JPH0124389Y2 - - Google Patents

Description

[Detailed Description of the Invention] [Industrial Application Field] The present invention relates to a fluid control device used in combination with a variable displacement hydraulic pump.

It is well known that piston pumps and vane pumps are variable displacement hydraulic pumps widely used in industrial machinery. The former controls the pump displacement per rotation of the rotor by the tilt angle of the swash plate, and the latter controls by the amount of eccentricity of the cam ring. Therefore, by variably controlling the tilt angle of the swash plate or the eccentricity of the cam ring in accordance with the pump discharge pressure, it is possible to flatten the change in the mechanical shaft input of the pump as much as possible with respect to the change in the discharge pressure. Approximate constant horsepower control can be performed.

The fluid control device of this invention is for performing approximate constant horsepower control of a variable displacement hydraulic pump such as the vane pump or piston pump.

[Prior Art] To explain the case of a piston pump as an example of a variable displacement hydraulic pump, generally speaking, in a variable displacement hydraulic pump, the pump is centered on an axis and parallel to the axis.
and a cylinder block having a plurality of pistons arranged in a circle, a drive shaft for driving the cylinder block to revolve around the central axis, and a swash plate (control plate) disposed at an angle with respect to the drive shaft. The swash plate regulates the position of the piston during its rotation so that each rotation of the cylinder block causes the piston to make one reciprocation, and further an inclination of the swash plate with respect to the axis. The stroke of the piston in one reciprocation can be changed by changing the angle, and the discharge amount of the pump is variably controlled by controlling the tilting angle of the swash plate.

As a control method for such variable displacement piston pumps, pressure compensator control has been developed in which the discharge volume suddenly drops to zero above a certain set pressure, but it requires a low pressure large flow rate and a high pressure small flow rate. In applications such as press machines, the pump shaft input changes significantly as the discharge pressure changes, and therefore, in this control system, it is necessary to use a pump drive motor with a large output capable of handling the maximum shaft input. Therefore, in order to make effective use of the motor's output, a pump control system that requires little change in shaft input in response to pressure changes is required. Approximate constant horsepower control is required.

One of the conventional examples of this approximate constant horsepower control system is called simple constant horsepower control, which is performed by a control device as shown in FIG.

FIG. 1 is a sectional view showing a main part of an example of a control device used for conventionally known simple constant horsepower control.

1 is a swash plate as a discharge amount control element of the piston pump described above, and this swash plate 1 is connected to a joint 3.
The pump is tilted about the fulcrum 2 by being driven by the operating piston 4 connected to the pump, and the pump discharge amount is determined by the tilting angle. The operating piston 4 is slidably housed in a cylinder 5, and a hollow first adjusting screw 6 is attached to the cylinder so that the screwing amount can be adjusted. A second adjusting screw 7 is also attached so that the screwing amount can be adjusted.
Two pressing springs 9 and 10 are interposed in series between the second adjusting screw 7 and the end surface of the operating piston 4 with an intermediate movable spring receiving member 8 interposed therebetween. The cylinder 5 forms a pressure chamber around the shoulder portion 12 of the piston 4 that is separated from the chamber in which the springs 9 and 10 are arranged, and a port 11 in the pressure chamber.
The pump discharge pressure is guided through the pump.

Now, when the pump discharge pressure increases, the operation piston 4 is moved against the pressure springs 9 and 10 due to the increase in pump discharge pressure acting on the shoulder 12, and this movement causes the operation piston 4 to move through the joint 3 to the swash plate. 1 is rotated clockwise in the figure, and this rotation reduces the tilt angle of the swash plate 1, shortens the reciprocating stroke of the plurality of pistons, and reduces the pump discharge amount. Moreover, when the pump discharge pressure decreases, the pump discharge amount increases, contrary to the above, and as a result, changes in the shaft input of the pump are flattened.

By the way, in the one in Fig. 1, the operating piston 4
Since it is for directly driving the swash plate 1, there is a premise that it cannot be made too thin from the viewpoint of strength. Therefore, the pressure receiving area of the operating piston 4 tends to be large, but since the discharge pressure is applied directly to the operating piston 4, if the pressure receiving area of the operating piston 4 is large, the counter spring 9,
Approximate constant horsepower control in the high pressure region will not be possible unless the engine 10 is made huge. For this reason, the operating piston 4 is provided with a pressure receiving part with a small area by a shoulder part 12, but this shoulder part 12 is usually machined to have an extremely small radial dimension, and therefore the machining accuracy is limited. If this is not sufficient, the coaxiality and parallelism between the large diameter part and the small diameter part of the operating piston 4 will be slightly distorted, and the machining accuracy of the sliding holes for these parts on the cylinder 5 side will also affect the sliding of the operating piston 4. Dynamic resistance tends to vary from product to product. Such sliding resistance of the operating piston 4 directly affects the approximate constant horsepower control operation during operation of the pump, and appears as a so-called hysteresis phenomenon in the pressure-flow characteristics. Therefore, with this method, different hysteresis phenomena occur depending on the product, and adjustment for each product is extremely difficult.

In addition, in the one shown in FIG. 1, the differential plunger is formed by the shoulder part 12 in order to reduce the area of the pressure receiving part of the operating piston 4, but even so, if the pump discharge pressure is high, the thrust of the operating piston 4 will inevitably increase. Therefore, the counter springs 9 and 10 need to have a large force. For this reason, a large space is required to accommodate the spring, which also has the drawback that the pump as a whole becomes large, especially in high-pressure applications.

In addition to using a differential plunger in which the pressure-receiving area of the operating piston 4 is constituted by a shoulder part, a piston having a small diameter is separately provided (Japanese Patent Laid-Open No. 49
-41902) is also known.

In this way, in the method of directly controlling the tilting angle of the swash plate using the pump discharge pressure using a differential plunger-type operation piston with a shoulder, the spring and its housing can be made as small as possible, and the pressure-flow rate can be reduced. If a characteristic without hysteresis is to be obtained, the operating piston and the sliding portion of its accommodation portion require highly precise machining, which is disadvantageous in that it is expensive.

In addition, in a piston pump, a rotational force (disturbance) that tries to change the inclination angle of the swash plate 1 due to the resultant force of the thrust of the piston (rotating part) due to the discharge pressure is always acting on the swash plate 1, so the operating piston 4
Since the thrust of This also applies to vane pumps.

Therefore, in order to obtain approximately constant horsepower characteristics, the conventional control device shown in Fig. 1, which directly opposes the spring force and discharge pressure and controls the swash plate in proportion to the increase in discharge pressure, To some extent, the spring becomes larger and the pump as a whole becomes larger, and the spring must be replaced to change the pressure-flow characteristics, and the pressure-flow characteristics also change as the rotation speed changes. The disadvantage that hysteresis tends to occur is unavoidable.

As another fluid control device used for simple constant horsepower control, in addition to the above, changes in the inclination angle of the swash plate are controlled by a machine (position ) is also known, but this type has a more complicated structure.

Furthermore, as disclosed in Japanese Patent Application Laid-Open No. 55-114894, which is used for approximate constant horsepower control, a flow control valve is provided in the main flow path from the pump to the load, and the opening degree is changed according to the discharge pressure. There is also a known system in which a load-sensitive control valve that is controlled by pressure is provided, and the pump discharge amount is controlled by this load-sensitive control valve so that the differential pressure before and after the flow rate control valve is constant. In order to provide the so-called full cut-off characteristic that limits the pump discharge pressure from exceeding a certain value when it increases, a separate valve mechanism is required, and the addition of this valve mechanism tends to complicate the device. In addition, since variable throttle control is performed on the main channel according to the discharge pressure, the discharge pressure is directly applied to one end of the spool for throttle control, and the valve spring that counters this has to be large. At the same time, the device is complicated because the discharge pressure is opposed to the spring via a small piston.Furthermore, the pump discharge flow rate is controlled so that it is almost inversely proportional to the pump discharge pressure, so as shown in Figure 2, constant horsepower is achieved. An unavoidable disadvantage is that the pressure range in which approximate constant horsepower control can be performed along the assumed characteristic curve a becomes narrow.

[Problem to be solved by the invention] The purpose of this invention is to develop a pilot control method that does not use a large piston or a counter spring, and therefore can take advantage of the fact that hysteresis is inherently less likely to occur in the pressure-flow characteristics. By combining valves with a simple structure, we are able to achieve approximate constant horsepower control over a wide pressure range, in which the pump discharge rate gradually decreases as the pressure increases, and the mechanism is made more compact, resulting in improved pressure-flow characteristics. It is easy to change, and has excellent stability of pressure-flow characteristics against changes in rotational speed.Furthermore, it has a so-called full cut-off characteristic that reduces the discharge amount to almost zero when the pump discharge pressure reaches a certain set pressure value. It is an object of the present invention to provide a fluid control device that can be easily adapted.

[Means for Achieving the Object] In order to achieve this object, the fluid control device of the present invention controls the pump by selectively guiding the discharge pressure of the variable displacement hydraulic pump to the discharge amount control element of the pump. In a fluid control device that controls the pressure vs. flow rate characteristic so as to approach a constant horsepower characteristic in which the discharge amount gradually decreases as the pressure increases, a differential pressure-sensitive throttle valve that is connected in parallel to the orifice and widens its opening in response to an increase in the differential pressure across the orifice; and one end face receives hydraulic pressure on the orifice inlet side of the passage. At the same time, the hydraulic pressure on the orifice outlet side is received by the other end face having a smaller pressure receiving area than the one end face, together with the spring force of an adjustable spring, and due to the difference in the pressure receiving area between the two end faces, the orifice exit side Pressure oil is supplied to the discharge amount control element of the pump so that the pump discharge amount changes such that the differential pressure across the orifice decreases (or increases) as the pressure oil pressure increases (or decreases). A compensator valve for controlling introduction (or extraction).

[Function] In the present invention, the compensator valve adjusts the hydraulic pressure (additional pressure) on the orifice outlet side to P,
Assuming that the spring force of the spring is F, the differential pressure before and after the orifice is ΔP, the pressure-receiving area of the one end face is A ₁ , and the smaller pressure-receiving area of the other end face is A ₂ , the following equation is obtained.

(P+ΔP)A ₁ =P・A ₂ +F ……(1) If we rearrange this equation (1) using ΔP, ΔP=F/A ₁ −(1−A ₂ /A ₁ )P ……(2) Here Since A ₁ >A ₂ >0, 1>(1-A ₂ /A ₁ )>0 holds true for the coefficient of the second item on the right side of equation (2). That is, since A ₂ /A ₁ does not become 1 in equation (2), the differential pressure ΔP across the orifice decreases as the load pressure P increases, and increases as the load pressure decreases.

On the other hand, the discharge amount Q of the pump is equal to the total flow rate passing through the orifice when the differential pressure sensitive throttle valve is closed, and this is calculated from the Bernoulli's theorem and the law of continuity between the differential pressure ΔP and the orifice opening area A.
It is uniquely determined by, and is as follows.

Q=CA (2g・ΔP/γ) ^1/2 ...(3) Here, γ is the specific weight of the hydraulic oil, C is the flow coefficient, g
is the gravitational acceleration.

From equations (2) and (3) above, it can be seen that, for example, as the pump discharge pressure increases, the discharge amount gradually decreases. The opening area A is substantially changed in accordance with the differential pressure ΔP by the differential pressure sensitive throttle valve that is parallel to the orifice, thereby bringing the pressure-flow characteristic closer to the ideal constant horsepower characteristic.

Further, the pressure oil acting on the other end face of the compensator valve on the side receiving the spring force is a load pressure P, and therefore a relief valve that opens when this load pressure P rises to a certain upper limit is provided. In addition, if a fixed throttle is provided on the compensator valve that creates a pressure difference between the front and rear sides due to the flow to the relief valve, the flow when the relief valve opens will displace the compensator valve against the spring. By utilizing this displacement, the pump discharge pressure can be directly applied to the discharge variable element to bring the pump into a full cut-off state. Such additional configurations can be easily implemented using conventional valve design techniques.

[Example] If this invention is explained in detail along with the example drawings, the third
In the figure, variable displacement hydraulic pump 2 to be controlled
1 decreases the discharge amount as the hydraulic pressure applied to the discharge amount control element 27 increases, and a passage 22 that sends the discharge pressure oil of the pump 21 to the load actuator has a pre-selected diameter. An orifice 26 is provided, and a differential pressure sensitive throttle valve 28 is provided in parallel with the orifice 26, which opens depending on the differential pressure before and after the orifice 26. Reference numeral 34 designates a compensator valve which receives the differential pressure before and after the parallel body of the orifice 26 and the throttle valve 28 as a pilot pressure, and guides the discharge pressure from the passage 22 to the discharge amount control element 27 of the pump 21 in accordance with this differential pressure. A specific example of the differential pressure sensitive throttle valve 28 and the compensator valve 34 is shown in FIG. 4, and in the example shown in FIG. Although a spool 23 and a valve spring 24 are incorporated, they may be constructed separately as shown in FIG. 7, which will be described later.

In the embodiment shown in FIG. 4, the compensator valve 34 includes a spool 38 which is slidable within a valve hole 37 formed in a valve body 41 and whose movement is limited in one direction by a restriction member 42, and a seal retainer 4.
a valve spring 36 which is adjustable by an adjustment screw 43 through 0 and biases the spool 38 against the limiting member 42;
It is made up of. The spool 38 has a structure in which it is divided into a small-diameter spring receiving guide part 45 on the right side in the figure and a hollow chamber 46 on the left side by a partition wall part provided through the fixed throttle 35 in the center. 4
The valve spring 36 is located within the valve spring 5 to form a spring chamber 47 which is adapted to drop from the pilot port 31 through the pilot relief valve 33 and from the tank port 32 into the tank.
Further, a throttle valve spool 23 is disposed within the hollow chamber 46 and is pressed by a spring 24 to a limiting member 48 fixed near the left end of the spool 38.
Passage 25 and 44 between opening 49 of spool 38
A throttle control section 20 is formed to control the opening degree of a bypass passage that communicates with the orifice 26, thereby forming a differential pressure sensitive throttle valve 28 that is sensitive to the differential pressure across the orifice 26. In this way, the spool 38 receives the pressure on the inlet side of the orifice 26 via the passage 25 on the end face on the restricting member 42 side in the pressure receiving area _A1 , and also receives the pressure on the outlet side of the orifice 26 from the passage 44 through the passage in the spool. 39 and a fixed throttle 35 provided in the spool, there is a pressure receiving area on the end face on the valve spring 36 side.
A ₂ is received, and in this case A ₁ > A ₂ . The spool 38 thus operates by applying the orifice outlet pressure to the other end of the pressure receiving area A ₂ and the spring force of the spring 36 against the orifice inlet pressure to one end of the pressure receiving area A ₁ . When the differential pressure of
The pressure is directed to a control element 27 of.

In this embodiment, the spring chamber 47 in which the valve spring 36 is disposed communicates with the hollow chamber 46 and the opening 39 to the passage 44, that is, the passage 22, to the orifice outlet side via the fixed throttle 35 of the partition wall inside the spool. A pilot relief valve 33 is further connected to this spring chamber 47 via a pilot port 31. This pilot relief valve 33
The poppet 50 is seated on the valve seat 55 at the end of the port 31 from the tank port 32 leading to the tank by the poppet spring 51, and the pressing force of the poppet 50 by the spring 51 is applied to the spring 51 through the seal retainer 52. The relief pressure can be set to a desired value by adjusting it with an adjustment screw 53 that is brought into contact with the pressure.

In a fluid control device having such a configuration, the orifice 2
6, and a pressure difference corresponding to the flow rate is generated before and after the orifice 26. The pressure oil on the inlet side of the orifice 26 passes through the passage 25 and acts on the end face of the spool 38, which has a large pressure receiving area A1, including the end face of the spool ₂₃ , while the pressure oil on the outlet side of the orifice 26 passes through the passage 44. , 39 and the hollow chamber 46 inside the spool 38 and the fixed throttle 3 provided therein.
5 to the spool 38 together with the spring force of the spring 36.
Acts on the end face of the smaller pressure-receiving area _A2 .

First, for convenience of explanation, the explanation will proceed assuming that the spool 23 of the differential pressure sensitive throttle valve 28 closes the throttle control section 20 thereof.

When the load pressure P increases, since the pressure receiving areas of the spool 38 are different, the spool 38 moves in the direction of opening the control section 54 (to the right in the figure) while deflecting the spring 36, and thereby As a result, the load pressure from the passage 44 is introduced into the port 29, and this load pressure acts on the control element of the pump 21, causing the pump 21 to reduce its discharge volume, thereby reducing the differential pressure across the orifice 26, thereby reducing the compensator. Equilibrate valve 34.

Conversely, when the load pressure P decreases, the spool 38 moves in a direction to close the control section 54 (to the left in the figure; FIG. 2 shows the spool at the leftward stroke end), This allows port 2
9 communicates from port 30 to the tank, pump 21 increases its displacement, and as a result, the differential pressure across orifice 26 increases to balance compensator valve 34.

The balanced equation of the compensator valve 34 is: P is the load pressure, F is the spring force of the spring 36, ΔP is the differential pressure before and after the orifice 26, A ₁ is the pressure-receiving area of the spool that receives the pressure on the inlet side of the orifice 26, and is the orifice 2
The pressure receiving area of the same spool that receives the pressure on the outlet side of 6 is
If A ₂ , then equation (1) is obtained.

(P+ΔP)A ₁ =P・A ₂ +F (1) When formula (1) is rearranged using the differential pressure (ΔP) of the orifice, formula (2) is obtained.

ΔP=F/A ₁ -(1-A ₂ /A ₁ )P ...(2) Here, since A ₁ >A ₂ >0, 1>(1-A ₂ /A ₁ )>0, As is clear from equation (2), the differential pressure ΔP decreases as the load pressure P increases.

This is illustrated in FIG. 5. That is, both pressure receiving areas A ₁ and A ₂ of the spool 38 of the compensator valve 34 are the areas before and after the throttle 26, as is clear from the equation (2) above, since A ₁ =A ₂ in a general pressure compensator valve. The differential pressure ΔP maintains a constant value determined by the spring force F and one pressure-receiving area A ₁ regardless of changes in the load pressure P. However, in this invention, the load pressure is reduced by setting A ₁ > A ₂ as described above. As P decreases, the differential pressure ΔP increases as shown in FIG.

By the way, the discharge amount of the pump 21 is controlled by the throttle control section 2.
If 0 is closed, it is the total flow rate through the orifice 26, and the total flow rate Q through the orifice 26 is
is the differential pressure ΔP before and after the orifice 26 as shown in equation (3).
Uniquely determined by

Q = CA (2g・ΔP/γ) ^1/2 ...(3) (γ is the specific weight of the hydraulic oil, C is the flow rate coefficient, g is the gravitational acceleration, and A is the opening area of the orifice 26) (2) and It can be seen from equation (3) that a pressure-flow rate characteristic as shown by curve a in FIG. 6 is obtained, in which the discharge amount decreases as the pump pressure increases. 6th
Compared to the ideal constant horsepower characteristic curve b, the curve a shown in the figure still has an opposite tendency, and the width of the pressure range showing the constant horsepower characteristic is narrow. This is a result of considering the differential pressure sensitive throttle valve 28 as fully closed, and is because the opening area A of the orifice 26 is a constant value in equation (3).

Here, in this invention, a differential pressure-sensitive throttle valve 28 is provided in parallel with the orifice 26, and its opening degree is changed according to the differential pressure ΔP, thereby converting the constant A in equation (3) above into a function of ΔP. , the flow rate corresponding to the increase in the differential pressure ΔP due to the decrease in the load pressure P is added to the flow rate passing through the orifice 26, thereby bringing the overall pressure-flow rate characteristic closer to the ideal constant horsepower characteristic.

That is, as shown in FIG. 5, as the load pressure P decreases, the differential pressure ΔP across the orifice increases according to equation (2), so the spring force of the spring 24 of the differential pressure sensitive throttle valve 28 should be adjusted appropriately. If the throttle control section 20 is set to open at an opening degree corresponding to an increase in the differential pressure ΔP (a decrease in the load pressure P), the flow rate passing through the throttle control section 20 will be as described in (2) above. According to the characteristic curve determined by the equation (characteristics shown in FIG. 5), the spring characteristics of the spring 24, the opening degree change characteristics of the throttle control section 20, etc., as the load pressure P changes, as shown by curve c in FIG. Changes to The flow rate of this curve c increases as the differential pressure ΔP increases, that is, the load pressure P decreases, and this flow is added to the flow rate passing through the orifice 26 shown in the curve A and flows to the load. Therefore, the overall pressure-flow characteristic becomes like curve d, which is the sum of curves a and c, and a characteristic closer to the ideal constant horsepower characteristic b can be obtained over a wide range of pressure. , a corresponding flattening of the pump shaft input can be achieved.

In addition, when the pilot relief valve 33 is connected to the spring chamber 47 of the compensator valve 34, when the load pressure P increases and exceeds the set relief pressure, the poppet 50 of the relief valve 33 is released from the spring 51.
If the spool 38 is opened against the spring 36, a flow is generated there, and a pressure difference is generated before and after the fixed throttle 35, and the pressure rise in the spring chamber 47 is stopped, so the spool 38 opens against the spring 36. As a result, the pressure oil in the passage 44 is guided from the port 29 to the control element 27 of the variable displacement hydraulic pump 21, and the pump discharge rate decreases rapidly, as shown in FIG. Full cut-off can be achieved at a set relief pressure as shown by curve e, and of course this full cut-off pressure can be arbitrarily adjusted using the adjustment screw 53.

FIG. 7 shows an example in which a differential pressure sensitive throttle valve is separated from the spool of the compensator valve 34 so that its spring can be independently adjusted from the outside using an adjusting screw 56. The same reference numerals are shown with dashes attached. By making it possible to adjust the spring force of the differential pressure sensitive throttle valve separately in this way, the pressure-flow characteristics of the pump change from curves f to g to h as the spring force is weakened, as shown in Figure 8.
It can be adjusted to stand upright.

[Effects of the invention] As described above, according to this invention, approximate constant horsepower control of a variable displacement hydraulic pump can be performed over a wide pressure range by simply providing a differential pressure sensitive throttle valve and a compensator valve. Moreover, since this compensator valve controls the pilot pressure of the pump's discharge amount control element, it can be constructed with a small diameter spool, so it can be constructed extremely compactly without increasing the size of the mechanism. It is easy to design and process. In addition, since it controls pilot pressure, it has the advantage that hysteresis is less likely to occur in the pressure-flow characteristics compared to systems that directly control the operating piston using the discharge pressure, and the characteristics can be changed. This is easy because it is only necessary to change the opening area of the orifice 26, and this predetermined characteristic is stable without being affected by changes in the rotational speed. Furthermore, the practical benefits of this invention are extremely large, as it can be easily adapted to the installation of a pilot relief valve and the addition of full cut-off characteristics is also simple.

[Brief explanation of the drawing]

Fig. 1 is a sectional view showing the main parts of a conventional fluid control device, Fig. 2 is a characteristic diagram obtained by the conventional approximate constant horsepower control method, and Fig. 3 is a hydraulic circuit diagram showing an embodiment of this invention. , FIG. 4 is a schematic structural diagram of an embodiment of this invention, FIG. 5 is a diagram showing the relationship between the load pressure P and the differential pressure ΔP before and after the orifice, and FIG. 6 is a diagram showing the relationship between the pressure -
FIG. 7 is a schematic structural diagram of another embodiment of this invention, and FIG. 8 is a pressure-flow characteristic diagram when the setting is variable. Explanation of symbols of main parts, 22... Variable displacement hydraulic pump, 27... Discharge rate control element, 26... Orifice, 28... Differential pressure sensitive throttle valve, 34...
compensator.

Claims (1)

[Claims for Utility Model Registration] By selectively guiding the discharge pressure of a variable displacement hydraulic pump to a discharge rate control element of the pump, the pressure vs. flow rate characteristics of the pump can be adjusted so that the discharge rate gradually increases as the pressure increases. A fluid control device that controls the fluid so as to approach a constant horsepower characteristic in which the pressure decreases to A differential pressure sensitive throttle valve that widens the opening degree in response to a rise in pressure, and a spring that receives hydraulic pressure on the orifice inlet side of the passageway at one end face and can adjust the hydraulic pressure on the orifice outlet side. The spring force is received by the other end surface having a smaller pressure-receiving area than the one end surface, and due to the difference in pressure-receiving area between the two end surfaces, the orifice a compensator valve that controls the introduction (or extraction) of pressure oil into the discharge amount control element of the pump so as to obtain a change in pump discharge amount that reduces (or increases) the differential pressure before and after the pump; A fluid control device characterized by: