JPH11292474A - Hydraulic control device - Google Patents

Abstract

PROBLEM TO BE SOLVED: To secure damping force of a hydraulic control device for a motor for swiveling a crane and the like when the motor is driven. SOLUTION: Two lines 6A, 6B linking a swiveling hydraulic motor 2 with a directional control valve 1 for swiveling are connected together with a solenoid proportional valve 9. Pressure sensors 10A, 10B provided on the two lines 6A, 6B and a speed sensor 11 11 provided on a swiveling body are connected to a control device 12, which calculates a target opening amount A of the solenoid proportional valve 9 based on high frequency element signals H1, H2 of a speed fluctuation element and a pressure fluctuation element respectively obtained from the speed sensor 11 and pressure sensors 10A, 10B and a differential signal S2 of line 6A, 6B, and a control signal A' corresponding to the target opening amount A is outputted to the solenoid proportional valve 9. By controlling the opening of the solenoid proportional valve 9 in such a manner, damping force acting on the motor 2 can be secured and vibration can be quickly lowered.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a hydraulic control device for ensuring a damping force during operation of a hydraulic actuator.

[0002]

2. Description of the Related Art Conventionally, various devices have been proposed for preventing a shock when an actuator is stopped when a fluid pressure actuator is driven in accordance with, for example, an operation amount of an operation lever. For example, Japanese Utility Model Laid-Open No. 5-40601 discloses a hydraulic pump 31, a hydraulic motor 33, a control valve 32 for switching the flow of hydraulic oil discharged from the hydraulic pump 31, and a control valve 32, as shown in FIG. Operating lever 37 for switching the hydraulic motor 33
An electromagnetic switching valve 43 capable of communicating with two pipes 34, 35 connected to each other, pressure sensors 41, 42 for detecting the pressures of the pipes 34, 35, and a detection result of the pressure sensors 41, 42 A hydraulic control device including a control unit 44 for switching the electromagnetic switching valve 43 has been proposed.

[0003] This hydraulic control device operates as follows. When the operation lever 37 is operated to the side A, the control valve 32 is switched to the position a, and the pressure P1
Occurs, and the hydraulic motor 33 rotates in one direction. In this state, when the operation lever 37 is operated to the neutral position,
The control valve 32 is switched to the position c, and the line 3
4 decreases, but the pressure P2 in the conduit 35 increases. Then, when the rising speed of the pressure P2 exceeds the reference value, the control unit 44 opens the electromagnetic switching valve 43 to connect the pipes 34 and 35, and reduces the pressure difference between the two pipes 34, 35. . Furthermore, the hydraulic motor 33
Is reversed and the pressure P1 in the pipeline 34 increases,
When the rising speed exceeds the reference value, the control unit 44 opens the electromagnetic switching valve 43 to connect the pipes 34 and 35, and reduces the pressure difference between the two pipes 34 and 35. Thereby, when the hydraulic motor 33 is stopped, it is possible to prevent the pressure in the pipelines 34 and 35 from suddenly rising, thereby preventing an impact from occurring in the device.

[0004]

However, in the device described in Japanese Utility Model Application Laid-Open No. H5-40601, when the hydraulic motor 33 is stopped, the pressures P1 and P2 in the pipelines 34 and 35 are prevented from sharply increasing. Although it is possible, the damping force acting on the hydraulic motor 33 is not always ensured. Actuators such as hydraulic motors vibrate at a specific frequency determined by the compressibility of pressure oil and the inertial mass driven by the actuator, etc., but if the damping force of the actuator against this vibration is small, the vibration is immediately converged. As a result, not only does the operator feel uncomfortable, but also it takes time to reach a stable operation of the actuator, and the operability deteriorates.

SUMMARY OF THE INVENTION It is an object of the present invention to secure a damping force for an actuator driven by hydraulic pressure, thereby effectively suppressing vibration, swingback, hunting, and the like of the actuator caused by insufficient damping force, and ensuring stable operation of the actuator. Another object of the present invention is to provide a hydraulic control device that does not cause a delay.

[0006]

A description will be given with reference to the drawings showing an embodiment. (1) As shown in FIG. 9, the invention of claim 1 provides a hydraulic pump 3, an actuator 2 driven by hydraulic oil discharged from the hydraulic pump 3, and a hydraulic oil supplied to the actuator 2 from the hydraulic pump 3. A control valve 1 for controlling the flow of
An operating lever 51 for switching the control valve 1 and instructing a target flow rate
This is applied to a hydraulic control device having: The pressure detectors 10A and 10B detect the pressures of the hydraulic oil in the two conduits 6A and 6B respectively connected to the inlet / outlet ports of the actuator 2 and output pressure signals, and the driving speed of the actuator 2 is detected. A speed detector 11; a correction flow rate calculating means 32 for calculating a correction flow rate based on the pressure signals detected by the pressure detectors 10A and 10B and the speed signal detected by the speed detector 11; The above object is achieved by providing control means 50A, 50B and 32 for controlling the control valve 1 with a flow rate obtained by subtracting the correction flow rate from the control flow rate. (2) According to the second aspect of the present invention, as shown in FIGS. 1 and 2, the hydraulic pump 3, the actuator 2 driven by hydraulic oil discharged from the hydraulic pump 3, and the hydraulic pump 3 are supplied to the actuator 2. Control valve 1 for controlling the flow of pressurized oil
And an operating lever 5 for switching the control valve 1. The pressure detectors 10A and 10B detect the pressures of the hydraulic oil in the two conduits 6A and 6B respectively connected to the entrance and exit ports of the actuator 2 and output pressure signals, and the speed for detecting the drive speed of the actuator. Detector 11 and each pressure detector 10A, 10B
And a correction flow rate calculating means 12 for calculating a correction flow rate based on the pressure signal detected by the speed detector 11 and the speed signal detected by the speed detector 11, and the correction flow rate calculation means 12 is inserted into a pipe communicating the two pipes 6A and 6B. The valve device 9 for adjusting the flow rate flowing through the pipeline and the bypass flow rate supplied from the high-pressure pipeline 6A (6B) to the low-pressure pipeline 6B (6A) bypassing the actuator 2 are calculated. So that the corrected flow rate
The above-mentioned object is achieved by providing the control means 12 for controlling the valve device 9. (3) The invention according to claim 3 is that the valve device 9 is an electromagnetic proportional valve 9 whose opening amount is controlled based on the corrected flow rate. (4) The invention according to claim 4 is characterized in that the corrected flow rate calculating means 12, 3
1 to 32 are frequency components of a pressure signal detected by the pressure detectors 10A and 10B which are equal to or higher than a predetermined value, and the speed detector 1
The correction flow rate is calculated on the basis of the frequency component of the pressure signal detected by (1) and higher than a predetermined value. (5) As shown in FIG. 10, the invention according to claim 5 is a hydraulic pump 3, an actuator 2 driven by hydraulic oil discharged from the hydraulic pump 3, and a hydraulic oil supplied to the actuator 2 from the hydraulic pump 3. Control valve 1 for controlling the flow of air
And an operating lever (5) for switching the control valve 1. The two conduits 6A and 6B connected to the entrance / exit ports of the actuator 2, respectively.
Pressure detectors 10A and 10B that respectively detect the pressure of the pressure oil and output a pressure signal, and the pressure detectors 10A and 10B
Flow rate K2 · H based on the pressure signal detected by
2 and the two pipe lines 6
A and 6B are inserted into a pipeline communicating with the pipeline, and a valve device 9 for adjusting a flow rate flowing through the pipeline, and a pipeline 6B on the low pressure side from the pipeline 6A (6B) on the high pressure side bypassing the actuator 2.
The above-described object is achieved by providing the control unit 33 that controls the valve device 9 so that the bypass flow rate supplied to (6A) becomes the calculated correction flow rate K2 · H2. (6) According to a sixth aspect of the present invention, the control valve 1 is a flow control type control valve having a flow controllability independent of a load.

In the section of the means for solving the above-mentioned problems, which explains the configuration of the present invention, the drawings of the embodiments of the present invention are used to make the present invention easy to understand. However, the present invention is not limited to this.

[0008]

Embodiments of the present invention will be described below with reference to the drawings. -First Embodiment- FIG. 1 is a circuit diagram showing a configuration of a hydraulic control device according to an embodiment of the present invention, and FIG. 2 is a detailed configuration of a control unit of the hydraulic control device according to the first embodiment. FIG. 3 is a side view showing a configuration of a crane in which the hydraulic control device according to the present embodiment is used. As shown in FIG. 3, the mobile crane includes a traveling body 61, a revolving revolving body 62 mounted on the traveling body 61, a boom 63 supported on the revolving body 62 so as to be able to undulate, and a boom 63. Sheave 64 provided at the tip
And a hook 65 connected to a wire rope via a sheave 64. A suspended load 66 is suspended from the hook 65.

As shown in FIG. 1, a hydraulic circuit for turning the revolving unit 62 of the mobile crane includes a hydraulic pump 3 driven by a prime mover 13 and a hydraulic pump driven by pressure oil discharged from the hydraulic pump 3. Directional control valve 1 for controlling the flow of pressurized oil supplied from hydraulic pump 3 to slewing hydraulic motor 2, operating lever 5 for the operator to input a slewing command, and operating lever 5 The pilot valves 14A and 14B to be operated and the pilot valve 1
A pilot hydraulic pressure source 7 for supplying pressure oil to 4A and 14B,
Hydraulic oil tank 8, relief valve 4, and two pipelines 6A, 6B connected to the inlet / outlet port of hydraulic motor 2 for turning.
An electromagnetic proportional valve 9 for communicating and shutting off the two pipes 6A and 6B via throttles, and a pressure sensor 10 for measuring the oil pressure of the pressure oil in the pipes 6A and 6B and outputting pressure signals P1 and P2.
A, 10B, a speed sensor 11 that detects the turning speed of the revolving unit 62 and outputs a plus speed signal S1 for forward rotation (A direction) and a minus speed signal S1 for reverse rotation (B direction), and pressure sensors 10A and 10B. And a controller 12 for controlling the valve opening of the proportional solenoid valve 9 in accordance with a detection value from the speed sensor 11.
Consists of

As shown in FIG. 2, the controller 12
A high-pass filter 21 for extracting a high-frequency component signal H1 having a predetermined value or more from the speed signal S1;
A multiplier 22 that multiplies 1 by a gain K1, a subtractor 23 that subtracts P2 from the pressure signal P1 (P1-P2) and calculates a difference signal S2, and a high-frequency component signal H2 of a predetermined value or more from the difference signal S2. A high-pass filter 24 to be taken out;
A multiplier 2 for multiplying the high frequency component signal H2 by a gain K2
5 and the high frequency component signals K1 · H1 and K2 multiplied by the gain
An adder 26 that adds H2 and calculates an addition signal KH3;
And the addition signal KH3 and the difference signal S2, the electromagnetic proportional valve 9
, A sign discriminator 28 for judging the sign of the difference signal S2, and a conversion table for outputting a control signal A 'corresponding to the target opening A to the electromagnetic proportional valve 9. 29A and 29B.

Here, the phenomenon of pressure oscillation will be described. FIGS. 4A and 4B are explanatory diagrams of pressure oscillation. In FIG. 4A, the hydraulic motor 2 is replaced with a hydraulic cylinder in order to simplify the configuration. When pressure oil having a flow rate Q flows from the turning direction control valve 1 into the hydraulic cylinder, the pressure P in the hydraulic cylinder increases due to the flow rate Q due to the compressibility of the hydraulic oil (compressibility as a whole system). Then, a thrust F is generated by the pressure P, and the object m causes an acceleration motion by the thrust. The speed of the object m is obtained by integrating this acceleration, and the speed increases the volume in the hydraulic cylinder, the rate of compressing the oil decreases, and the pressure P decreases. Then, a vibration phenomenon occurs due to the repetition of these. In order to quickly converge such vibration phenomena, it is necessary to increase the viscous resistance such as friction acting on the hydraulic cylinder to ensure a sufficient damping force. FIG. 4B is a diagram in which FIG. 4A is replaced with a vibration model, and pressure vibration is considered to be similar to vibration of a general spring-mass system including a spring and a mass (mass). Can be

Next, this phenomenon will be described with reference to the turning hydraulic motor 2. The turning direction control valve 1 includes an operation lever 5
Is supplied to the turning hydraulic motor 2 according to the input. Actually, a speed reducer, a bearing, a sliding surface (not shown), etc. (not shown) act as frictional resistance during movement on the turning hydraulic motor 2, so that the differential pressure (P
1-P2) and the speed of the revolving superstructure whose driving is controlled by the front-rear differential pressure (P1-P2) statically have a smooth waveform without vibration as shown in FIG. However, when the damping force of the revolving structure is not sufficient, the above-mentioned vibration phenomenon is added to the static characteristics of FIG. 5, and the differential pressure across the hydraulic motor 2 for revolving and the speed of the revolving structure are as shown in FIG. It becomes a very oscillating waveform. Such an oscillating waveform causes a turning operation such as hunting and swinging back, and causes a delay until reaching a stable turning operation.

Hereinafter, the operation of the first embodiment will be described. When the operating lever 5 is operated to switch the turning direction control valve 1 to the a position or the b position, pressure oil is supplied from the hydraulic pump 3 to the hydraulic motor 2, and the turning hydraulic motor 2 rotates. When the operation lever 5 is stopped to switch the turning direction control valve 1 to the position c, the hydraulic pump 3
The hydraulic oil is shut off, and the hydraulic motor 2 stops. When the hydraulic motor 2 starts or stops, the pressure of the pressure oil in the pipelines 6A and 6B fluctuates due to insufficient damping force acting on the hydraulic motor 2 and is detected by the pressure sensors 10A and 10B. Pressure difference (P1-
P2) and the turning speed detected by the speed sensor 11 have an oscillating waveform having a high frequency component as shown in FIG.

The pressure signals P1 and P2 are input to a differentiator 23 of the controller 12 as shown in FIG.
2 is input to the high-pass filter 24, and the high-frequency component signal H2 exceeding the predetermined value f1 among the difference signals is filtered out. The predetermined value f1 is f2 <f1 <f0 in consideration of, for example, a response frequency f2 (for example, 1 Hz) of a lever operation performed by an operator, a natural frequency f0 uniquely determined by an inertial mass of the revolving unit, a hydraulic system, and the like. Is set to a value that satisfies the condition FIG. 7 is an example of the high-frequency component signal H2. Since the data has passed through the high-pass filter 24, the steady-state value has been removed, that is, the pressure fluctuation based on the static pressure waveform of FIG. . In FIG. 7, a pressure larger than the static pressure waveform is plus (H2 ≧ 0), and a smaller pressure is minus (H2 <0), and the plus and minus pressure waveforms appear alternately. Such a high-frequency component signal H2 is multiplied by a predetermined gain K2 in a multiplier 25, and a signal of K2 · H2 is output from the multiplier 25.

The speed signal S1 is input to a high-pass filter 21, and a high-frequency component signal H1 exceeding a predetermined value f1 'is filtered and taken out. The predetermined value f1 'in this case is, for example, the predetermined value f1 described above.
Is set to the same value as Although the high-frequency component signal H1 has a time delay with respect to the high-frequency component signal H2 due to pressure fluctuation, it has a waveform substantially similar to that of FIG.
Plus (H1 ≧ 0) and minus (H1 <0) speed waveforms appear alternately based on the static speed waveform. Then, the high-frequency component signal H1 is converted into a predetermined gain K by the multiplier 22.
The multiplier 22 outputs a signal of K1 · H1.

K1 · H1 and K thus output
Each signal of 2 · H2 is added by the adder 26, and the adder 26
Outputs an addition signal KH3. This addition signal KH3
Is a signal calculated in consideration of both the speed fluctuation component of the inertial body and the pressure fluctuation component in the pipeline, and is necessary to remove these fluctuation components by increasing the damping force acting on the hydraulic motor. Corresponding to the target flow rate. In the opening amount calculator 27, a calculation represented by the following equation (I) is executed based on the vibration component KH3 and the absolute value | S2 | of the difference signal S2, and the target opening amount A is calculated.

A = C1 · KH3 / √ | S2 | where C1: a constant (I) The above equation (I) is a general orifice equation and is given by the following equation (I)
The target opening amount A of the electromagnetic proportional valve 9 is determined based on the addition signal KH3 corresponding to the flow rate Q of the electromagnetic proportional valve 9 and the differential pressure signal S2 corresponding to the orifice differential pressure ΔP. Is done.

Q = C2 · A ＝ (2 · △ P / ρ) where C2: constant ΔP: differential pressure across orifice ρ: density (II) The sign discriminator 28 discriminates the sign of the difference signal S2. , 0
If ≤S2, the target aperture A is input to the conversion table 29A, and if S2 <0, the target aperture A is converted to the conversion table 29A.
B is input. Each of the conversion tables 29A and 29B outputs a control signal A 'to the electromagnetic proportional valve 9 based on the magnitude of the opening amount A calculated by the opening amount calculator 27.

The operation of such a hydraulic control device will be described more specifically. When the operating lever 5 is actuated to rotate the revolving superstructure forward (direction A in FIG. 1), the turning direction control valve 1 is switched to the position b in accordance with the operation amount, and the pressure from the hydraulic pump 3 is changed. Oil is supplied to line 6A. As a result, the driving speed of the inertial body and the differential pressure (P1-P2) in the pipelines 6A and 6B increase while fluctuating, and the speed signal S1 is increased.
≧ 0, and the difference signal S2 ≧ 0. Here, due to the pressure fluctuation, the differential pressure (P1-P2) in the pipelines 6A and 6B exceeds the static pressure waveform (H2 ≧ 0), and the speed fluctuation component also becomes positive (H1 ≧ 0). , The addition signal KH3 is KH3 ≧
When the value becomes 0, the control signal A ′ from the conversion table 29A is output to the electromagnetic proportional valve 9. The electromagnetic proportional valve 9 has an opening corresponding to the signal A ′, and a part of the pressure oil in the pipe 6A bypasses the hydraulic motor 2 and passes through the electromagnetic proportional valve 9 to flow into the pipe 6B. Further, the differential pressure (P1-P2) in the pipelines 6A and 6B falls below the static pressure waveform (H2 <0), and the speed fluctuation component similarly becomes minus (H1 <0), and the addition signal KH3
, The control signal A ′ = 0 is output from the conversion table 29A, the electromagnetic proportional valve 9 is closed, and the bypass of the pressure oil is stopped.

As described above, when the addition signal KH3 ≧ 0, a part of the pressure oil in the high-pressure side line 6A is removed from the low-pressure side line 6B.
By virtue of this, the damping force acts greatly, and pressure fluctuations in the pipeline and speed fluctuations of the revolving structure are quickly prevented, and there is no delay before reaching a stable turning operation, and smooth Acceleration operation can be realized. When the addition signal KH3 <0, the pressure oil cannot be bypassed from the low-pressure side line 6B to the high-pressure side line 6A. An extra bypass to 6B is prevented,
The responsiveness at the time of driving the hydraulic motor can be improved.

On the other hand, when the operation lever 5 is stopped to stop the revolving body that is rotating forward, the turning direction control valve 1 is switched from the position b to the position c, and the hydraulic oil from the hydraulic pump 3 is pressed. Supply is stopped. As a result, the driving speed of the inertial body and the differential pressure (P1-P2) in the pipelines 6A and 6B are obtained.
And decrease (increase in the negative direction) while fluctuating. At this time, since the hydraulic motor 2 is rotated forward by the inertial force of the inertial body, the speed signal S1 ≧ 0, and the difference signal S2 <0.
Becomes Here, due to the pressure fluctuation, the differential pressure (P1-P2) in the pipelines 6A and 6B falls below a static pressure waveform (H2 <
0), the speed fluctuation component also becomes minus (H1 <), and when the addition signal KH3 becomes KH3 <0, the control signal A ′ from the conversion table 29B is output to the electromagnetic proportional valve 9. The electromagnetic proportional valve 9 has an opening degree corresponding to the signal A ′, and a part of the pressure oil in the pipe 6B passes through the electromagnetic proportional valve 9 and the pipe 6A
To prevent pressure fluctuations. In addition, pipe 6A,
6B exceeds the static pressure waveform (H2 ≧ 0), and the speed fluctuation component is also positive (H1 ≧ H2).
0) and the addition signal KH3 satisfies KH3 ≧ 0,
Control signal A ′ = 0 is output from conversion table 29b. As a result, the electromagnetic proportional valve 9 is closed and the line 6 is closed.
An extra bypass from B to line 6A is prevented.

When the operating lever 5 is actuated to reverse the revolving structure (direction B in FIG. 1), the revolving direction control valve 1 is switched to the position a, and the pressure oil from the hydraulic pump 3 is released. It is supplied to the pipe 6B. As a result, the drive speed of the inertial body and the differential pressure (P1-P2) in the pipelines 6A and 6B decrease (increase in the negative direction) while fluctuating, and the speed signal S1 <0 and the difference signal S2 <0. . Here, due to the speed fluctuation of the revolving superstructure and the pressure fluctuation in the pipeline, the addition signal K is obtained.
When H3 becomes KH3 <0, the control signal A ′ from the conversion table 29B is output to the electromagnetic proportional valve 9, and the electromagnetic proportional valve 9
Is an opening in accordance with the signal A ', and a part of the pressure oil in the pipe 6B bypasses the hydraulic motor 2 and passes through the electromagnetic proportional valve 9 to flow into the pipe 6A. Then, the addition signal KH3 becomes KH3.
When ≧ 0, the control signal A ′ =
0 is output, the electromagnetic proportional valve 9 is closed, and the bypass of the pressure oil is stopped.

Further, when the operating lever 5 is stopped to stop the revolving revolving structure, the revolving direction control valve 1 is switched from the position a to the position c, and the pressure oil from the hydraulic pump 3 is released. The supply is stopped. As a result, the driving speed of the inertial body and the differential pressure (P1-P
2), it increases while fluctuating, and the speed signal S1 <0 and the difference signal S2 ≧ 0. Here, the addition signal KH3 is KH3 ≧ 0 due to the speed fluctuation of the revolving superstructure and the pressure fluctuation in the pipeline.
Then, the control signal A ′ from the conversion table 29A is output to the electromagnetic proportional valve 9 and a part of the pressure oil in the pipe 6B bypasses the hydraulic motor 2 and passes through the electromagnetic proportional valve 9 to pass the pipe 6A
Flows into. When the addition signal KH3 becomes KH3 <0, the control signal A ′ = 0 is output from the conversion table 29A, the electromagnetic proportional valve 9 is closed, and the bypass of the pressure oil is stopped.

As described above, according to the first embodiment, the valve opening of the electromagnetic proportional valve 9 is controlled based on the pressure fluctuation component and the speed fluctuation component, and the predetermined flow rate is changed from the high pressure side to the low pressure side. Since the bypass is made or the bypass is prohibited, the damping force acting on the hydraulic motor 2 can be compensated without deteriorating the responsiveness of the turning operation, and the vibration of the turning body can be suppressed promptly. Further, high-frequency components equal to or higher than a predetermined value are extracted using the high-pass filters 21 and 24, and the opening of the electromagnetic proportional valve 9 is controlled so as to remove the high-frequency components. It does not adversely affect the input of the low-frequency operator.

Second Embodiment FIG. 8 is a diagram showing a detailed configuration of a control unit of a hydraulic control device according to a second embodiment. The same parts as those in FIGS. 1 and 2 used in the first embodiment are denoted by the same reference numerals, and the differences will be mainly described below. As shown in FIG. 8, the controller 31 of the second embodiment has
The high-pass filters 21 and 24 of the third embodiment are not provided. Therefore, the multiplier 22 multiplies the speed signal S1 detected by the speed sensor 11 by the gain K1, and the multiplier 25 multiplies the difference signal S2 between the pressure signals P1 and P2 detected by the pressure gauges 10A and 10B by the gain K.
Multiply by two. The adder 26 adds the high-frequency component signals K1 · S1 and K2 · S2 after multiplication of the gains K1 and K2 to calculate an addition signal KS3. A target opening amount A is calculated based on the addition signal KS3 and the difference signal S2, and the drive of the electromagnetic proportional valve 9 is controlled by a control signal A 'corresponding to the target opening amount A.

In the second embodiment constructed as described above, similarly to the first embodiment, the electromagnetic proportional valve is controlled based on the pressure fluctuation components in the pipes 6A and 6B and the speed fluctuation component of the revolving unit. 9 is controlled, an effective damping force for eliminating these fluctuations can be applied to the hydraulic motor 2. Further, in the second embodiment, since the high-pass processing is not required, the configuration of the controller 31 can be simplified. However, in the second embodiment, since the high-pass processing of the speed signal S1 and the differential pressure signal S2 is not performed, the electromagnetic proportional based on the pressure fluctuation and the speed fluctuation in all the frequency bands including the stationary data. Valve 9
Is calculated. As a result, for example, when the revolving structure is rotating at a constant speed without vibration in the normal rotation direction (A direction), in the first embodiment, both the speed signal S1 and the differential pressure signal S2 are positive. There is
Since the signals H1 and H2 that have passed through the high-pass filters 21 and 24 are both zero, the addition signal KH3 = 0 is set. However, in the second embodiment, the addition signal KS3> 0
It is said. As a result, the bypass flow rate KS3 passes through the electromagnetic proportional valve 9 even at the time of constant-speed turning without vibration.

Third Embodiment FIG. 9 is a diagram showing a detailed configuration of a control unit of a hydraulic control device according to a third embodiment. The same parts as those in FIGS. 1 and 2 used in the first embodiment are denoted by the same reference numerals, and the differences will be mainly described below. As shown in FIG. 9, in the third embodiment, when the rotation is forward,
An operation lever 51 for outputting an operation signal S3 corresponding to a negative operation amount at the time of reverse rotation, and an electromagnetic proportional pilot valve 50A for driving the turning direction control valve 1 whose valve opening is controlled by a control signal from a controller 32; 50B,
The electromagnetic proportional valve 9 as in the first embodiment is omitted. The controller 32 is different from the controller 12 of the first embodiment in that a multiplier 52 for multiplying an operation signal S3 representing a target flow rate from an operation lever 51 by a gain K3, and a multiplication value S3 · K3, Addition signal K calculated based on detection values from pressure gauges P1 and P2
A subtractor 53 for subtracting S3 and calculating a subtraction signal QCV;
Conversion table 5 for outputting control signals PA and PB corresponding to the subtraction signal QCV to the electromagnetic proportional pilot valves 50A and 50B.
4A and 54B.

The operation of the third embodiment configured as described above will be described. When the operation lever 51 is activated to drive the revolving superstructure in the normal rotation direction (A direction),
An operation signal S3 corresponding to the operation amount is multiplied by a gain K3. Immediately after startup, the differential pressure across the hydraulic motor 2 (P1-
Since P2) and the swing speed of the swing body are both zero, the addition signal KS3 = 0, and the operation signal K3 · S3 after multiplication by the gain K3 is directly output as the subtraction signal QCV (> 0). This subtraction signal QCV is output from the conversion table 54
A converts the control signal PA into a control signal PA, and the control signal PA controls the valve opening (secondary pressure) of the electromagnetic proportional pilot valve 50A. As a result, the pilot port 1a of the turning direction control valve 1 is supplied with a pilot pressure in which the primary pressure from the pressure source 7 is reduced according to the control signal PA, and the turning direction control valve 1 is rotated in the forward direction. The switching is performed, and the hydraulic motor 2 is driven in the normal rotation direction (A direction). Since the subtraction signal QCV> 0 at the start of forward rotation, the electromagnetic proportional pilot valve 50B
A control signal PB (= 0) from the conversion table 54B is output to the pilot port 1b of the turning direction control valve 1.
Is not supplied with pilot pressure.

When the revolving body starts revolving, the hydraulic motor 2
The differential pressure (P1-P2) and the swing speed of the swing body
It increases while fluctuating as shown in FIG. The controller 32 includes a speed sensor 11 and pressure gauges 10A, 10A.
An addition signal KS3 is calculated based on the detection value from B,
In the differentiator 53, a subtraction signal QCV obtained by subtracting the addition signal KS3 from the operation signal K3 · S3 after multiplying by the gain K3.
Is calculated. In accordance with the subtraction signal QCV, one of the electromagnetic proportional pilot valves 50A and 50B is controlled in the same manner as described above, and the drive of the hydraulic motor 2 is controlled.

As described above, according to the third embodiment, the addition signal KS3 is calculated based on the pipe pressure and the turning speed, and the operation signal K3 · S corresponding to the operation amount of the operation lever 51 is calculated.
3, the subtraction signal QCV obtained by subtracting the addition signal KS3 from
Since the turning direction control valve 1 is drive-controlled to control the flow of the pressure oil in the pipelines 6A and 6B, it is possible to compensate for the lack of the damping force acting on the hydraulic motor 2 and to reduce the pressure fluctuation and Speed fluctuation can be suppressed promptly. Further, since only a required flow rate is allowed to pass through the pipelines 6A and 6B without bypassing a predetermined flow rate of hydraulic oil from the electromagnetic proportional valve 9 as in the first embodiment, energy loss is also reduced. Reduced.

When the revolving structure is in a constant speed revolving state without vibration, as described above, the speed signal S1 and the pressure signal S2 are both positive, and the addition signal KS3> 0. Here, if the controller 52 is not provided with the multiplier 52, the operation signal S3
Is directly subtracted from the sum signal KS3.
CV becomes smaller than the operation signal S3 (QCV <S3), and as a result, the actual turning speed becomes lower than the turning speed corresponding to the operation signal S3. Such an operation signal S3
The controller 32 is provided with a multiplier 52 in order to prevent a deviation between the rotation speed and the turning speed, that is, to make the subtraction signal QCV and the operation signal S3 equal (QCV = S3). Operation signal S3
Is multiplied by a predetermined gain K3 (K3 · S3), S3 <
K3 · S3, and the difference signal 53 subtracts the operation signal KS3 from the signal K3 · S3 (K3 · S3-KS3).
QCV = S3. As a result, at the time of constant speed turning, the turning body is turned at a speed corresponding to the operation amount of the operation lever 51.

The operation of the revolving superstructure at the time of reverse rotation or at the time of stop is described with reference to the sign of the operation signal S3 and the addition signal KS3,
The basic operation is the same as that described above except that the sign of the subtraction signal Q is inverted, and a description thereof will be omitted.

Fourth Embodiment FIG. 10 is a diagram showing a detailed configuration of a control unit of a hydraulic control device according to a fourth embodiment. The same parts as those in FIGS. 1 and 2 used in the first embodiment are denoted by the same reference numerals, and the differences will be mainly described below. As shown in FIG. 10, in the fourth embodiment, the speed sensor 11
Is omitted. Therefore, only the signals P1 and P2 from the pressure sensors 10A and 10B are taken into the controller 33, and all the terms of the velocity components are ignored. The controller 33 filters a high-frequency component signal H2 of a predetermined value or more from the pressure difference signal S2, and multiplies the filtered signal by a gain K2 to calculate a pressure signal K2 · H2. The target opening A of the electromagnetic proportional valve 9 is calculated according to the pressure signals K2 and H2, and the drive of the electromagnetic proportional valve 9 is controlled by a control signal A 'corresponding to the target opening A. In this way, a damping force acts on the hydraulic motor 2 so as to reduce pressure fluctuations in the pipelines 6A and 6B.

As described above, according to the fourth embodiment, since the pressure signals K2 and H2 are calculated from the detection values of the pressure gauges 10A and 10B to control the drive of the electromagnetic proportional valve 9, a simple configuration is provided. Thus, the damping force acting on the hydraulic motor 2 is compensated, and the pressure fluctuation component can be effectively removed. In this case, if the gain K2 is set large,
The flow rate that bypasses the hydraulic motor 2 increases, the damping force increases, and the high-frequency vibration component can be greatly attenuated. However, as shown by the broken line in FIG. 11, the response of pressure increase / decrease is deteriorated by the increase in the setting of the gain K2. Therefore, as shown by the alternate long and short dash line in FIG. 11, from the viewpoint of effectively attenuating the high-frequency vibration component without deteriorating the response, as described in the first to third embodiments. In addition, it is desirable to take into account not only the pressure fluctuation component but also the speed fluctuation component, and to appropriately determine the gains K1 and K2 to compensate for the damping force.

In the above-described embodiment, the valve openings of the electromagnetic proportional valves 9, 50A, 50B are controlled based on the speed signal and the pressure signal. Also, it is possible to apply a damping force that quickly removes speed fluctuations and pressure fluctuations.

In the first and second embodiments, the aperture signal calculator 27 calculates the addition signals KH3,
The target opening amount A is calculated from KS3, and a control signal A ′ corresponding to the target opening amount A is output to the electromagnetic proportional valve 9, but as shown in FIG. 12, which is a modification of the first embodiment. In addition, the conversion signals 30A and 30B determine the addition signal K
H3 and KS3 may be directly converted into a control signal A 'corresponding to the target opening amount A. Accordingly, there is no need to calculate the target opening amount A, and the configuration of the controller 12 is simplified.

Further, in the above-described embodiment, an example has been described in which the present invention is applied to the hydraulic circuit for driving the turning hydraulic motor 2 for driving the upper turning body 62 of the crane. However, the present invention is not limited to this. Instead, the present invention can be applied to any circuit using an actuator driven by hydraulic pressure. Furthermore, in a swivel device or the like having a swiveling direction control valve 1 having a flow control characteristic such that a flow rate of a control valve stroke does not depend on a load, such as a load compensating function, the swirling speed and pressure may fluctuate. Although easy, the present invention is suitable for such a turning device, and the effect in that case can be greatly expected.

The present invention can also compensate for the damping force in a so-called reverse lever, in which the state of the hoisting operation is immediately switched to the state of the lowering operation via the neutral position. Further, an on-off type electromagnetic valve is provided instead of the electromagnetic proportional valve 9, and the opening / closing time of the electromagnetic valve is controlled to control the bypass amount of the pressure oil from the pipe 6A (6B) to the pipe 6B (6A). You may do so. Furthermore, the line 6A (6B) on the high pressure side is connected to the line 6B on the low pressure side.
Instead of bypassing the pressurized oil to (6A), the pressurized oil may be directly returned to the tank from the high-pressure side pipeline 6A (6B).

In the correspondence between the above-described embodiment and the claims, the hydraulic motor for turning 2 is an actuator, the pressure sensors 10A and 10B are pressure detectors, the speed sensor 11 is a speed detector, and the electromagnetic proportional valve 9 is a solenoid valve. Connect the valve device to the controller 1
2, 31 to 33 constitute control means and correction flow rate calculation means, respectively.

[0038]

As described in detail above, according to the first aspect of the present invention, the drive of the control valve is controlled by the corrected flow rate based on the pressure signal and the speed signal. According to this configuration, since the correction flow rate is bypassed from the high-pressure side line to the low-pressure side line by drive control of the valve device, a damping force can be effectively applied to the actuator, and pressure fluctuation and speed fluctuation can be achieved. Can be suppressed promptly. Furthermore, according to the third aspect of the present invention, since the flow rate is controlled by the electromagnetic proportional valve, it is possible to prevent vibration caused by a shock at the time of valve switching. Furthermore, according to the invention of claim 4,
Since the correction flow rate is calculated based on a high-frequency component equal to or more than a predetermined value, it is possible to reliably prevent only high-frequency vibrations that impact the apparatus without having any adverse effect on operator input or the like. Further, according to the invention of claim 5, since the present invention is applied to a flow control type control valve in which vibration is easily generated, vibration can be greatly reduced.

[Brief description of the drawings]

FIG. 1 is a hydraulic circuit diagram according to an embodiment of the present invention.

FIG. 2 is a diagram showing a detailed configuration of a control unit according to the first embodiment.

FIG. 3 is an overall configuration diagram of a crane to which the present invention is applied.

FIG. 4 is a diagram for explaining pressure oscillation.

FIG. 5 is a diagram showing an example of static characteristics of speed and pressure when a hydraulic motor is driven.

FIG. 6 is a diagram illustrating an example of dynamic characteristics of speed and pressure when a hydraulic motor is driven.

FIG. 7 is a diagram showing a state where a pressure signal is filtered.

FIG. 8 is a diagram showing a detailed configuration of a control unit according to the second embodiment.

FIG. 9 is a diagram showing a detailed configuration of a control unit according to the third embodiment.

FIG. 10 is a diagram showing a detailed configuration of a control unit according to a fourth embodiment.

FIG. 11 is a diagram showing another example of the dynamic characteristics of the pressure during the turning drive.

FIG. 12 is a diagram illustrating a detailed configuration of a control unit as a modification of the first embodiment.

FIG. 13 is a circuit diagram of a conventional hydraulic control device.

[Explanation of symbols]

 DESCRIPTION OF SYMBOLS 1 Turning direction control valve 2 Hydraulic motor 3 Hydraulic pump 5, 51 Operating lever 6A, 6B Pipe line 9 Electromagnetic proportional valve 10A, 10B Pressure sensor 11 Speed sensor 12, 31-33 Controller 50A, 50B Electromagnetic proportional pilot valve

 ──────────────────────────────────────────────────の Continuing on the front page (72) Inventor Jun-ichi Narusawa 650, Kandamachi, Tsuchiura-shi, Ibaraki Hitachi Construction Machinery Co., Ltd. Tsuchiura Plant (72) Inventor Kazuhisa Ishida 650, Kandamachi, Tsuchiura-shi, Ibaraki Hitachi Construction Machinery Inside the Tsuchiura Plant, Ltd. (72) Koji Funato, Inventor 650, Kandate-cho, Tsuchiura City, Ibaraki Prefecture Inside the Tsuchiura Plant, Hitachi Construction Machinery Co., Ltd.

Claims (6)

[Claims] A hydraulic pump; an actuator driven by hydraulic oil discharged from the hydraulic pump; a control valve for controlling a flow of hydraulic oil supplied from the hydraulic pump to the actuator; and a control valve for switching the control valve. A hydraulic pressure control device comprising: an operation lever for commanding a target flow rate; a pressure detector for detecting pressures of pressure oils of two pipelines respectively connected to the inlet / outlet ports of the actuator and outputting a pressure signal; A speed detector that detects a driving speed of the actuator; a correction flow rate calculation that calculates a correction flow rate based on a pressure signal detected by each of the pressure detectors and a speed signal detected by the speed detector. Means for controlling the control valve with a flow rate obtained by subtracting the correction flow rate from the target flow rate. Apparatus. 2. A hydraulic pump, an actuator driven by hydraulic oil discharged from the hydraulic pump, a control valve for controlling a flow of hydraulic oil supplied from the hydraulic pump to the actuator, and switching the control valve. A hydraulic pressure control device comprising: an operation lever; a pressure detector that detects a pressure of pressure oil in two pipes respectively connected to an entrance port of the actuator and outputs a pressure signal; and a driving speed of the actuator. A speed detector for detecting the pressure signal, a pressure signal detected by each of the pressure detectors, and a corrected flow rate calculating means for calculating a corrected flow rate based on the speed signal detected by the speed detector. A valve device inserted into a conduit communicating with the conduit to regulate a flow rate flowing through the conduit; and Bypass flow supplied to the low pressure side of the conduit is such that a the calculated corrected flow rate, the hydraulic control apparatus characterized by comprising a control means for controlling the valve device. 3. The hydraulic control device according to claim 2, wherein the valve device is an electromagnetic proportional valve whose opening amount is controlled based on the corrected flow rate. 4. The correction flow rate calculating means according to claim 1, wherein: a frequency component of the pressure signal detected by each of the pressure detectors is equal to or more than a predetermined value; The hydraulic control device according to any one of claims 1 to 3, wherein the correction flow rate is calculated based on the following formula. 5. A hydraulic pump, an actuator driven by hydraulic oil discharged from the hydraulic pump, a control valve for controlling a flow of hydraulic oil supplied from the hydraulic pump to the actuator, and switching the control valve. A hydraulic pressure control device comprising: an operation lever; a pressure detector that detects a pressure of pressure oil in each of two pipes connected to an entrance port of the actuator and outputs a pressure signal; Correction flow rate calculation means for calculating a correction flow rate based on the pressure signal detected by the valve; a valve device inserted into a pipe connecting the two pipes to adjust a flow rate flowing through the pipes; and the actuator The valve device is controlled so that the bypass flow rate supplied from the high pressure side pipe line to the low pressure side pipe line by bypassing the calculated correction flow rate becomes equal to the calculated correction flow rate. Hydraulic control apparatus characterized by comprising a control means. 6. The hydraulic control apparatus according to claim 1, wherein the control valve is a flow control type control valve having a flow controllability independent of a load.