JPH06107174A - Vibration controller of rolling stock - Google Patents

Abstract

PURPOSE:To provide a vibration controller of rolling stock that is able to do good vibration control even to a vibration of medium frequency consisting in several hertzes after compensating any operation delay in a proportional control valve. CONSTITUTION:This is a vibrocontroller of rolling stock, consisting of a fluid actuator installed in space between a car body and a truck, a proportional flow control valve driving this fluid actuator, a detector gage detecting a vibration in the car body, and a controller determining the extent of control input into the proportional control valve from output of this detector gage, and having such a function as actively controlling any vibration being produced in the car body. In addition, it is added with a pressure compensator 23 which detects the extent of pressure in the fluid actuator 3, and compensates for a drop in vibro-controlling performance attributable to an operation delay in the proportion control valve by its detection output. With this constitution, good vibration control over rolling stock can be done with such a proportional control valve as more inexpensive than a servo valve.

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 for suppressing the vibration generated in the car body of a railroad vehicle, a fluid actuator is installed between the car body and a bogie in accordance with the vibration direction, and a control force having an opposite phase to the vibration of the car body is exerted. It is common to let them do it. As a conventional vibration control device for a railway vehicle, Japanese Patent Laid-Open No. 56-1775
No. 4 “Vehicle vibration control device” and the like are known. An example thereof is shown in FIG. A fluid actuator 3 installed between the vehicle body 1 and the carriage 4 is driven by a servo valve 5. The control input to the servo valve 5 is performed by a controller 7 including an integrator 9, a phase compensation element 10, and a gain element 11 shown in FIG. 13 by using the output of the acceleration detector 6 installed in the vehicle body 1. It is determined. The transfer function of the controller 7 is expressed by the following first equation. However, in the formula, s is a Laplace operator, K is a gain, and T _{1 to} T ₄ are time constants for phase compensation.

[0003]

[Equation 1]

The servo valve 5 has a structure as shown in FIG. When the torque motor 12 rotates due to the input current,
The back pressure of the nozzle 14 increases or decreases due to the movement of the flapper 13,
The spool 15 moves. On the other hand, the displacement of the spool 15 is changed by the feedback motor 16 through the torque motor 12
As a result, the spool 15 eventually moves in proportion to the input current to control the supply and discharge of fluid.

As described above, since the servo valve has a servo mechanism inside the valve, it is possible to perform pressure control and flow rate control with high accuracy and high speed. However, the servo valve is processed with high precision, has a high sensitivity to contamination, and has a complicated structure, so that it is expensive and requires a great deal of effort to maintain its performance.

In contrast to the servo valve described above, there are proportional control valves such as the proportional flow rate control valve shown in FIG. 15 and the proportional pressure control valve shown in FIG. These proportional control valves are relatively simple in structure and inexpensive in price. However, the proportional control valve is
Since the response of the proportional solenoid 20 that drives the motor 5 is not so high, the response time of the valve is 0.5 to 10 ms of the servo valve.
It is slower than 30 to 100 ms. Therefore, as compared with the case where the servo valve is used, the controllability with respect to the vibration of the medium frequency of about 3 to 10 Hz is deteriorated.

[0007]

As described above, the servo valve can perform pressure control and flow rate control with high accuracy and high speed.
The structure is complicated and economically disadvantageous. On the other hand, the proportional control valve is advantageous in terms of economical efficiency and maintainability, but there is a problem in that the controllability with respect to the vibration of the medium frequency due to the delay of the operation of the valve is deteriorated.

In view of the above situation, the present invention compensates for the operation delay of the proportional control valve and controls the vibration of a railway vehicle equipped with pressure compensation so that good vibration control can be performed even for medium frequency vibrations. A device is provided.

[0009]

To achieve the above object, a vibration control device for a railway vehicle according to the present invention comprises a fluid actuator installed between a vehicle body and a bogie, a proportional control valve for driving the fluid actuator, and the vehicle body. Of a railway vehicle having a function of actively controlling the vibration generated in the vehicle body, which is composed of a detector that detects the vibration of the vehicle and a controller that determines the control input to the proportional control valve from the output of the detector. In the control device, a pressure compensator is added to detect the pressure in the fluid actuator and compensate the decrease in the vibration control performance due to the operation delay of the proportional control valve or the like by the detection output.

[0010]

The concept of pressure compensation in the present invention will be described with reference to FIG. The proportional control valve 21 attached to the fluid actuator 3 controls the inflow / outflow flow rate of the fluid to / from the fluid actuator 3 so that the control force based on the differential pressure ΔP between the internal pressures P ₁ and P ₂ of the fluid actuator 3. Generate. At this time, without pressure compensation, the internal pressures P ₁ and P ₂ of the fluid actuator 3 do not follow the pressure commands P _R2 and P _R2 due to the operation delay of the proportional control valve 21.

Therefore, the pressure gauge 22 is used to detect the internal pressures P ₁ and P ₂ of the fluid actuator 3, and the pressure command P
Deviations P _E1 and P _E2 from _R1 and P _R2 are obtained. Then, the pressure compensator 23 determines the compensation inputs U ₁ and U ₂ to the proportional control valve 21 so that this deviation becomes zero. In this way, the operation delay of the proportional control valve 21 can be compensated.

In designing the pressure compensator, it is necessary to satisfy mutually contradictory specifications of improving response and ensuring stability. Here, how to apply the H ^■ control theory is a known control theory as an example for designing the pressure compensator to satisfy the above specifications. When the pressure compensator of FIG. 2 is rewritten into a block diagram relating to the pressure difference ΔP between the internal pressures P ₁ and P ₂ of the fluid actuator 3, a feedback control system as shown in FIG. 3 is obtained.
Here, the control target 24 is a collection of the dynamic characteristics of the fluid actuator 3 and the proportional control valve 21 of FIG.

In FIG. 3, the transfer function of the controlled object 24, that is, the transfer function from the compensation input U to the differential pressure ΔP of the internal pressure of the fluid actuator is G (s), and the transfer function of the pressure compensator 23, that is, the pressure command ΔP. _Let K (s) be the transfer function from the control deviation ΔP _E between _R and the pressure difference ΔP of the internal pressure of the fluid actuator to the compensating input U, and the sensitivity S (s) from the following second equation is complemented by the third equation. Define the sensitivity T (s).

[0014]

[Equation 2]

[0015]

[Equation 3]

The sensitivity S (s) is a transfer function from the pressure command ΔP _R to the control deviation ΔP _E , and the smaller this is, the better the response to the pressure command is. On the other hand, complementary sensitivity T (s)
Is a transfer function from the pressure command ΔP _R to the differential pressure ΔP, and the smaller this is, the more stable the feedback control system of FIG. 3 is to disturbance such as noise. However, the sensitivity S (s) and the complementary sensitivity T (s) have the relationship shown in the following fourth equation, and therefore both cannot be reduced at the same time.

Equation 4 S (s) + T (s) = 1

When it is noted that the response is a problem in the low frequency region and the stability is a problem in the high frequency region, the sensitivity S is as follows.
It is possible to reduce (s) and complementary sensitivity T (s). At low frequencies, | S (jω) | is small, and | T (jω) |
Is almost 1 at high frequencies, | S (jω) | is almost 1, and | T (jω) |
Is small, where ω is the angular frequency and | A (jω) | is the transfer function A
It is the gain of (s). However, A represents S or T. The above specification, by introducing what is referred to as H ^■ norm ‖ ● ‖ _■, is formulated as shown in Equation 5 below.

[0019]

[Equation 4]

H of transfer function A (s) ^■ norm ‖A (s)
‖ _■ the gain | A (jω) | Since corresponding to the maximum values for the omega, when the fifth equation is satisfied, in all of the angular frequency, which consists sixth equation of the following.

[0021]

[Equation 5]

Therefore, the weighting functions W _S (s) and W _T
(S) is set to have a gain characteristic shown in FIG. 4, by obtaining the pressure compensator K that satisfies the Equation 5 (s) is at H ^■ control design algorithm, the sensitivity S (s) and the complementary sensitivity T (S) is as shown in FIG. 5, and it can be seen that a pressure compensator satisfying the above specifications could be designed. As mentioned above,
By compensating the operation delay of the proportional control valve, the pressure in the fluid actuator can be controlled with high accuracy and at high speed.

[0023]

【Example】

Embodiment 1 FIG. 1 shows an example in which a vibration control device for a railway vehicle according to the present invention is implemented for the purpose of suppressing yawing that occurs around the center of a vehicle body as shown in FIG. In this embodiment, an air spring 25 is used instead of the spring 2 in the vibration control device shown in FIG. 12, and a double-acting pneumatic cylinder 26 and proportional flow rate control valves 27 to 30 are used for the fluid actuator and the proportional control valve. There is. The subscript f attached to the reference numeral in the drawing indicates the front bogie side, and the r indicates the rear bogie side.
Further, in order to reduce the load on the fluid actuator, an oil damper 32 is provided in parallel with the double-acting pneumatic cylinder 26, and a lateral accelerometer 6 and a relative displacement meter 33 are provided as detectors.
And the pressure gauge 22 is used.

The lateral accelerations Y _1f , Y _1r , relative displacements Y _2f , Y _2r between the vehicle body and the bogie and the pressures P _1f , P _2f , P _1r , P _2r detected by the above-mentioned detector are A / D converters. It is converted into a digital value at 34 and input to the control computer 35. In the control computer 35, the yaw component calculator 36 first determines the yaw acceleration Y _y1 and the yaw relative displacement Y _y2 by the following seventh equation.

Formula 7 Y _y1 = (Y _1f -Y _1r ) / 2 Y _y2 = (Y _2f -Y _2r ) / 2

Next, the yaw controller 37 obtains the control pressure command P _y generated by the double-acting pneumatic cylinder 26, and the pressure compensator 23 calculates the double-acting pneumatic pressure from the pressure value detected by the pressure gauge. Compensation inputs U _1f , U _2f to cylinder 26,
U _1r and U _2r are _calculated by the following eighth formula.

Eq. 8 U _1f = K (s) (P ₀ + P _y −P _1f ) U _2f = K (s) (P ₀ −P _y −P _2f ) U _1r = K (s) (P ₀ − _{P y -P 1r) U 2r =} K (s) (P 0 + P y -P 2r)

Here, K (s) is the transfer function of the pressure compensator 23, and P ₀ is the reference pressure. Finally, the control valve input calculator 38 determines the input to each proportional flow rate control valve 27, 28, 29, 30 by the following formula (9).

[0029]

[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 39, and then input to each proportional flow rate control valve.

[0031] applying the H ^■ control theory to the design of the yaw controller 37. The control system shown in FIG.
Let G (s) be the transfer function from to the yaw acceleration Y _y1 .
Next, determine the yaw controller that satisfies the tenth equation with H ^■ control design algorithm. Note that W (s) in the tenth formula
Is a weighting function, and a function such as the following Expression 11 was used.

[0032]

[Equation 7]

[0033]

[Equation 8]

At this time, it is known that the tenth expression is equivalent to the following twelfth expression, and it is understood that G (s) can be defined by W (s).

[0035]

[Equation 9]

On the other hand, the pressure compensator is a proportional gain element shown in the following thirteenth expression.

Equation 13 K (s) = K _P

Pressure step response R without pressure compensation
FIG. 8 shows a comparison of the pressure step response R _B when _A and the pressure compensation (K _P = 1) are performed. From FIG. 8, it can be seen that the response is improved by the pressure compensation. Further, the gain diagram of the transfer function from the trajectory disturbance to the yaw acceleration is shown in FIG. 9 in comparison with G _O without control, G _A with control (without pressure compensation) and G _B with control (with pressure compensation). . It can be seen from FIG. 9 that the response of the control valve is improved by the pressure compensation, so that the control effect appears over a wider frequency range.

The design of the pressure compensator K (s) in Example 1, illustrating the case of applying the method of applying the H ^■ control theory described above. The control target in FIG. 3 is represented by the transfer function G _P (s) of the following fourteenth expression in which the dead time is added to the second-order delay in the actual target. On the other hand, in the design, a model represented by the transfer function G _M (s) of the 15th equation with no dead time is used.

[0040]

[Equation 10]

[0041]

[Equation 11]

Therefore, the designed pressure compensator K
(S) must operate sufficiently stably even with respect to G _P (s) of the 14th equation. In order to achieve this goal, the weight W _T (s) for the complementary sensitivity in Equation 5 may be determined as follows. That is, when the deviation Δ (s) between the model and the actual object is defined as the 16th equation, | Δ (j
ω) | <| W _T (jω) |, and W _T (s) · G _M
W _T (s) may be determined so that (s) is proper (the denominator of the transfer function and the order of the numerator are the same). Specifically, as W _T (s), a quadratic lead like the following Expression 17 is used. Further, the first-order lag of the 18th equation is used as the weight W _S (s) for the sensitivity function in the 5th equation.

[0043]

[Equation 12]

[0044]

[Equation 13]

[0045]

[Equation 14]

[0046] Using these weights, the pressure compensator K satisfying Equation 5 (s) is determined by H ^■ control design algorithm. _Let this be K _H. FIG. 10 compares the pressure step response R _B when using the proportional gain K _{P of the} first embodiment and the pressure step response R _C when using the above K _H. Figure 10
Therefore, compared with the pressure step response R _B when the proportional gain K _P is used, the pressure step response R _C when K _H is used is
It can be seen that the response is faster and more stable. Also, the gain diagram of a transfer function from track disturbance to the yaw acceleration, no control G _O, CONTROLLED (pressure compensation K _P) G _B, control There FIG compared with the (pressure compensation K _H) G _C 11 Shown in. From this result, it is understood that the control effect appears over a wider frequency range due to the improved response.

[0047]

According to the present invention, by adding the pressure compensator, it is possible to compensate for the deterioration of the vibration control performance due to the operation delay of the proportional control valve, and to use the proportional control valve which is cheaper than the servo valve. Therefore, good vibration control of railway vehicles can be performed.

[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 an explanatory view showing the concept of pressure compensation in the present invention.

FIG. 3 is a block diagram showing the pressure compensation of FIG. 2 as a feedback control system related to a pressure difference.

FIG. 4 is a gain diagram of a weighting function for sensitivity and complementary sensitivity.

FIG. 5 is a gain diagram of sensitivity and complementary sensitivity.

FIG. 6 is an explanatory diagram of yawing generated on the vehicle body.

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

8 is a graph comparing with the case R _B in the case of no pressure compensation pressure step response R _A and there pressure compensation (proportional gain K _P).

[9] In gain diagram of a transfer function from track disturbance to the yaw acceleration, the case of no control G _O, if G _A control Yes (without pressure compensation), the control there (pressure compensation K _P) when G _B It shows by comparing with.

10 is a graph comparing with the case R _C of Yes pressure compensating pressure step response (proportional gain K _P) there when R _B and pressure equalization (H ^■ compensator K _H).

FIG. 11 is a gain diagram of a transfer function from an orbital disturbance to a yaw acceleration, G _O without control, G _B with control (pressure compensation K _P ), G with control (pressure compensation K _H ). Shown in comparison with _C.

FIG. 12 is an explanatory diagram showing an example of a conventional vibration control device for a railway vehicle.

FIG. 13 is an explanatory view showing an example of a controller in a conventional vibration control device for a railway vehicle.

FIG. 14 is a sectional view showing a main part of a servo valve.

FIG. 15 is a sectional view showing a main part of a proportional flow rate control valve.

FIG. 16 is a sectional view showing a main part of a proportional pressure control valve.

[Explanation of symbols]

 1 Body 2 Spring 3 Fluid Actuator 4 Bogie 5 Servo Valve 6 Accelerometer 7 Controller 8 Fluid Source 9 Integrator 10 Phase Compensation Element 11 Gain Element 12 Torque Motor 13 Flapper 14 Nozzle 15 Spool 16 Feedback Spring 17 Supply Port 18 Discharge Port 19 Load Port 20 Proportional Solenoid 21 Proportional Control Valve 22 Pressure Gauge 23 Pressure Compensator 24 Controlled Object 25 Air Spring 26 Double Acting Pneumatic Cylinder 27, 28, 29, 30 Proportional Flow Control Valve 31 Air Source 32 Oil Damper 33 Displacement Meter 34 A / D converter 35 Control computer 36 Yaw component calculator 37 Yaw controller 38 Control valve input calculator 39 D / A converter Subscript f Front bogie side Subscript r Rear bogie side

Claims (1)

[Claims] 1. A fluid actuator installed between a car body and a bogie of a railway vehicle, a proportional control valve for driving the fluid actuator, a detector for detecting vibration of the car body, and an output from the detector to the proportional control valve. In a vibration control device for a railway vehicle having a function of actively controlling the vibration generated in the vehicle body, the vibration control device comprising a controller for determining a control input, detects pressure in the fluid actuator, and outputs the proportionality by the detection output. A vibration control device for a railway vehicle, characterized in that a pressure compensator is added to compensate for a decrease in vibration control performance caused by a control valve operation delay.