JPH0128314Y2 - - Google Patents

Description

[Detailed Description of the Invention] [Industrial Application Field] This invention relates to a control device for 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 amount of eccentricity of the cam ring according to the pump discharge pressure, changes in the mechanical shaft input of the pump can be made as flat as possible with respect to changes in the discharge pressure. So-called approximate constant horsepower control can be performed.

The 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 a variable displacement hydraulic pump as an example, a variable displacement piston pump generally has a plurality of pistons arranged in a circle parallel to a central axis. a cylinder block having a cylinder block, a drive shaft that drives the cylinder block to revolve around the central axis, and a swash plate (discharge rate control element) disposed so as to be tiltable with respect to the central axis.
The swash plate regulates the position of the piston during its rotation, so that each rotation of the cylinder block causes the piston to reciprocate once, and furthermore, the inclination angle of the swash plate with respect to the central axis is adjusted. The amount of stroke of the piston in one reciprocation is made variable, and the discharge amount of the pump is variably controlled by controlling the tilting angle of the swash plate.

The control method for such a variable displacement piston pump is to operate the pump at a swash plate tilt angle corresponding to the set discharge flow rate, and when the discharge pressure reaches a certain set pressure, a portion of the discharge pressure oil is Pressure compensator control is known in which the discharge amount is suddenly brought to zero by tilting the swash plate. At the same time, the pump shaft input changes significantly, and therefore, in this control system, it is necessary to use a pump drive motor with a large output to accommodate 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. Constant horsepower control is required.

This constant horsepower control theoretically draws a hyperbolic curve in the flow rate vs. pressure characteristic curve of the pump, but in reality it is simply approximated by a straight line within a certain range of use, and the pressure rise is The inverse proportional slope characteristic in which the flow rate gradually decreases as the engine speed increases is called simple constant horsepower control, and the problem was how to optimize this slope.

Conventionally, as representative systems for performing this simple constant horsepower control, the system shown in FIG. 1 and the system shown in FIG. 2 have been known.

First, the one shown in FIG. 1 is a type in which the pump discharge pressure is directly opposed to the spring force acting on the operating piston 4 of the swash plate 1, which is the discharge amount variable element of the variable displacement piston pump. The figure shows only the main parts around the operating piston. Swash plate 1
is capable of tilting about a fulcrum 2, and this tilting is performed by driving an operating piston 4 connected by a joint 3, 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 adjustment screw 6 is attached to the cylinder so that the screwing amount can be adjusted. A second adjustment screw 7 is 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 face of the operating piston 4 with an intermediate movable spring receiving member 8 interposed therebetween. The cylinder 5 includes:
As a pressure chamber for the operating piston 4, a cylinder chamber 13 is formed around the shoulder portion 12 of the piston 4, which is separated from the chamber in which the springs 9 and 10 are arranged, and the pump discharge pressure is applied to the cylinder chamber 13. It is designed to be introduced via port 11.

Now, when the discharge pressure of this pump increases, the operating piston 4 is moved in the cylinder chamber 13 in a direction in which it is compressed by the pressure springs 9 and 10 due to the increase in the pump discharge pressure acting on its shoulder portion 12. Due to the movement, the swash plate 1 is rotated clockwise in the figure via the joint 3, and this rotation reduces the tilt angle of the swash plate 1 with respect to the central axis, thereby shortening the reciprocating stroke of the plurality of pistons. , the pump discharge amount decreases. Conversely, 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 directly drives the swash plate 1, the force required to drive the operating piston 4 not only overcomes the frictional resistance of the tilting of the swash plate itself but also opposes the discharge pressure, so it depends on the rating of the pump. is a relatively large force, and the operating piston itself cannot be made very thin in terms 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, the operating piston 4
If the pressure receiving area of is too large, the counter spring 9,1
Approximate constant horsepower control in the high pressure region will not be possible unless 0 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. Therefore, in this one,
As a result, different hysteresis phenomena appear for each product, making adjustment for each individual product extremely difficult.

In addition, in the one shown in FIG. 1, although the differential plunger is constructed by the shoulder section 12 in order to reduce the pressure receiving area of the operating piston 4, the thrust of the operating piston 4 is still inevitable in a pump with a high discharge pressure. Since the springs 9 and 10 are large in size, large springs with a large force are required as the springs 9 and 10. 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, especially in the case of a piston pump, a rotational force (disturbance) that attempts to change the tilting angle of the swash plate 1 is constantly acting on the swash plate 1 due to the last resultant force of the piston part (rotating part) due to the discharge pressure. Therefore, the thrust force of the operating piston 4 needs to be large enough to overcome this, so the pressure-receiving area of the operating piston 4 cannot be made extremely small. Since 9 and 10 must be used, there is a limit to the miniaturization of the spring. If the pressing force of the springs 9 and 10 is insufficient, the operating piston will be moved immediately due to the increase in discharge pressure, and approximate constant horsepower characteristics will not be obtained, but characteristics that are not much different from those of the pressure compensator control described above will be obtained. . This also applies to vane pumps.

In addition to using the differential plunger type operating piston described above to suppress the increase in the size of the spring, there is also known a method in which a small diameter piston or push rod is provided separately from the operating piston (Japanese Patent Laid-Open No. 49-41902). , but this also contains similar problems.

Therefore, in order to obtain approximately constant horsepower characteristics, the discharge pressure is made to directly oppose the spring force applied to the operating piston, and the swash plate is controlled by the operating piston in accordance with the increase in the discharge pressure, as shown in Fig. 1. In conventional control devices such as this, the spring is large to some extent, and the pump as a whole is also large, and when changing the pressure vs. flow rate characteristics, the spring must be replaced with another one, and the pressure vs. The unavoidable drawbacks are that the flow rate characteristics tend to change depending on the pump rotational speed, and hysteresis tends to occur.

On the other hand, the system shown in FIG. 2 does not have the drawbacks of the above-mentioned systems, but it does have other problems.

In other words, the method shown in FIG.
This method, which is known in Japanese Patent Application No. 55-114894, is provided with a flow control valve 16 that changes its opening depending on the discharge pressure in the main flow path 15 from the pump 14 to the load, and is controlled by the differential pressure before and after the flow control valve 16. A load-sensitive compensator valve 17 is provided in the pilot line to the variable discharge amount control element 18 of the pump, so that the differential pressure across the flow control valve 16 is maintained constant by the load-sensitive compensator valve 17. The hydraulic pressure to the variable discharge amount control element 18 of the pump 14 is adjusted and cut-off control is performed, and the flow rate control valve 16 in the main flow path 15 controls the discharge flow rate so that it is almost inversely proportional to the discharge pressure. .

In the method shown in FIG. 2, a flow control valve 16 is provided in the main flow path 15 and approximate constant horsepower characteristics are obtained through variable throttling control according to the load pressure. 16 has a large diameter and has a large spool in order to control a large flow rate, which inevitably increases the size of the device. Furthermore, in this system, control is performed so that the pump discharge flow rate is almost linearly inversely proportional to the pump discharge pressure, so the disadvantage that the pressure range in which approximately constant horsepower control can be performed is narrowed is unavoidable.

The purpose of this invention is to eliminate the need for a large flow rate control valve without using a large piston or counter spring, and without sacrificing the advantage of the pilot control method that hysteresis phenomenon is essentially less likely to occur in the pressure vs. flow rate characteristics. It is an object of the present invention to provide an approximate constant horsepower control device using a compensator valve, which can be easily adapted to the addition of a full cut-off characteristic within the same valve structure by combining a valve with a simple structure and a spring mechanism.

[Means for Solving the Problems] A control device for a variable displacement hydraulic pump according to this invention controls the pressure vs. flow rate characteristics of the pump by selectively guiding pump discharge pressure to a discharge amount control element of the pump. This control is performed so that the discharge volume approaches constant horsepower characteristics in which the discharge volume gradually decreases as the pressure increases.In particular, in order to achieve the above-mentioned objective, the pressure oil discharged from the pump is controlled in the passageway through which the discharge pressure oil flows toward the load. a fixed orifice disposed in the passage; one end face receiving hydraulic pressure on the inlet side of the fixed orifice of the passageway, and the other end face receiving hydraulic pressure on the outlet side of the fixed orifice together with the spring force of a valve spring; Has a spool made of
a compensator valve that controls the introduction and discharge of discharge pressure oil to the control element so that the differential pressure across the fixed orifice follows the value determined by the spring force by the spool; a control spring mechanism that presses the valve spring toward the spool and changes the amount of deflection of the valve spring in accordance with load pressure.

[Operation] In the control device of this invention, a fixed orifice is disposed in the passage through which the discharge pressure oil of the pump flows toward the load, so a pressure difference is generated before and after the orifice in accordance with the discharge flow rate. The spool of the compensator valve receives the spring force of the valve spring on one end face, and also receives the differential pressure before and after the orifice between both end faces, and in balance with the spring force of the valve spring, the differential pressure before and after the fixed orifice increases with the spring force. The introduction and derivation of the discharge pressure oil to the control element is controlled so as to follow the value determined by . The valve spring of the compensator valve is pressed toward the spool by a control spring mechanism via a piston, and the control spring mechanism changes the amount of deflection of the valve spring in accordance with the load pressure. For example, when load pressure increases, a control spring mechanism displaces the piston away from the spool in response, thereby reducing the amount of deflection of the valve spring. This causes the spool of the compensator valve to increase the introduction of pump discharge pressure to the control element, which in turn relatively reduces the pump discharge flow rate. When the load pressure decreases, the discharge flow rate increases by the opposite operation.

In the device of this invention, the piston is displaced according to the load pressure, so when the load pressure reaches a certain upper limit, the piston opens a passage to release the load pressure to the low pressure line. It can be done. This causes a flow of pressure oil in the load pressure introduction passage to the piston. Therefore, if a fixed throttle is provided in this introduction passage, a pressure difference will be generated before and after the throttle, and this will be transferred to the spool of the compensator valve. It becomes possible to make it work easily. In this way, the spool moves against the valve spring due to the differential pressure generated before and after the fixed throttle at the upper limit of the load pressure, and the pump discharge pressure is rapidly introduced into the control element to bring the pump into a full cut-off state. You will be able to do this.

As described above, in the device of the present invention, the same compensator valve portion can also have the control function of bringing the pump into the full cut-off state at a desired discharge pressure value, if necessary.

There is another way to achieve a full cut-off function, and instead of releasing the load pressure to the low pressure line at the upper limit, the spool is subjected to the force of another weak spring opposite to the valve spring.
In this case, the other spring has a lower force than the valve spring, but also has a force greater than the weight of the spool. In this way, when the load pressure increases and reaches a certain upper limit, the piston is displaced away from the spool, and when the force of the valve spring applied to the spool decreases, another spring acts on the piston. The spool fully opens the discharge pressure introduction passage to the control element, thereby placing the pump in a full cut-off state. The full cut-off pressure in this case is determined by the spring force of the other spring.

An embodiment of this invention will be described below with reference to the drawings.

[Embodiment] FIG. 3 is a schematic configuration diagram showing an embodiment of this invention, in which the variable displacement hydraulic pump 21 decreases its discharge volume in response to an increase in hydraulic pressure guided by its discharge volume control element 27. This is the type of thing that does. A fixed orifice 26 having a selected opening area is disposed within the main passage 22 that leads the discharged pressure oil from the pump 21 to a load actuator (not shown).
The compensator valve 20 has a spool 28 slidably disposed in a valve hole 34 provided in a valve body 39, and an orifice connected from a branch passage 23 to an oil chamber 44 in which the distal end surface is located via an inlet port 30. 26 is introduced into the spring chamber 43 where the tail end face on the opposite side is located, the valve spring 35 is applied to the tail end face, and the spring force elastically urges the spool 28 toward the front end face. The orifice 2 is connected to another branch passage 24 through the load pressure port 33 into the spring chamber 43.
6, the pilot port 31 is switched from the tank port 32 to the oil chamber 44 according to the balance between the differential pressure before and after the orifice and the spring force of the valve spring 35, and the pump discharge pressure is transferred to the branch passage 2.
3 to the pump discharge amount control element 2 via the oil chamber 44, the pilot port 31, and the pilot passage 25.
It is configured to be installed in 7.

The main structure of the control spring mechanism 40 is slidably disposed within the valve hole 34 so as to receive the tail end side of the valve spring 35 of the compensator valve 20 and substantially seal the end of the spring chamber 43. It consists of a piston 29 and a control spring 36 that presses the tail end of the piston via a spring push 37.

The control spring mechanism 40 receives the pressure on the outlet side of the orifice 26 led to the spring chamber 43, that is, the load pressure, on the tip surface of the piston 29, thereby causing the piston 29 to be loaded against the spring force of the control spring 36. The displacement of the piston 29 causes the valve spring 35 of the compensator valve 20 to be displaced in response to the pressure.
The amount of deflection is changed according to the load pressure.

In the example shown in FIG. 3, a fixed throttle 42 is installed in the branch passage 24, and a port 41 is installed to drop the spring chamber 43 into the tank when the piston 29 moves back by a predetermined amount of displacement due to an increase in load pressure.
is provided. This allows the spring chamber 43 to communicate with the tank via the port 41 when the load pressure reaches a certain maximum regulation value, thereby creating a pressure oil flow in the branch passage 24, and this flow causes The spool 28 of the compensator valve 20 is moved against the valve spring 35 by the differential pressure generated before and after the fixed throttle 42, the oil chamber 44 is suddenly opened to the port 31, and the discharge amount of the pump 21 is suddenly reduced. A so-called full cut-off operation is performed.
This maximum regulation pressure is set using the adjustment screw 38 of the control spring 36. Now, in FIG. 3, pump 21
Pressure oil discharged from the pump is sent to a load (not shown) through an orifice 26. Pressure oil on the inlet side of the orifice 26 enters the oil chamber 44 of the compensator valve 20 through the branch passage 23 and acts in a direction to press the spool 28 against the valve spring 35. Further, the pressure oil on the outlet side of the orifice 26 enters the spring chamber 42 from the branch passage 24, and the spool 28 enters the spring chamber 42.
It acts in the direction of pressing against the hydraulic pressure of No. 4. As a result, the spool 28 is connected to the orifice 26.
The pump stops at a position where the differential pressure before and after and the force of the valve spring 35 are balanced, resulting in a balanced state, and the pressure oil to the pump discharge amount control element 27 is controlled.

When the load pressure increases from this balanced state,
The piston 29 is activated by the control spring 3 according to its load pressure.
6, and the amount of deflection of the valve spring 35 is reduced by the amount of movement. This allows spool 2
8 is out of balance, the spool 28 moves to further bend the valve spring 35 in order to restore balance, the controlled opening from the oil chamber 44 to the port 31 is opened, and the pump 21 is therefore Reduce discharge rate. Due to this decrease in discharge amount, the differential pressure across the orifice 26 also decreases, and the spool 28 returns to a balanced state.

In this way, the pump discharge amount decreases as the load pressure increases, resulting in an approximate constant horsepower characteristic as shown in FIG. In this case, if the valve spring 35 has non-linear spring characteristics, the obtained approximate constant horsepower characteristic can be brought closer to the ideal constant horsepower characteristic. In FIG. 4, the characteristic curve suddenly drops to zero due to the increase in pressure, but this is due to the full cut-off operation by the port 41 and fixed throttle 42 described above.

FIG. 5 shows another embodiment of this invention, in which parts having the same effect as those in FIG. 3 are given the same reference numerals. The difference from the one in Figure 3 is that instead of the port 41 to the tank and the fixed throttle 42, the valve spring 3
Another point is that another spring 45, which is weaker than 5 and has enough force to just cancel out the weight of the spool 28, is arranged in the oil chamber 44. In this case, when the load pressure increases and the pump discharge amount gradually decreases in the same manner as described above, when the piston 29 is displaced upward against the control spring 36, the force that the valve spring 35 has applied to the spool 18 gradually decreases, and when it reaches zero, the other spring 45 pushes up the spool 28 against its own weight, and the valve chamber 44 fully opens to the port 31. As a result, a pressure versus flow rate characteristic as shown by the dashed line in FIG. 4 is obtained. The full cut-off pressure of the pump in this case can be varied within this range by selecting the spring force of the other spring 45 to be weaker than the valve spring 35 and sufficient to cancel out the dead weight of the spool 28. That is, the characteristic shown by the chain line in FIG.
This corresponds to the case when the size is selected to cancel out the dead weight of the spring 8. If this spring is made a little stronger, the full cut-off pressure will shift to the lower pressure side as shown by the solid line in Fig. 4. It should be noted that instead of using the separate spring 45 in the example of FIG. 5, the spool 28 may be moved back integrally with the piston 29 when the backward movement of the piston 29 against the control spring 36 exceeds a certain amount. Alternatively, the oil chamber 44 may be fully opened to the port 31.

FIG. 6 shows an example of application using the embodiment of this invention shown in FIG. 3, and parts having the same effect as those in FIG. 3 are given the same reference numerals. In the example shown in FIG. 6, a differential pressure sensitive throttle valve 46 is further added to the one shown in FIG. 3, so that the pressure versus flow rate characteristics can be brought closer to the ideal constant horsepower characteristics.

That is, the load pressure sensitive throttle valve 46 is provided in parallel with the orifice 26 and forms a bypass flow path 49 that changes the flow path area according to the load pressure.
This throttle valve 46 includes a spool 47 that controls the opening degree of a bypass flow path 49 that is parallel to the orifice 26.
6 and press the spool 47 toward the limiting member 50 with the spring 48.
The spring force of this spring 48 can be variably set using an adjustment screw 51.

Now, the differential action of the compensator valve 20 when the load pressure rises is the same as in the case of FIG. 3. In other words, as the load pressure increases, the spring force of the valve spring 35 acting on the spool 28 decreases, and the differential pressure across the orifice 26 also decreases accordingly. Since the differential pressure before and after the orifice 26 acts on the spool 47 of the differential pressure sensitive throttle valve 46 in balance with the spring 48, the degree of opening of the bypass passage 49 by the spool 47 also depends on the differential pressure caused by the increase in load pressure. It will decrease as the value decreases. Here, the discharge flow rate of the pump 21 is the sum of the flow rate passing through the orifice 26 and the flow rate passing through the bypass passage 49 whose flow area is controlled by the differential pressure sensitive throttle valve 46. Therefore, the spring force of the spring 48 of the differential pressure sensitive throttle valve 46 is appropriately selected so that the spool 47 opens the bypass passage 49 at an opening degree corresponding to the increase in the differential pressure (decrease in the load pressure). As shown in FIG. 7, the pressure versus flow rate characteristic in this case is the characteristic when the bypass passage 49 remains closed (the
In contrast to A (same as the solid line curve in the figure), the discharge amount increases by the amount flowing through the bypass flow path of the throttle valve 46 in the low load pressure region, resulting in characteristics a, b, and c.
In this way, constant horsepower characteristics are brought closer to the ideal.

Incidentally, as the set load of the spring 48 is reduced by the adjustment screw 51, the characteristic changes from a to b to c in FIG.

[Effects of the invention] As stated above, according to this invention, only a fixed orifice is placed in the main passage;
The differential pressure before and after the fixed orifice is applied to the spool of the compensator valve, and the spring force of the valve spring that presses and biases the spool is changed according to the load pressure, so the advantages of the compensator valve system can still be utilized. , approximate constant horsepower control of variable displacement pumps can be performed stably over a wide pressure range,
Furthermore, 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, and it is easy to add a pump full cut-off control function to the same compensator valve.

[Brief explanation of the drawing]

Fig. 1 is a sectional view showing a conventional control device, Fig. 2 is a hydraulic circuit diagram showing another conventional example, Fig. 3 is a schematic configuration diagram showing an embodiment of this invention, and Fig. 4 is the same as the previous figure. A diagram illustrating the pressure-flow characteristics of the example, FIG. 5 is a schematic configuration diagram showing another embodiment of this invention,
FIG. 6 is a schematic block diagram showing an application example using the embodiment of this invention, and FIG. 7 is a diagram illustrating the pressure-flow characteristics of the example shown in the previous figure. Explanation of symbols of main parts, 20: Compensator valve, 21: Variable displacement hydraulic pump, 26: Fixed orifice, 27: Pump discharge amount control element, 28:
Spool, 29: Piston, 35: Valve spring, 3
6: control spring, 40: control spring mechanism.

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 fixed orifice disposed in a passage through which pressure oil discharged from the pump flows toward the load, and a pressure oil pressure on the inlet side of the fixed orifice of the passage. has a spool configured to receive on one end face and the hydraulic pressure on the outlet side of the fixed orifice together with the spring force of the valve spring on the other end face,
a compensator valve that controls the introduction and discharge of discharge pressure oil to the control element so that the differential pressure across the fixed orifice follows the value determined by the spring force by the spool; A control device for a variable displacement hydraulic pump, comprising: a control spring mechanism that presses the valve spring toward the spool and changes the amount of deflection of the valve spring in accordance with load pressure.