JPH06107173A - Vibration-controller of rolling stock - Google Patents

Abstract

PURPOSE:To provide a vibration-controller of rolling stock suited to control over a vibration being produced in a car body. CONSTITUTION:Using a double-acting type pneumatic cylinder 21 as a fluid actuator set up in space between a car body 1 and a truck 2, four proportional flow control valves 22 to 25 driving this cylinder 21, a longitudinal accelerometer 13 as a detector gage detecting a vibration in the car body, a displacement gage 27 and a pressure gage 28, this vibro-controller consists of a multivariable digital controller determining the extent of control input into each control valve from output of the detector gage, having such a function as actively controlling any vibration being produced in the car body. In addition, it adds a high-pass filter 31 for reducing any effect of centrifugal acceleration to be produced in the car body at the time of curve passing, and frequency dependency is imparted to the controller, and it is formed into such a structure as checking the control input into the control valve at a low frequency area where the effect of the centrifugal acceleration is large enough. With this constitution, even in time of the curve passing, vibro-controlling performance almost equivalent to a straight-line passing time is maintainable.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a vibration control device for a railway vehicle, which is suitable for suppressing the vibration generated in the vehicle body of the railway vehicle.

[0002]

2. Description of the Related Art As a method of suppressing the vibration generated in a railroad vehicle, a fluid actuator is installed between a vehicle body and a bogie in accordance with the vibration direction, and a control force having an opposite phase to the vibration of the vehicle body is generated. , For example, JP-A-56-1775
No. 4 “Vehicle vibration control device” and the like are known.

On the other hand, when a railway vehicle passes a curve,
As shown in FIG. 12, since the centrifugal acceleration Ac is applied to the vehicle body, it is necessary to provide a means for reducing this influence and preventing the vibration control capability from decreasing.

An example of such a railway vehicle vibration control device is the "vehicle vibration control device" of Japanese Patent Application Laid-Open No. 57-11163. Examples thereof are shown in FIGS. 10 and 11. This device suppresses vibration by driving a fluid actuator 5 installed between the vehicle body 1 and the bogie 2 with a servo valve 6.
The control input to the servo valve 6 is determined as follows. First, the control input for vibration is determined from the output of the vertical accelerometer 8 installed on the vehicle body 1 using the vibration controller 9 including the integrator 10, the phase advance element 11, and the gain element 12.

On the other hand, from the output of the lateral accelerometer 13 installed on the vehicle body 1, the frequency 0.3 corresponding to the centrifugal acceleration is obtained.
Low-pass filter 1 for detecting acceleration components below Hz
5. Determine the control input for the centrifugal acceleration by using the centrifugal acceleration compensator 14 including the phase lead element 16 for compensating the phase delay in the low pass filter 15 and the gain element 17. By configuring the vibration control device as described above, it is possible to simultaneously reduce the vibration that occurs when passing through a curve.

[0006]

As described above, even when the conventional vibration control device is used, when a curve is passed through at a high speed or when a pneumatic actuator having a slow response is used as the fluid actuator, as shown in FIG. As shown, centrifugal acceleration remains at the entrance and exit of the curve. Therefore, the centrifugal acceleration is included in the acceleration input to the vibration controller, and the vibration control performance is deteriorated.

The present invention provides a vibration control device for a railway vehicle in which the influence of centrifugal acceleration generated on a vehicle body when passing a curve is reduced in order to improve the drawbacks found in the above-mentioned prior art.

[0008]

In order to achieve the above object, a vibration control device for a railway vehicle according to the present invention is a fluid actuator installed between a vehicle body of a railway vehicle and a bogie, a control valve for driving the fluid actuator, and a vehicle body. And a multi-variable digital controller that determines the control input to the control valve from the output of the detector, and further reduces the influence of centrifugal acceleration generated on the vehicle body when passing a curve. A high-pass filter is added and the controller is frequency-dependent so that the control input to the control valve is suppressed in the low frequency range where the influence of centrifugal acceleration is large.

Alternatively, in the vibration control device for a railway vehicle, a high-pass filter for reducing the influence of centrifugal acceleration generated on the vehicle body when passing through a curve is added, and the controller is provided with frequency dependence so that the centrifugal acceleration is increased. In the low frequency range, where the displacement is large, the displacement control is prioritized over the acceleration control.

[0010]

The above-mentioned vibration controller having frequency dependence can be realized as follows by applying the ^H4 control theory which is one of the known control theories. The vibration control system is represented by a feedback control system as shown in FIG. Here, the controlled object 18 is a model of the equation of motion of the railway vehicle, the dynamic characteristics of the fluid actuator, and the like. A trajectory disturbance W is applied to the controlled object, and a vibration acceleration Y ₁ of the vehicle body and a relative displacement Y ₂ between the vehicle body and the bogie are detected. In the feedback control system shown in FIG. 7, the vibration acceleration Y ₁ , the relative displacement Y _2, and the control input U from the orbital disturbance W.
Transfer functions to G ₁ (s), G ₂ (s) and G ₃ (s)
(S is Laplace operator) noted to obtain the H ^■ controller 19 to satisfy the evaluation of the first equation below.

[0011]

[Equation 1]

[0012] Here, ‖ ● ‖ _■ in an amount called H ^■ norm, H ^■ norm ‖X transfer function X (s) (s) || _■ the gain of X (s) │X (jω) │ ( j is an imaginary unit,
ω represents the maximum value of the angular frequency). Therefore, when the first expression holds, the second expression holds for all angular frequencies ω.

[0013] The second equation _{│G 1 (jω) │ <│W} 1 (jω) │ -1 │G 2 (jω) │ <│W 2 (jω) │ -1 │G 3 (jω) │ <│W ₃ (jω) │ ^-1

[0014] As can be seen from the second equation, the transfer function G
₁(S) ~ G₃(S) is a weighting function W₁(S) ~ W₃To (s)
Therefore, it can be shaped on the frequency. Use this fact
Of the low-frequency component of the acceleration, which is equivalent to centrifugal acceleration,
It is possible to design a vibration controller that is less susceptible to noise. For example, billing
To design the vibration controller as shown in item 1, the weight function
A few W₁(S) ~ W₃(S) as in the following third to fifth equations
You can set it to.

[0015]

[Equation 2]

Equation 4 W ₂ (s) = K ₂

[0017]

[Equation 3]

FIG. 8 shows a gain diagram of W ₁ (s) to W ₃ (s). If the weighting function is set in this way, control that keeps the vibration acceleration low near the resonance point of vibration works, and if the frequency is lower than that, the control input is kept low and the control is affected by the low-frequency component of acceleration equivalent to centrifugal acceleration. A difficult vibration controller can be designed. Further, in order to design the vibration controller as set forth in claim 2, the weighting functions W ₁ (s) to W
₃ (s) may be designed as follows.

[0019]

[Equation 4]

[0020]

[Equation 5]

Equation 8 W ₃ (s) = K ₃

FIG. 9 shows a gain diagram of W ₁ (s) to W ₃ (s). If the weighting function is set in this way, control that keeps vibration acceleration low near the resonance point of vibration works, and control that keeps vibration displacement low at lower frequencies gives priority to acceleration equivalent to centrifugal acceleration. It is possible to design a vibration controller that is unlikely to be affected by the low frequency component of.

The vibration controller thus designed is generally a multivariable and high-order controller, but if it is digitized, it can be mounted on a computer in the form of software and control can be easily executed. You can

[0024]

【Example】

Embodiment 1 An example of a vibration control device for a railway vehicle according to the present invention intended to suppress yawing of a vehicle body is shown in FIG.
Shown in. In this embodiment, an air spring 20 is used instead of the spring 3 in the vibration control device shown in FIG. 10, and a double-acting pneumatic cylinder 2 is used as a fluid actuator and a control valve.
1 and proportional flow control valves 22 to 25 are used. Further, a lateral accelerometer 13, a displacement gauge 27 and a pressure gauge 28 are used as detectors.

_Lateral accelerations Y _1f and Y _1r detected by these detectors, relative displacements Y _2f and Y _2r between the vehicle body and the carriage, and pressures P _1f , P _2f and P _1r in the double-acting pneumatic cylinder,
P _2r is converted into a digital value by the A / D converter 29,
It is input to the control computer 30. In the control computer 30, the high-pass filter 31 is first used to remove the centrifugal acceleration. For the high pass filter,
A Butterworth filter with a cutoff frequency of 0.2 Hz was digitized with a control cycle of 10 msec.

Next, the yaw component calculator 32 calculates the yaw acceleration Y _y1 and the yaw relative displacement Y _y2 by the following equation (9).

Formula 9 Y _y1 = (Y _1f -Y _1r ) / 2 Y _y2 = (Y _2f -Y _2r ) / 2

Next, the yaw controller 33 obtains the control pressure command P _y generated by the double-acting pneumatic cylinder 21, and the pressure compensator 34 calculates the following tenth formula from the pressure value detected by the pressure gauge. Is calculated, and the double-acting pneumatic cylinder 2
Compensation inputs U _1f , U _2f , U _1r , U _2r to 1 are obtained.

Formula 10 U _1f = K _P (P ₀ + P _y -P _1f ) U _2f = K _P (P ₀ -P _y -P _2f ) U _1r = K _P (P ₀ -P _y -P _1r ) U _2r = K _P (P ₀ + P _y −P _2r ).

Here, K _P is a pressure compensation gain, and P ₀ is a reference pressure. Finally, the control valve input calculator 35 inputs the inputs to the proportional flow rate control valves 22, 23, 24, 25 to the following eleventh
Determined by the formula.

[0031]

[Equation 6]

However, U _i0 and U _o0 are reference inputs to the air supply valve and the exhaust valve, respectively. The input value to each proportional flow rate control valve is converted into an analog value by the D / A converter 36, and then input to each proportional flow rate control valve.

The design of the yaw controller 33 is such that the transfer functions G _y1 (s), G _y from the trajectory disturbance in the yaw direction to the yaw acceleration Y _y1 , the total yaw displacement Y _y2 and the control pressure command P _y.
y2 (s) and G _y3 to (s) performed with H ^■ control design algorithm to satisfy the evaluation of the equation (12), and digitized in the control period 10 msec.

[0034]

[Equation 7]

Here, the weighting functions W _y1 (s) to W _y3 (s)
In consideration of the fact that the yaw resonance frequency of the device of the present invention is about 1 Hz, and in order to see the effect of the invention of claim 1, in the case of the weight 1 shown in the thirteenth equation, the fourteenth equation, and the fifteenth equation, Then, two kinds of setting in the case of the weight 2 shown in the 16th, 17th and 18th equations were performed.

[0036]

[Equation 8]

Equation 14 W _y2 (s) = 1.0

Fifteenth Expression W _y3 (s) = 0.002

[0039]

[Equation 9]

Expression 17 W _y2 (s) = 1.0

[0041]

[Equation 10]

As can be seen from the above equation, W _y1 (s) is set so that the effect of acceleration control is maximized at the resonance frequency of yawing, and W _y3 (s) at weight 2 is set.
Is designed to keep the control pressure command low below the resonance frequency of yawing.

FIG. 2 is a comparison of the gain diagrams of the transfer function from the yaw acceleration Y _y1 to the control pressure command P _y in the yaw controller when the weight is 1 for G _a and the weight is 2 for G _b. Indicates. By using the weight 2, it can be seen that the control pressure command is suppressed low at 1 Hz or less.

FIG. 3 shows the waveform of the yaw acceleration when a curve passage experiment is conducted by the vibration control device of this embodiment. When a yaw controller designed with a weight of 1 is used, the vibration control performance is deteriorated due to the influence of the centrifugal acceleration generated at the entrance and the exit of the curve. On the other hand, when the yaw controller designed with a weight of 2 is used, the vibration control performance is less deteriorated.

Embodiment 2 As another embodiment of the present invention, FIG. 4 shows a vibration control device for suppressing lateral movement and rolling of a vehicle body.
In this embodiment, a vertical accelerometer 8 is further added as a detector to the device of the first embodiment, and the accelerations Y _3f , Y _4f , Y ₃ r and Y _4r detected by these are _digitalized by the A / D converter 29. Is input to the control computer 30.

In the control computer, instead of the yaw component calculator 32 of the first embodiment, left-right movement. Using the roll component calculator 37, the lateral movement acceleration Y _r1 , the lateral movement relative displacement Y _r2, and the roll acceleration Y _r3 are obtained by the following calculations.

Formula 19 Y _r1 = (Y _1f + Y _1r ) / 2- (h / b) Y _r3 Y _r2 = (Y _2f + Y _2r ) / 2 Y _r3 = (Y _3f -Y ₄ f + Y _3r -Y _4r ) / 4

However, h is the distance from the center of gravity of the vehicle body to the lateral accelerometer, and b is half the distance between the vertical accelerometers.
Is.

Next, the yaw controller 33 of the first embodiment is moved in the left / right direction. Using the roll controller 38, the control pressure command P _r generated by the double-acting pneumatic cylinder 21 is obtained, and the calculation in the pressure compensator 34 is changed as follows.

[0050] The 20 equation _{U 1f = K P (P o} + P r -P 1f) U 2f = K P (P o -P r -P 2f) U 1r = K P (P o + P r -P 1r) U _2r = K _P (P _o −P _r −P _2r )

The design of the left-right movement / roll controller 38 is as follows.
From the trajectory disturbance in the left / right direction, the left / right acceleration Y _r1 , the left / right relative displacement Y _r2 , the roll acceleration Yr _3, and the control pressure command P _r.
The transfer function G _r1 up _{(s), G r2 (s} ), digitized with a control period 10msec with H ^■ control design algorithm to satisfy the following evaluation to G _r3 (s) and G _r4 (s) I went.

[0052]

[Equation 11]

Here, the weighting functions W _r1 (s) to W _r4 (s)
Has a resonance frequency of about 0.5H for the lower-center rolling of this device (low-frequency vibration due to the combined vibration of left-right movement and rolling).
In view of the effect of the invention of claim 2 of this application in consideration of z, the case of weight 3 shown in the following formulas 22 and 23 and the formulas 24, 25, 26 Two settings were made for the case of weight 4 shown in.

[0054]

[Equation 12]

Formula 23 W _r2 (s) = 1.0 W _r3 (s) = 0.3 W _r4 (s) = 0.002

[0056]

[Equation 13]

[0057]

[Equation 14]

Expression 26 W _r3 (s) = 0.3 W _r4 (s) = 0.002

As can be seen from the equations representing the weights 3 and 4, W _r1 (s) is set so that the effect of acceleration control is maximized at the resonance frequency of the lower center rolling, and W _{r2 at} weight 4 is set. (S) is set so that the displacement control is prioritized at the resonance frequency of the lower center rolling or lower. FIG. 5 shows a gain diagram of the transfer function from the lateral motion acceleration Yr ₁ to the control pressure command P _r in the lateral motion / roll controller by using G _{c in} the case of weight 3 and G _{d in} the case of weight 4. A comparison is shown. 0.5H by using a weight of 4
As a result of prioritizing the displacement control below z, it can be seen that the control pressure command is kept low.

FIG. 6 shows the waveform of the lateral motion acceleration when a curve passing experiment is performed by the vibration control device of this embodiment. When a left / right motion / roll controller designed with a weight of 3 is used, the vibration control capability is affected by the centrifugal acceleration generated by the curve. On the other hand, when the left / right motion / roll controller designed with a weight of 4 is used, the vibration control performance is less deteriorated.

[0061]

According to the present invention, a high-pass filter is installed to remove centrifugal acceleration, and in order to reduce the influence on the vibration control of the centrifugal acceleration component that cannot be removed by the high-pass filter, the vibration controller has frequency dependence. As a result, the control input is kept low in the low frequency range, or the displacement control is prioritized over the acceleration control. As a result, vibrations that can maintain the same vibration control performance when passing a curve as when passing a straight line A control device can be provided.

[Brief description of drawings]

FIG. 1 is an explanatory diagram of a railway vehicle vibration control device according to a first embodiment of the present invention.

FIG. 2 is a gain diagram of the yaw controller used in the first embodiment of the present invention.

3A and 3B are graphs showing the time waveform of the yaw acceleration obtained in the curve passing experiment by the vibration control device according to the first embodiment of the present invention. FIG. 3A shows the shape of the curve, and FIG. 3B shows weight 1 and weight 2. The yaw acceleration waveforms with and without the designed yaw controller are shown.

FIG. 4 is an explanatory diagram of a railway vehicle vibration control device according to a second embodiment of the present invention.

FIG. 5 is a gain diagram of the lateral movement / roll controller used in Embodiment 2 of the present invention.

6A and 6B are graphs showing the time waveform of the lateral acceleration obtained by the curve passing experiment by the vibration control device according to the second embodiment of the present invention, FIG. 6A showing the shape of the curve, and FIG. Waveforms for the case of using the left-right movement / roll controller designed in 1. and without control are shown.

FIG. 7 is a block diagram of a vibration control system.

[8] The vibration controller of the claims 1 of the present invention H ^■
It is a gain diagram of a weighting function when designing using a control theory.

[9] The vibration controller of the claims 2 of the present invention H ^■
It is a gain diagram of a weighting function when designing using a control theory.

FIG. 10 is an explanatory diagram showing an example of a conventional vibration control device for a railway vehicle.

11 is a block diagram of the vibration control device shown in FIG.

12A and 12B are explanatory diagrams of centrifugal acceleration generated in a vehicle body when passing through a curve. FIG. 12A shows a curved shape, FIG. 12B shows centrifugal acceleration, and FIG. 12C shows centrifugal acceleration A generated in the vehicle body on a curve.
c is shown.

FIG. 13 is a graph showing an example in which the centrifugal acceleration generated in the vehicle body when passing through a curve cannot be completely compensated, FIG. 13a shows the shape of the curve, and FIG. 13b shows the centrifugal acceleration before and after compensation.

[Explanation of symbols]

1 vehicle body 2 bogie 3 spring 4 oil damper 5 fluid actuator 6 servo valve 7 fluid source 8 vertical accelerometer 9 vibration controller 10 integrator 11, 16 phase advance element 12, 17 gain element 13 lateral accelerometer 14 centrifugal acceleration compensation 15 Low-pass filter 18 Control target of vibration control system 19 H ^■ Controller 20 Air spring 21 Double-acting pneumatic cylinder 22-25 Proportional flow control valve 26 Air source 27 Displacement meter 28 Pressure gauge 29 A / D converter 30 For control Computer 31 High-pass filter 32 Yaw component calculator 33 Yaw controller 34 Pressure compensator 35 Control input calculator 36 D / A converter 37 Left / right movement / roll component calculator 38 Left / right movement / roll controller

Claims (2)

[Claims] 1. A fluid actuator installed between a vehicle body and a bogie of a railway vehicle, a control valve for driving the fluid actuator, a detector for detecting vibration of the vehicle body, and a control input from the output of the detector to the control valve. In a vibration control device for a railway vehicle, which is composed of a multi-variable digital controller that determines the vibration and has a function of actively controlling the vibration generated in the vehicle body, a high-pass for reducing the influence of centrifugal acceleration generated in the vehicle body when passing a curve. A vibration control device for a railroad vehicle, characterized by having a structure in which a filter is added and the controller has frequency dependence to suppress the control input to the control valve in a low frequency region where the influence of centrifugal acceleration is large. . 2. A fluid actuator installed between a car body and a bogie of a railway vehicle, a control valve for driving the fluid actuator, a detector for detecting vibration of the car body, and a control input from the output of the detector to the control valve. In a vibration control device for a railway vehicle, which is composed of a multi-variable digital controller that determines the vibration and has a function of actively controlling the vibration generated in the vehicle body, a high-pass for reducing the influence of centrifugal acceleration generated in the vehicle body when passing a curve. A vibration control device for a railway vehicle, characterized in that a filter is added and the controller has frequency dependency so that displacement control is prioritized over acceleration control in a low frequency range where centrifugal acceleration has a great influence.