JP4191013B2 - Damping device and vehicle with damping function - Google Patents

Description

  The present invention relates to a vehicle having a function of suppressing shaking, and more particularly, to a vibration control device that suppresses vertical vertical vibration generated in the vehicle and a vehicle with a vibration control function having such a vibration control function.

  In recent years, with the speeding up of railway vehicles, various ideas have been made to improve riding comfort. As a factor that affects the ride comfort, the vehicle shakes. This vehicle shake can be broadly divided into longitudinal vibration that occurs in the longitudinal direction of the vehicle and lateral vibration that occurs in the lateral direction.

  Of these, many countermeasures have been conventionally taken to reduce the lateral vibration, and a lateral vibration damping damper and a damping system for a railway vehicle are known.

  However, due to the recent increase in the speed of railways and the accompanying reduction in weight of vehicles, the elastic vibration of the vehicle body in the vertical direction, which seems to be caused by a decrease in rigidity, has become a problem. The elastic vibration (especially the primary bending vibration) of this body is often located near the frequency band of 8 to 11 Hz, and this band is close to the band of 4 to 8 Hz where humans feel the vertical vibration most sensitively. It becomes a factor to deteriorate comfort.

  As a countermeasure against this elastic vibration, it is possible to reduce the coupling force between the carriage and the vehicle body. However, if the suspension is softened in the vertical direction in order to reduce the coupling force, in high-speed running vehicles, the vehicle body leans excessively due to the centrifugal force in the left-right direction when turning a curve, or while traveling in a place with a height difference Since it is predicted that low-frequency excitation due to the centrifugal force in the vertical direction will cause resonance of the air spring between the carriage and the vehicle body, it is not preferable to make the suspension extremely soft.

  Therefore, conventionally, as a countermeasure against vertical vibration, a method of increasing the structural damping of the vehicle body, a method of causing the cart to act as a dynamic damper, and an active suspension or semi-active suspension with a control actuator between the vehicle body and the cart are proposed. Has been.

For example, a “railway vehicle anti-sway control system” disclosed in the following Non-Patent Document 1 is known as one that suppresses the elastic vibration by installing a full active suspension between a vehicle body and a carriage.
"Development of active vibration control system for upper and lower systems", 1998 Railway Technology Union Symposium, Kenjiro Uebayashi and others

FIG. 15 is an image diagram schematically showing a system configuration of the upper and lower active vibration control apparatus disclosed in Non-Patent Document 1. As shown in the figure, in a vehicle composed of a vehicle body 100, a carriage frame 200, and a shaft box 300, an air spring 400 and a hydraulic actuator 410 are installed between the vehicle body and the carriage, and a shaft is provided between the carriage frame and the axle box. A spring 500 is installed. The hydraulic actuator 410 includes a hydraulic unit including a servo motor 411, a hydraulic pump 412, a servo valve 413, a bypass valve 414, and a shut-off valve 415. The vehicle body 100 is provided with a vehicle body vertical accelerometer 600, and the vehicle frame 200 is provided with a vehicle vertical accelerometer 700. The hydraulic unit is configured to control the hydraulic actuator 410 so as to suppress the vertical vibration between the vehicle body and the carriage based on the measurement values of the two accelerometers 600 and 700.

Also, for example, one disclosed in Patent Document 1 below is known as a semi-active suspension installed between a vehicle body and a carriage to suppress vertical elastic vibration and the like.
JP 2003-72544 A

FIG. 16 is a diagram schematically showing a configuration of a bogie part of a railway vehicle having the semi-active vertical suspension system disclosed in Patent Document 1. In FIG. As shown in the figure, a variable damping damper built-in air spring 410 is installed between a vehicle body (not shown) and the carriage frame 200, and a shaft spring 500 is provided between the carriage frame 200 and the axle box 300. The axle box 300 supports an axle 330, and the axle 330 includes a wheel 310 and an axle 320. In addition, at least three acceleration sensors are installed on a vehicle body (not shown). And it is comprised from the output of this sensor to control the damping force of the variable damping damper (built-in air spring 410) so as to suppress the vertical vibration of the vehicle body.

  However, since the vibration damping system disclosed in Non-Patent Document 1 and Patent Document 1 is equipped with a vibration-damping actuator or a variable damping damper between the vehicle body and the carriage, the effect of the vibration control is reduced by the carriage frame. Depending on the vibration state of the vehicle, there may be a case where the vertical vibration of the vehicle body cannot be sufficiently suppressed. In particular, it may be difficult to sufficiently obtain the effect of vibration suppression control at a location where the vibration of the bogie frame becomes large.

  The present invention has been made in view of such problems, and an object thereof is to provide a vibration damping device and a vehicle with a vibration damping function that can sufficiently suppress vertical vibrations of a vehicle body.

In order to solve the above problems, a vehicle with a vibration damping function according to the present invention includes a vehicle body, a plurality of carriage frames that support the vehicle body, and an axle box and a wheel shaft that support the carriage frame. Based on the acceleration sensor provided in each cart frame for detecting the vertical vibration of each cart frame, the variable damping damper provided between the cart frame and the axle box, and the output of each acceleration sensor And a control means for controlling the damping force of each variable damping damper so as to reduce the vertical vibration of each carriage frame .

Further, the vehicle with a vibration damping function according to the present invention is a vehicle including a vehicle body, a plurality of carriage frames that support the vehicle body, an axle box and a wheel shaft that supports the carriage frame, and the vertical direction of each carriage frame. An acceleration sensor provided on each carriage frame for detecting the vibration of the carriage, and between the carriage frame and the axle box for generating a forcible driving force to reduce the vertical vibration of the vehicle body. And a control means for controlling the driving force of each actuator so as to reduce the vertical vibration of each carriage frame based on the output of each acceleration sensor. And

The vibration damping device according to the present invention is a vehicle including a vehicle body, a plurality of carriage frames that support the vehicle body, and an axle box and a wheel shaft that supports the carriage frame, in order to reduce vertical vibrations of the vehicle body. a damping device, an acceleration sensor in which the respectively provided to each bogie frame in order to detect the vibration in the vertical direction of the bogie frame, the bogie frame - a variable damping damper disposed between the axle box And control means for controlling the damping force of each variable damping damper so as to reduce the vertical vibration of each carriage frame based on the output of each acceleration sensor.

The vibration damping device according to the present invention is a vehicle including a vehicle body, a plurality of carriage frames that support the vehicle body, and an axle box and a wheel shaft that supports the carriage frame, in order to reduce vertical vibrations of the vehicle body. of a vibration damping device, wherein an acceleration sensor provided respectively to the each bogie frame in order to detect the vibration in the vertical direction of the truck frame, the vertical vibration of the vehicle body by generating a forcible driving force In order to reduce the driving force of each actuator so as to reduce the vertical vibration of each bogie frame based on the actuator provided between the bogie frame and the axle box and the output of each acceleration sensor. And a control means for controlling.

  According to the vibration damping device and the vehicle with the vibration damping function according to the present invention, it is possible to sufficiently suppress the vertical vibration of the vehicle body.

  Hereinafter, embodiments of the present invention will be described in detail with reference to the drawings. FIG. 1 is a diagram schematically showing a configuration of a railway vehicle according to the present embodiment. As shown in the figure, the railway vehicle is mainly composed of a vehicle body 10 and a carriage 20. The carriage 20 is provided with a front carriage 20-1 and a rear carriage 20-2 before and after the vehicle body 10.

  First, a description will be given of a conventional cart configuration that is the basis of a vehicle according to the present embodiment. In the conventional cart configuration, a shaft damper 59 is installed instead of the position of the variable damping damper 51 in FIG. Each bogie 20 includes a bogie frame 21, an air spring 41 that connects the bogie frame 21 and the vehicle body 10, a wheel shaft 31 that includes wheels and axles, and a shaft box that is a housing of a main bearing that transmits the weight of the vehicle body / bogie to the wheel shaft 31. 30, a shaft spring 50 and a shaft damper 59 for connecting the bogie frame 21 and the shaft box 30 are provided. The axle box 30 is provided with front axle boxes 30-1 and 30-3 on the front side of each carriage 20, and rear axle boxes 30-2 and 30-4 on the rear side. The shaft damper 59 is also provided corresponding to each shaft box 30. Since this figure is a view seen from the right side of the traveling direction, only the axle box 30, the axle spring 50 and the axle damper 59 on the right side of the vehicle are shown, but the axle box and the axle spring are similarly shown on the left side of the vehicle. And a shaft damper. The vehicle body 10 and the bogie frame 21 are also connected by a not-shown yaw damper and traction link. The air spring 41 acts to reduce the longitudinal vibration between the vehicle body and the carriage frame, and the yaw damper acts to suppress the horizontal rotation (yawing) of the carriage 20. Further, the tow link plays a role of transmitting the driving force in the front-rear direction of the carriage 20 to the vehicle body 10. The shaft damper 59 acts to attenuate the vibration in the vertical direction and improve the riding comfort.

  Subsequently, the configuration of the vehicle according to the present embodiment shown in FIG. 1 will be described. As described above, the vehicle is provided with the variable damping damper 51 in place of the shaft damper 59. The variable damping damper 51 has a variable damping force, and the damping force is controlled so as to improve riding comfort, as will be described in detail later. The variable damping damper 51 is installed between the carriage frame 21 and the axle box 30 to reduce the vertical vibrations of the carriage 20 and the vehicle body 10 as a feature of the railway vehicle according to the present embodiment. Yes. Note that conventional railway vehicles include vehicles equipped with the above-described shaft dampers and vehicles that are not. Therefore, when the present invention is applied to a conventional vehicle, the vehicle is replaced with a variable damping damper for a vehicle equipped with a shaft damper, and additionally variable for a vehicle not equipped. A damping damper may be installed.

As the variable damping damper 51, a type using an electromagnetic proportional relief valve, a type using a high-speed switching electromagnetic valve, or the like is used. FIG. 2 is a diagram illustrating a hydraulic circuit when a variable damping damper is configured by a biflow circuit using an electromagnetic proportional relief valve. The throttle valve 53 gives flow resistance to the fluid flowing inside, and the pressure difference before and after the throttle valve 53 is controlled by the electromagnetic proportional relief valve 52. Since the electromagnetic proportional relief valve 52 is a relief valve to the last, the pressure cannot be made higher than the pressure difference generated by the fluid flowing through the throttle valve 53. The damping force of the variable damping damper is obtained by multiplying this pressure difference by the pressure receiving area of the cylinder 55. Specifically, if the pressure and pressure receiving area on the rod side of the cylinder 55 are P ₁ and A ₁ , and the pressure and pressure receiving area on the opposite rod side are P ₂ and A ₂ , the damping force is | A ₂ * P ₂ − A ₁ * P ₁ | The direction of the force is opposite to the extension direction of the damper.

  The two electromagnetic proportional relief valves 52 are connected to a control unit (not shown), the electromagnetic proportional relief valve 52-2 controls the damping force when the variable damping damper contracts, and the electromagnetic proportional relief valve 52-1 The variable damping damper is configured to control a damping force when the damper is extended. The accumulator 54 is used to supply / recover the excess / deficiency fluid when the damper expands / contracts to cause excess / deficiency of the fluid. According to the variable damping damper configured by such a biflow circuit, there is an advantage that the response speed can be improved. When always acting as a one-effect damper, it is possible to adopt a configuration in which the two electromagnetic proportional relief valves 52-1 and 52-2 are controlled by a single solenoid. Specifically, the damping force control characteristic is configured so that the damping force is reversed in the stroke of the extension and contraction of the damper, and when the damping force on the expansion side is variable with a large value, the damping side damping force is changed. When the damping force on the contraction side is variable with a large value, the expansion side damping force is fixed at a small value. As described above, when a valve that can control both damping forces by one solenoid is used, a variable damping damper can be configured at low cost.

  FIG. 3 is a diagram showing a hydraulic circuit when a variable damping damper is configured by a uniflow circuit using an electromagnetic proportional relief valve. The operations of the electromagnetic proportional relief valve 52 and the throttle valve 53 are the same as those of the biflow circuit type shown in FIG. However, in the uniflow circuit type, damping force control of the extension and contraction of the damper is performed by the common electromagnetic proportional relief valve 52. In order to define the direction of force generation, the expansion side unload valve 57-1 and the compression side unload valve 57-2 are used. If the extension side unload valve 57-1 is turned on while the damper is extended, the pressure difference between the rod side and the non-rod side of the cylinder 55 becomes zero, and no damping force is generated. On the other hand, when turned OFF, the pressure difference before and after the throttle valve 53 is controlled by the electromagnetic proportional relief valve 52. The operation of the compression side unload valve 57-2 is the same. The variable damping damper using this uniflow circuit is not suitable when a quick response is required, but the same proportional proportional relief valve is used for both the expansion and contraction strokes. There is an advantage that there is no difference in characteristics and adjustment is easy.

  FIG. 4 is a diagram showing a hydraulic circuit of a variable damping damper configured to control the damping force by changing the combination of the throttle valves 53 by using the high-speed switching electromagnetic valve 56. By switching the three high-speed switching electromagnetic valves 56, six combinations of throttle valves 53 can be made. Therefore, the optimum throttle valve 53 combination is selected on the controller side based on the damper piston speed and the damping force command value to be described later, and the corresponding high-speed switching solenoid valve 56 is driven to perform pressure control. The operation of the unload valve 57 is the same as that of the uniflow type shown in FIG. 4, and the relief valve 52-5 is used to define the maximum damping force. However, this type is disadvantageous in terms of cost because it requires a stroke sensor (not shown) for measuring the stroke of the damper.

  Note that the variable damping damper 51 may be one that exhibits the characteristics of an existing passive damper (shaft damper) when the control unit is powered off (stopped). For example, even if an abnormality occurs in the control system of the variable damping damper 51, it is possible to maintain the same riding comfort as that of the current passive damper by turning off the power of the control unit. The variable damping damper configured by the uniflow circuit or the biflow circuit can be realized by using an electromagnetic proportional relief valve whose relief pressure has a passive characteristic when the valve current is zero. Further, it can be realized by a variable damping damper that is provided with a switching valve that can switch between a damping force variable valve and a passive valve and that is switched to the passive valve when the power is turned off.

  Furthermore, in the railway vehicle according to the present embodiment, at least one of the bogie frame 21 and the vehicle body 10 is provided with an acceleration sensor for measuring the vertical acceleration of the bogie frame 21 or the vehicle body 10. Since the installation position of the acceleration sensor differs depending on the control law for controlling the variable damping damper 51, the installation position of the acceleration sensor will be described together in the description of the control law described later. From the output of the acceleration sensor, the vertical translation mode, pitching mode, primary bending mode, etc. of the bogie frame 21 and the vehicle body 10 can be calculated.

  Here, the vertical translation mode, the pitching mode, and the primary bending mode will be described. FIG. 5 is a diagram conceptually showing the up / down translation mode, the pitching mode, and the primary bending mode, as viewed from the side of the vehicle. FIG. 5A shows a vertical translation (bouncing) mode, which represents a motion parallel to the vertical direction. FIG. 5B shows a pitching mode, in which the front and rear of the vehicle body move up and down in opposite phases. FIG. 5C shows a primary bending mode and shows a primary bending motion with the center of the vehicle body as an abdomen.

  Next, a control law for controlling the variable damping damper from the output of the acceleration sensor in the railway vehicle having the above configuration will be described. In the present embodiment, as an example, there are two types of control law 1 that performs semi-active control based only on information on bogie frame vibration and control law 2 that performs semi-active control based on information on both bogie frame vibration and vehicle body vibration. The simulation was performed by the control law.

  The variable damping damper (1) increases the damper internal pressure in proportion to the square of the fluid flow speed in the damper. (2) Controls the electromagnetic proportional relief valve etc. against this damper internal pressure. Assuming an electromagnetic proportional relief valve type that satisfies the following conditions: (3) The generated force is expressed by the product of the damper internal pressure and the pressure receiving area.

The proportional electromagnetic relief valve, to the command u _rij, the operation u _dij of the valve, it is assumed to have a primary delay characteristic comprising the following formula (1). However, i = 1 (front cart), 2 (rear cart), j = 1 (front shaft), and 2 (rear shaft). The dots in the equation mean first-order differentiation (d / dt) by time t, and the same applies to the following equations. If there are two dots, it means the second derivative (d ² / dt ² ).

The damping force of the variable damping damper rises in proportion to the square of the oil flow speed (piston speed) v _dij and is ideally relieved by the electromagnetic proportional relief valve with the delay of the above equation (1). It is said. Assuming that the opening of the throttle valve is c _d and the inclination of the proportional relief is k _d , the damping force f _dij is expressed by the following equation (2).

(1) Control law 1 (control based only on bogie frame vibration information)
In this method, only the vertical acceleration of the carriage frame is measured, and the vertical vibration control of the carriage frame is performed for each carriage. In this control method, the vertical vibration of the vehicle body is reduced by reducing the vertical vibration of the bogie frame that is a transmission path of the vertical vibration to the vehicle body. Various specific control rules can be used, but in this embodiment, skyhook control is used. Since this control law does not consider the vibration mode of the vehicle body, an acceleration sensor for measuring the vibration of the vehicle body is not required, and the circuit configuration of the control unit is simplified, which is advantageous in cost reduction.

  FIG. 6 is a diagram schematically showing the arrangement of the acceleration sensors 60-1 and 60-2 and the control units 70-1 and 70-2 when the control law 1 is applied. The acceleration sensor 60, the control unit 70, and the variable damping damper 51 constitute a vibration damping device. In the figure, the right side is the traveling direction.

As shown in the figure, acceleration sensors 60-1 and 60-2 are installed in the central portions of the front bogie frame 21-1 and the rear bogie frame 21-2, respectively. The acceleration sensor 60 detects the acceleration d ² z _Ti / dt ² in the vertical direction of the carriage frame. The vehicle body 10 is provided with two control units 70-1 and 70-2 for the front bogie frame 21-1 and the rear bogie frame 21-2. The output of the acceleration sensor 60-1 is sent to the control unit 70-1, and the output of the acceleration sensor 60-2 is sent to the control unit 70-2. The control unit 70-1 calculates a damping force command value u _r1j to the variable damping damper 51 provided between the front bogie frame 21-1 and the axle boxes 30-1 and 30-2 from the output value, Transmit to the variable damping damper 51. Similarly, the control unit 70-2 calculates a damping force command value _ur2j to the variable damping damper 51 provided between the rear carriage frame 21-2 and the axle boxes 30-3 and 30-4, and each variable damping damper. To 51.

Each variable damping damper 51 controls the damping force based on the damping force command value _urij . Thereby, the vertical vibration of the bogie frame is reduced, and thus the ride comfort in the vehicle body can be improved.

In the control unit 70, first, the vertical acceleration of the carriage frame 21 is integrated from the output value to obtain the vertical translation speed dz _Ti / dt of the carriage frame. Here, the subscript T means a value related to the carriage. Next, this velocity value is substituted into the following equation (3) to calculate the damping force command value u _rij .

Here, C _s is a so-called skyhook gain. The damping force command value u _rij is calculated independently for the front bogie frame 21-1 and the rear bogie frame 21-2 by the control units 70-1 and 70-2. The same damping force command value _ur1j is given to the four variable damping dampers 51 installed between the front bogie frame 21-1, the front axle box 30-1 and the rear axle box 30-2. The same damping force command value _ur2j is given to the four variable damping dampers 51 installed between the rear bogie frame 21-2, the front axle box 30-3, and the rear axle box 30-4.

  Here, one acceleration sensor is arranged in each bogie frame, but by locating more sensors in the bogie frame, accelerations of other motion modes can be obtained, and these other motion modes are also taken into consideration. Control becomes possible. For example, if one acceleration sensor is arranged before and after each bogie frame, vibrations in the vertical translation mode and pitching mode of the bogie frame can be obtained. As a control law at this time, for example, the above-described skyhook control may be applied to the pitching mode and combined with the skyhook gain in the vertical translation mode. In particular, when the height of the mounting position of the traction link or the yaw damper that connects the vehicle body and the bogie frame is greatly different from the height of the pitching center of the bogie frame, by controlling the pitching mode in this way, It is possible to reduce vertical vibration.

(2) Control law 2 (control based on information on both bogie frame vibration and vehicle body vibration)
Next, a method for controlling by measuring the vibration acceleration of the vehicle body in addition to the bogie frame vibration acceleration will be described. As for the specific control law, various control laws can be applied as in the case of the control law 1. However, in this embodiment, the controller is designed by the optimal control method, and this is changed to semi-active control. Applied.

  FIG. 7 is a diagram schematically showing the arrangement of the acceleration sensor 60 and the control unit 70 when the control law 2 is applied. The acceleration sensor 60, the control unit 70, and the variable damping damper 51 constitute a vibration damping device. In the figure, the right side is the direction of travel.

As shown in the figure, acceleration sensors 60-1 and 60-2 are arranged at the center of each bogie frame 21, and acceleration sensors 60-3 are respectively provided at the front side, the center, and the rear side of the vehicle body floor. , 60-4, 60-5 are arranged. Further, output values from all the acceleration sensors 60 are collected in the control unit 70 provided in the vehicle body 10. The control unit 70 calculates a damping force command value _urij to the variable damping damper 51 provided between the carriage frame 21 and the axle box 30 from this output value, and transmits it to each variable damping damper 51. Each variable damping damper 51 controls the damping force based on the damping force command value _urij , whereby the vertical vibration of the vehicle body 10 is attenuated and the riding comfort in the vehicle body 10 can be improved.

In calculating the damping force command value in the control unit 70, a linear system having 32 state quantities, 8 disturbance inputs, and 4 control inputs, obtained by modeling this railway vehicle. The equation of state (the following equation (4): the calculation method will be described later) was used. However, since it is actually difficult to observe (estimate) all the state variables, the influence of the longitudinal coupling characteristics between the vehicle body and the bogie frame is ignored, and the effect on the vertical vibration of the vehicle body is large. Optimal feedback using the 10th-order reduced-order state equation (Equation (5)) consisting of the vertical and horizontal translation modes of the bogie frame, the vertical translation mode of the vehicle body, the pitching mode, and the primary bending mode. Asked.

Here, the vector x _Wij is the wheel shaft longitudinal displacement, z _Wij is the wheel shaft vertical displacement, x _Ti is the cart frame longitudinal displacement, z _Ti is the cart frame vertical displacement, and θ _Ti is Pitching displacement of the carriage frame, x _Di is the longitudinal displacement of the yaw damper, x _B is the longitudinal displacement of the vehicle body, z _B is the vertical displacement of the vehicle body, q _B is the displacement of the primary bending mode of the vehicle body, f _dij is the generated force of the variable damping damper. As described above, i = 1 (front carriage), 2 (rear carriage), j = 1 (front axle), and 2 (rear axle).

を最小にするような状態フィードバック（式(7)）、

によって可変減衰ダンパへの指令ｕを計算する。すなわち、ベクトルｕは、[ｕ_r11,ｕ_r12,ｕ_r21,ｕ_r22]^Tとなる。なお、状態ベクトルｘ_rのなかに観測できない状態変数がある場合には、適当な状態オブザーバを設計して推定すれば良い。 And in the control unit, the evaluation function represented by the following equation (6) for the system represented by the state equation (5),

State feedback (equation (7)) that minimizes

To calculate the command u to the variable damping damper. That is, the vector u is a _{[u r11, u r12, u} r21, u r22] T. When there is a state variable which can not be observed among the state vector x _r it may be estimated by designing a suitable state observer.

  According to the control law 2 as described above, the variable damping damper is controlled in consideration of the primary bending mode of the vehicle body, and the riding comfort on the vehicle body can be further improved. Needless to say, as in the case of the control law 1, the arrangement configuration of the acceleration sensor and the like can be appropriately changed.

  Here, a method of calculating the state equation (4) will be described. FIG. 8 is a model diagram obtained by modeling the railway vehicle in this embodiment for simulation. The right side in the figure is the traveling direction of this model, which is the x axis in the traveling direction, the z axis in the vertical direction in the figure, and the y axis in the direction perpendicular to the paper surface in the figure. In this model, in addition to vertical vibrations of the vehicle body 10 as a beam and the bogie frame 21 as a rigid body, the degree of freedom such as the vehicle body front and rear, the vehicle frame front and rear and pitching, and the wheel shaft 31 front and rear are handled. This is a model that can handle the influence of the traction link 43 and the yaw damper 42, which are the coupling elements of.

  Further, in order to handle a semi-active suspension (variable damping damper), a time history simulation is required. For this reason, the damping element is expressed using viscous damping, not a complex spring.

車両のモデリングに関して、ａ．車体のモデル、ｂ．結合力のモデル、ｃ．運動方程式の順で説明する。 First, referring to FIG. 8, the main symbols used in this modeling are shown in Table 1 below.

Regarding vehicle modeling: a. Body model, b. A model of binding force, c. It explains in order of the equation of motion.

a. Model of the car body Assuming that the car body is a uniform elastic beam free at both ends, and considering the inertia, internal viscosity, and bending rigidity, the vehicle body is subject to a small car body vertical displacement z (x, t) at the position in the longitudinal direction x. The following equation (8) holds.

Here, f _k (t) is an external force applied in the vertical direction of the vehicle body at the position l _k , and δ (x) is a delta function.
Equation (8) is developed as the following equation (9) by the mode analysis method with the displacement of each mode of the vehicle body as q _Bm (m ≧ 1).

但し、ｆ_Aizは、位置ｌ_TCiでの空気ばねによる支持力であり、また、ω，ζは、下式(13)で定義される。 However, the vertical translational displacement of the vehicle body is z _B , and the pitching displacement is θ _B. Here, if the higher order modes (m> 3) than the primary bending of the vehicle body are ignored, the following equations (10) to (12) are established for the vehicle body in each mode (m = 1, 2, 3). .

However, f _Aiz is a supporting force by the air spring at the position l _TCi , and ω and ζ are defined by the following expression (13).

また、上式(12)の曲げモード固有関数は、下式(14)で表される。

ここで、Ｘ＝λｘ／ｌであり、λは、振動数方程式である下式(15)の１次曲げモードに対応する根である。

The bending mode eigenfunction of the above equation (12) is expressed by the following equation (14).

Here, X = λx / l, and λ is a root corresponding to the primary bending mode of the following equation (15) which is a frequency equation.

b. Coupling force modeling For each axle box support, vehicle body support, traction link, and yaw damper, the transmission force is formulated into a formula. Here, in the vertical direction of the shaft box support portion, the shaft spring rigidity k _Wijz and the shaft damper damping c _Wijz are considered. However, at the time of semi-active control, c _Wijz = 0 and the generated force of the semi-active damper is f _dij . In the longitudinal direction, the longitudinal spring stiffness k _Wijx and the damping c _Wijx of the shaft spring are taken into consideration.

The front and rear generation force f _Wijx the axle box support, the upper and lower generation force f _Wijz is the following formula, respectively (16), expressed in (17) (composite, in the order of j = 1, 2) .

Further, regarding the vehicle body support portion, the longitudinal stiffness and damping of the air spring are sufficiently small as compared with the traction link and the yaw damper, and are ignored and only the vertical direction is considered. When the air spring is a two-element model composed of spring stiffness k _Aiz and damping c _Aiz , the vertical direction generated force f _Aiz of the vehicle body support portion is expressed by the following equation (18) (composite is i = 1). , 2 in this order).

The longitudinal transmission force f _Li of the traction link is expressed by the following equation (19) in consideration of the rigidity k _LRi and the viscous damping c _LRi of the mounting rubber.

Further, the vehicle body side longitudinal force f _DBi and the vehicle side longitudinal force f _DRi of the yaw damper are respectively _expressed as x _DBi and x _DRi and the displacements of the body side and the vehicle side of the yaw damper are modeled as shown in FIG. 20) to (24) (composites are in the order of i = 1, 2).

c. Equation of motion Next, the equation of motion is derived. First, the following formula (25) holds for the front and rear of the wheel shaft.

In addition, the following formulas (26) to (28) are established for the front and rear, the top and bottom, and the pitching of the bogie frame.

Further, the following formulas (29) to (32) are established for the front and rear, the top and bottom, the pitching, and the primary bending of the vehicle body (composites are in the order of i = 1 and 2).

  Based on these equations of motion and the equation of motion of the yaw damper (24), the linearity having 32 state quantities, 8 disturbance inputs and 4 control inputs expressed by the above equation (4). A system is obtained.

Next, real-time simulation results based on the control law 1 and the control law 2 will be described. As the vertical displacement and speed disturbance to the wheel axis, the shaft box acceleration of the Shinkansen train was integrated to obtain the shaft box vertical speed and displacement, and this was input with a phase difference corresponding to a traveling speed of 300 [km / h]. The main parameters were set as shown in Table 2 below.

As for the variable damping damper, c _d = 1.57 × 10 ⁶ [N / (m / s) ² ], k _d = 4.905 [N / (m / s)], and maximum damping force 1.08 × 10 ⁴ per wheel axle. [N]. This damping force corresponds to the damping force per axis when the piston speed is 0.15 [m / s] in an existing passive damper, and is a realistic setting. The response delay of the valve was 10 [ms].

As the present simulation results, using the ride level L _T which is generally used as an index for evaluating the riding comfort of the railway vehicle. In this method, the vehicle body vibration acceleration is weighted using human vibration sensation characteristics (riding comfort filter), and the effective value is normalized with a threshold value and displayed in dB. For example, it is explained in detail in Japan Railway Technology Association "Report on Ride Comfort Management Standards" (1979-1981).

As the value of the ride level L _T is small, ride comfort is good. In addition, about the vibration of the vertical system, 4 to 8 [Hz] is most heavily weighted, and it is important to reduce the vibration in this frequency band. Therefore, in this embodiment, the controller is designed with attention to this characteristic.

Then, the control law 1 with skyhook control of the bogie frame, parameters since only skyhook gain C _s, bogie vibration is a value that can be reduced to some extent. In this simulation, C _s = 254,800 [N / (m / s)].

In Control Law 2 using the optimal control law (LQR (Linear Quadratic Regulator)), the primary bending vibration of the vehicle body with a large power in the vicinity of 4 to 8 [Hz] and the vertical vibration of the bogie frame are emphasized. The vectors Q and R, which are parameters of the equation (6), are represented by the following equations (33) and (34).

First, the riding comfort improvement effect will be described based on the simulation results. FIG. 10 is a diagram showing vertical acceleration PSD (Power spectral density) on the vehicle body floor surface directly above the front carriage. FIG. 11 is a diagram showing the vertical acceleration PSD on the center floor of the vehicle body. Both figures show the results of control law 1 and control law 2 and, as a comparison target, the results of the case without control (passive damper). Further, in the figure legends, it is shown together L _T value. In FIG. 10, L _T = 86.9 [dB] without control, L _T = 84.7 [dB] with control law 1, and L _T = 83.7 [dB] with control law 2. . In FIG. 11, L _T = 86.4 [dB] without control, L _T = 83.5 [dB] with control law 1, and L _T = 82.8 [dB] with control law 2. . FIG. 12 is a diagram showing the vertical translational acceleration PSD in the front bogie frame.

Looking at the simulation result of the control law 1, it can be seen that the vertical vibration acceleration of the bogie frame generally decreases in the vicinity of 3 to 10 [Hz] as shown in FIG. As a result, as shown in FIGS. 10 and 11, it can be seen that the vertical acceleration of the vehicle body is similarly reduced in the same band. Then, the ride comfort level L _T, as compared with the case without control, the vehicle body center at 3 [dB] or so, and then 2 [dB] much better in the previous carriage just above, the control law 1, to the extent that can experience It can be seen that ride comfort is improved.

Further, looking at the simulation result of the control law 2, in the frequency band of 5 to 7 [Hz], the vertical vibration at the center of the vehicle body is slightly larger than that without control (see FIG. 11). This is due to the effects of modeling errors such as the fact that the actuator is a variable damping damper (semi-active damper) and that a low-order controller is used, and not all the state quantities in equation (4) are used. It is believed that there is. However, has been greatly reduced 9 [Hz] around which is emphasized by the ride comfort level (frequency band having a large power body of the first bending mode), as a result, L _T value in the case of no control Compared to the control law 1, the improvement is about 3.5 [dB] and 3 [dB] just above the front carriage.

Here, in order to determine which frequency band to improve the ride comfort level L _T is contributing, multiplied by the weight of the ride filter acceleration PSD, integrating the results for each octave band (vibration power) in FIG. 13 Show. FIG. 13A is a diagram showing the vibration power distribution on the vehicle body floor surface directly above the front carriage, and FIG. 13B is a diagram showing the vibration power distribution on the vehicle body center floor surface. As shown in the figure, the _LT value without control is dominated by vibration power in the 8 [Hz] band, and this is particularly noticeable at the center of the vehicle body. On the other hand, in the case of control law 1, the vibration power in the 8 [Hz] band can be reduced to half, and in the case of control law 2, the vibration power in the 8 [Hz] band is reduced to about 1/3. are made, it can be seen that this contributes to the reduction of L _T values.

In addition, although the operation state of the variable damping damper when the control law 2 was applied was verified, the generating force of the variable damping damper was about 5 [kN] per one, and it was found to be a realistic value. .
As described above, the railway vehicle with the vibration damping function according to the present embodiment has been described in detail. However, in order to improve the riding comfort in the vehicle body by installing a variable damping damper between the bogie frame and the axle box, According to the present embodiment in which the damping force is controlled, the riding comfort can be improved as compared with a conventional railway vehicle. In particular, when the control law 1 is applied, the ride comfort can be improved with a simple configuration, and when the control law 2 is applied, the ride comfort level can be further improved than the control law 1. .

The present invention is not limited to the above-described embodiment, and various modifications can be made without departing from the gist of the present invention.
For example, it goes without saying that the control law of the variable damping damper can be appropriately changed to a predetermined control law. In Control Law 2, the LQR controller is applied as it is, but it is necessary to change to a control law that takes into account the nonlinearity of the variable damping damper to some extent, or to change to a control law that further improves performance for lowering the dimensions. Can do.

In the present embodiment, a railway vehicle has been described as an example. However, the present invention is not limited to this, and the present invention can be applied to any vehicle including a vehicle body, a carriage frame, and a axle box.

  In this embodiment, a semi-active suspension (variable damping damper) is used, but an active suspension that is more advantageous in terms of vibration reduction effect can be used instead. The semi-active suspension only controls the damping force of the variable damping damper, while the active suspension uses an actuator that generates driving force, and controls the generating force of the actuator according to the vibration situation of the vehicle body, It is intended to actively reduce vibration. As the actuator, for example, a linear motor, a hydraulic cylinder, an electro-hydraulic actuator, a pneumatic cylinder, or the like is used. Since this active suspension can use a linear control law, there is a merit that the prospect of the control system design is easier than the semi-active suspension.

  Hereinafter, the configuration when the active suspension is used will be described in detail. The basic railcar configuration itself is the same as the semi-active suspension described above. However, in the case of an active suspension, there may be a case where the vehicle body is vibrated on the contrary if an abnormality occurs in the control system. It is necessary to configure.

FIG. 14 schematically shows an example of a configuration between the carriage frame 21 and the axle box 30 when the active suspension is employed. FIG. 14A shows a configuration in which a shaft damper 59 is arranged in addition to the actuator 58 and the shaft spring 50 between the carriage frame and the shaft box. As the actuator 58, a linear motor or the like is used. When an abnormality occurs in the active suspension, the control unit (not shown) is turned off so that the actuator 58 does not generate control force. In this case, since the shaft damper 59 functions as a passive damper, it is possible to maintain the same riding comfort as the current situation.

FIG. 14B shows a configuration in which an electro-hydraulic actuator as the actuator 58 and the shaft spring 50 are arranged between the carriage frame and the shaft box. This electro-hydraulic actuator also has a function as a damper, and is configured to be able to switch between acting as an actuator and acting as a shaft damper by switching a built-in valve.
If a member that acts as a shaft damper is arranged when the actuator is turned off as shown in FIG. 14 , even if an abnormality occurs in the control system of the actuator, a stable riding comfort can be obtained. It is possible to maintain.

  Next, some examples of active suspension control rules will be briefly described. First, as in the case of the control law 1 of the semi-active suspension described above, a method of controlling for each carriage frame by paying attention only to the vibration of the carriage frame can be mentioned. As shown in FIG. 6, when one acceleration sensor is installed in each bogie frame, only the vertical translation of the bogie frame can be controlled. When two acceleration sensors are installed in front of and behind each bogie frame, vertical translation and pitching of the bogie frame can be controlled.

  Further, as in the case of the control law 2 for the semi-active suspension described above, a control method that considers both the bogie frame and the vibration of the vehicle body is also included. As shown in FIG. 7, when one acceleration sensor is installed on each bogie frame and three acceleration sensors are installed on the vehicle body floor surface, the bogie frame is vertically translated, the vehicle body is vertically translated, and the pitching is performed. The primary bending mode can be detected and controlled.

  As a specific control law, vibration acceleration is detected and fed back in any of the above cases, but a desired control law such as skyhook control, optimal control, and H∞ optimal control can be used as appropriate. It is. In the case of the active suspension, as in the case of the semi-active suspension described above, the ride comfort can be improved by suppressing the vertical vibration of the vehicle body.

1 is a diagram schematically showing a configuration of a railway vehicle according to an embodiment of the present invention. It is a figure which shows the biflow type hydraulic circuit which concerns on embodiment of this invention. It is a figure which shows the uniflow type hydraulic circuit which concerns on embodiment of this invention. It is a figure which shows the uniflow type hydraulic circuit which concerns on embodiment of this invention. It is a figure which shows notionally the up-down translation mode, pitching mode, and primary bending mode which concern on embodiment of this invention. It is a figure which shows roughly the structure of the damping device in the case of applying the control law 1 which concerns on embodiment of this invention. It is a figure which shows roughly the structure of the damping device in the case of applying the control law 2 which concerns on embodiment of this invention. It is a simulation model figure of a railway vehicle concerning an embodiment of the invention. It is a simulation model figure showing the yaw damper between the vehicle body of a railway vehicle and a bogie according to an embodiment of the present invention. It is a figure which shows the simulation result of the vertical direction acceleration PSD in the vehicle body floor surface just above the front trolley | bogie which concerns on embodiment of this invention.

It is a figure which shows the simulation result of the vertical direction acceleration PSD in the vehicle body center floor based on Embodiment of this invention. It is a figure which shows the simulation result of the vertical translation acceleration PSD in the front bogie frame which concerns on embodiment of this invention. It is a figure which shows the distribution result of the vibration power by the simulation which concerns on embodiment of this invention. It is a figure which shows roughly the structure between a trolley | bogie frame and an axle box at the time of employ | adopting the active suspension which concerns on the modification of embodiment of this invention. It is a figure which shows schematically the system configuration | structure of the conventional vertical system active vibration suppression control apparatus. It is a figure which shows schematically the structure of the trolley | bogie part of the rail vehicle which has the conventional semi-active type vertical suspension system.

Explanation of symbols

DESCRIPTION OF SYMBOLS 10 Car body 20 Bogie 21 Bogie frame 30 Shaft box 31 Wheel 32 Wheel shaft 41 Air spring 42 Yaw damper 43 Traction link 50 Shaft spring 51 Variable damping damper 58 Actuator 60 Acceleration sensor 70 Control unit

Claims (6)

In a vehicle including a vehicle body, a plurality of carriage frames that support the vehicle body, and an axle box and a wheel shaft that support the carriage frame,
An acceleration sensor provided respectively to the each bogie frame in order to detect the vibration in the vertical direction of the respective bogie frame,
A variable damping damper provided between the carriage frame and the axle box;
Control means for controlling the damping force of each variable damping damper so as to reduce the vertical vibration of each bogie frame based on the output of each acceleration sensor. Vehicle with function. In order to detect the primary bending vibration of the vehicle body, the vehicle further comprises at least three acceleration sensors installed on the vehicle body,
Said control means based on the output of the acceleration sensor installed in the vehicle body, further control of claim 1, wherein the controller controls the variable attenuation damper to reduce the primary bending vibration of the vehicle body Vehicle with vibration function. In a vehicle including a vehicle body, a plurality of carriage frames that support the vehicle body, and an axle box and a wheel shaft that support the carriage frame,
An acceleration sensor provided respectively to the each bogie frame in order to detect the vibration in the vertical direction of the respective bogie frame,
In order to reduce the vertical vibration of the vehicle body by generating a forced driving force, an actuator provided between the carriage frame and the axle box,
Control means for controlling the driving force of each actuator so as to reduce the vertical vibration of each carriage frame based on the output of each acceleration sensor. Vehicle with function. In order to detect the primary bending vibration of the vehicle body, the vehicle further comprises at least three acceleration sensors installed on the vehicle body,
Said control means based on the output of the acceleration sensor installed in the vehicle body, so as to further reduce the primary bending vibration of the vehicle body, the damping of claim 3, wherein the controller controls the actuator Vehicle with function. In a vehicle comprising a vehicle body, a plurality of carriage frames that support the vehicle body, and an axle box and a wheel shaft that support the carriage frame, a vibration damping device for reducing vertical vibrations of the vehicle body,
An acceleration sensor provided respectively to the each bogie frame in order to detect the vibration in the vertical direction of the respective bogie frame,
A variable damping damper provided between the carriage frame and the axle box;
Control means for controlling the damping force of each variable damping damper so as to reduce the vertical vibration of each bogie frame based on the output of each acceleration sensor. apparatus. In a vehicle comprising a vehicle body, a plurality of carriage frames that support the vehicle body, and an axle box and a wheel shaft that support the carriage frame, a vibration damping device for reducing vertical vibrations of the vehicle body,
An acceleration sensor provided respectively to the each bogie frame in order to detect the vibration in the vertical direction of the respective bogie frame,
In order to reduce the vertical vibration of the vehicle body by generating a forced driving force, an actuator provided between the carriage frame and the axle box,
And a control means for controlling the driving force of each actuator so as to reduce the vertical vibration of each carriage frame based on the output of each acceleration sensor.

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