JPH01141175A - Power steering - Google Patents

Abstract

PURPOSE:To enhance the extent of rectilinear safety at the time of high speed traveling by installing a flow regulating valve with a valve body, which moves according to a pressure differential equivalent to the discharge and force of a spring with a nonlinear spring characteristic, at the discharge side of a hydraulic pump feeding a power cylinder with pressure oil. CONSTITUTION:A power steering is provided with a directional control valve 2 which selects the feeding direction of pressure oil out of a hydraulic pump P to a steering auxiliary hydraulic cylinder according to a steering torque and a steering direction at the time when performing its steering operation. In this case, a flow control valve 1 is installed in a front stage of the directional control valve 2 in a discharge passage 4 of the hydraulic pump P. This flow control valve 1 is provided with a valve body 12 which is moved according to a pressure differential equivalent to discharge of the hydraulic pump P at both upper and down streams of a contracted diametral part 14a, and an energizing force of a spring 15 being added to the reverse direction to the acting direction of this pressure differential pressure, and it is constituted to discharge the excessive oil from a discharge port 11b according to movement of this valve body 12. As for the spring 15, such one that has a characteristic for producing a nonlinear displacement to the load is used.

Description

[Detailed Description of the Invention] [Field of Industrial Application] The present invention relates to a hydraulic control valve that switches the supply direction of pressure oil to a hydraulic cylinder for steering assistance according to the operating direction of a steering wheel, and The present invention relates to a hydraulic power steering device equipped with a restraint device that restrains switching operations.

[Prior art]

A hydraulic power steering device uses a hydraulic cylinder installed in the steering mechanism to generate steering assist force to reduce the force required to operate the steering wheel. The hydraulic pump driven by the steering wheel is connected via a rotary directional control valve that switches the oil passage in response to rotation of the steering wheel.

This directional control valve has an input shaft that rotates in conjunction with the steering wheel,
The output shaft, which rotates around the axis in conjunction with the steering mechanism, is coaxially connected via a torsion bar, and the connecting end of the input shaft is provided with a long groove extending in the axial direction on its outer peripheral surface. A plurality of grooves are formed equally spaced to form a valve body, while a cylindrical casing having long grooves of the same number of trees as the aforementioned long grooves formed equally spaced on the inner peripheral surface is attached to the connecting end of the output shaft. The valve body is rotatably fitted into the casing, and a relative angular displacement is generated between the valve body and the casing in response to the torsion generated in the torsion bar as the steering wheel is operated. The long grooves of the casing are alternately communicated with both oil chambers of the hydraulic cylinder, and the long grooves of the valve body are alternately communicated with the hydraulic pump, which is a high pressure source, and an oil tank maintained at a low pressure. It is.

Therefore, when the steering wheel is rotated in one direction and a relative angular displacement occurs between the valve body and the casing as described above, the long groove of the valve body communicating with the hydraulic pump and the circumferentially both sides thereof The area of the gap between the long grooves of adjacent casings increases on one side and decreases on the other according to the relative angular displacement, and the pressure in the long groove of the casing on the side of the gap where the area has increased and the area As a result, there is a difference between the side where the
The hydraulic cylinder generates a steering assist force corresponding to this pressure difference.

Now, the force required to operate the steering wheel varies depending on the vehicle speed, and while driving at low speeds and when stopped, a large operating force is required to operate the steering wheel, while when driving at high speeds, a small operating force is required to operate the steering wheel. be able to. Therefore, in a power steering system, in order to reduce the driver's labor burden when driving at low speeds and when stopping, the hydraulic cylinder for steering assistance is required to have characteristics that allow it to generate as large a steering assistance force as possible. If the same steering assist force is generated when driving at high speeds as when driving at low speeds or when stopped, steering will be performed with a small amount of operational force applied to the steering wheel when the vehicle is traveling straight, resulting in a deterioration of straight-line stability. When traveling at high speeds, the hydraulic cylinder is required to have a property of generating almost no steering assist force in order to impart appropriate rigidity to the steering wheel. As a power steering device that realizes these contradictory characteristics, there is a power steering device disclosed in Japanese Patent Application Laid-Open No. 61-200063.

In this power steering device, when in a steering situation where a large steering assist force is not required, such as when driving at high speed, the relative angular displacement between the valve body and the casing, in other words, the relative angular displacement between the human power shaft and the output shaft. The device is equipped with a restraint device that restrains the switching operation of the oil passage in the directional control valve by restraining the relative angular displacement between the valve body and the connecting end of the input shaft. The connecting end of the output shaft is formed into a cylindrical shape, and the extended portion of the input shaft is fitted into the cylindrical portion, and the cylindrical portion is radially penetrated through the cylindrical portion. A plunger is fitted into each of the plurality of circular holes so as to be slidable in the axial direction of the circular hole, and a plunger is fitted around the outer periphery of the cylindrical portion in such a manner that the circular holes communicate with each other. Pressure oil having a pressure according to the steering situation is introduced into the formed annular oil chamber, and the pressing force of this pressure oil presses the tip of the plunger against the outer peripheral surface of the extension of the input shaft, thereby connecting the output shaft and the input shaft. It is configured to restrict relative angular displacement with the shaft, and an oil path from the hydraulic pump to the directional control valve is branched into the annular oil chamber, passing through a variable orifice and a fixed orifice in this order. The pressure between the two throttles is guided in the branched oil path leading to the tank.

The pressure in the annular oil chamber, in other words, the restraint force generated by the restraint device, is determined by automatically adjusting the opening degree of the variable throttle based on the detection results of physical quantities indicative of steering conditions such as vehicle speed and steering angle. It is adjusted by changing the pressure between the throttle and the fixed throttle. For example, when the opening degree of the variable throttle is increased or decreased depending on the size of the detection chain, the pressure between the variable throttle and the fixed throttle is
The restraint force generated by the restraint device increases or decreases depending on the vehicle speed, and as a result, the restraint force generated by the restraint device increases or decreases depending on the vehicle speed.
When driving at high speeds, a large restraining force is obtained, and stiffness corresponding to this restraining force is imparted to the steering wheel, improving straight-line stability. On the other hand, when driving at low speeds, the switching operation of the directional control valve is hardly restrained and is small. The operating force can cause a relative angular displacement between the valve body and the casing, and a large steering assist force can be obtained.

[Problem that the invention seeks to solve]

As mentioned above, the purpose of the restraint device is to provide rigidity to the steering wheel in order to improve straight-line stability.
Maximum restraint force generation is required during straight-line driving at high speeds. However, in the conventional power steering device configured as described above, the pressure within the annular oil chamber that contributes to the restraining force, that is, the pressure between the variable throttle and the fixed throttle in the middle of the branch oil path, is
While it is determined by the opening degree of the variable throttle and the pressure upstream of the branch point of the branch oil passage, the pressure upstream of this branch point is determined according to the resistance of the oil passage on the side connected to the directional control valve, and this resistance is mainly the flow resistance inside the directional control valve, so when driving straight ahead when there is no relative displacement between the valve body and the casing of the directional control valve, high pressure does not occur upstream of the branch point and there is no pressure inside the annular oil chamber. As a result of not introducing high pressure, the restraint device cannot generate a large restraint force and cannot provide sufficient rigidity to the steering wheel.

This difficulty can be solved by using a sufficiently large capacity hydraulic pump, or by providing another hydraulic pump exclusively for supplying hydraulic pressure to the restraint device, as disclosed in the above-mentioned Japanese Patent Application Laid-Open No. 61-200063. However, in this case, the horsepower consumption of the hydraulic pump increases, which increases the fuel consumption of the engine, and also increases the amount of heat generated by the hydraulic pump and increases the noise generated. A new problem arises.

The present invention was made in view of the above circumstances, and provides sufficient rigidity to the steering wheel when traveling straight at high speed without increasing the fuel consumption of the engine, the amount of heat generated by the hydraulic pump, and the noise generated. can be granted,
An object of the present invention is to provide a power steering device that can improve straight-line stability.

[Means for solving problems]

The power steering device according to the present invention is driven by an engine. A direction control valve is provided for switching the direction of supply of pressure oil from the hydraulic pump to the steering assist hydraulic cylinder in accordance with the operating direction of the steering wheel, and the hydraulic oil is supplied from the hydraulic pump with reduced pressure depending on the steering condition. In the power steering device, the power steering device is equipped with a restraining device that restrains the switching operation of the directional control valve with a force corresponding to the pressure of the pressure oil. A control valve that moves in response to a pressure difference corresponding to the discharge amount of the hydraulic pump and a biasing force of a spring that is applied in a direction opposite to the acting direction of the pressure difference and causes a nonlinear displacement with respect to the load. The valve body is characterized by comprising a body and an excess oil discharge hole whose opening area changes according to the movement of the valve body.

[Effect]

In the present invention, the discharge oil of the hydraulic pump driven by the engine is branched and supplied to the discharge hole, the restraint device, and the direction control valve according to the movement of the valve body of the flow control valve, and further, The valve body exhibits a non-linear movement mode due to the biasing force of the spring, and the amount of pressure oil supplied to the restraint device and the directional control valve increases when the vehicle speed is high. The highest oil pressure available in the device is high when the vehicle speed is high.

〔Example〕

The present invention will be described in detail below based on drawings showing embodiments thereof. FIG. 1 is a schematic diagram showing the configuration of a power steering device according to the present invention, together with a longitudinal cross-sectional view of a restraint device and a direction control valve, and a longitudinal cross-sectional view of a flow control valve that is a feature of the present invention.

The power steering device according to the present invention includes a hydraulic pump P driven by an engine (not shown), a hydraulic cylinder S for steering assistance configured in the steering mechanism, and disposed on the discharge side of the hydraulic pump P, A flow rate control valve l that controls the discharge amount of the pump P
, a direction control valve 2 that switches the feeding direction of pressure oil to the hydraulic cylinder S in accordance with the operating direction of a steering wheel (not shown), and a restraint device that is installed in parallel with the direction control valve 2 and restrains the switching operation thereof. It is composed of 3 etc. The discharge oil passage 4 of the hydraulic pump P passes through the flow rate control valve l and is connected to the first oil passage 41.
The first oil passage 41 is connected to the pressure oil introduction hole 22 of the directional control valve 2 via the fixed throttle 5. The second oil passage 42 communicates with the oil tank T via a variable throttle 6 and a fixed throttle 7.

The directional control valve 2 is similar to that disclosed in Japanese Patent Application Laid-Open No. 61-200063, and is coaxially rotatably supported within a cylindrical housing 80 and connected via a torsion bar 81. The input shaft 8 and the output shaft 9 are connected to each other. The input shaft 8 is a hollow shaft whose protruding end from the housing 80 is connected to a steering wheel (not shown), and rotates as the steering wheel rotates.
A pinion 90 formed on the outer periphery of the middle part is engaged with a rack shaft 91 in the steering mechanism, rotates as the rank shaft 91 moves in the axial direction, and connects to the input shaft 8. It is a solid shaft formed by molding into a cylindrical shape.

The valve body 20 of the directional control valve 2 has a plurality of long grooves 20a extending in the axial direction on the outer circumferential surface of the input shaft 8 in the middle.

20a... are formed equally spaced in the circumferential direction, and the casing 21 of the directional control valve 2 has the same number of long grooves 21a extending in the axial direction as the long grooves 20a, 20a... , 21a... are formed on the inner peripheral surface thereof at equal intervals in the circumferential direction, and three annular grooves 21b, 21
b, 21b are formed on its outer peripheral surface, and is rotatably fitted into the housing 80, and its rotation is restrained by a fixing pin 21c at the connecting end of the output shaft 9. It is installed by hand. The direction control valve 2 has the long groove 20a.
, 20a... and long grooves 21a, 21a... are positioned alternately so that adjacent ones communicate with each other through a small gap in the circumferential direction, and the valve body 20 is inserted into the casing 21. The valve body 20 whose rotation is restrained by the steering wheel and the casing 21 whose rotation is restrained by the rack shaft 91 correspond to the torsion generated in the torsion bar 81 as the steering wheel is operated. There should be no relative angular displacement between them.

In the directional control valve 2 configured in this way, the length of the casing 21 is set as above. n 21a, 21a... and valve body 2
A plurality of spaces are formed between the long grooves 20a, 20a, . A similar space is also formed in between. The former space is alternately communicated with the annular grooves 21b, 21b on both sides in the axial direction of the three annular grooves 21b, 21b, 21b through communication passages passing through the casing 21 in the radial direction, Half of the latter spaces, which are located every other space, are communicated with the central annular groove 21b via communication passages formed similarly to the communication passages, and the remaining half are arranged so that the valve body 20 is radially connected to the annular groove 21b. The input shaft 8 is communicated with the hollow portion of the input shaft 8 via a communication path penetrating through the input shaft 8 . On the other hand, the central annular groove 21b
The first oil passage 41 is connected to the first oil passage 41 through an oil introduction hole 22 that is opened to the outside of the annular groove 21 on both sides.
b and 21b are connected to both oil chambers of the hydraulic cylinder S through oil outlet holes 23 and 24 opened to the outside of the housing 80, respectively. The housing 80 also has an oil discharge hole 25 connected to the oil tank T.
The kernel oil discharge hole 25 communicates with the hollow portion of the input shaft 8 within the housing 80 .

Pressure oil from the hydraulic pump P flows through the discharge oil passage 4 and the first oil passage 41, and enters the long groove 20a of the valve body 20 through the oil introduction hole 22 and the central annular groove 21b. Then, it passes through the gaps on both sides of the long groove 20a, and is introduced into the long grooves 21a, 21a of the casing 21 adjacent thereto. When no operating force is applied to the steering wheel and no relative angular displacement occurs between the valve body 20 and the casing 21, the gap areas on both sides of the long groove 20a are equal, and the adjacent long groove 21 a * Since no pressure difference occurs between the long grooves 21a, the pressure oil introduced into the long grooves 21a, 21a is transferred to the adjacent long groove 2 on the opposite side from the long groove 20a.
0a, 20a, further passes through the hollow part of the input shaft 8, and returns to the oil tank T through the oil discharge hole 25, so that no pressure difference is generated between both arm chambers of the hydraulic cylinder S. , the cylinder S does not generate any steering assist force. This is a straight-ahead running condition.

On the other hand, when an operating force is applied to the steering wheel to perform steering, a relative angular displacement occurs between the valve body 20 and the casing 21 in accordance with the accompanying twisting of the torsion bar 81. The long groove 20a into which pressure oil is introduced
The gap area on both sides increases on one side and decreases on the other. As a result, the pressure in the long groove 21a adjacent to the long groove 20a on the side where the area increased is higher than the pressure in the long groove 21a adjacent to the other side. Also, a pressure difference is created between both sleeve chambers of the hydraulic cylinder S which communicates with these long grooves 21a and 21a through the oil outlet holes 23 and 24, respectively, and the pressure of the cylinder S is increased according to this pressure difference. The operation provides a steering assist force in a direction corresponding to the operating direction of the steering wheel.

Further, as shown in FIG. 1 and FIG. 2, which is an enlarged sectional view taken along the line ■-■, the restraint device 3 is arranged equidistantly in the circumferential direction on the cylindrical portion of the output shaft 9. Four insertion holes 30, 30... formed in a manner that penetrates in the radial direction.
A short cylindrical plunger 31.31.
... are fitted so as to be slidable in the axial direction, and each plunger 31, 31... is fitted on the outer periphery of the input shaft 8 at a portion that matches the mounting position of the plunger 31, 31... in the axial direction.・
It has a semicircular axial cross-sectional shape that matches the inner end of the hemisphere, and is constructed by forming four concave portions 32, 32, . . . equally spaced. The insertion hole 30.30...
Each of the annular oil chambers 33 formed on the outer periphery of the output shaft 9
The annular oil chamber 33 receives hydraulic pressure between the variable throttle 6 and the fixed throttle 7 in the second oil passage 42 from a hydraulic pressure introduction hole 34 opened on the outer periphery of the housing 80. It has been introduced.

Therefore, the plungers 31, 31... are inserted into the hydraulic insertion hole 34.
The oil pressure introduced into the annular oil chamber 33 is received by the outer end surface of the annular oil chamber 33, and it slides toward the axis of the input shaft 8 along the respective insertion holes 30, 30... is engaged with the recessed portion of the input shaft 8, thereby causing a relative angular displacement of the input shaft 8 with respect to the output shaft 9, that is, the casing 2 of the valve body 20 of the directional control valve 2.
Constrain relative displacement with respect to 1.

The magnitude of the restraint force generated by the restraint device 3 operating in this manner corresponds to the magnitude of the oil pressure in the annular oil chamber 33, and the magnitude of this oil pressure corresponds to the magnitude of the opening degree of the variable throttle 6. The opening degree of the variable diaphragm 6 is automatically adjusted according to an opening/closing signal given to it from a diaphragm control section 60, and the diaphragm control section 60 receives a vehicle speed signal input thereto from a vehicle speed sensor (not shown). Depending on the vehicle speed, the opening/closing signal is generated to increase the opening degree when the vehicle speed is high, and to decrease the opening degree when the vehicle speed is low. Therefore, the restraint force generated by the restraint device 3 increases or decreases depending on the vehicle speed, and when an operating force is applied to the steering wheel while driving at high speed, the large restraint force generated by the restraint device 3 causes the valve body 20 and the casing to 21, the hydraulic cylinder S for steering assistance does not generate any steering assistance force, and the rotation of the input shaft 8 is directly transmitted to the output shaft 9 via the restraint device 3. This is further converted into axial movement of the rank shaft 91 via a pinion 90 formed on the output shaft 9 for steering, and rigidity corresponding to the restraining force is imparted to the steered wheels.

The maximum restraint force that can be generated by the restraint device 3 operating in this manner corresponds to the pressure obtained between the variable throttle 6 and the fixed throttle 7 when the variable throttle 6 is fully opened.
This corresponds to the pressure at the branch point between the oil passage 41 and the second oil passage 42 . Therefore, in order to obtain sufficient restraining force, it is required to increase the pressure at the branch point, which is
Since the resistance on the oil path 42 side is constant, this can be achieved by increasing the discharge amount of the hydraulic pump P. However, if the discharge amount of the hydraulic pump P is increased unconditionally,
There is a possibility that the operating speed of the hydraulic cylinder S, which is operated by pressure oil supplied from the hydraulic pump P, becomes excessive, and a comfortable steering feeling cannot be obtained. During this period, a constant amount of oil is discharged, and as the output shaft 9 rotates via the restraint device 3, the movement direction is changed between the binion 90 and the rack shaft 91, and the movement of the hydraulic cylinder S is controlled. When steering is performed, it is required to discharge a sufficiently large amount of oil.

In the power steering device according to the present invention, the above-mentioned discharge amount characteristics are applied to the hydraulic pump P by the flow rate control valve l disposed on the upstream side of the branch point of the discharge oil passage 4 of the hydraulic pump P. give. As shown in FIG. 1, this flow control valve 1 consists of a bottomed cylindrical casing 1O1 having a valve hole 11 having a circular axial cross-sectional shape formed at its axial center position. The valve body 12 is comprised of a valve body 12 fitted so as to be slidable in a direction, a lid member 13 screwed and fixed to the inner periphery of the opening of the valve hole 11, and the like.

Two annular grooves 11a and llb are formed on the inner peripheral surface of the valve hole 11.
An annular groove 11a located on the opening side communicates with the hydraulic pump P, and an annular groove 11b located on the bottom side communicates with the oil tank T. Further, the lid member 13 is a cylindrical member with a discharge hole 14 having a reduced diameter portion 14a formed in the inner part of the casing 10 at its axial center position, and when screwed and fixed as described above. , as shown in the figure, its inner end surface is substantially aligned with the side wall surface on the opening side of the annular groove 11a. Thus, the pressure oil from the hydraulic pump P flows into the valve hole 11 from the annular groove 11a, passes through the diameter-reduced portion 14a, and flows into the discharge hole 14.
The oil is introduced into the discharge oil passage 4 connected to the discharge hole 14.

Further, the valve body 12 is applied with a pressing force toward the lid member 13 by a compression spring 15 interposed between it and the bottom of the casing 10, and the end of the valve body 12 on the lid member 13 side is is a thin short cylindrical shape having an inner diameter larger than the opening diameter of the discharge hole 14 into the casing 10 and an outer diameter sufficiently larger than the outer diameter of the other part of the valve body 12, and A protruding part 16 formed with notches 16a, 16a, . . . formed in a plurality of locations is continuously provided. Therefore, as shown in Figure 1,
The valve body 12 is pressed against the lid member 13 by the pressing force of the compression spring 15.
Even when the valve body 12 and the lid member 13 are slid toward each other, the tip of the protrusion 16 first contacts the inner end surface of the lid member 13 and restrains the valve body 12 from sliding thereafter.
A space corresponding to the axial length of the protrusion 16 is formed between the two, and after the pressure oil in the annular groove 11a flows into this space, it flows into the plurality of notches 16a, 16a.
. . , and flows out into the discharge oil passage 4 via the discharge hole 14 as described above.

As shown in FIG. 3, which is a partially enlarged view of FIG. 1, the compression spring 15 has a small pitch portion 15a with a small pitch between the coils;
This is an uneven pitch spring having a large pitch portion 15b coaxially with the large pitch portion 15b, and before the gap between the coils in the small pitch portion 15a becomes O due to compression, the small pitch portion 15a
It has a spring constant that is a combination of the spring constant of the large pitch part 15b and the spring constant of the large pitch part 15b, and after the gap becomes 0, it has the spring constant of only the large pitch part 15b. , as shown in Fig. 4, the compression amount is a predetermined amount N (b
After reaching point ), the increase in the pressing force relative to the compression amount becomes larger than before.

Further, a back pressure chamber 17 formed between the valve body 12 and the bottom of the casing 10 is connected to the lid member 1 through a communication oil passage 18.
A ring formed around the outer circumference of 3? The pressure on the downstream side of the reduced diameter portion 14a is introduced into the back pressure chamber 17. Therefore, the pressure on the downstream side of the reduced diameter portion 14a acts on the surface of the valve body 12 on the back pressure chamber 17 side, and the pressure on the upstream side of the reduced diameter portion 14a acts on the surface on the lid member 13 side. Therefore, as is well known, the pressure difference that occurs before and after the reduced diameter portion 14a is the flow rate of the pressure oil passing through the reduced diameter portion 14a, in other words, the pressure difference that occurs before and after the reduced diameter portion 14a. Since the pressure corresponds to the flow rate, the valve body 12 is pressed in the direction toward the front pressure chamber 17 by a pressing force corresponding to the flow rate of the pressure oil introduced into the discharge oil passage 4 .

Now, the flow rate control operation of the flow rate control valve 1 configured as above will be explained next.

As described above, the pressing force of the compression spring 15 is constantly acting on the valve body 12 of the flow rate control valve 1 toward the lid member 13, while pressure oil from the hydraulic pump P is applied to the valve body 12 through the annular groove 11a. When the pressurized oil passes through the diameter-reduced portion 14a, a pressing force due to a pressure difference acts in a direction opposite to the pressing force, that is, in a direction toward the back pressure chamber 17. Also,
As described above, the hydraulic pump P is driven by the engine, and the amount of oil introduced from the hydraulic pump P increases as the engine speed increases, so the pressure difference that occurs before and after the reduced diameter portion 14a. increases as the engine speed increases. Therefore, in response to an increase in the pressing force caused by this pressure difference, the valve body 12 compresses the compression spring 15 while compressing the back pressure chamber 17.
According to this movement, the annular groove 11b communicating with the oil tank T is opened, and a part of the oil introduced from the hydraulic pump P is sent from the discharge hole 14 to the discharge oil path 4. The valve element 12 operates to flow back to the oil tank T through the annular groove 11b, and the flow rate is controlled by the operation of the valve element 12.

FIG. 5 is a graph showing changes in the amount of pressure oil supplied to the discharge oil passage 4 with respect to changes in engine speed.
Before reaching the formation position, the entire amount of the pressure oil introduced into the flow rate control valve l is fed to the discharge oil passage 4, so that the hydraulic pump increases as the engine speed increases while the pressure oil is being fed. It increases in response to an increase in the amount of P discharged.

Furthermore, as the engine speed increases and the amount of movement of the valve body 12 increases, and after the movement position reaches the position where the annular groove 11b is formed, the annular groove 11b opens in accordance with the movement of the valve body I2. However, a part of the pressure oil introduced into the flow control valve 1 is returned to the oil tank T via the annular groove 11b, and the remaining part is sent to the discharge oil passage 4. Now, the amount of oil returned to the oil tank T at this time corresponds to the opening area of the annular groove 11b, and while this opening area corresponds to the amount of movement of the valve body 12, as described above, the spring constant of the compression spring 15 changes before and after the above-mentioned point, and the compression spring 15 causes the valve body 1 to
Since the pressing force applied to the annular groove 11b changes as shown in FIG.
The rate of increase in the opening area of , that is, the rate of increase in the amount of return to the oil tank T is different before and after the amount of compression of the compression spring 15 reaches the above-mentioned point, and the amount of return to the amount of movement of the valve body 12 after that is different. The rate of increase will be smaller than that before. In the compression spring 15, a composite spring constant of the small pitch portion 15a and the large pitch portion 15b is set so that the return amount corresponds to an increase in the amount of pressure oil introduced into the flow control valve 1, In the compression range after the above point where the spring constant is larger than this, the amount returned to the oil tank T does not correspond to the amount of increase in the introduced pressure oil, and the latter becomes larger than the former, resulting in discharge As shown in FIG. 5, the amount of pressure oil fed to the oil passage 4 gradually increases as the engine speed increases. Note that point b in FIG. 5 corresponds to point b in FIG. 4.

Due to such an operation of the flow rate control valve 1, the amount of pressure oil fed to the discharge oil passage 4 exhibits a changing state as shown in FIG.
When the engine speed is high, in other words when the vehicle speed is high, a sufficiently high oil pressure is obtained at the branch point between the first oil passage 41 and the second oil passage 42, as described above, and the restraint device 3 A sufficiently large restraining force can be generated at 6th
The figure is a graph showing the restraining force in the power steering device according to the present invention as a solid line, and the restraining force in a conventional power steering device not equipped with the flow control valve 1 as a broken line, with vehicle speed taken on the horizontal axis. As shown in this figure, the power steering device according to the present invention can obtain a larger restraining force in a high-speed driving range than the conventional power steering device, and can provide sufficient rigidity to the steering wheel. It is possible to have it granted.

In this embodiment, an unequal spring as shown in FIG. 3 was used as the compression spring 15 in order to produce a non-linear displacement with respect to the load. Non-linear spring characteristics can also be achieved by combining springs of class 2[r] or higher with different natural lengths, and as shown in Fig. 7, the winding diameter gradually increases in the axial direction. Similar nonlinear spring characteristics can be achieved with conical coil springs, and these springs can also be used as the compression spring I5.

〔effect〕

As detailed above, in the power steering device according to the present invention, the flow rate control valve provided on the discharge side of the hydraulic pump has a pressure difference corresponding to the discharge amount of the hydraulic pump, and a spring having nonlinear spring characteristics. The valve body is equipped with a valve body that moves according to the urging force from the valve body, and discharges excess oil according to the movement of the valve body, so that the amount of oil introduced from the hydraulic pump to the flow control valve reaches a predetermined amount. , until the amount of movement of the valve body reaches a predetermined amount, approximately -
A fixed amount of oil is supplied to the restraint device and the directional control valve, the restraint device only generates a small restraint force, and the steering is controlled by the operation of the hydraulic cylinder by the pressure oil supplied through the directional control valve. This reduces the steering wheel operation force when running at low speeds and when stopping, and if the amount of movement of the valve body further increases in response to an increase in the discharge volume of the hydraulic pump due to an increase in engine speed, the restraint device The amount of pressure oil supplied to the directional control valves increases as the engine speed increases, or in other words, as the vehicle speed increases, and high hydraulic pressure becomes available in the restraint system, so that sufficient pressure is supplied to the steering wheels during high-speed driving. Rigidity can be imparted, and straight-line stability can be improved. Further, the hydraulic pump only needs to be equipped with one unit having a capacity to drive the hydraulic cylinder at a moderately constant speed. Furthermore, the present invention has excellent effects such as avoiding an increase in fuel consumption of the engine.

[Brief explanation of the drawing]

The drawings show one embodiment of the present invention, and FIG. 1 is a schematic diagram showing the configuration of a power steering device according to the present invention, and FIG. 2 is a cross-sectional view of the restraint device taken along line nn in FIG. 1. 3 is a partially enlarged view of the flow control valve, FIG. 4 is a graph showing the relationship between the amount of compression of the compression spring in the flow control valve and the pressing force exerted on the valve body by the compression spring, and FIG. 5 6 is a graph showing the relationship between the engine speed and the amount of pressure oil fed from the flow control valve, and FIG. 6 shows the restraint force obtained in the restraint device between the power steering device according to the present invention and the conventional power steering device. FIG. 7 is a sectional view showing an example of another spring that can be used as a compression spring. l...Flow control valve 2...Direction control valve 3...
・Restriction device 4...Discharge oil path 6...Variable throttle -7...Fixed throttle 8...Input shaft 9...
・Output shaft 11a, llb... annular groove 12...
Valve body 14...Discharge hole 14a...Reduced diameter part 15
...Compression spring 15a...Small pitch portion 15b
... Large pitch part 17 ... Back pressure chamber 18 ... Communication oil passage 31 ... Plunger 33 ... Annular oil chamber
41...First oil passage 42...Second oil passage 60
... Throttle control section P... Hydraulic pump S...
Hydraulic cylinder patent Applicant Koyo Seiko Co., Ltd. Agent Patent attorney Noboru Kono Figure 2 Figure 7 Compression amount Figure 4 Figure 5 Figure 6

Claims (1)

[Scope of Claims] 1. A direction control valve that switches the feeding direction of pressure oil from a hydraulic pump driven by an engine to a hydraulic cylinder for steering assistance in accordance with the direction in which the steering wheel is operated; In the power steering device, the power steering device includes a restraint device that restrains the switching operation of the directional control valve with a force corresponding to the pressure of the pressure oil that is supplied from the pump after being depressurized according to the steering situation. A flow control valve is provided on the discharge side, and the flow control valve applies a pressure difference corresponding to the discharge amount of the hydraulic pump in a direction opposite to the acting direction of the pressure difference, and causes a nonlinear displacement with respect to the load. 1. A power steering device comprising: a valve body that moves in response to the biasing force of a spring that generates an excess oil discharge hole whose opening area changes in accordance with the movement of the valve body.