JP2011201476A - Railroad vehicle, and method of suppressing vibration of vehicle body of the same - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a railroad vehicle and a method of suppressing the vibration of a vehicle body of the same capable of obtaining an excellent vibration suppressing effect without performing complicated adjustment.SOLUTION: A railroad vehicle 1 includes: the vehicle body 10; and a mass body fitted to the vehicle body through spring elements and damping elements so as to vibrate relative to the vehicle body. The natural frequency of the mass body is set lower than that of the vehicle body in elastic vibration, and a damping capacity of the mass body is set larger than that of the vehicle body in elastic vibration.

Description

  The present invention relates to a railway vehicle and a method for suppressing vibration of a railway vehicle body, and more particularly to an apparatus for suppressing vibration using an additional system elastically supported with respect to the vehicle body.

  Generally, a railway vehicle body for passengers is formed in a substantially hexahedron shape including a roof structure, a floor structure, a side structure, and a wife structure, and has a shape that extends long along the traveling direction of the track. . Such a car body exhibits various vibration modes such as vertical bending vibration that resembles a single beam or vibration that accompanies shear deformation of the car body cross section due to the vertical vibration force from the carriage during traveling. It has been known. Such elastic vibration of the vehicle body is perceived by passengers as being inferior in riding comfort, and is therefore required to be suppressed.

Conventionally, as a technique for suppressing elastic vibration, it is known to use a dynamic vibration absorber (dynamic damper) in which a mass is connected to a vibration body to be controlled via a spring element and a damping element.
For example, Non-Patent Document 1 describes that an underfloor device of a railway vehicle is suspended by a spring and tuned to act as a dynamic damper.
Patent Document 1 discloses a dynamic vibration absorber that suppresses bending vibration generated in a running railway vehicle, including a central portion in the longitudinal direction of the railway vehicle, at a predetermined position on the floor of the railway vehicle or above the floor. It is described that a weight having a mass of 1 is attached via an elastic body.

"Reducing body bending vibration by dynamic damper" Ryutaro Ishikawa et al. Proceedings of the 68th General Meeting of the Japan Society of Mechanical Engineers (Vol.C) March 4, 1991

JP 2002-029420 A

In general, a fixed point theory of a two-degree-of-freedom system is widely known as a design method for a dynamic vibration absorber. In other words, utilizing the fact that there are two fixed points in the response curve of the main system to which a dynamic vibration absorber is attached, the response curves are designed to have maximum values at the two fixed points and to have the same height. To do. Or, more simply, it is set so that the natural frequency of the additional system matches the natural frequency (peak frequency) of the damping target, and the damping is adjusted to optimize the damping effect. . In either case, fine adjustment of the spring constant and damping constant is required according to the mass of the additional system.
However, in the case of a railway vehicle, the input such as track height irregularities has characteristics distributed over a wide frequency band, and even if it is limited to a frequency of 20 Hz or less, which has a great influence on the ride comfort, there are a plurality of inherent characteristics within that range. Has a frequency. For this reason, even if the magnitude of the frequency response (FRF) is reduced only in the vicinity of the peak frequency by the tuned dynamic vibration absorber designed based on the above fixed point theory, if the frequency response is increased at other frequencies, the overall vibration is reduced. It may not be. In addition, since a plurality of modes may affect the ride comfort, even if only one mode is controlled, the overall ride comfort may not be improved. Furthermore, it is complicated to adjust the attenuation of the dynamic vibration absorber so that the vibration damping effect is optimized. Further, it is conceivable that the natural frequency and damping of the additional system change due to various causes during use. However, in the above-described dynamic vibration absorber, the vibration damping performance is greatly deteriorated.
In view of the above-described problems, an object of the present invention is to provide a railway vehicle and a vibration suppression method for a railway vehicle body that can obtain a good vibration suppression effect without complicated adjustment.

  In order to solve the above-described problems, a railway vehicle according to the present invention is a railway vehicle that includes a vehicle body and a mass body that is attached to the vehicle body via a spring element and a damping element and that vibrates relative to the vehicle body. The natural frequency of the mass body is set lower than the natural frequency of the elastic vibration of the vehicle body, and the damping capacity of the mass body is set to be larger than the damping capacity of the elastic vibration of the vehicle body. Features.

As described above, in the usual dynamic vibration absorber design method, the peak height in the frequency response characteristics can be lowered, but a new peak is formed on both sides of the railway vehicle so that the ride comfort deteriorates. In some cases, it was not always possible to obtain optimum results when considering the application to the above.
In addition, in a dynamic vibration absorber with the natural frequency of the additional system matched to the natural frequency of the vibration suppression target, the vibration suppression effect varies greatly depending on the support rigidity and damping characteristics of the additional system. It is necessary to adjust the rigidity and damping capacity accordingly, and the design and production are complicated.
On the other hand, according to the present invention, the natural frequency of the mass body constituting the additional system is set lower than the natural frequency in the elastic vibration of the vehicle body, and the damping capacity (damping ratio) is set to the elastic vibration of the vehicle body. By setting a large value for the damping capacity (damping ratio), it is possible to achieve a good damping effect even for vibrations other than the natural frequency, and to adjust the rigidity and damping capacity of the additional system strictly. Even if it is not necessary, it is possible to ensure the vibration control effect which is not much different from the optimum case.

In the present invention, a plurality of mass bodies can be dispersed and attached to the vehicle body.
According to this, a plurality of vibration modes can be effectively suppressed, and the riding comfort of the railway vehicle can be further improved.

Moreover, in this invention, the said mass body can be set as the structure which is an apparatus suspended under the floor of the said vehicle body.
According to this, the weight of the vehicle is not increased as in the prior art in which a dedicated weight is added as a dynamic vibration absorber, and the spring characteristics and the damping characteristics of the support portion of the existing equipment are adjusted by adjusting the spring characteristics and the damping characteristics of the present invention. A vibration control effect can be obtained.

In addition, the method for suppressing vibration of a vehicle body for a railway vehicle according to the present invention attaches a mass body capable of relative vibration to the vehicle body via a spring element and a damping element, and sets the natural frequency of the mass body to elastic vibration of the vehicle body. And the damping capacity of the mass body is set to be larger than the damping capacity of the elastic vibration of the vehicle body.
In the present invention, a plurality of mass bodies can be dispersed and attached to the vehicle body.
Moreover, in this invention, the said mass body can be set as the structure which is an apparatus suspended under the floor of the said vehicle body.

  As described above, according to the present invention, it is possible to provide a railway vehicle and a vibration suppression method for a railway vehicle body that can obtain a good vibration suppression effect without complicated adjustments.

It is a model figure at the time of making a main system and an additional system into a simple mass point supported by a spring and viscous damping, respectively. 2 is a graph showing a frequency response when the mass ratio is 0.06, the main system damping ratio is 0.03, and the natural frequency ratio of the additional system and the main system is 0.6 in the model of FIG. 2 is a graph showing a frequency response when the mass ratio is 0.06, the main system damping ratio is 0.03, and the natural frequency ratio between the additional system and the main system is 0.8 in the model of FIG. 2 is a graph showing a frequency response when the mass ratio is 0.06, the main system damping ratio is 0.03, and the natural frequency ratio of the additional system and the main system is 1.0 in the model of FIG. 2 is a graph showing a frequency response when the mass ratio is 0.06, the main system damping ratio is 0.03, and the natural frequency ratio of the additional system and the main system is 1.2 in the model of FIG. 2 is a graph showing a frequency response when the mass ratio is 0.20, the main system damping ratio is 0.03, and the natural frequency ratio between the additional system and the main system is 0.6 in the model of FIG. 2 is a graph showing a frequency response when the mass ratio is 0.20, the main system damping ratio is 0.03, and the natural frequency ratio between the additional system and the main system is 0.8 in the model of FIG. 2 is a graph showing a frequency response when the mass ratio is 0.20, the main system damping ratio is 0.03, and the natural frequency ratio of the additional system and the main system is 1.0 in the model of FIG. 2 is a graph showing a frequency response when the mass ratio is 0.20, the main system damping ratio is 0.03, and the natural frequency ratio of the additional system and the main system is 1.2 in the model of FIG. FIG. 2 is a contour diagram showing an area ratio of a frequency response function when the natural frequency of the additional system with respect to the natural frequency of the main system and the damping ratio of the additional system with respect to the damping ratio of the main system are used as parameters in the model of FIG. The data for a mass ratio of 0.06 and a main system damping ratio of 0.03 are shown. FIG. 2 is a contour diagram showing an area ratio of a frequency response function when the natural frequency of the additional system with respect to the natural frequency of the main system and the damping ratio of the additional system with respect to the damping ratio of the main system are used as parameters in the model of FIG. The data for a mass ratio of 0.20 and a main system damping ratio of 0.03 are shown. It is a figure which shows the vehicle vibration analysis model of the railway vehicle which is a bogie car which has a 2-axis 1st trolley | bogie and a 2nd trolley | bogie. 13 is a graph showing the frequency response of acceleration at the center of the vehicle body when the wheel shaft is vibrated in the model of FIG. It is a graph which shows the frequency response of the acceleration directly on the trolley | bogie at the time of wheel-axis vibration in the model of FIG.

Hereinafter, a railway vehicle to which the present invention is applied and an elastic vibration suppressing method for a railway vehicle body will be described.
First, the study on the basic characteristics of the dynamic vibration absorber performed by the inventor of the present invention will be described.
A model diagram of the dynamic vibration absorber used for the study is shown in FIG.
As shown in FIG. 1, this model is composed of a main system to be controlled and an additional system that functions as a dynamic vibration absorber. The main system is a mass body having a mass m ₀ supported by a spring element having a spring constant k ₀ and a damping element having a damping coefficient c ₀ . The additional system is obtained by attaching a mass body having a mass m _{1 to} the main system via a spring element having a spring constant k ₁ and a damping element having a damping coefficient c ₁ .

2 to 9 show a mass ratio (m ₁ / m ₀ ) obtained by applying force excitation to the main system using the model of FIG. 1, and dividing the mass m ₁ of the additional system by the mass m ₀ of the main system, Dimensionless when the frequency ratio kf obtained by dividing the natural frequency of the additional system by the natural frequency of the system and the damping capacity ratio kz obtained by dividing the damping ratio of the additional system by the damping ratio of the main system are changed as parameters. It is a graph which shows the frequency response function (FRF) gain with respect to a frequency.
In each figure, the horizontal axis indicates the dimensionless frequency normalized by the natural frequency of the main system, and the vertical axis indicates the response magnification (FRF gain) of the main system. The response magnification is obtained by dividing the displacement X of the main system by the static displacement Xst = F / k ₀ of the main system alone. Here, F is an excitation force acting on the main system.

In each of FIGS. 2 to 9, the frequency response when the damping capacity ratio kz is 1, 2, 4, 8, and 10 when there is no additional system is shown by different line types.
2 to 5 show cases where the mass ratio of the additional system to the main system is 0.06 and the damping ratio of the main system is 0.03, and the natural frequency of the main system has the natural frequency of the additional system. The FRF gain when the frequency ratio kf to the number is 0.6, 0.8, 1.0, 1.2 is shown.
6 to 9 show cases where the mass ratio of the additional system to the main system is 0.20 and the damping ratio of the main system is 0.03. The FRF gain when the frequency ratio kf to the number is 0.6, 0.8, 1.0, 1.2 is shown.

As shown in FIGS. 4 and 8, when the natural frequency of the additional system is matched with the natural frequency of the main system (frequency ratio = 1), the natural frequency of the main system (non-dimensional vibration) It can be seen that the peak height in number = 1) is low, but there are frequency regions on both sides of which the vibration is amplified.
In addition, since the dependence of the damping effect on the damping ratio of the additional system is large, and the damping ratio deviates from the optimum value, the vibration reducing effect is deteriorated, so that it is necessary to adjust the damping.
Further, as shown in FIGS. 5 and 8, even when the natural frequency of the additional system is made higher than the natural frequency of the main system, the vibration reduction effect may be deteriorated.

  On the other hand, when the natural frequency of the additional system is made lower than the natural frequency of the main system, especially when the frequency ratio kf is reduced to about 0.6, as shown in FIGS. It can be seen that the damping effect can be exerted in a wide frequency band including the natural frequency of the main system, and there is almost no adverse effect such that new peaks are formed on both sides of the peak frequency. In this case, even when the rigidity and damping ratio of the additional system are changed, the influence on the vibration damping effect is negligible, so that it is not necessary to adjust the rigidity and damping ratio strictly. .

10 and 11 are contour diagrams showing the area ratio of the FRF gain diagram area when the additional system is applied to the FRF gain diagram area when the additional system is not applied in the model of FIG. The area that is the basis for calculating this area is the area in the range from 0 to 3 in the dimensionless frequency in FIGS.
10 and 11, the horizontal axis represents the frequency ratio kf, which is the ratio of the natural frequency of the additional system to the natural frequency of the main system, and the vertical axis represents the ratio of the damping ratio of the additional system to the damping ratio of the main system. The damping capacity ratio kz is shown. FIG. 10 shows the case where the mass ratio, which is the ratio of the mass of the additional system to the mass of the main system, is 0.06, and FIG. 11 shows the case where the mass ratio is 0.20.

10 and 11, the area ratio is less than a predetermined value (for example, 0.85 or less, 0.9 or less, etc.) when the natural frequency of the additional system is set lower than the natural frequency of the main system. It can be seen that the range is wide.
Further, in both FIG. 10 and FIG. 11, it can be seen that the area ratio is wider than the predetermined value when the attenuation ratio of the additional system is increased.
Therefore, in order to obtain a good damping effect over a wide frequency range, it is effective to add a vibration body with a high damping capacity having a natural frequency lower than that of the object to be damped as a dynamic vibration absorber. I understand that.

Next, application of an additional system (dynamic vibration absorber) having a low natural frequency and a high damping capacity to a vibration suppression target will be described.
FIG. 12 is a diagram showing a vehicle vibration analysis model of a railway vehicle that is a bogie car having a two-axis first-position carriage and a second-position carriage.
As shown in FIG. 12, the railway vehicle 1 in this model includes a vehicle body 10, a first cart 20, a second cart 30, and the like.

The vehicle body 10 is modeled as a uniform beam having a length L in consideration of elastic deformation in the vertical direction (y-axis direction in the drawing) and rigid body displacement in the vertical and front-rear directions.
The first and second carts 20 and 30 are two-axis carts arranged along the traveling direction of the vehicle 1.
The first cart 20 includes a cart frame 21, a first wheel shaft 22, a second wheel shaft 23, and the like.

The bogie frame 21 is a frame-like structure to which the first wheel shaft 22 and the second wheel shaft 23 are attached via a shaft box support device.
The first wheel shaft 22 and the second wheel shaft 23 have wheels fixed to both ends of the axle, and are sequentially arranged from the front side in the traveling direction. At both ends of the first wheel shaft 22 and the second wheel shaft 23, shaft boxes having bearings, a lubricating device, and the like are provided.

A primary spring system including a shaft spring 24 and a shaft damper 25 is provided between the axle box of the first wheel shaft 22 and the carriage frame 21 and between the axle box of the second wheel shaft 23 and the carriage frame 21. ing.
The shaft spring 24 generates a spring reaction force according to the relative displacement of the first wheel shaft 22 and the second wheel shaft 23 in the vertical direction (direction orthogonal to the raceway surface) with respect to the carriage frame 21.
The shaft damper 25 generates a damping force according to the vertical relative speed of the first wheel shaft 22 and the second wheel shaft 23 with respect to the carriage frame 21.

An air spring 26 that is a pillow spring constituting a secondary spring system is provided between the carriage frame 21 and the floor structure of the vehicle body 10.
The air spring 26 generates a spring reaction force according to the relative displacement of the carriage frame 21 relative to the vehicle body 10 in the vertical direction. The air spring 26 generates a damping force according to the relative speed in the vertical direction of the carriage frame 21 with respect to the vehicle body 10. Therefore, the air spring 26 can be expressed as a spring element and a damping element arranged in parallel in the model shown in FIG.

Further, a traction device 27 and a yaw damper 28 are provided between the carriage frame 21 and the vehicle body 10.
The traction device 27 includes a link device that transmits a longitudinal force between the carriage frame 21 and the vehicle body 10, and buffer rubber having a predetermined viscoelastic property is provided at both ends of the link for vibration isolation. Yes.
In an actual vehicle, the yaw damper 28 suppresses yawing (oscillation around the vertical axis) of the carriage frame 21 with respect to the vehicle body 10. Since this vehicle model does not consider yawing of the bogie and the vehicle body, it is modeled as generating a damping force according to the relative speed of the vehicle body and the bogie in the longitudinal direction. Further, buffer rubber having predetermined viscoelastic characteristics is provided at both ends of the yaw damper 28 for vibration isolation.

The second cart 30 is arranged on the rear side in the traveling direction with respect to the first cart 20.
The second carriage 30 is substantially the same as the carriage frame 21, first wheel shaft 22, third wheel shaft 23, shaft spring 24, shaft damper 25, air spring 26, traction device 27, and yaw damper 28 of the first truck 20. A frame 31, a third wheel shaft 32, a fourth wheel shaft 33, a shaft spring 34, a shaft damper 35, an air spring 36, a traction device 37, a yaw damper 38, and the like are provided.

Moreover, the meanings of the symbols in FIG. 12 are as follows.
z _{1 to} z ₄ : Excitation input to the first wheel shaft to the fourth wheel shaft (trajectory up and down irregularity)
u _{w1 to} u _w4 : Track direction (traveling direction) response displacement of the first wheel shaft to the fourth wheel shaft u _t1 , u _t2 : Track frame response displacement of the bogie frame of the first cart and the cart frame of the second cart w _t1 , w _t2: bogie frame ₁ of the bogie, the vertical displacement response theta _t1 of the bogie frame 2 of the _bogie, theta _t2: bogie frame ₁ of the bogie, the pitching angle of the bogie frame 2 of the bogie u _c: vehicle track direction response Displacement w _c (x, t): vertical response displacement at time phase t at a predetermined position on the x-axis (trajectory direction) of the vehicle body

FIGS. 13 and 14 are graphs showing the frequency response of the vehicle body acceleration to the vibration displacement when the displacement vibration of the same magnitude (4-axis in-phase vibration) is simultaneously applied to all the wheel shafts in the model of FIG. is there. FIG. 13 shows data at the center of the vehicle body, and FIG. 14 shows data immediately above the carriage. 13 and 14, the horizontal axis indicates the frequency, and the vertical axis indicates the frequency response function gain.
Such 4-axis in-phase excitation is different from the actual running, but was adopted because the peak of the vehicle body bending vibration becomes clear.
Here, the mass of the vehicle body 10 was 25.4 tons, and the natural frequency of the primary bending vibration as shown in FIG. 12 was 10 Hz.

Moreover, in FIG.13 and FIG.14, each line type represents the following data, respectively.
Solid line: No additional mass Dotted line: Natural frequency of additional mass 10 Hz, damping ratio 0.3
Dash-dot line: natural frequency of additional mass 10 Hz, damping ratio 0.5
Two-dot chain line: natural mass of added mass 6 Hz, damping ratio 0.3
Three-dot chain line: natural mass of added mass 6 Hz, damping ratio 0.5
The additional mass is 900 kg supported on the bottom surface of the floor structure at the center of the vehicle body via a spring element and a viscous damping element.
As such additional mass, various devices necessary for the operation of the railway vehicle can be used. According to this, it is not necessary to mount a new weight to configure the dynamic vibration absorber, and a vibration damping effect can be obtained without increasing the weight of the vehicle.
As such devices, for example, various control devices, electrical devices, pneumatic devices, hydraulic devices, air conditioners, and other devices can be used.

As shown in FIGS. 13 and 14, when the natural frequency of the additional mass is set to 10 Hz that synchronizes with the vibration control target peak (dotted line and alternate long and short dash line), as in the design of a general dynamic vibration absorber. Although the reduction amount of the peak height is large, there is a frequency region in which vibration increases on both sides of the peak frequency as compared with the case where no additional mass is attached. In addition, there is a damping ratio that maximizes the amount of vibration reduction, and if the damping ratio is different from the value, the vibration reduction effect is deteriorated, so that it is necessary to adjust the damping.
For example, in the example of FIG. 13 and FIG. 14, there is an optimum attenuation ratio between 0.2 and 0.3, and the attenuation ratio is 0.5 (dashed line) compared to the attenuation ratio of 0.3 (dotted line). The peak height has increased.
Although not shown in FIGS. 13 and 14, in the case of an attenuation ratio smaller than 0.2, two peaks are generated on both sides of 10 Hz that is a vibration target peak, and the peak decreases as the attenuation ratio decreases. Therefore, the vibration reduction effect is also reduced.

  On the other hand, when the natural frequency of the additional mass is set to 6 Hz which is smaller than the peak frequency to be controlled (10 Hz) (two-dot chain line and three-dot chain line: embodiments of the present invention), the damping ratio is increased. As a result, the peak height decreases, but no noticeable increase in vibration is observed in the frequency regions on both sides of the peak frequency. Although not shown in FIGS. 13 and 14, in the range up to about 0.7, the peak height decreases almost monotonously with the increase of the attenuation ratio, and peaks when the attenuation ratio is about 0.4. The change in height is small. Therefore, if the damping ratio is increased to some extent (for example, 0.4 or more), a vibration reduction effect that is not much different from the optimum case can be expected without adjusting the damping ratio.

In the excitation condition corresponding to the running of the railway vehicle, for example, the input such as the irregularity of the track height has frequency characteristics, so even if the frequency response function gain decreases only in the vicinity of the peak frequency, it will increase at other frequencies. As a result, vibration as a whole may not be reduced, and riding comfort may deteriorate. The gain of the frequency response function expresses the ease of response of the vehicle body to the input, but especially on the Shinkansen, the input from the track may be large in the frequency range of 10 to 20 Hz, and therefore an increase in vibration in that frequency range is preferable. Absent.
In this regard, when an additional system having a low natural frequency and high attenuation is used for a vehicle body to be controlled as in the present invention, the damping effect in a narrow frequency region near the peak to be controlled is a normal tuning type. Although inferior to a dynamic vibration absorber, a good vibration damping effect can be exhibited over a wide frequency band, and the vibration reduction effect as a whole can be improved.
Further, if the additional system has a damping ratio of a certain level or more, it is possible to obtain a damping effect which is not much different from the optimum case without particularly adjusting the rigidity and damping.

(Other embodiments)
In addition, this invention is not limited only to above-described embodiment, Various application and deformation | transformation can be considered.
For example, in the embodiment described above, only one additional system that functions as a dynamic vibration absorber is provided at the center of the vehicle body, but a plurality of additional systems may be provided in a distributed manner. According to this, vibration control in a plurality of vibration modes is possible. Even in such a case, according to the present invention, it is possible to obtain a multi-mode damping effect without particularly adjusting the rigidity and damping of each additional system.
Further, the position where the additional system is attached is not limited to the lower part of the floor structure, and may be attached to the upper part of the floor structure, the roof structure, the side structure, the wife structure, or the like.

DESCRIPTION OF SYMBOLS 1 Railcar 10 Car body 20 1st place bogie 21 Bogie frame 22 1st wheel shaft 23 2nd wheel shaft 24 Shaft spring 25 Shaft damper 26 Air spring 27 Pulling device 28 Yaw damper 30 2nd bogie 31 Bogie frame 32 3rd wheel shaft 33 4th wheel shaft 34 Shaft spring 35 Shaft damper 36 Air spring 37 Traction device 38 Yaw damper

Claims (6)

The car body,
A rail vehicle that is attached to the vehicle body via a spring element and a damping element and that vibrates relative to the vehicle body,
The natural frequency of the mass body is made lower than the natural frequency in the elastic vibration of the vehicle body, and the damping capacity of the mass body is set to be larger than the damping capacity in the elastic vibration of the vehicle body. Railway vehicle. The railway vehicle according to claim 1, wherein a plurality of the mass bodies are dispersed and attached to the vehicle body. The railway vehicle according to claim 1 or 2, wherein the mass body is a device that is suspended under the floor of the vehicle body. A mass body capable of relative vibration with respect to the vehicle body is attached via a spring element and a damping element, the natural frequency of the mass body is made lower than the natural frequency in the elastic vibration of the vehicle body, and the mass body is attenuated. The vibration suppression method for a railway vehicle body is characterized in that the performance is set to be greater than the damping capacity in the elastic vibration of the vehicle body. The vibration suppression method for a railway vehicle body according to claim 4, wherein a plurality of mass bodies are dispersed and attached to the vehicle body. The vibration suppression method for a railway vehicle body according to claim 4 or 5, wherein the mass body is a device that is suspended below the floor of the vehicle body.

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