A random detuning track vibration reduction system and its design method

Abstract

The present invention discloses a random detuning track vibration reduction system and a design method thereof. The structure includes rail members and vibration reduction units. The rail members are arranged in groups and at intervals on the track plate. Multiple groups of vibration reduction units are arranged on the inner and outer sides of the rail members or on the track plate. The multiple groups of vibration reduction units on the same side of the rail member correspond to the longitudinal quasi-periodic distribution of the rail member. The vibration reduction units on both sides of the rail member are arranged symmetrically or staggered. The main frequency of the vibration reduction unit is close to the band gap frequency of the track structure, and the vibration mode of the vibration reduction unit is coupled with the vibration mode of the track structure, so that the resonance structure of the vibration reduction unit and the ballastless track structure produce a coupling effect, amplifying the band gap range, and the mass, stiffness and installation position parameters of the vibration reduction unit are designed to be randomly distributed. The random detuning design method proposed in the present invention can further broaden the vibration reduction range, will not cause the vibration of the track structure itself to increase, and does not need to replace the vibration reduction unit as the service time increases or the environment changes, effectively providing vibration reduction capabilities.

Description

Technical Field

[0001] The present invention belongs to the technical field of rail transit vibration reduction and noise reduction, and in particular relates to a random detuning track vibration reduction system and a design method thereof. Background Art

[0002] With the rapid development of urban rail transit, train speeds have increased, and the impact on environmental vibration and noise has become increasingly significant. The demand for urban rail transit track vibration and noise reduction is prominent. Current track vibration reduction measures mainly focus on vibration isolation and dynamic vibration absorption. Vibration isolation measures, such as vibration-damping fasteners and floating slab roadbeds, reduce the transmission of vibration to the outside world by inserting elastic layers between track structure layers. However, this also increases the vibration of the track structure. For example, vibration-damping fasteners amplify rail vibration, and floating slab roadbeds amplify roadbed structure vibration. Dynamic vibration absorption measures, such as rail dynamic vibration absorbers and track slab dynamic vibration absorbers, can transfer track structure vibration energy to additional devices. However, these absorbers typically operate at a single frequency and have a narrow vibration reduction frequency band. As their service life increases, the performance of the elastic layer deteriorates, and the vibration reduction effect decreases. Therefore, it is necessary to propose a measure that can achieve broadband vibration reduction of the track structure, achieving both environmental vibration control and reducing track structure vibration.

[0003] In order to enhance the vibration reduction effect of the vibration reduction measures, relevant researchers and technicians have carried out extensive optimization design based on the existing vibration reduction fasteners, steel spring / vibration reduction pad floating plates, rail vibration absorbers, and floating plate vibration absorbers. This is mainly reflected in the following aspects:

[0004] Frequency-adjustable rail dynamic vibration absorbers: Existing technologies pre-design the absorber's mass or stiffness to create a resonant frequency that matches the rail's vibration frequency, thereby transferring and dissipating the vibration energy of the wheel-rail primary vibration system. However, during train operation, the wheel excitation of the rail is a continuously changing broadband excitation. Rail grinding or track reconstruction, for example, can cause changes in the wheel-rail excitation frequency. Existing rail tuned vibration absorbers often fail to achieve secondary frequency modulation after installation, are cumbersome and complex to replace on-site, and have limited rail noise reduction effectiveness.

[0005] Chinese patent CN114411456A discloses a mass-adjustable rail vibration absorber, which changes the operating frequency of the rail vibration absorber by replacing an adjusting member of a specific mass, thereby improving the working efficiency of secondary frequency modulation.

[0006] Chinese patent CN211009410U discloses a clamp for installing a frequency-modulated rail vibration absorber, which facilitates the disassembly and assembly of the dynamic vibration absorber.

[0007] Chinese patent CN117847130A discloses a dynamic vibration absorber with adjustable frequency and mass. The stiffness is adjusted by compressing a rubber layer in the vibration absorber, and different vibration absorption frequencies are obtained by installing mass blocks of different masses.

[0008] Chinese patent CN116516736A discloses an adjustable rail vibration absorber, which changes the operating frequency of the vibration absorber by adjusting the number of counterweights.

[0009] Multi-frequency vibration absorber structure: To achieve vibration control in a wider frequency range, the existing vibration absorber structure has a single operating frequency. Chinese patent CN111778781A discloses a rigidly connected wide-band rail dynamic vibration absorption device, in which vibration absorbers are arranged on the left and right sides of the rail, which can control at least two or more vibration frequencies. Chinese patent CN213328472U discloses a multi-stage rail vibration absorber, which adjusts the operating frequency by combining the mass block with the horizontal beam and adjusting the position of the mass block. Chinese patent CN201020257075.5 discloses a rail waist vibration absorber, in which three metal mass blocks are embedded in the rail waist rubber part, and the three metal blocks have different masses to improve the vibration reduction effect.

[0010] Bandgap vibration damping structure based on phononic crystal theory: The phononic crystal structure is a structure with periodic characteristics. Due to its periodic characteristics, the vibration propagation in the phononic crystal has a bandgap effect, that is, the vibration within a specific frequency range cannot be transmitted in the phononic crystal.

[0011] Chinese patent CN110528340A discloses a phononic crystal floating plate isolator, which suppresses the transmission of floating plate vibration to the substrate by designing the steel spring into a phononic crystal structure with periodic characteristics, thereby achieving vibration reduction.

[0012] Chinese patent CN218294321U discloses a phononic crystal rail dynamic vibration absorber, which has a similar principle to that of a dynamic vibration absorber. It uses a mass block and an elastic layer to achieve vibration reduction, and increases the vibration reduction frequency band by setting hollow steel columns of different heights.

[0013] Chinese patent CN111778783A proposes a dynamic vibration absorber with an embedded phononic crystal structure. By setting a periodically distributed multi-layer structure on one side of the rail waist and then installing a mass body on the outside, a vibration reduction structure is formed. However, this structure still belongs to the design of a vibration absorber.

[0014] Existing vibration reduction methods for ballastless track structures rely on dynamic vibration absorption, elastic vibration isolation, and phononic crystal band gap designs. However, dynamic vibration absorbers have a narrow frequency band and rely primarily on adjusting the mass to change the frequency. Vibration isolation methods under rails or floating slabs can lead to increased vibration in the rail or trackbed structure. Vibration reduction designs based on phononic crystal band gaps require that the vibration transmission path align with the periodic arrangement of the structure, and the band gap width is generally narrow. To achieve vibration reduction across a wider frequency range while avoiding increased vibration in the ballastless track structure due to vibration reduction measures, new vibration reduction design methods are needed.

[0015] Therefore, how to provide a track vibration reduction system with a wide-band vibration reduction effect and a design method thereof is an urgent problem that those skilled in the art need to solve. Summary of the Invention

[0016] In light of this, the present invention provides a random detuning track vibration reduction system and its design method. Existing ballastless track vibration reduction designs, such as those using dynamic vibration absorbers, suffer from the drawbacks of a single control frequency and diminishing vibration reduction effectiveness with installation conditions, environmental conditions, or service life. By introducing a random detuning system into the ballastless track structure and coupling it with the periodic characteristics of the ballastless track structure, a broadband vibration reduction band gap is generated, achieving ballastless track vibration and noise control over a wide frequency range.

[0017] In order to achieve the above-mentioned object, the present invention adopts the following technical solution: a random detuning track vibration reduction system, comprising:

[0018] Rail members, the rail members are arranged in groups and at intervals on the track plate, and a fastener system is arranged between the rail members and the track plate;

[0019] A vibration damping unit, wherein the vibration damping units are provided in multiple groups, and the multiple groups of the vibration damping units are arranged on the inner and outer sides of the rail or on the track plate. The multiple groups of vibration damping units on the same side of the rail are distributed quasi-periodically in the longitudinal direction of the rail. The vibration damping units on both sides of the rail are arranged symmetrically or staggered. The main frequency of the vibration damping unit is close to the band gap frequency of the track structure, and the vibration mode of the vibration damping unit is coupled with the vibration mode of the track structure, so that the resonance structure of the vibration damping unit and the ballastless track structure produce a coupling effect, thereby enlarging the band gap range.

[0020] The beneficial effects of the present invention are as follows: the present invention does not weaken the integrity and strength of the track structure, the present invention is easy to install, and the vibration energy dissipation of the track structure is achieved by using the vibration reduction unit. At the same time, the vibration reduction unit exhibits a nearly periodic distribution characteristic in the longitudinal direction of the corresponding rail member, avoiding the generation of new resonant frequencies after the introduction of the vibration reduction measure, while broadening the vibration control frequency range of the vibration reduction measure. The present invention is based on the concept of random detuning design and combines it with the band gap theory of phononic crystals. The band gap of the elastic wave of the phononic crystal is coupled with the band gap of the ballastless track structure itself to achieve vibration control of the track structure. The random detuning parameters of the vibration reduction structure are further designed to produce a coupling effect between the resonant structure of the vibration reduction unit and the ballastless track structure, suppressing the resonance peak and broadening the vibration reduction frequency band, thereby maximizing the vibration effect. The present invention breaks the idea of ​​periodic arrangement in the existing track structure vibration reduction design and proposes a random detuned vibration reduction system design method suitable for ballastless track structure. Combining the periodic characteristics of the track structure itself, the present invention arranges approximately periodic detuned vibration reduction units in the track structure to broaden the vibration reduction frequency band, reduce the vibration of the track structure itself, and reduce problems such as rail corrugation and resonance fracture of fastener spring bars caused by vibration, while reducing the transmission of track structure vibration to the outside world.

[0021] Preferably, the vibration damping unit includes a first type of vibration damping unit and a second type of vibration damping unit, the first type of vibration damping unit is arranged on the rail member, and the second type of vibration damping unit is arranged on the track plate.

[0022] The resulting technical effect is that the vibration reduction unit of the present invention actually has two structural forms. The first structural form is suitable for installation on the rail member, and the second structural form is suitable for installation on the track plate. Regardless of the structural form, the vibration control of the track structure is achieved, the resonance peak is suppressed and the vibration reduction frequency band range is widened.

[0023] Preferably, the first type of vibration reduction unit includes a viscoelastic body, a constraint layer and a scatterer, the viscoelastic body is fixed to the side of the rail, the constraint layer is fixed to the outside of the viscoelastic body away from the rail, and a plurality of grooves are provided inside the viscoelastic body, and the plurality of grooves pass through both ends of the viscoelastic body, and the scatterers are in multiple groups and are respectively installed in the grooves.

[0024] The resulting technical effect is that the viscoelastic body, the constraining layer and the scatterer are attached to the rail member, thereby generating a resonance effect to absorb and dissipate the vibration energy of the rail member, thereby reducing the outward transmission of the vibration energy of the track structure.

[0025] Preferably, the viscoelastic body is rubber, polyurethane or a viscoelastic damping material, the scatterer is a metal structure, and the constraining layer is a hard material plate.

[0026] Preferably, the viscoelastic body is divided into multiple substructures in the axial direction, and the cross section of the viscoelastic body is divided into a first half zone and a second half zone, the first half zone is close to the rail side, and the second half zone is close to the constraint layer side, the first half zone is provided with at least a first slot cavity, a second slot cavity and a third slot cavity in sequence from top to bottom, and the second half zone is provided with at least a fourth slot cavity, a fifth slot cavity and a sixth slot cavity in sequence from top to bottom, the scatterer includes at least a first scatterer, a second scatterer, a third scatterer, a fourth scatterer, a fifth scatterer and a sixth scatterer and are respectively installed in the first slot cavity, the second slot cavity, the third slot cavity, the fourth slot cavity, the fifth slot cavity and the sixth slot cavity, a plurality of the second scatterers and a plurality of the sixth scatterers connect the multiple substructures in series, the first scatterer, the third scatterer, the fourth scatterer and the fifth scatterer correspond to the substructures, and the first scatterer, the third scatterer, the fourth scatterer and the fifth scatterer in adjacent substructures are periodically spaced in the axial direction.

[0027] The resulting technical effect is that the second scatterer and the sixth scatterer are actually continuous structures, the purpose of which is to connect multiple sub-structures in series, and the continuous structure is used to guide the action of elastic waves, effectively transferring the vibration of the rail to the additional vibration damping unit, and then the scatterers and viscoelastic bodies on the vibration damping unit vibrate to dissipate energy.

[0028] Preferably, at least two groups of vibration-damping units on the same side of the rail constitute an installation group, and the vibration-damping units in the same installation group are connected in series through the second scatterer or the sixth scatterer. One side of the rail has multiple installation groups, and the second scatterers between adjacent installation groups abut against each other.

[0029] The resulting technical effect is: in actual implementation, several groups of vibration reduction units are usually combined into a mounting structure, multiple second scatterers in a mounting structure are fixedly connected, and then the second scatterers between adjacent mounting structures are abutted.

[0030] Preferably, the second type of vibration damping unit includes a frame structure, an elastic layer and a mass body, the frame structure is a long strip structure and is fixed to the track plate in the length direction of the corresponding rail member, a plurality of square grooves are arranged side by side on the frame structure, the mass body is fixed in the square grooves, and the elastic layer is arranged between the mass body and the side walls of the square grooves.

[0031] The resulting technical effect is: on the basis of the setting of the first type of vibration damping unit, in order to improve the vibration damping effect, a second type of vibration damping unit is usually arranged on the track plate, thereby improving the vibration damping effect of the overall structure.

[0032] Preferably, a vibration gap of 10 mm is reserved between the elastic layer, the mass body and the track plate respectively.

[0033] The resulting technical effect is that maintaining a gap between the mass, elastic layer, and track plate is a prerequisite for the vibration reduction unit to achieve its vibration reduction effect. Furthermore, the shear deformation of the elastic layer is utilized to achieve lower support stiffness, thus achieving track plate vibration control in a lower frequency range.

[0034] Preferably, the mass bodies in the plurality of square slots in the same row have different sizes, the masses of the mass bodies in the adjacent square slots corresponding to the same row satisfy a linear variation rule, and the masses of the mass bodies in the corresponding square slots in adjacent rows are the same or different.

[0035] The resulting technical effect is to prevent resonance between the attached vibration-damping structure and the track structure. This can be understood as each substructure controlling vibrations within a specific frequency band. If all substructures were arranged identically, the frequency band would not be widened. Therefore, the mass or stiffness of adjacent substructures is randomly detuned, shifting the control frequency bands of adjacent substructures to widen the vibration control range.

[0036] The present invention also discloses a design method for a random detuning track vibration reduction system, which comprises the following steps:

[0037] Step 1: Identify the vibration band gap characteristics of the ballastless track

[0038] Conduct an analysis of the ballastless track structure's periodic characteristics and attenuation domain. Based on the track structure's periodic characteristics, fastener parameters, and ballastless track bed structural parameters, calculate the elastic wave band gap frequency range of the ballastless track structure.

[0039] Step 2: Design the basic vibration damping unit

[0040] The vibration reduction unit is a local resonance structure unit. It consists of a scatterer, a viscoelastic body and a constraint layer. The initial design parameters of the vibration reduction unit, including mass, elastic layer modulus and installation position, are designed based on the control target frequency and the mass-spring system. The formula is used.

[0041]

[0042] Preliminary design values ​​for the mass of each scatterer and the elastic modulus of the elastic layer are given. The parameters of each mass body are obtained through dynamic calculations, with f being the control target frequency and k being the elastic coefficient. A refined finite element model of the rail-vibration damping unit is then established, and the effects of parameters such as mass, stiffness, and installation position on the control frequency band are analyzed to determine the initial design parameters of the vibration damping unit. The main frequency of the vibration damping unit is close to the band gap frequency of the track structure, so that the vibration damping unit structure attached to the rail member or track plate has a coupling effect with the ballastless track structure, maximizing the band gap range.

[0043] Step 3: Design the structural parameters of the detuned vibration reduction unit

[0044] The parameters of the vibration reduction unit structure are randomly detuned, including the weight of the scatterer, the stiffness of the viscoelastic body, and the installation position. These parameters follow a uniform distribution with a mean of A1 and a coefficient of variation of δ1. The mean is the initial design value before detuning is considered. The coefficient of variation is obtained in the second step and is empirically determined to be 0.1 to 0.5. The detuning parameters and degree of detuning are determined by objective functions such as avoiding structural resonance after installing the vibration reduction unit and increasing the vibration reduction frequency range.

[0045] Step 4: Add other control bands that require attention

[0046] According to the target control frequency band, the second or third type of vibration reduction unit is set. The second or third type of vibration reduction unit also adopts a random detuning design, and the key parameters of the uniform distribution are set to A2 or A3 and the coefficient of variation is δ2 or δ3;

[0047] Step 5: Evaluate the vibration reduction effect of the vibration reduction system

[0048] The vibration attenuation domain and vibration response characteristics of the ballastless track random detuning vibration reduction system are analyzed, and the vibration reduction effect of the random detuning vibration reduction system is predicted.

[0049] The beneficial technical effect of the present invention is: based on the random detuning design concept and combined with the phononic crystal band gap theory, the phononic crystal elastic wave band gap is coupled with the band gap of the ballastless track structure itself to achieve track structure vibration control, and further design the random detuning parameters of the vibration reduction structure to suppress the resonance peak and widen the vibration reduction frequency band range.

[0050] Preferably, in the step one, based on the phononic crystal elastic wave theory, the elastic wave band gap range of the ballastless track structure itself is calculated, and the frequency and vibration mode at the elastic wave band gap boundary are obtained; in the step two, the main frequency of the vibration reduction unit is close to the band gap frequency of the track structure, and the vibration mode of the vibration reduction unit is coupled with the significant vibration mode of the track structure, so that the additional local resonance structure and the ballastless track structure produce a coupling effect, thereby maximizing the band gap range; in the step five, first, the vibration attenuation domain range of the ballastless track random detuning vibration reduction system is analyzed to evaluate the vibration reduction range, and then the vibration response of the track structure under the action of simple harmonic load is calculated to evaluate that no obvious resonance response occurs in the structure, and finally, a dynamic coupling analysis model of the vehicle-track-random detuning vibration reduction system is established, and the vibration reduction effect of the random detuning vibration reduction system is predicted through the vibration response of the track structure and the vibration response of the structure under the track plate. BRIEF DESCRIPTION OF THE DRAWINGS

[0051] Figure 1 This is a structural diagram of a random detuning track vibration reduction system of the present invention;

[0052] Figure 2 This is a partial enlarged view of a random detuning track vibration reduction system of the present invention;

[0053] Figure 3 This is a structural diagram of a vibration reduction unit of a random detuning track vibration reduction system according to the present invention;

[0054] Figure 4 This is a diagram showing the layout of scatterers inside a vibration reduction unit of a random detuning track vibration reduction system according to the present invention;

[0055] Figure 5 This is an application diagram of the second type of vibration reduction unit of a random detuning track vibration reduction system of the present invention;

[0056] Figure 6 This is the structural diagram of the traditional ballastless track;

[0057] Figure 7 This is the vibration attenuation range of traditional ballastless track;

[0058] Figure 8The figure shows a comparison of the vibration reduction effects of the present invention and a conventional dynamic vibration absorber.

[0059] 1 Rail, 2 Fastener system, 3 Track plate, 4 Vibration damping unit, 41 First type vibration damping unit, 411 Viscoelastic body, 4111 First slot cavity, 4112 Second slot cavity, 4113 Third slot cavity, 4114 Fourth slot cavity, 4115 Fifth slot cavity, 4116 Sixth slot cavity, 412 Constraint layer, 413 Scatterer, 4131 First scatterer, 4132 Second scatterer, 4133 Third scatterer, 4134 Fourth scatterer, 4135 Fifth scatterer, 4136 Sixth scatterer, 42 Second type vibration damping unit, 421 Frame structure, 422 Elastic layer, 423 Mass body. DETAILED DESCRIPTION

[0060] The following will clearly and completely describe the technical solutions in the embodiments of the present invention in conjunction with the accompanying drawings. Obviously, the described embodiments are only part of the embodiments of the present invention, not all of the embodiments. Based on the embodiments of the present invention, all other embodiments obtained by ordinary technicians in this field without making creative efforts are within the scope of protection of the present invention.

[0061] See the attached Figures 1 to 8 According to an embodiment of the present invention, a random detuning track vibration reduction system includes:

[0062] Rail members 1, which are actually steel rails, are arranged in groups and at intervals on a track slab 3, which is a ballastless track slab. A fastener system 2 is arranged between the rail members 1 and the track slab 3;

[0063] The vibration reduction unit 4 is actually an additional local vibration reduction unit. There are multiple groups of vibration reduction units 4, which are arranged on the inner and outer sides of the track member 1 or the track plate. In other words, they can be arranged on the track member, on the track plate, or on both the track member and the track plate.

[0064] The multiple groups of vibration-damping units 4 on the same side of the rail member 1 correspond to the longitudinal (i.e., axial) quasi-periodic distribution of the rail member 1, which can be understood as an approximately periodic distribution. This is to avoid resonance between the additional local vibration-damping units and the periodic track structure. The vibration-damping units on both sides of the rail member 1 are arranged symmetrically or staggered. This requires consideration of the structural form of the rail member. When the fastener is a symmetrical structure, the vibration-damping units on both sides of the rail member can be staggered or symmetrically arranged; the main frequency of the vibration-damping unit 4 is close to the band gap frequency of the track structure, and the vibration mode of the vibration-damping unit is coupled with the vibration mode of the track structure, so that the resonance structure of the vibration-damping unit and the ballastless track structure produce a coupling effect, amplify the band gap range, and reduce the transmission of the track structure vibration to the outside world.

[0065] Specifically, the vibration damping unit 4 includes a first type of vibration damping unit 41 and a second type of vibration damping unit 42 . The first type of vibration damping unit 41 is arranged on the rail member 1 , and the second type of vibration damping unit 42 is arranged on the track plate 3 .

[0066] The first type of vibration reduction unit 41 includes a viscoelastic body 411, a constraint layer 412 and a scatterer 413. The viscoelastic body 411 is fixed on the side of the rail 1. It can be connected by adhesive material or fixed on the rail using a clamp. The constraint layer 412 is fixed on the outside of the viscoelastic body 411 away from the rail 1. The interior of the viscoelastic body 411 is provided with multiple grooves, and the multiple grooves pass through both ends of the viscoelastic body 411. There are multiple groups of scatterers 413, which are respectively installed in the grooves.

[0067] The viscoelastic body is rubber, polyurethane or viscoelastic damping material, the scatterer is a metal structure, and the constraint layer is a hard material plate. The function of the constraint layer is to constrain the viscoelastic body and improve the damping energy dissipation of the viscoelastic body. In addition, since the viscoelastic body usually has low stiffness, the high stiffness of the constraint layer can provide conditions for the integrity of the structure and the installation of the fixture. The fixture can be installed through the bottom of the rail and will not affect the normal operation of the train.

[0068] More specifically, the viscoelastic body 411 in each vibration reduction unit is divided into five substructures in the axial direction, and the cross section of the viscoelastic body 411 is divided into a first half zone and a second half zone, the first half zone is close to the side of the rail 1, and the second half zone is close to the side of the constraint layer 412, and the first half zone is provided with a first slot cavity 4111, a second slot cavity 4112 and a third slot cavity 4113 in sequence from top to bottom in the first half zone, and the fourth slot cavity 4114, the fifth slot cavity 4115 and the sixth slot cavity 4116 in sequence from top to bottom in the second half zone, and the scatterers include a first scatterer 4131, a second scatterer 4132, a third scatterer 4133, a fourth scatterer 4134, a fifth scatterer 4135 and a sixth scatterer 4136, and are respectively installed in the first slot cavity 4111, the second slot cavity 4112, the third slot cavity 4113 and the fourth slot cavity 4111 in a one-to-one correspondence. 4. In the fifth slot cavity 4115 and the sixth slot cavity 4116, multiple second scatterers 4132 connect multiple substructures in series. The first scatterer 4131, the third scatterer 4133, the fourth scatterer 4134, and the fifth scatterer 4135 correspond to each other in the substructures. The first scatterers 4131, the third scatterer 4133, the fourth scatterer 4134, and the fifth scatterer 4135 in adjacent substructures are periodically spaced in the axial direction. Although each vibration reduction unit is divided into five substructures, these five substructures are designed as one body, just to distinguish that the scatterers other than the second scatterer and the sixth scatterer have discrete characteristics. It can be understood that the second scatterer and the sixth scatterer are continuous structures. On the one hand, they connect the substructures in series, and on the other hand, they guide the vibration of the transmission rail to the vibration reduction unit for energy consumption.

[0069] During specific implementation, three groups of vibration-damping units 4 on the same side of the rail 1 constitute an installation group (such as three-span fasteners, 1.8m), and the vibration-damping units in the same installation group are connected in series through the second scatterer 4132. One side of the rail 1 has multiple installation groups, and adjacent installation groups are connected through the second scatterer 4132.

[0070] In other embodiments, the second type of vibration damping unit 42 includes a frame structure 421, an elastic layer 422 and a mass body 423. The frame structure 421 is a long strip structure and is fixed on the track plate 3 corresponding to the length direction of the rail 1. A plurality of square grooves are arranged side by side on the frame structure 421. The mass body 423 is fixed in the square grooves, and the elastic layer 422 is arranged between the mass body 423 and the side walls of the square grooves.

[0071] It should be noted that a vibration gap of 10 mm is reserved between the elastic layer 422 , the mass body 423 and the track plate 3 .

[0072] The mass bodies 423 in the multiple square slots on the same row have different sizes. The masses of the mass bodies 423 in the adjacent square slots in the same row satisfy a linear variation rule. The masses of the mass bodies 423 in the corresponding square slots in adjacent rows are the same or different.

[0073] Specific case description, Case 1:

[0074] Ordinary ballastless track structure such as Figure 6-7 As shown in Figure 1, it consists of rails, fastener systems, and track plates. The existing structural vibration reduction capacity is mainly determined by the fasteners, but when the fastener stiffness decreases, it will cause the rail vibration response to be too large, causing rail corrugation. In order to improve the vibration reduction effect of the track structure, the proposed random detuning rail vibration reduction system is shown in Figure 1. Figure 1 As shown, vibration damping units are added to the left and right sides of the rail. The vibration damping units are arranged approximately periodically along the longitudinal direction of the rail (not completely periodic, to avoid resonance problems between the additional vibration damping structure and the periodic track structure), and for two adjacent unit groups on the same side, their initial spacing remains consistent because they are arranged between two fasteners, and the fastener spacing of the track structure is the same. It is also possible to introduce detuning design parameters (i.e., random detuning of the installation position) to randomly adjust the installation positions of adjacent unit groups (including the substructures in each unit) to broaden the vibration damping range and avoid resonance caused by structural periodicity.

[0075] The vibration reduction structure within the fastener spacing is a basic unit, and the left and right sides of the rails are staggered, such as Figure 2As shown, multiple outer and inner basic units are arranged in a staggered pattern. This staggered arrangement is intended to avoid structural resonances caused by a completely symmetrical structure, similar to a detuned design approach. If the structure were completely periodic and symmetrical, the structural vibration would be significantly enhanced at the resonant frequency.

[0076] There are 5 substructures inside each basic unit, such as Figure 3 and Figure 4 As shown in Figure 2 , the basic vibration damping unit consists of a constraining layer, a viscoelastic body, and a scatterer. The vibration damping unit is attached to the rail through its side and can be fixed by gluing or by clamping on the constraining layer. The scatterers within the basic unit are composed of two types of structures: one is a continuous structure (the second scatterer and the sixth scatterer), and the other is a discrete structure (the first scatterer, the third scatterer, the fourth scatterer, and the fifth scatterer). Both the continuous and discrete structures are metal structures. The continuous structure can keep adjacent vibration damping units connected, and the discrete structure makes the multiple substructures within a vibration damping unit also approximately periodic structures.

[0077] In order to broaden the vibration reduction effect of the vibration reduction system, the elastic wave band gap range of the track structure itself is first calculated as follows: Figure 7 As shown. It can be seen that without vibration reduction measures, the ballastless track structure only has an extremely narrow bandgap attenuation domain at low frequencies below 150 Hz and near 1000 Hz, and most structural vibrations can propagate freely. Therefore, by designing the parameters of the second and sixth scatterers of the continuous structure, including their cross-sectional inertia moment, cross-sectional shape, and mass density, the additional vibration reduction system can produce a coupling effect with the bandgap near 1000 Hz, widening the vibration bandgap in this frequency range. After the coupling bandgap, in order to further control other frequencies, such as the mid-frequency vibration of 300-400 Hz, the mass, position, and shape of the first, third, fourth, and fifth scatterers of the discrete structures can be designed so that the main frequency of their substructure is close to the target frequency. At this time, the track structure vibration can be transmitted to the additional vibration reduction system and dissipated in the scatterers.

[0078] Layout relationship and shape influence of the first to sixth scatterers:

[0079] The masses of the discrete structures are randomly detuned and designed in diamond or rectangular shapes to enhance their shear viscoelastic properties, thereby dissipating energy. Furthermore, the cross-sectional layout is divided into two sections (the first and second halves). The left section, located away from the rails, controls lateral bending and torsional vibrations and is therefore designed as a sharp-cornered diamond scatterer. The right section, located closer to the rails, primarily controls vertical vibrations and requires higher mass, so it is designed as a rectangular body.

[0080] At the same time, based on the idea of ​​random detuning, the parameters of the first, third, fourth, and fifth scatterers of adjacent substructures and adjacent basic units are optimized. Their initial design quality parameters are m1, m2, m3, and m4, respectively; and their initial position design parameters are d1, d2, d3, and d4.

[0081] The above parameters are designed to be uniformly distributed with a coefficient of variation of δ1, that is, the mass of the first scatterer is

[0082] Location is Therefore, the first scatterers in the rail vibration reduction system are randomly distributed along the rail longitudinal direction, including random mass and random position. Considering random detuning, the coefficient of variation can be taken as 0.1 to 0.5 according to experience.

[0083] In order to verify the rationality of the design method, with the goal of matching the pinned-pinned resonance frequency (bandgap boundary frequency) of the track structure near 1000Hz, the initial design mass of the four scatterers is set to m1=1.5kg, the elastic layer stiffness modulus is 1.2Mpa, and the scatterer mass variation coefficient δ1=0.3. Under the premise of the same total added mass, the vibration reduction effect of the traditional single-degree-of-freedom vibration absorber and the random detuning vibration reduction system designed by this method are compared. Figure 8 It can be seen that the additional vibration damping structure is coupled with the band gap of the track structure itself, and under the random detuning design, the attenuation range is further widened.

[0084] Case 2: Random Detuning Vibration Reduction System for Ballastless Track Structure

[0085] Based on Case 1, in order to further control the transmission of low-frequency vibration to the lower foundation, a local resonance structure with random detuning characteristics is added to the track plate.

[0086] The structure has the following characteristics: a basic unit consists of a frame structure, an elastic layer, and a mass body. The frame structure is connected to the track plate, and a 10mm gap is reserved between the elastic layer and the mass body and the track plate to reserve space for scatterer vibration. The vibration reduction system on the track plate consists of two rows of local resonance structures. Each mass body and elastic layer in each row is different, and their sizes follow a linear variation law. From one end to the other, the minimum mass body is 10kg, and the mass of the mass body gradually increases in steps of 2kg. The corresponding elastic layer also gradually changes. The designed shear stiffness is a minimum of 1.2e5N / m and a maximum of 4e6N / m. The other row of mass bodies is characterized by a large mass in the middle and a small mass at both ends.

[0087] By matching the local resonance frequencies of the mass bodies, the frequencies of adjacent local resonance units are controlled at intervals of 10 Hz, achieving low-frequency vibration control in the range of 10 to 100 Hz.

[0088] At the same time, the track structure vibration is effectively controlled through the comprehensive control of the rail random detuning vibration reduction system and the track plate random detuning vibration reduction system.

[0089] It is understandable that each substructure can control the vibration of a frequency band. If all substructures are the same, then only this frequency band can be controlled. Therefore, the mass or stiffness of adjacent substructures is randomly detuned, that is, the control frequency bands of adjacent substructures are staggered, thereby widening the vibration control range. Therefore, when designing, the mass of the mass body is linearly increased, and the elastic layer is also changed accordingly to achieve the widening of the control frequency band. The purpose of setting the other row of structures to be large in the middle and small in the two sections is also to widen the vibration control range, and the control frequency bands of the two rows of structures are inconsistent and have overlap, thus covering 10 to 100 Hz.

[0090] The present invention also discloses a design method for a random detuning track vibration reduction system, which comprises the following steps:

[0091] Step 1: Identify the vibration band gap characteristics of the ballastless track

[0092] Conduct an analysis of the ballastless track structure's periodic characteristics and attenuation domain. Based on the track structure's periodic characteristics, fastener parameters, and ballastless track bed structural parameters, calculate the elastic wave band gap frequency range of the ballastless track structure.

[0093] Step 2: Design the basic vibration damping unit

[0094] The vibration reduction unit is a local resonance structure unit. It consists of a scatterer, a viscoelastic body and a constraint layer. The initial design parameters of the vibration reduction unit are designed, including mass, elastic layer modulus and installation position. First, based on the control target frequency and the mass-spring system, the formula is used.

[0095]

[0096] Initial design values ​​for the mass of each scatterer and the elastic modulus are given, with f being the target control frequency and k being the elastic coefficient. A refined finite element model of the rail-vibration damping unit is then established, analyzing the impact of parameters such as mass, stiffness, and installation position on the control frequency band to determine the initial design parameters for the damping unit. The damping unit's main frequency is close to the track structure's bandgap frequency, allowing the damping unit attached to the rail or slab to couple with the ballastless track structure, maximizing the bandgap range.

[0097] Step 3: Design the structural parameters of the detuned vibration reduction unit

[0098] The parameters of the vibration reduction unit structure are randomly detuned, including the weight of the scatterer, the stiffness of the viscoelastic body, and the installation position. Based on the initial design parameters, these parameters follow a uniform distribution with a mean of A1 (i.e., the initial design parameters) and a coefficient of variation of δ1, which can be between 0.1 and 0.5. The detuning parameters and degree of detuning are determined by the objective function of not causing structural resonance and increasing the vibration reduction frequency range after installing the vibration reduction unit.

[0099] Step 4: Add other control bands that require attention

[0100] According to the target control frequency band, the second or third type of vibration reduction unit is set. The second or third type of vibration reduction unit also adopts a random detuning design, and the key parameters of the uniform distribution are set to A2 or A3 and the coefficient of variation is δ2 or δ3;

[0101] Step 5: Evaluate the vibration reduction effect of the vibration reduction system

[0102] The vibration attenuation domain and vibration response characteristics of the ballastless track random detuning vibration reduction system are analyzed, and the vibration reduction effect of the random detuning vibration reduction system is predicted.

[0103] Specifically, in step one, based on the phononic crystal elastic wave theory, the elastic wave band gap range of the ballastless track structure itself is calculated, and the frequency and vibration mode at the elastic wave band gap boundary are obtained; the ballastless track structure has the characteristic of being periodically arranged along the longitudinal direction of the line, and its basic unit is a structure within the length of the track plate or the length of the fastener spacing. The track structure itself is a periodic structure and has a band gap characteristic. Based on the phononic crystal elastic wave theory, the elastic wave band gap range of the ballastless track structure itself is calculated, and the frequency and vibration mode at the elastic wave band gap boundary are obtained. The purpose of this step is to determine the vibration characteristics of the track structure itself, so that when designing the additional vibration reduction system, the vibration mode of the vibration reduction system is coupled with the vibration mode of the track structure, especially to produce a coupling effect at the elastic wave boundary of the track structure. At this time, the band gap range can be effectively expanded to maximize the vibration effect;

[0104] In step 2, the main frequency of the vibration reduction unit is close to the band gap frequency of the track structure, and the vibration mode of the vibration reduction unit is coupled with the significant vibration mode of the track structure, so that the additional local resonance structure and the ballastless track structure produce a coupling effect, thereby maximizing the band gap range.

[0105] To achieve significant main frequency coupling between the local resonant structure and the track structure, the direction of local vibration movement should be the same as that of the rail or track plate. Furthermore, to further transmit the track structure vibration to the additional local resonant structure, a continuous structure is inserted into the vibration reduction unit design, introducing a phononic crystal elastic wave passband mechanism. This continuous structure effectively transmits vibration, thereby guiding the track structure vibration to the vibration reduction system and transmitting it longitudinally along the line, while simultaneously dissipating the vibration energy through the vibration of the local resonant unit. Therefore, the vibration reduction unit consists of two types: a continuous structure (which generates an elastic wave passband, guiding the track vibration to the vibration reduction system) and a local resonant structure (which generates an elastic wave bandgap, consuming vibration energy).

[0106] In step five, the vibration attenuation domain of the ballastless track random detuning vibration reduction system is first analyzed to evaluate the vibration reduction range. Then, the vibration response of the track structure under simple harmonic load is calculated to evaluate whether there is any obvious resonance response in the structure. Finally, a dynamic coupling analysis model of the vehicle-track-random detuning vibration reduction system is established. The vibration reduction effect of the random detuning vibration reduction system is predicted based on the vibration response of the track structure and the vibration response of the structure under the track slab.

[0107] The core principle of this method is to design a random detuned track vibration reduction system based on the passband and bandgap mechanism of phononic crystal elastic waves and taking into account the periodic characteristics of the track structure itself.

[0108] First, the basic unit of the random detuning vibration reduction system consists of a discrete local resonance unit and a continuous guiding structure. Therefore, based on these two types of structures, the track structure vibration can be guided through the continuous structure to the attached vibration reduction system based on the passband mechanism, achieving the guidance and propagation of the track structure vibration. Then, based on the bandgap mechanism, the vibration is transferred to the local oscillator through the discrete local resonance structure and dissipated.

[0109] Furthermore, to address the limited frequency band of existing vibration reduction measures, some localized resonant units in the additional structure are coupled with the band gap of the periodic track structure itself, thereby expanding the track structure's band gap. Vibrations within the band gap are difficult to propagate within the track structure itself and must instead propagate into the additional structure.

[0110] Furthermore, the random detuning structure amplifies the vibration reduction effect of the localized resonant unit, overcoming the limitations of traditional dynamic vibration absorbers that rely on single-frequency control. Simultaneously, by randomizing the stiffness, mass, and mounting position parameters, the resonance of the periodic structure itself is avoided, achieving better vibration reduction while eliminating the structural resonance issues associated with traditional supplementary vibration damping systems.

[0111] As for the devices and methods of use disclosed in the embodiments, since they correspond to the methods disclosed in the embodiments, the description is relatively simple, and the relevant parts can be referred to the description of the methods.

[0112] The above description of the disclosed embodiments is intended to enable one skilled in the art to implement or use the present invention. Various modifications to these embodiments will be readily apparent to one skilled in the art, and the general principles defined herein may be implemented in other embodiments without departing from the spirit or scope of the present invention. Therefore, the present invention is not limited to the embodiments shown herein but is intended to conform to the widest scope consistent with the principles and novel features disclosed herein.

Claims

1. A random detuning track vibration reduction system, characterized in that: include: Rail members (1), the rail members (1) are arranged in groups and at intervals on a track plate (3), and a fastener system (2) is arranged between the rail members (1) and the track plate (3); A vibration damping unit (4), wherein the vibration damping unit (4) comprises a plurality of groups, and the plurality of groups of the vibration damping units (4) are arranged on both inner and outer sides of the rail member (1) or on the track plate, and the plurality of groups of the vibration damping units (4) on the same side of the rail member (1) are quasi-periodically distributed in the longitudinal direction of the rail member (1), and the vibration damping units on both sides of the rail member (1) are symmetrically arranged or staggered, and the main frequency of the vibration damping unit (4) is close to the band gap frequency of the track structure, and the vibration mode of the vibration damping unit is coupled with the vibration mode of the track structure, so that the resonance structure of the vibration damping unit and the ballastless track structure produce a coupling effect, thereby amplifying the band gap range; The vibration reduction unit (4) includes a first type of vibration reduction unit (41) and a second type of vibration reduction unit (42). The first type of vibration reduction unit (41) includes a viscoelastic body (411), a constraint layer (412) and a scatterer (413). The viscoelastic body (411) is fixed to the side of the rail member (1). The viscoelastic body (411) is divided into a plurality of substructures in the axial direction. The cross section of the viscoelastic body (411) is divided into a first half zone and a second half zone. The first half zone is close to the side of the rail member (1). The second half zone is close to the side of the rail member (1). The second half region is close to the side of the constraint layer (412), the constraint layer (412) is fixed on the outer side of the viscoelastic body (411) away from the rail member (1), the interior of the viscoelastic body (411) is provided with a plurality of grooves, and the scatterers (413) are provided in a plurality of groups and are respectively installed in the grooves; the first half region is provided with at least a first groove (4111), a second groove (4112) and a third groove (4113) in sequence from top to bottom, and the second half region is provided with at least a fourth groove (4111) in sequence from top to bottom. 4), a fifth slot cavity (4115) and a sixth slot cavity (4116), wherein the scatterers at least include a first scatterer (4131), a second scatterer (4132), a third scatterer (4133), a fourth scatterer (4134), a fifth scatterer (4135) and a sixth scatterer (4136), and are respectively installed in the first slot cavity (4111), the second slot cavity (4112), the third slot cavity (4113), the fourth slot cavity (4114), the fifth slot cavity (4115) and the sixth slot cavity ( 4116), a plurality of the second scatterers (4132) and a plurality of the sixth scatterers (4136) connect a plurality of substructure parts in series, the first scatterer (4131), the third scatterer (4133), the fourth scatterer (4134) and the fifth scatterer (4135) are correspondingly located in the substructure parts, and the first scatterer (4131), the third scatterer (4133), the fourth scatterer (4134) and the fifth scatterer (4135) in adjacent substructure parts are periodically spaced in the axial direction.

2. The random detuning track vibration reduction system according to claim 1, characterized in that: The first type of vibration damping unit (41) is arranged on the rail member (1), and the second type of vibration damping unit (42) is arranged on the track plate (3).

3. The random detuning track vibration reduction system according to claim 1, characterized in that: The plurality of groove cavities all pass through both ends of the viscoelastic body (411).

4. The random detuning track vibration reduction system according to claim 1, characterized in that: At least two groups of vibration-damping units (4) on the same side of the rail member (1) constitute an installation group, and the vibration-damping units in the same installation group are connected in series via a second scatterer (4132) or a sixth scatterer (4136). One side of the rail member (1) has a plurality of installation groups, and the second scatterers (4132) between adjacent installation groups abut against each other.

5. The random detuning track vibration reduction system according to claim 2, characterized in that: The second type of vibration reduction unit (42) comprises a frame structure (421), an elastic layer (422) and a mass body (423); the frame structure (421) is a long strip structure and is fixed on the track plate (3) in the length direction of the corresponding rail member (1); a plurality of square grooves are arranged side by side on the frame structure (421); the mass body (423) is fixed in the square grooves; and the elastic layer (422) is arranged between the mass body (423) and the side walls of the square grooves.

6. The random detuning track vibration reduction system according to claim 5, characterized in that: A vibration gap of 10 mm is reserved between the elastic layer (422), the mass body (423) and the track plate (3).

7. The random detuning track vibration reduction system according to claim 5, characterized in that: The mass bodies (423) in the plurality of square slots on the same row have different sizes, the masses of the mass bodies (423) in the adjacent square slots corresponding to the same row satisfy a linear variation law, and the masses of the mass bodies (423) in the square slots corresponding to adjacent rows are the same or different.

8. A design method for a random detuning track vibration reduction system according to any one of claims 1 to 7, characterized in that: The following steps are involved: Step 1: Identify the vibration band gap characteristics of the ballastless track Conduct an analysis of the ballastless track structure's periodic characteristics and attenuation domain. Based on the track structure's periodic characteristics, fastener parameters, ballastless trackbed structure, and track slab support stiffness parameters, calculate the elastic wave band gap frequency range of the ballastless track structure. Step 2: Design the basic vibration damping unit The vibration reduction unit is a local resonance structure unit. It consists of a scatterer, a viscoelastic body and a constraint layer. The initial design parameters of the vibration reduction unit, including mass, elastic layer modulus and installation position, are designed based on the control target frequency and the mass-spring system. The formula is used. The initial design values ​​of the mass of each scatterer and the elastic modulus of the elastic layer are given. The parameters of each mass body are obtained through dynamic calculations, with k being the elastic coefficient. A refined finite element model of the rail-vibration damping unit is then established. The effects of parameters such as mass, stiffness, and installation position on the control frequency band are analyzed to determine the initial design parameters of the vibration damping unit. The main frequency of the vibration reduction unit is close to the band gap frequency of the track structure, so that the vibration reduction unit structure attached to the rail member or track plate produces a coupling effect with the ballastless track structure, maximizing the band gap range; Step 3: Design the structural parameters of the detuned vibration reduction unit The parameters of the vibration reduction unit structure are randomly detuned, including the weight of the scatterer, the stiffness of the viscoelastic body, and the installation position. These parameters follow a uniform distribution with a mean of A1 and a coefficient of variation of δ1. The mean is the initial design value before detuning, obtained in the second step. The coefficient of variation can be taken as an empirical value of 0.1-0.

5. The detuning parameters and degree of detuning are determined by the objective function of not causing structural resonance and increasing the vibration reduction frequency range after installing the vibration reduction unit. Step 4: Add other control bands that require attention According to the target control frequency band, the second or third type of vibration reduction unit is set. The second or third type of vibration reduction unit also adopts a random detuning design, and the key parameters of the uniform distribution are set to A2 or A3 and the coefficient of variation is δ2 or δ3; Step 5: Evaluate the vibration reduction effect of the vibration reduction system The vibration attenuation domain and vibration response characteristics of the ballastless track random detuning vibration reduction system are analyzed, and the vibration reduction effect of the random detuning vibration reduction system is predicted.

9. The design method of a random detuning track vibration reduction system according to claim 8, characterized in that: In the first step, based on the phononic crystal elastic wave theory, the elastic wave band gap range of the ballastless track structure itself is calculated, and the frequency and vibration mode at the elastic wave band gap boundary are obtained; in the second step, the main frequency of the vibration reduction unit is close to the band gap frequency of the track structure, and the vibration mode of the vibration reduction unit is coupled with the significant vibration mode of the track structure, so that the additional local resonance structure and the ballastless track structure produce a coupling effect, thereby maximizing the band gap range; in the fifth step, the vibration attenuation domain range of the ballastless track random detuning vibration reduction system is first analyzed to evaluate the vibration reduction range, and then the vibration response of the track structure under the action of simple harmonic load is calculated to evaluate that no obvious resonant response occurs in the structure. Finally, a dynamic coupling analysis model of the vehicle-track-random detuning vibration reduction system is established, and the vibration reduction effect of the random detuning vibration reduction system is predicted based on the vibration response of the track structure and the vibration response of the structure under the track plate.

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