypass line Lb is opened, most of the hydraulic oil discharged from the hydraulic pump 2 flows through the bypass line Lb and is returned to the inlet side 2i of the hydraulic pump 2.
In addition, since the orifice 17 is interposed in the bypass line Lb, the amount of pressure oil supplied to the power cylinder 3 via the pressure oil line L can be ensured by the amount of pressure loss caused by the orifice 17. .

FIG. 2 shows a change mode (FIG. 2: characteristic A) of the steering wheel operation angle when a lane change (for example, a lane change to the right lane) is performed during traveling.
FIG. 3 shows changes in the internal pressures (B ₃₃ , B ₃₄ ) of the left chamber 33 and the right chamber 34 of the power cylinder 3 due to the lane change of FIG. 2, and shows the characteristics of the bypass line “when open”. .
FIG. 4 shows changes in the internal pressures (B ₃₃ , B ₃₄ ) of the left chamber 33 and the right chamber 34 of the power cylinder 3 due to the lane change of FIG. 2, and shows the characteristic of the bypass line “when closed”. .
FIG. 5 shows the opening characteristic (characteristic B) of the flow rate adjusting valve 16 and the bypass flow characteristic (characteristic C) in the lane change (lane change in which the change in the steering operation angle is shown in FIG. 2).
FIG. 6 shows the drive power characteristic (characteristic D) of the hydraulic pump 2 in accordance with the lane change (the change in the steering wheel operation angle is shown in FIG. 2).

In FIG. 2, first, the vehicle is traveling straight (A1 region), and the steering wheel 15 is once steered to the right to perform a lane change (the steering angle in the right direction increases slightly in the region A2).
Subsequently, the steering wheel 15 is returned toward the neutral position (the steering angle is decreased in the region A3).
When the neutral point Pc (82 position in the valve unit 8) is passed, the steering wheel 15 is now steered to the left (the steering angle in the left direction increases slightly in the area A4).
Next, the steering wheel 15 is returned again toward the neutral position (the steering angle is reduced in the region A5).
When the lane change is completed, the vehicle returns to the straight traveling state (area A6).

In FIG. 3 showing the pressure characteristics in the power cylinder 3 when the bypass line Lb is opened, first, the internal pressure (characteristic B ₃₄ ) in the right chamber 34 rises, then maintains a constant value, and falls at the neutral point Pc. After the neutral point Pc continues, the internal pressure (characteristic B ₃₃ ) of the left chamber 33 increases, maintains a constant value, and decreases immediately before the end of the lane change.
In FIG. 3, since the bypass line Lb is opened, the inlet pressure B of the valve unit 8 is not sufficient, and the power cylinder 3 cannot obtain the required hydraulic pressure.
A broken line in FIG. 3 indicates a pressure (operating pressure) Pa at which the flow rate adjusting valve 16 is closed.

Here, at the time of steering, it is desired to close the bypass line Lb as shown in FIG. 4 to sufficiently increase the pressure in the power cylinder 3.
However, when the flow regulating valve 16 is simply operated in proportion to the internal pressure of the power cylinder 3, the bypass line Lb instantaneously “opens” in the region near the turning point Pc (the portion circled by the broken line in FIG. 4). In the reverse steering, the hydraulic pressure is insufficient and the responsiveness deteriorates.

On the other hand, in the first embodiment, when turning back the steering, the internal pressure of the power cylinder 3 decreases near the turning point Pc, and the signal pressure of the flow rate adjusting valve 16 changes from the line Ls5 to the inlet side of the valve unit 8. When exiting (decreasing pressure), the pressure decreasing speed is delayed by the action of the orifice 193 interposed in the line Ls5 (characteristic B4y in FIG. 4).
Furthermore, even when the internal pressure of the power cylinder 3 decreases just before the end of the lane change, the rate of decrease in the signal pressure of the flow rate adjustment valve 16 is delayed in the same manner as in the vicinity of the turning point Pc (characteristic B3y).
Also in FIG. 5, as indicated by the C characteristic (opening characteristic), the flow rate adjusting valve 16 starts to open slowly (characteristic Cy <b> 1) when turning back after the first right steering (near the turning point Pc). However, left steering is immediately performed, and the internal pressure of the left chamber 33 of the cylinder body 30 increases, so that the flow regulating valve 16 is “closed” again.

In FIG. 5, the broken line B indicates the signal pressure (pilot pressure) of the flow rate adjustment valve 16. The opening degree characteristic C of the flow regulating valve 16 is given by the signal pressure (pilot pressure) of the flow regulating valve 16 having the characteristic B indicated by the broken line.
In FIG. 5, characteristics Cx1 and Cx2 indicate the opening characteristics of the flow regulating valve 16 when the orifice 193 is not provided. On the other hand, characteristics Cy <b> 1 and Cy <b> 2 indicate opening characteristics of the flow rate adjustment valve 16 when the operation effect of the orifice 193 is achieved.
According to FIG. 5, when the orifice 193 is provided, the opening characteristics Cy1 and Cy2 of the flow rate adjusting valve 16 are in the middle of the lane change, that is, in the vicinity of the rudder turning point Pc (see FIG. 2) or just before the end of the lane change. As compared with the characteristics Cx1 and Cx2 when the orifice 193 is not provided, the rate of change of the opening is obviously slow.

FIG. 6 shows a change (characteristic D) of the driving power of the hydraulic pump 2 when the lane change described above with reference to FIGS.
FIG. 6 also shows a characteristic Do as a comparison object when the bypass line Lb is not provided.
According to FIG. 6, for example, when driving straight, the driving power is reduced by ΔWp in the illustrated embodiment as compared to the case where the bypass line Lb is not provided.

According to the illustrated first embodiment, by providing the bypass line Lb with the flow rate adjusting valve 16, the flow rate adjusting valve 16 is opened when the load of the power cylinder 3 is low, such as when the vehicle is traveling straight. The discharge port 2o and the suction port 2i of the hydraulic pump 2 are communicated with each other.
As a result, the discharge port 2o and the suction port 2i of the hydraulic pump 2 are communicated with each other, so that power consumption for circulating the pressure oil is reduced and wasteful power consumption is reduced.
That is, the auxiliary machine drive power of the engine 1 is reduced and the fuel consumption is also improved.

  On the other hand, when the load of the power cylinder 3 is high as in steering, the flow rate adjustment valve 16 is closed and the bypass line Lb is also closed. As a result, the pressure oil discharged from the hydraulic pump 2 is reliably supplied to the valve unit 8 of the power cylinder unit 3, and a sufficient steering assist force can be generated during steering.

Further, according to the first embodiment, the bypass line Lb having the flow rate adjusting valve 16 interposed therebetween, the lines Ls1 to Ls4 communicating the left chamber 33 and the right chamber 34 of the power cylinder 3 and the flow rate adjusting valve 16, and the valve unit. It is only necessary to add the line Ls5 communicating with the upstream region to the existing power steering mechanism.
Therefore, the existing parts of the vehicle that already has the power steering mechanism can be utilized to easily carry out the existing vehicle by so-called “retrofitting”.

Further, in the first embodiment, for example, when shifting from steering to straight traveling, the power cylinder 3 shifts from a high load state to a low load state, and in the left chamber 33 or the right chamber 34 of the cylinder 3. The high signal pressure (pilot pressure) that has been applied to the flow rate adjustment valve 16 until then is transmitted to the upstream region of the valve unit 8 from the branch point B5 via the line Ls5.
Here, in the first embodiment, since the orifice 193 is provided in the line Ls5, a high signal pressure (pilot pressure) of the flow rate adjusting valve 16 is transmitted to the upstream region of the valve unit 8, and the signal pressure (pilot pressure) is transmitted. ) Will cause a slight delay before the pressure drops.

  For example, when changing lanes, the vehicle is steered and instantaneously goes straight, and then steers in the opposite direction again. When going straight ahead instantaneously, that is, when the load of the power cylinder 3 becomes low, when the flow control valve 16 immediately shifts to the open state and the pressure of the power cylinder 3 is reduced, the reverse direction is then reached. When steering, no steering assist force is generated and the steering is very heavy.

  In contrast, in the illustrated first embodiment, when the load of the power cylinder 3 shifts from a high state to a low load state, the flow rate adjustment valve 16 shifts from a closed state to an open state. A slight delay is caused by the orifice 193. Therefore, the pressure of the power cylinder 3 is not immediately reduced during such instantaneous straight traveling, and a state in which the steering becomes suddenly heavy is avoided.

Similarly, in the illustrated first embodiment, since the orifice 17 is interposed in the bypass line Lb, the flow rate adjustment valve 16 is opened, and the discharge port 2o and the suction port 2i of the hydraulic pump 2 are bypassed. However, the amount of pressure oil supplied to the power cylinder 3 via the pressure oil line L can be ensured by the amount of pressure loss at the orifice 17.
Therefore, it is possible to prevent a situation in which the pressure in the power cylinder 3 immediately disappears and the steering becomes too heavy when steering is performed thereafter.

Next, a second embodiment will be described with reference to FIG.
The steering mechanism according to the second embodiment shown in FIG.
In the first embodiment of FIG. 1, there is one flow rate control valve interposed in the bypass line Lb, whereas in the second embodiment of FIG. 7, two flow rate control valves 16A, 16B is interposed. Further, the configuration of the line for applying the signal pressure (pilot pressure) to the flow control valves 16A and 16B is also different from that of the first embodiment.
Hereinafter, the second embodiment will be described mainly with respect to differences from the first embodiment.

In FIG. 7, a first flow control valve 16A and a second 16B are interposed in the bypass line Lb.
The first flow control valve 16A is interposed in a region on the first branch point B1 side in the bypass line Lb. And the 2nd flow control valve 16B is interposed in the field by the side of the 2nd confluence G2 in bypass line Lb.
An orifice 17 is interposed in a region between the second flow rate adjustment valve 16B and the second junction G2 in the bypass line Lb.
The signal pressure (pilot pressure) input side of the first flow control valve 16A communicates with the right chamber 34 of the power cylinder 3 via the first signal pressure line Ls6. On the other hand, the signal pressure (pilot pressure) input side of the second flow control valve 16B communicates with the left chamber 33 of the power cylinder 3 via the second signal pressure line Ls7.

A first bypass operation control valve 20 is interposed in the first signal pressure line Ls6.
The first bypass operation control valve 20 includes a first signal pressure line Ls6, two branch points B61 and B62 formed in the line Ls6, an internal bypass line Lb6, a signal transmission check valve 201, and a residual pressure return valve. A check valve 202 and an orifice 203 are provided.

  The branch point B61 is formed in a region on the right chamber 34 side in the first signal pressure line Ls6, and the branch point B62 is formed in a region on the flow rate adjustment valve 16A side in the first signal pressure line Ls6.

The internal bypass line Lb6 connects the branch points B61 and B62, and bypasses the signal transmission check valve 201 interposed in the first signal pressure line Ls6.
The signal transmission check valve 201 is interposed in a region between the branch points B61 and B62 in the first signal pressure line Ls6.
The residual pressure return check valve 202 is interposed in the internal bypass line Lb6. The orifice 203 is interposed in a region between the branch point B61 and the residual pressure return check valve 202 in the internal bypass line Lb6.

A second bypass operation control valve 21 is interposed in the second signal pressure line Ls7.
The second bypass operation control valve 21 includes a second signal pressure line Ls7, two branch points B71 and B72 formed in the line Ls7, an internal bypass line Lb7, a signal transmission check valve 211, and a residual pressure return valve. A check valve 212 and an orifice 213 are provided.
The branch point B71 is formed in a region on the left chamber 33 side in the second signal pressure line Ls7, and the branch point B72 is formed in a region on the flow rate adjustment valve 16B side in the second signal pressure line Ls7.

The internal bypass line Lb7 connects the branch points B71 and B72, and bypasses the signal transmission check valve 211 of the second signal pressure line Ls7.
The signal transmission check valve 211 is interposed in a region between the branch points B61 and B62 in the second signal pressure line Ls7.
The residual pressure return check valve 212 is interposed in the internal bypass line Lb7. The orifice 213 is interposed in a region between the branch point B71 and the residual pressure return check valve 212 in the internal bypass line Lb7.

According to the second embodiment of FIG. 7, by providing the bypass line Lb interposing the flow rate adjusting valves 16A, 16B, the load on the power cylinder 3 is low as in the case of straight traveling of the vehicle, and the left chamber 33, the right chamber When both of the pressures of 34 are low, both the flow rate adjusting valves 16A and 16B are opened, and the discharge port 2o and the suction port 2i of the hydraulic pump 2 are communicated.
As a result, the discharge port 2o and the suction port 2i of the hydraulic pump 2 are communicated with each other, so that power consumption for circulating the pressure oil is reduced and wasteful power consumption is reduced.
And the auxiliary machine drive power of the engine 1 is reduced and the fuel consumption is also improved.

  On the other hand, when the load of the power cylinder 3 is high as during steering, the pressure in either the left chamber 33 or the right chamber 34 of the cylinder 3 increases, so that the left chamber 33 or the right chamber 34 is increased in pressure. Any of the flow rate adjusting valves 16A, 16B communicating with either of them is closed, and the bypass line Lb is also closed. As a result, the pressure oil discharged from the hydraulic pump 2 is reliably supplied to the valve unit 8 of the power cylinder 3 without flowing through the bypass line Lb, and a sufficient steering assist force can be generated during steering.

According to the second embodiment, the bypass line Lb including the two flow rate adjusting valves 16A and 16B, the two flow rate adjusting valves 16A and 16B, the left chamber 33 of the power cylinder unit 3, and the right chamber 34 are provided. The two signal pressure lines Ls6 and Ls7 communicating with each other can be implemented simply by adding to the existing power steering mechanism.
Therefore, the existing parts of the vehicle that already has the power steering mechanism can be used easily for the existing vehicle by so-called “retrofitting”, and the introduction cost can be reduced.

In the second embodiment of FIG. 7, as in the case of transition from steering to straight travel, the power cylinder 3 shifts from a high load state to a low load state, and the left chamber 34 or the right chamber 33 of the power cylinder 3. When the pressure at the flow control valve 16 becomes lower, the higher pressure on the flow rate adjusting valve 16A or 16B side is set at the left chamber 34 of the power cylinder 3 via any of the internal bypass lines Lb6 and Lb7 in either of the bypass operation control valves 20 and 21. Alternatively, it is transmitted to any one of the right chambers 33 with a time difference.
Here, since the orifices 203 and 213 are interposed in the internal bypass lines Lb6 and Lb7, a slight delay occurs until the high pressure on the flow rate adjusting valves 16A and 16B side is transmitted to the cylinder 3.

  For example, when changing lanes, the vehicle is steered and instantaneously goes straight, and then steers in the opposite direction again. When going straight ahead instantaneously, that is, when the load of the power cylinder 3 becomes low, when the flow control valve immediately shifts to the open state and the pressure of the power cylinder 3 is reduced, the steering is then reversed. When doing so, steering assist force is not generated, and the steering is very heavy.

  In contrast, in the second embodiment of FIG. 7, when the load of the power cylinder 3 is shifted from a high load state to a low load state, the flow rate adjustment valve 16A or 16B shifts from a closed state to an open state. However, at that time, a slight delay occurs in the pressure oil flowing to the power cylinder 3 side by either of the orifices 203 and 213. Therefore, even if the vehicle goes straight ahead instantaneously due to a lane change or the like, the pressure of the power cylinder 3 is not immediately reduced, and a state in which the steering becomes suddenly heavy is avoided.

Further, in the second embodiment of FIG. 7, since the orifice 17 is interposed in the bypass line Lb, the flow rate adjusting valves 16A and 16B are opened, and the discharge port 2o and the suction port 2i of the hydraulic pump 2 are bypassed. Even if this is done, the amount of pressure oil supplied to the power cylinder 3 via the pressure oil line L can be ensured by the amount of pressure loss at the orifice 17.
Therefore, it is possible to prevent the pressure in the power cylinder 3 from being lost and the steering from becoming too heavy when steering is performed thereafter.

  Other configurations and operational effects in the second embodiment of FIG. 7 are the same as those of the first embodiment of FIGS.

  The illustrated embodiment is merely an example, and is not intended to limit the technical scope of the present invention.

DESCRIPTION OF SYMBOLS 1 ... Engine 2 ... Hydraulic pump 3 ... Power cylinder unit / power cylinder 4 ... Gear box 5 ... Oil tank 6 ... Flow control valve 7 ... Pressure control valve 8 ... Valve unit 9 Pitman arm 10 First link 11 Knuckle arm 12 Second link 13 Frame 14 Steering column 15 Steering wheel 16 16A, 16B ... Bypass flow control valve 17 ... Orifice 18 ... Bypass flow control valve drive circuit 19 ... Bypass operation control valve 20, 21 ... Bypass operation control valve

Claims (6)

In a power steering mechanism having a hydraulic pump directly connected to an engine, a power cylinder unit, a reservoir tank, and a hydraulic oil line communicating the discharge port of the hydraulic pump with the valve unit of the power cylinder, the discharge port of the hydraulic pump The bypass line communicates with the suction port, and a flow rate adjusting valve is interposed in the bypass line, and the pressure in the cylinder section partitioned by the piston of the power cylinder unit is a signal pressure in the flow rate adjusting valve. It is closed when the signal pressure is high, and closes when the signal pressure is low. The signal pressure line communicates with each of the cylinder sections partitioned by the piston. , A line communicating with the signal pressure input side of the flow control valve, a line communicating with the valve unit upstream side (hydraulic pump discharge side) of the pressure oil line, a line communicating with the signal pressure input side and the valve unit upstream A power steering mechanism comprising a line that communicates a branch point of a line that communicates with a side region and a merge point of a line that communicates with each of the cylinder sections. In a power steering mechanism having a hydraulic pump directly connected to an engine, a power cylinder unit, a reservoir tank, and a hydraulic oil line communicating the discharge port of the hydraulic pump with the valve unit of the power cylinder, the discharge port of the hydraulic pump The bypass line communicates with the suction port, and a flow rate adjusting valve is interposed in the bypass line, and the pressure in the cylinder section partitioned by the piston of the power cylinder unit is a signal pressure in the flow rate adjusting valve. And is closed when the signal pressure is high, and opens when the signal pressure is low. Two flow rate adjusting valves are provided in the bypass line, and the two flow rates are The adjusting valve is characterized in that the pressure in each of the cylinder sections partitioned by the piston of the power cylinder unit is input as a signal pressure. Power steering mechanism that. Each of the lines communicating with each of the cylinder sections partitioned by the piston is provided with a check valve in an area closer to the cylinder side than the merging point, and the line communicating the merging point and the branching point is a branching point. 2. The power steering mechanism according to claim 1, wherein a check valve is communicated with a region closer to the junction point, and a check valve and a pressure loss mechanism are interposed in a line communicating with a region upstream of the valve unit. Each of the cylinder sections partitioned by the piston and the signal pressure input side of each of the flow rate regulating valves are communicated with each other via a signal pressure line, and each of the signal pressure lines increases the pressure of the cylinder section. A first path for transmitting the pressure rise to the signal pressure input side of the flow regulating valve, and a high pressure on the flow regulating valve side is transmitted to the section with a time difference when the pressure of the cylinder section drops. The power steering mechanism according to claim 2, further comprising a second path. A check valve that transmits the signal pressure only toward the flow rate adjustment valve is interposed in the first path, and a check valve that transmits the signal pressure only toward the cylinder section and the second path. The power steering mechanism according to claim 4, wherein a pressure loss mechanism is interposed. The power steering mechanism according to any one of claims 1 to 5, wherein a pressure loss mechanism is interposed in the bypass line.

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