CN112575630A - Track impact damping vibration absorber - Google Patents
Track impact damping vibration absorber Download PDFInfo
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- CN112575630A CN112575630A CN202011497758.2A CN202011497758A CN112575630A CN 112575630 A CN112575630 A CN 112575630A CN 202011497758 A CN202011497758 A CN 202011497758A CN 112575630 A CN112575630 A CN 112575630A
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- rail
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- motion guide
- elastic damping
- damping body
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- E—FIXED CONSTRUCTIONS
- E01—CONSTRUCTION OF ROADS, RAILWAYS, OR BRIDGES
- E01B—PERMANENT WAY; PERMANENT-WAY TOOLS; MACHINES FOR MAKING RAILWAYS OF ALL KINDS
- E01B19/00—Protection of permanent way against development of dust or against the effect of wind, sun, frost, or corrosion; Means to reduce development of noise
- E01B19/003—Means for reducing the development or propagation of noise
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Abstract
The invention relates to a track impact damping vibration absorber which comprises an elastic damping body (3), wherein a plurality of motion guide grooves (7) are arranged in the elastic damping body (3), a gap is formed between each motion guide groove (7) and a track waist surface (2) or an upper surface (8) of a track foot, at least one resonant mass body (5) is arranged in the gap, the resonant mass body (5), the motion guide grooves (7) and the elastic damping body (3) form a vibration assembly (6), and the vibration assembly (6) is closely attached to the track waist surface (2) or the upper surface (8) of the track foot. Compared with the prior art, the invention has the advantages of simple structure, realization of simultaneous vibration absorption in the vertical direction and the horizontal direction, good vibration absorption effect and the like.
Description
Technical Field
The invention belongs to the technical field of rail transit, and particularly relates to a rail impact damping vibration absorber technology which is used for weakening steel rail vibration and wheel rail rolling noise generated in the running process of a rail vehicle.
Background
The rail transit has the advantages of high transportation capacity, high timeliness, low pollution, low energy consumption and the like, and promotes the improvement of social productivity and the economic development while bringing convenience to people for going out. Along with the vigorous popularization and popularity of rail transit, the problem of vibration noise caused by rail transit is increasingly prominent. The vibration noise problem of track traffic not only can reduce the life and the operating mass of masses along the line, also can reduce passenger's in the car comfort level of taking.
In the running process of a train, the rail structure is excited by the interaction of wheel and rail to generate vibration and spread to the environment, which is a main source of rail transit vibration noise. The vibration noise control is the most effective method for vibration and noise reduction of rail transit from the source.
According to multiple researches, the rail vibration noise has the characteristics of wide frequency band and multiple frequency bands. The adoption of the track impact damping vibration absorber is one of effective methods for solving/reducing the problem of track system vibration noise from the source.
Patent document ZL201921890225.3 discloses a dynamic vibration absorber, the key structure includes a mass with through holes, a spring, a bolt, a damping element, and a steel plate, the structure is complicated, and the frequency band range is not considered. In the same section range, vibration in only a single direction can be absorbed.
Patent document ZL201710149926.0 discloses a multistage shear type steel rail dynamic damping vibration absorber. The main structure comprises a multi-stage resonance assembly consisting of a resonance mass block and an elastic damping layer, and vibration energy is consumed by only depending on the movement of the resonance mass block in the damping layer.
Disclosure of Invention
The invention aims to overcome the defects in the prior art and provide the track impact damping vibration absorber which has a simple structure, can realize simultaneous vibration absorption in the vertical direction and the horizontal direction and has a good vibration absorption effect.
The purpose of the invention can be realized by the following technical scheme: a vibration absorber with damping track impact is composed of an elastic damping body with multiple moving guide slots, at least one resonant mass body in said moving guide slots, and multiple resonant masses arranged in said moving guide slots in parallel1>0.01mm, the largest cross-sectional area of the resonant mass body and the largest gap formed by the steel rail, the elastic damping body and the motion guide grooveThe ratio u of the cross-sectional areas satisfies 0.01<<u<1. At least 4 resonance mass bodies are arranged in the elastic damping body with the same section, wherein two resonance mass bodies are arranged in the gap symmetrical to the rail waist surface on the two sides of the steel rail, and the other two resonance mass bodies are arranged in the gap symmetrical to the rail foot upper surface on the two sides of the steel rail. The resonance mass body, the motion guide groove and the elastic damping body form a vibration assembly, and the vibration assembly is closely attached to the rail waist surface or the upper surface of the rail foot.
The resonance mass body is composed of a plurality of independent mass blocks, and is arranged along the length direction of the steel rail to form a multi-stage resonance assembly of the multi-degree-of-freedom spring mass system along the track direction.
When the steel rail generates transverse and vertical vibration, each resonant mass block makes reciprocating motion between the steel rail and the motion guide groove, and when only the motion guide groove acts on the elastic damping body, each order of resonant frequency omegaaEquivalent stiffness K to elastic damping body1Equivalent modal mass m of motion guide slotcThe relationship is as follows:
when the resonant mass body is in contact with the motion guide groove and acts together with the elastic damping body, the resonant frequency omega of each orderbEquivalent stiffness K to elastic damping body2Equivalent modal mass m of a resonant masshEquivalent modal mass m of motion guide slotcThe relationship is as follows:
the resonant mass body and the motion guide groove transmit the vibration energy of the steel rail to the elastic damping body, and consume the vibration energy to form the impact damping vibration absorber.
Further, the resonant mass m of each resonant masssModal mass m corresponding to the peak vibration value of the railrRatio of mu to1＝ms/mrSatisfies 0.1<<μ1<1; resonance of each motion guideMass mgModal mass m corresponding to the peak vibration value of the railrRatio of mu to2＝mc/mgSatisfies 0.1<<μ2<The steel rail, the resonant mass body, the elastic damping body and the motion guide groove form a multi-degree-of-freedom spring mass resonant system.
Further, the elastic damping body mass mtModal mass m with railrRatio of mu to3＝mt/mrSatisfy 0.01<<μ3<1, the damping loss factor range of the elastic damping body is 0.01-0.5.
By adding the equivalent modal mass m of a single resonant masshEquivalent modal mass m of motion guide slotcThe suppression of the low-frequency vibration of the steel rail can be realized.
The resonance frequency band range of the multi-stage resonance assembly consisting of the resonance mass block, the motion guide groove and the elastic damping body is a broadband or a plurality of segmented frequency bands with certain bandwidths.
Furthermore, the outer side of the vibration assembly is provided with a limiting constraint part, the limiting constraint part can be connected and positioned with the steel rail through a metal buckle, the limiting constraint part and the elastic damping body are fixedly connected into a whole in a vulcanization or bonding mode, or the limiting constraint part is lapped on the elastic damping body, and the material of the limiting constraint part can be metal materials or non-metal materials such as 304 stainless steel.
Furthermore, the elastic damping body is in a partial contact mode of not completely covering the upper surface of the rail foot of the steel rail and the waist surface of the steel rail, the contact surface can be provided with grooves, and the grooves are net-shaped or nail column-shaped.
Furthermore, the resonance mass bodies are distributed along the length direction of the steel rail, and the maximum sectional area of the resonance mass bodies is smaller than the sectional area of the elastic damping body, the motion guide groove and the steel rail in the direction vertical to the length direction of the steel rail;
furthermore, the density and the hardness of the material of the resonance mass block can be larger than, smaller than or equal to those of the material of the steel rail, and the shape of the resonance mass block is circular, oval, rectangular or polygonal.
Furthermore, the motion guide groove is firmly embedded in the elastic damping body in an adhesion or vulcanization mode, and the material of the motion guide groove can be metal materials such as 304 stainless steel and the like or non-metal materials such as plastics and nylon and the like. The cross section of the motion guide groove is of a structure with an opening at one end, and the structure comprises a C shape, an A shape, an F shape, a Pi shape, an L shape or an I shape, and the equivalent stiffness of the motion guide groove is greater than that of the elastic damping body.
The impact damping vibration absorber is fixed on the steel rail through an elastic metal buckle or is connected with the steel rail in a bonding mode.
Compared with the prior art, the invention has the following beneficial effects:
1. the vibration combination body is simple in structure, only comprises the resonance mass body and the motion guide groove, and is easy to realize. In the same section, the vibration absorption in the vertical direction and the horizontal direction can be realized simultaneously. The resonance mass body and the motion guide groove can play a role in absorbing vibration, so that the vibration absorption effect of the vibration absorber is enhanced, and the application range of the vibration absorber is easy to adjust.
2. The applicable frequency ranges of the invention are: the transverse vibration of the steel rail is 200Hz-1500Hz, and the vertical vibration of the steel rail is 200Hz-1500 Hz. The expected vibration speed level of the steel rail is reduced by 10dB-15dB, and the noise radiation level caused by the corresponding steel rail vibration is reduced by 3dB (A) -8dB (A).
Drawings
FIG. 1 is a structural sectional view of embodiment 1 of the present invention;
FIG. 2 is a front perspective structural view of the vibrating assembly of embodiment 1;
FIG. 3 is a side view of the structure of example 1 containing elastic fasteners;
FIG. 4 is an isometric structural view of example 1 containing elastic fasteners;
FIG. 5 is a front perspective structural view of the vibration composite body of embodiment 2 including a plurality of movement guide grooves;
FIG. 6 is a structural sectional view of embodiment 3;
FIG. 7 is a structural sectional view of embodiment 4.
1. A steel rail; 2, rail waist surface; 3. an elastic damping body; 4. a limiting restraint part; 5 a resonant mass; 6 vibrating the combination; 7. a motion guide groove; 8. the upper surface of the rail foot; 9. a metal buckle; 10. a metal separator; 11. vibrating assembly (under rail); 12. a resonant mass block; 13. a plurality of resonant masses.
Detailed Description
The invention is described in detail below with reference to the figures and specific embodiments.
Example 1:
as shown in fig. 1, an orbital impact-damped vibration absorber includes an elastic damping body 3. The elastic damping body 3 is internally provided with a plurality of motion guide grooves 7, the motion guide grooves 7 are firmly embedded in the elastic damping body 3 in a vulcanization mode, and the material of the motion guide grooves 7 can be 304 stainless steel. A gap is formed between each motion guide groove 7 (the structure is in a C shape) and the rail waist surface 2 or the rail foot upper surface 8, at least one resonant mass body 5 is arranged in the gap, in the embodiment, 4 resonant mass bodies 5 are arranged in the elastic damping body 3 with the same section, two of the resonant mass bodies are arranged in the gap symmetrical to the rail waist surface 2 at two sides of the steel rail 1, and the other two resonant mass bodies are arranged in the gap symmetrical to the rail foot upper surface 8 at two sides of the steel rail. The resonance mass body 5, the motion guide slot 7 and the elastic damping body 3 form a vibration assembly 6, the vibration assembly 6 is closely attached to the rail waist surface 2 or the upper surface 8 of a rail foot, the outer side of the vibration assembly 6 is provided with a limit restraint part 4, the limit restraint part 4 can be connected and positioned with a steel rail through a metal buckle 9, the limit restraint part 4 and the elastic damping body 3 are fixedly connected into a whole in a bonding mode, and the material of the limit restraint part can be 304 stainless steel. The elastic damping body 3 in the vibration assembly 6 is in a partial contact mode of not completely covering the upper surface 8 of the rail foot of the steel rail 1 and the waist surface 2 of the steel rail, and the contact surface is provided with a reticular groove.
As shown in fig. 2, the resonant masses 5 are arranged along the length of the rail to form a multi-degree-of-freedom spring mass system having a wide frequency band or a plurality of segmented frequency bands with a certain bandwidth. When the rail vibrates vertically and transversely, each resonant mass 5 reciprocates among the rail 1, the motion guide groove 7 and the elastic damping body 3. When the resonant mass 5 does not collide with the motion guide groove 7, the motion guide groove 7 can also pull and press the elastic damping body 3, and the resonant frequency omega of each stepaEquivalent stiffness K to the elastic damping body 31Motion guide groove 7 equivalent modal mass mcThe relationship is as follows:
when the resonant mass 5 collides with the motion guide groove 7 and presses the elastic damping body 3 together with the motion guide groove 7, the resonant frequency ω of each stepbEquivalent stiffness K to the elastic damping body 32Equivalent modal mass m of the resonant mass 5hMotion guide groove 7 equivalent modal mass mcThe relationship is as follows:
both of these movements dissipate the vibration energy of the rail and inhibit the propagation of vibrations along the length of the rail.
The non-steel rail contact surface in the outside of the vibration assembly 6 is provided with the limiting and restraining part 4, the vertical and transverse vibration displacement of the assembly can be limited, the wheel cannot be influenced through production, the vibration reduction effect of the impact damping vibration absorber cannot be influenced, and meanwhile the limiting and restraining part 4 can also play a role in protecting the vibration assembly 6.
As shown in fig. 3-4, the impact-damping vibration absorber is fixed on the steel rail 1 through a metal buckle 9 with elasticity, and is connected and positioned with the steel rail.
The resonant mass m of each resonant mass 5sModal mass m corresponding to the peak vibration value of the rail 1rRatio of mu to1＝ms/mrSatisfies 0.1<<μ1<1; resonant mass m of each motion guide 7gModal mass m corresponding to the peak vibration value of the rail 1rRatio of mu to2＝mg/mrSatisfies 0.1<<μ2<The steel rail, the resonant mass body, the elastic damping body and the motion guide groove form a multi-degree-of-freedom spring mass resonant system.
By adding the equivalent modal mass m of a single resonant mass 5hEquivalent modal mass m of motion guide groove 7cAnd the low-frequency vibration of the steel rail can be restrained.
The mass, stiffness and number of the resonator mass 5, motion guide 7 and elastic damping body 3 can be designed according to the above requirements, e.g.: taking the equivalent modal mass m of the resonant mass 5h2kg, equivalent modal mass m of the motion guide channel 7cEquivalent stiffness K of the elastic damping body 3 of 4kg2＝1╳107N/m, resonant frequency ωb205.6 Hz. The prior art does not set reciprocating motion and the equivalent modal mass is mhThe resonance mass block is only completely wrapped by the elastic damping body and has the equivalent modal mass of mcThe movement guide groove. By adopting two equivalent modal masses, the suppression of low-frequency vibration of the steel rail, such as vibration with the frequency of 200Hz, can be realized.
The invention can realize the inhibition of the transverse and vertical vibration of the steel rail at 200Hz-1500Hz, and compared with the prior art, the invention widens the inhibition range of the vibration frequency of the steel rail. The expected vibration speed level of the steel rail is reduced by 10dB-15dB under the structure of the embodiment, and the noise radiation level caused by the corresponding steel rail vibration is reduced by 3dB (A) -8dB (A).
Example 2:
this application example is substantially the same as the above example 1, and as shown in fig. 5, the difference is that the motion guide groove inside the vibration assembly 6 is not a completely through single structure in example 1, and includes a plurality of motion guide grooves, and the cross section of the motion guide groove may be C-shaped, a-shaped, F-shaped, pi-shaped, L-shaped, i-shaped, and the like, and has a structure with one open end. The number of the motion guide grooves can be any integer from 1 to 100. The number of the resonance bodies in each motion guide groove can be any integer from 1 to 500, a plurality of resonance mass bodies are arranged in parallel in the motion guide groove, and the adjacent resonance mass bodies are arranged between the resonance mass bodiesLeaving a gap epsilon1>0.01mm, the ratio u of the maximum sectional area of the resonant mass body 5 to the maximum sectional area of the gap formed by the steel rail 1, the elastic damping body 3 and the motion guide groove 7 meets the requirement that 0.01 < ><1. The relative height and horizontal relative distance of the motion guide groove can be adjusted within the range of 0.01mm-300 mm.
This configuration can adjust the resonant frequency range of the damper. An elastic damping body is arranged between the adjacent motion guide grooves, and relative motion is generated between the motion guide grooves, so that the elastic damping body generates shear deformation, and the elastic damping body provides elastic restoring force for the motion guide grooves.
Example 3:
this application example is basically the same as the above example 1, and as shown in fig. 6, the difference is that the vibration assembly 11 having a similar structure to that of example 1 is mounted on the bottom of the rail and closely contacts with the bottom of the rail 1, and the metal clip 9 having elasticity closely contacts the vibration assembly 6 with the rail web surface 2 and the vibration assembly 11 under the rail foot with the rail bottom without any gap. The metal partition 10 is arranged in the middle of the vibration assembly (below the rail foot) 11, and the metal partition 10 is aligned with the middle of the rail. The vibration assembly 11 is provided with an L-shaped motion guide groove, and the resonance mass body 5 can generate vertical and horizontal motions in a space formed by the L-shaped motion guide groove, the metal partition body 10 and the steel rail 1, so that the control of the freedom degree vibration of the steel rail 2 is realized.
The vibrating assembly 6 contains a plurality of resonant masses 12, and the resonant masses 12 are arranged along the length direction of the rail, and the cross section of the resonant masses 12 can be circular, rectangular, oval, polygonal and the like. The number of resonator masses may be any integer from 0 to 100. The resonance mass block can generate vertical and horizontal movements in the vibration combination body, and the control of two-degree-of-freedom vibration of the steel rail is realized.
Example 4:
this applied embodiment is substantially the same as embodiment 3 described above, with the difference that there are a plurality of cross-sectional shapes and sizes of the resonant mass bodies 13 in the motion guide channel, as shown in fig. 7, and the number thereof may be any integer from 1 to 500. By varying the number of resonant masses, the frequency range of action of the impact damper can be adjusted.
Claims (10)
1. The utility model provides a track shock-damping vibration absorber, includes elastic damping body (3), characterized in that, elastic damping body (3) in be equipped with a plurality of motion guide slots (7), each motion guide slot (7) and rail waist face (2) or with the above-mentioned (8) of rail foot constitute the space, be equipped with at least one resonance mass body (5) in this space, resonance mass body (5), motion guide slot (7) and elastic damping body (3) constitute vibration assembly (6), vibration assembly (6) laminate closely on rail waist face (2) or with above-the-mentioned (8) of rail foot.
2. The vibration absorber as claimed in claim 1, wherein the plurality of resonant masses (5) and the motion guide channel (7) form a plurality of resonant frequencies, each resonant frequency being ω when only the motion guide channel (7) acts on the elastic damping body (3)aEquivalent stiffness K to the elastic damping body (3)1Equivalent modal mass m of motion guide groove (7)cThe relationship is as follows:
3. a vibration absorber with damping of orbital impacts according to claim 1 characterized by the fact that the multiple resonant masses (5) and the motion guide (7) form a multi-step resonance frequency, the resonance frequency ω being each step when the resonant masses (5) are in contact with the motion guide (7) and co-act with the elastic damping body (3)bEquivalent stiffness K to the elastic damping body (3)2Equivalent modal mass m of the resonant mass (5)hEquivalent modal mass m of motion guide groove (7)cThe relationship is as follows:
4. the rail impact damping vibration absorber of claim 1 wherein eachResonance mass m of a resonance mass (5)sModal mass m corresponding to the peak value of vibration of the rail (1)rRatio of mu to1＝ms/mrSatisfies 0.1 & lt mu1Less than 1; resonant mass m of each motion guide slot (7)gModal mass m corresponding to the peak value of vibration of the rail (1)rRatio of mu to2＝mc/mgSatisfies 0.1 & lt mu2＜1。
5. The rail impact-damped vibration absorber of claim 1 wherein the elastic damping body (3) has a mass mtModal mass m of steel rail (1)rRatio of mu to3＝mt/mr0.01 < mu3The damping loss factor range of the elastic damping body (3) is 0.01-0.5, and the resonant mass body (5), the steel rail (1), the elastic damping body (3) and the motion guide groove (7) form a multi-degree-of-freedom spring mass resonant system.
6. The shock-absorbing and damping device for track according to claim 1 wherein the vibration assembly (6) is provided with a position-limiting restraining member (4) at its outer side, and is fixedly connected to the rail (1) by a metal clip (9), and the position-limiting restraining member (4) and the elastic damping body (3) are integrally fixed by vulcanization or adhesion, or the position-limiting restraining member (4) is lapped on the elastic damping body (3).
7. The vibration absorber as claimed in claim 1, wherein the contact surface of the elastic damping body (3) with the rail waist surface (2) or the rail foot upper surface (8) is provided with grooves, and the grooves are net-shaped or stud-shaped.
8. The shock-absorbing device for track impact according to claim 1 wherein the resonant mass (5) is distributed along the length of the rail, the maximum cross-sectional area of the resonant mass (5) being smaller than the cross-sectional area of the elastic damping body (3), the motion guide groove (7), and the gap formed by the rail (1) in the direction perpendicular to the length of the rail; the cross section of the resonance mass body (5) is circular, elliptical, rectangular or polygonal.
9. The shock absorber of claim 1 wherein the motion guide channel (7) is firmly embedded in the elastic damping body (3) by bonding or vulcanization, and the cross-sectional shape of the motion guide channel (7) is a structure with an open end, including a C-shape, an a-shape, an F-shape, a pi-shape, an L-shape or an i-shape, and the equivalent stiffness of the motion guide channel is greater than the equivalent stiffness of the elastic damping body (3).
10. The shock-absorbing absorber of claim 1 wherein the shock-absorbing absorber is fixed to the rail (1) by means of a resilient metal clip (9) or is adhesively attached to the rail (1).
Priority Applications (1)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| CN202011497758.2A CN112575630A (en) | 2020-12-17 | 2020-12-17 | Track impact damping vibration absorber |
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| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| CN202011497758.2A CN112575630A (en) | 2020-12-17 | 2020-12-17 | Track impact damping vibration absorber |
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| Publication Number | Publication Date |
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| CN112575630A true CN112575630A (en) | 2021-03-30 |
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| Application Number | Title | Priority Date | Filing Date |
|---|---|---|---|
| CN202011497758.2A Pending CN112575630A (en) | 2020-12-17 | 2020-12-17 | Track impact damping vibration absorber |
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| CN (1) | CN112575630A (en) |
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2020
- 2020-12-17 CN CN202011497758.2A patent/CN112575630A/en active Pending
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