Trolley and trailer vibration reduction method

Abstract

The invention discloses a trailer and a trailer damping method.The trailer is applied to a wheel track type belt conveyor and used for supporting and dragging a conveying belt on a track of the wheel track type belt conveyor and walking along with the conveying belt, the trailer comprises a frame and a plurality of wheel assemblies distributed on the two sides of the frame, and the frame is provided with mounting seats corresponding to the wheel assemblies one to one; the wheel assembly comprises a wheel body, a wheel axle and a bearing, the wheel axle is in transmission connection with the wheel body through the bearing, at least part of the area of the wheel axle penetrates through a mounting hole formed in the mounting base and is in clearance fit with the mounting hole, and an elastic structure located between the mounting base and the wheel axle is arranged in the mounting hole. When the trailer transports a conveying belt and materials on the track, material loads and wheel-rail force are subjected to energy storage and energy consumption when passing through the elastic structure, so that the overall vibration of the trailer is attenuated, the vibration reduction effect of the trailer during operation on the track can be greatly improved, the peak value of a bearing vibration curve can be reduced, the peak clipping effect is achieved, and the service life of the bearing is prolonged.

Description

Technical Field

[0001] The present application relates to the technical field of belt conveyors, and in particular to a trolley and a trolley vibration reduction method. Background Art

[0002] Compared with the traditional roller conveyor that uses rollers to support the conveyor belt, the wheel-rail conveyor is a new type of belt conveyor. The conveyor uses a trolley to support the conveyor belt, and the conveyor belt is used to place materials. The trolley runs on the track. There is rigid contact between the track and the wheels, and the wheels of the trolley run with a small rolling friction. During operation, the trolley and the conveyor belt are basically in a relatively static state, eliminating the roller depression resistance and the material wear resistance caused by the wave motion of the conveyor belt in the roller conveyor, saving a lot of energy during operation. Its energy consumption is saved by 60%, and the carrying capacity is increased by about 40%, which is in line with the national green mine development strategy. Moreover, the wheel-rail belt conveyor has a large single carrying capacity and a long single-machine transportation distance, and is more suitable for long-distance continuous mineral resource transportation in field environments.

[0003] Existing medium-track vehicles, such as high-speed railways, use wheelsets for running and steering on rails. High-speed railways use wheelsets because they can provide stable support and guidance functions, ensuring that the train maintains good stability when running at high speeds. In addition, wheelsets can effectively reduce frictional resistance and improve operating efficiency. Compared with other forms of transportation, wheelsets perform exceptionally well on high-speed railways and can meet the speed, stability, and efficiency requirements of high-speed trains. Therefore, when designing trolleys for wheel-rail belt conveyors, the wheels were modeled after the wheelsets of existing medium-speed rail vehicles. However, due to stiffness requirements, it is difficult for trolleys to take into account the noise reduction performance of wheelsets running on rails. Moreover, compared with the smoothness and linearity of high-speed rail tracks, the material transportation routes used in mining are complex, with more turning and ramp sections, steeper slopes, and more horizontal and vertical turns and three-dimensional turns. The tracks are installed at height differences on both sides of the turns. Therefore, when trolleys are running on these special sections, the forces acting on the wheels are more complex, and the requirements for reasonable force distribution on the wheels are higher. Derailment is more likely to occur due to defects in force distribution, and the noise during operation will also increase. This makes the traditional wheelset system no longer suitable for the needs of trolleys for wheel-rail belt conveyors in specific transportation scenarios.

[0004] The trolley of a wheel-rail belt conveyor is usually composed of basic elements such as a bracket, a frame, and wheels. The bracket is installed on the frame to support the conveyor belt, and the wheels are installed on the frame through bearings. However, when the trolley is running on the track, its overall high rigidity makes it perform poorly in terms of vibration. The vibration will generate huge operating noise and affect its own service life and the smoothness of material transportation. Therefore, how to reduce the vibration of the trolley itself without making large-scale modifications to the trolley has become an industry problem.

[0005] Currently, the most commonly used vibration reduction method for road vehicles is to add a spring-damper suspension system. If trailers also adopt a similar structure, it will undoubtedly increase the overall size of the trailer and the manufacturing and maintenance costs. It will also increase the cost of modifying old trailers, which does not have a clear advantage from a practical and practicability perspective. Based on the material load transmission path of the wheel-rail belt conveyor, the conveyor belt, bracket, frame, wheel axle, bearing, wheel body and track are the following: the wheel-rail belt conveyor has a higher stiffness requirement than the traditional roller conveyor, which makes the trailer highly incompatible with the spring-damper suspension system. In the mechanical field, bearings themselves are a kind of vulnerable parts. In the wheel-rail belt conveyor, they serve as a link in the load transmission, which significantly increases the probability of bearing damage. Therefore, in the trailer design process, how to ensure that the trailer has sufficient stiffness to ensure the normal transmission operation and service life of the bearings, while also ensuring that the trailer has excellent vibration reduction performance? The balance and trade-off between these two has become a difficult problem that plagues the improvement of production quality of wheel-rail conveyors. Summary of the Invention

[0006] The present application aims to solve the technical problems in the above-mentioned technologies at least to a certain extent. In order to overcome the vibration problem of the existing medium-wheel-rail belt conveyor trolley during its operation on the track and the technical problems such as the short service life of the bearings, and to solve the shortcomings and deficiencies in the existing technology, the present application provides a trolley and a trolley vibration reduction method.

[0007] The technical solutions adopted in this application are:

[0008] A trolley is used for a wheel-rail type belt conveyor, and is used to support and pull the conveyor belt on the track of the wheel-rail type belt conveyor and move along with the conveyor belt. The trolley includes a frame and a plurality of wheel assemblies distributed on both sides of the frame. The frame is provided with mounting seats corresponding to the wheel assemblies one by one. The wheel assembly includes a wheel body, a wheel axle and a bearing. The wheel axle is transmission-connected to the wheel body through the bearing. At least a portion of the wheel axle is inserted into the mounting hole provided on the mounting seat and is gap-fitted with the mounting hole. An elastic structure is provided in the mounting hole and is located between the mounting seat and the wheel axle.

[0009] The trailer provided in this application also includes the following additional technical features:

[0010] The elastic structure includes one or a plurality of elastic bodies arranged at intervals along the axial direction of the wheel axle, and the elastic body is configured as a closed structure extending around the circumference of the wheel axle.

[0011] The elastic body is compressed between the mounting seat and the wheel axle, and a compression rate λ of the elastic body is ≤30%.

[0012] The elastic body is configured as an O-ring with a circular, polygonal or irregular cross-section.

[0013] The elastomer is configured as a single material, which includes an elastic material that can be elastically compressed; or, the elastomer is configured as a composite material, which includes a support skeleton and an elastic material that can be elastically compressed and surrounded by the support skeleton.

[0014] The elastomer is configured as an O-ring with a circular cross-section, and an annular groove adapted to a single O-ring is provided on the outer peripheral surface of the wheel shaft. The cross-sectional area S2 of the annular groove and the cross-sectional area S1 of the O-ring satisfy: 0.6≤S1 / S2≤1.2, wherein the cross-sectional area S1 is the standard area of ​​the O-ring in the initial uncompressed state, and the depth h of the annular groove along the radial direction of the mounting hole or the wheel shaft and the cross-sectional diameter d of the O-ring satisfy: 0.5<h / d<1.

[0015] The annular groove is configured as a structure with a rectangular or conical cross section.

[0016] The mounting seat is constructed as an integrally formed structure, and the mounting hole passes through both ends of the mounting seat; or, the mounting seat includes a first mounting body and a second mounting body arranged at intervals, and the mounting hole includes a first through hole passing through the first mounting body and a second through hole passing through the second mounting body, the first mounting body is close to the wheel body, and the second mounting body is located on the side of the first mounting body away from the wheel body, and the elastic structure is distributed between the first mounting body and the wheel axle and / or between the second mounting body and the wheel axle.

[0017] The mounting seat is welded to the vehicle frame; or, the mounting seat is connected to the vehicle frame via threaded fasteners; or, the mounting seat is riveted to the vehicle frame.

[0018] The mounting seat is provided with a conical guide slope at the end of the mounting hole close to one end of the wheel body, and the conical guide slope is used to guide the elastic structure to be installed into the mounting hole; and / or, the mounting seat is provided with a mounting groove on the inner wall of the mounting hole, and the elastic structure is installed in the mounting groove.

[0019] The present application discloses a method for reducing vibration of a trailer, which is applied to the trailer as described in at least some of the technical features above. The method includes: setting the stiffness of the O-ring to k1, and predicting the stiffness of the O-ring with the compression rate. Specifically, the stiffness of the O-ring k1 = e bλ N / mm, where b is the transmission coefficient between λ and k1, 26≤b≤30.

[0020] The method for reducing the vibration of the trailer also includes a method for determining the number of O-rings, which is specifically, the number of O-rings in, is the original dynamic load on the O-ring, m1 is the total mass borne by the O-ring, g is the acceleration due to gravity, a is the acceleration of the trailer, d is the cross-sectional diameter of the O-ring, and h is the depth of the annular groove.

[0021] The stiffness k1 of the O-ring and the resonant frequency f, the total mass m1 borne by the O-ring and the number n of O-rings satisfy:

[0022] Among them, 0 <f≤50Hz。

[0023] The compression rate λ and resonance frequency f of the O-ring, the total mass m1 borne by the O-ring and the number n of O-rings satisfy:

[0024] Among them, 0 <f≤50Hz。

[0025] Due to the adoption of the above technical solution, the technical effects achieved by this application include at least:

[0026] 1. In the present invention, the wheel assemblies are distributed on both sides of the frame, and the frame is provided with mounting seats corresponding to the wheel assemblies one by one, so that each wheel assembly is independently mounted on the frame through the mounting seats. Compared with the wheelset, in this solution, each wheel assembly can independently respond to the changes in the track surface, which is beneficial to maintaining the traction and adhesion of the trailer, so that the force distribution of the wheel assembly when the trailer is subjected to various forces during the operation of the trailer on complex track sections is more superior. This superiority can improve the smoothness and flexibility of the wheel assembly running on the track, and provide an excellent structural foundation and application environment for the relevant design of the trailer on the balance and trade-off between stiffness and vibration reduction. At least part of the wheel axle is passed through the mounting hole provided in the mounting seat and is in clearance fit with the mounting hole. An elastic structure is provided in the mounting hole between the mounting seat and the wheel axle. First, when the trolley is running on the track, the material load is transmitted through the conveyor belt, bracket, frame, mounting seat, elastic structure, wheel axle, bearing, wheel body and track in sequence. Among them, the bracket, frame and mounting seat are rigidly connected in series, and the wheel axle, bearing and wheel body are transmission connected in series. The wheel body rolls on the track, and the material load and wheel-rail force experience energy storage and energy consumption when passing through the elastic structure, so that the overall vibration of the trolley is attenuated, which is reflected in the ability to greatly improve the The vibration reduction effect of the trailer when running on the track can at least be reflected in the ability to significantly reduce the peak value of the bearing vibration curve at the microscopic level, thereby achieving a peak-cutting effect and improving the bearing life. Secondly, because the elastic structure is located between the mounting seat and the axle, and the wheel assemblies are located on both sides of the frame, the frame is configured to reduce vibration at both ends and bear material loads in the middle. In this state, the two wheel assemblies arranged opposite each other on the frame can cooperate with each other to achieve a secondary distribution of the torque load, thereby reducing stress concentration, vibration peaks and noise caused by unbalanced torque loads (such as torque loads generated under conditions such as turning and height differences between the tracks on both sides).

[0027] 2. The elastic structure includes one or more elastic bodies arranged at intervals along the axial direction of the wheel axle. The elastic body is constructed as a closed structure extending circumferentially around the wheel axle. On the one hand, the elastic body is integrally sleeved on the outer peripheral surface of the wheel axle, which is convenient for installation and radial positioning. On the other hand, the load on the wheel axle in each radial direction can be stored and consumed through the elastic body, so that the elastic structure can improve its response ability to all-directional forces when dealing with the force distribution when the trailer passes through complex sections, thereby improving the force balance of the wheel assembly and the adhesion of the wheel body to the track, so that the trailer has better vibration reduction and vibration peak reduction effects when dealing with complex sections.

[0028] 3. The compression rate of the elastomer after being compressed between the mounting seat and the wheel axle is a key factor in balancing the rigidity and vibration reduction performance of the trailer structure. If the compression rate of the elastomer is too large, the rigidity of the elastomer itself and even the trailer structure will be too large, while the vibration reduction performance will be weakened. If the compression rate of the elastomer is too small, the vibration reduction performance of the trailer will be improved while losing the most favorable strength environment for the operation of the bearing. Therefore, the present invention obtains the optimal solution for the compression rate after comprehensively considering various factors, that is, the compression rate of the elastomer λ≤30%, which satisfies the excellent vibration reduction performance of the trailer and improves the noise reduction effect while ensuring that the bearing operates in the most favorable strength environment, thereby effectively improving the bearing life.

[0029] 4. The O-ring and the annular groove cooperate. On the one hand, the side walls on the left and right sides of the annular groove can be used to form axial limits for the O-ring to prevent the O-ring from shifting along the axial direction of the wheel axle and affecting the working performance. On the other hand, the elastomer can also abut against the side walls on the left and right sides of the annular groove to store and consume energy of the load on the wheel axle in the axial direction, thereby forming axial vibration reduction of the wheel assembly. In addition, a reasonable match between the sizes of the O-ring and the annular groove is an important guarantee for improving the vibration reduction effect and extending the service life of the O-ring. This application reasonably optimizes the size matching of the O-ring and the annular groove by making the cross-sectional area S2 of the annular groove and the cross-sectional area S1 of the O-ring satisfy: 0.6≤S1 / S2≤1.2, and making the depth h of the annular groove along the radial direction of the mounting hole or the wheel axle and the cross-sectional diameter d satisfy: 0.5≤h / d≤1, thereby effectively avoiding the O-ring being too large or too small. From a macro perspective, through this size design, the inner and outer sides of the O-ring are in stable and reliable contact with the wheel axle and the mounting seat respectively, and have sufficient deformation space and stable deformation ability, ensuring long-term and effective vibration reduction. From a micro perspective, through this size design, it helps to further reasonably optimize important vibration-related factors including the pre-compression rate of the O-ring when unloaded, the compression rate after loading, stiffness, quantity and resonance.

[0030] 5. The trailer vibration reduction method disclosed in the present application further integrates the prediction of O-ring stiffness with compression rate, calculation of the number of O-rings, trailer oscillation system resonance prevention strategy, and bearing service life with compression rate prediction method, so that without large-scale modification of the overall trailer structure, a good vibration reduction effect can be achieved, the risk of trailer resonance is greatly reduced, and the service life of O-rings and bearing wearing parts is increased as much as possible. Moreover, the present method can improve the vibration reduction effect by changing the compression rate, number, and material of the O-rings according to actual requirements, and the O-rings are easy to maintain and replace, the overall structure is simple, the stability is good, and it has strong practicality. BRIEF DESCRIPTION OF THE DRAWINGS

[0031] The drawings described herein are used to provide a further understanding of the present application and constitute a part of the present application. The illustrative embodiments of the present application and their descriptions are used to explain the present application and do not constitute an improper limitation on the present application. In the drawings:

[0032] Figure 1 A schematic diagram of the structure of a trailer provided in an embodiment of the present application;

[0033] Figure 2 Schematic diagram of the partial structure of the trailer provided in the embodiment of the present application Figure 1 ;

[0034] Figure 3 Schematic diagram of the structure of the trolley on the track provided in the embodiment of the present application Figure 1 ;

[0035] Figure 4 A force analysis diagram of a trolley on a straight section of a track provided in an embodiment of the present application;

[0036] Figure 5 for Figure 4 Force analysis diagram of the structure at point A;

[0037] Figure 6 Force analysis of the trailer in the turning section of the track provided in the embodiment of the present application Figure 1 ;

[0038] Figure 7 Force analysis of the trailer in the turning section of the track provided in the embodiment of the present application Figure 2 ;

[0039] Figure 8 Schematic diagram of the structure of the trolley on the track provided in the embodiment of the present application Figure 2 ;

[0040] Figure 9 A force analysis diagram of a trailer on a turning section of a track when the frame and axle provided in an embodiment of the present application are rigidly connected;

[0041] Figure 10 This is a force analysis diagram of a trailer on a turning section of a track when the frame and axle provided in an embodiment of the present application are connected by an elastic body;

[0042] Figure 11 A cross-sectional view of a partial structure of a trailer provided in one embodiment of the present application;

[0043] Figure 12 A cross-sectional view of a partial structure of a trailer provided in another embodiment of the present application;

[0044] Figure 13 A schematic structural diagram of a mounting base provided in one embodiment of the present application.

[0045] List of parts and reference numerals:

[0046] 11 frame, 12 mounting seat, 121 mounting hole, 1211 first through hole, 1212 second through hole, 122 first mounting body, 123 second mounting body, 124 tapered guide slope, 13 wheel body, 14 axle, 141 annular groove, 15 bearing, 16 bracket, 17 elastic body, 18 positioning baffle, 19 locking member, 20 elastic shock-absorbing pad;

[0047] 2 tracks. DETAILED DESCRIPTION

[0048] In order to more clearly illustrate the overall concept of the present application, a detailed description is given below in an illustrative manner in conjunction with the accompanying drawings.

[0049] In the following description, many specific details are set forth to facilitate a full understanding of the present application. However, the present application may also be implemented in other ways different from those described herein. Therefore, the scope of protection of the present application is not limited to the specific embodiments disclosed below.

[0050] In addition, in the description of the present application, it should be understood that the terms "upper", "lower", "top", "bottom", "inside", "outside", "axial", "radial", "circumferential", "lateral", "longitudinal", etc., indicating orientations or positional relationships, are based on the orientations or positional relationships shown in the accompanying drawings, and are only for the convenience of describing the present application and simplifying the description, and do not indicate or imply that the device or element referred to must have a specific orientation, be constructed and operated in a specific orientation, and therefore should not be understood as a limitation on the present application.

[0051] In this application, unless otherwise expressly specified or limited, terms such as "installed," "connected," "connect," and "fixed" should be understood in a broad sense. For example, they may refer to fixed connection, detachable connection, or integration; mechanical connection, electrical connection, or communication; direct connection or indirect connection through an intermediate medium; and internal communication between two components or interaction between two components. Those skilled in the art will understand the specific meanings of the above terms in this application based on specific circumstances.

[0052] In this application, unless otherwise expressly specified and limited, a first feature "above" or "below" a second feature may be that the first and second features are in direct contact, or the first and second features are in indirect contact through an intermediate medium. In the description of this specification, the description with reference to the terms "one embodiment", "some embodiments", "example", "specific example", or "some examples" means that the specific features, structures, materials or characteristics described in conjunction with the embodiment or example are included in at least one embodiment or example of the present application. In this specification, the schematic representation of the above terms does not necessarily refer to the same embodiment or example. Moreover, the specific features, structures, materials or characteristics described may be combined in an appropriate manner in any one or more embodiments or examples.

[0053] In the embodiments of this application, a trailer is provided. For ease of explanation and understanding, the following content provided herein is based on the illustrated product structure. Of course, those skilled in the art will understand that the above structure is merely a specific example and schematic illustration and does not constitute a specific limitation of the technical solution provided herein. Furthermore, this application also provides a trailer vibration reduction method.

[0054] A trailer provided by this application, such as Figures 1 to 13 As shown, it is applied to a wheel-rail type belt conveyor, and is used to support and traction the conveyor belt on the track 2 of the wheel-rail type belt conveyor and travel along with the conveyor belt. Figure 3 Schematically depicts an embodiment of a trolley located on a track 2, wherein the trolley includes a frame 11 and a plurality of wheel assemblies distributed on both sides of the frame 11, the frame 11 is provided with mounting seats 12 corresponding one to one with the wheel assemblies, the wheel assembly includes a wheel body 13, a wheel axle 14 and a bearing 15, the wheel axle 14 is transmission-connected to the wheel body 13 via the bearing 15, at least a portion of the wheel axle 14 is inserted into a mounting hole 121 provided in the mounting seat 12 and is clearance-fitted with the mounting hole 121, and an elastic structure is provided in the mounting hole 121 between the mounting seat 12 and the wheel axle 14.

[0055] In the present invention, the wheel assemblies are distributed on both sides of the frame 11, and the frame 11 is provided with mounting seats 12 corresponding to the wheel assemblies one by one, so that each wheel assembly is independently mounted on the frame 11 through the mounting seat 12. Compared with the wheelset, in this solution, each wheel assembly can independently respond to the changes in the track surface of the track 2, which is beneficial to maintaining the traction and adhesion of the trailer, so that the force distribution of the wheel assembly when the trailer is running on complex sections of the track 2 such as turning sections, slope sections, and vertical U-turn sections is more superior. This superiority can improve the smoothness and flexibility of the wheel assembly running on the track 2, and provide an excellent structural foundation and application environment for the relevant design of the trailer to balance and make trade-offs between stiffness and vibration reduction.

[0056] Therefore, in order to more fully utilize the structural foundation and application environment provided by the independent installation of the wheel assemblies on both sides of the trailer through the mounting seat 12, the present invention allows at least a portion of the wheel axle 14 to pass through the mounting hole 121 provided on the mounting seat 12 and to fit the mounting hole 121 with a clearance, and an elastic structure located between the mounting seat 12 and the wheel axle 14 is provided in the mounting hole 121. First, during the operation of the trailer on the track 2, the material load transmission path is the conveyor belt, the bracket 16 (the structure installed on the frame 11 for supporting the conveyor belt), the frame 11, the mounting seat 12, the elastic structure, the wheel axle 14, the bearing 15, the wheel body 13 and the track 2 in sequence, wherein the bracket 16, the frame 11 and the mounting seat 12 are rigidly connected in series, the wheel axle 14, the bearing 15 and the wheel body 13 are transmission connected in series, and the wheel body 13 is on the track. 2, the material load and the wheel-rail force experience energy storage and energy consumption when passing through the elastic structure, so that the overall vibration of the trolley is attenuated. Macroscopically, it can greatly improve the vibration reduction effect when the trolley runs on the track 2. Microscopically, it can at least significantly reduce the peak value of the vibration curve of the bearing 15, play a peak-cutting effect, and improve the service life of the bearing 15. Secondly, because the elastic structure is located between the mounting seat 12 and the wheel axle 14, and the wheel assembly is located on both sides of the frame 11, the frame 11 is configured to be in a working state of vibration reduction at both ends and bearing material loads in the middle. In this state, the two wheel assemblies arranged opposite to each other on the frame 11 can cooperate with each other to realize the secondary distribution of the torque load, thereby reducing the stress concentration, vibration peak and noise caused by unbalanced torque loads (such as torque loads generated under conditions such as turning and height difference between the tracks on both sides).

[0057] Specifically, the entire shaft section of the axle 14 exposed outside the bearing 15 can be passed through the mounting hole 121, or a part of the axle 14 can be passed through the mounting hole 121 and the other part can be passed through the mounting hole 121, so that the axle 14 and the mounting seat 12 can be positioned outside the mounting hole 121 through the positioning structure.

[0058] The following is a force analysis of the specific scenario of the trolley running on the straight section of track 2 in Example 1 and a force analysis of the specific scenario of the trolley running on the curved section of track 2 in Example 2. Both embodiments mainly focus on the analysis of the force distribution of the elastic structure in the four-point positioning support when a single wheel assembly is simultaneously subjected to radial force, shear force, and axial force, so as to more intuitively judge the advantages of setting the elastic structure between the wheel axle 14 and the mounting seat 12.

[0059] Example 1: Figure 4 The figure shows the force analysis of the trailer's running state on the straight section of track 2. When the trailer travels in a straight line, the trailer is subjected to gravity G and tension F. l , friction force Ff , and the wheel-rail vertical force F v Among them, G includes the total weight of the material, conveyor belt (not shown), and trailer; the pulling force F l It is the combined tension of the conveyor belt, the wire rope tension (forward), and the wire rope drag force (backward), wherein the wire rope (not shown) is a traction structure connected to the frame 11 for pulling the trolley along the track 2.

[0060] The force on a single wheel assembly in a straight line is: Figure 5 It can be seen that the connection between the wheel axle 14 and the frame 11 bears the gravity G1 and the support force F v1 , tension F l1 , resistance F z1 , lateral force F a1 , axial force F a '1. Where G1 includes the total weight of the material, conveyor belt, and frame 11; F v1 Provided by bearing 15; F l1 Transmitted from the frame 11 to the axle 14; F z1 From wheel body 13; F a1 is the lateral slip force, which is caused by the uneven force on the conveyor belt and the frame 11 in the lateral direction; F' a1 The force from the lateral friction of the track 2 and the force from the wheel flange to the side of the track 2. Similarly, the wheel body 13 (including the bearing 15) bears the gravity G2, the wheel-rail vertical force F v2 , tension F l2 , friction resistance F f2 , lateral force F a2 , axial force F′ a2 Wherein, G2 includes the total weight of the material, conveyor belt, frame 11, and axle 14; F l2 Transmitted from the axle 14 to the wheel body 13; F a2 is the lateral slip force, caused by the uneven force on the trailer in the lateral direction; F' a2 It is the resultant force of the lateral friction of track 2 and the force acting on the side of the wheel flange and track 2.

[0061] From the above analysis, it can be seen that when the trailer is traveling in a straight section, the elastic structure is installed, and the connection between the wheel axle 14 and the mounting seat 12 is subjected to both shear force and axial force, with shear force being dominant and axial force being relatively small. In addition, the elastic structure has a vibration reduction effect in the axial, radial, and circumferential directions, which reduces the vertical force f of the track 2 for the vulnerable part bearing 15. v2 and the axial force F′ a2 The dynamic amplitude is reduced to achieve the effect of reducing vibration acceleration and vertical and lateral wheel-rail force peaks, thereby reducing the equivalent dynamic load of bearing 15 and extending the life of bearing 15.

[0062] Advantages of Four-Point Positioning: Assuming the track 2 has a flat surface, if the wheel axle 14 and mounting base 12 are rigidly connected without an elastic structure, one of the wheel assemblies will inevitably derail, as the three wheel assemblies define a plane at three points. This puts greater pressure on each wheel assembly in the three-point support. However, by adding an elastic structure between the wheel axle 14 and mounting base 12, the wheel assembly that was originally prone to derailment will contact the track surface under the influence of the material and the conveyor belt's gravity, resulting in a four-point support structure for the four wheel assemblies. This four-point support provides more balanced force on the wheel assemblies, stronger adhesion of the wheel body 13, and smoother operation.

[0063] Example 2: Figure 6 、 Figure 7 and Figure 8 The figure shows the force analysis of the running state of the trailer on the turning section of track 2. When the trailer enters the turning section, it is subjected to the centrifugal force F c and the centripetal force F generated by the conveyor belt bending x Influence, forming a medial elevated state.

[0064] in, Where m is the total mass, v is the conveyor belt speed, and R is the radius of gyration. Obviously, the centripetal force F x It is proportional to the tension and inversely proportional to the radius of gyration R. Therefore, it is assumed that: Where T is the conveyor belt tension (longitudinal tension). Since the centripetal force generated by the conveyor belt tension is greater than the centrifugal force, it is necessary to raise the inner track 2 so that gravity produces separation in the direction of the centrifugal force. Let the raising angle be α, according to the mechanical equilibrium condition, we have

[0065]

[0066] Further:

[0067]

[0068] The conveyor belt tension T can be calculated according to the conveyor belt design manual and will not be elaborated here.

[0069] Force on a single wheel in a turning section: From the above analysis, it can be seen that when the trailer is turning, the axial force on the wheel will increase significantly due to the influence of tension. Under the action of the elastic structure, part of the lateral tension from the belt is buffered. Therefore, the lateral vibration or displacement of the wheel body 13 will be reduced, and the lateral impact on the bearing 15 will also be reduced.

[0070] Combined with the above force analysis of the trailer in the straight section and the turning section, it can be seen that the elastic structure adopted has the function of bearing shear force and normal pressure at the same time.

[0071] Forces on the two wheels in the turning section: As mentioned above, due to the centripetal force and centrifugal force, the axial force on the wheel assembly in the turning section increases significantly and becomes more unstable. Figure 9 As shown, if the frame 11 and the wheel axle 14 are rigidly connected, the two wheel assemblies on both sides of the frame 11 will move synchronously, including axial displacement and radial displacement. This increases the overturning moment of the entire trailer and is likely to cause derailment. Figure 10 As shown, after adding an elastic structure between the axle 14 and the mounting seat 12, the flexibility of the axles 14 at both ends of the frame 11 is improved, and a certain degree of deflection is allowed in the axial, radial and circumferential directions, thereby reducing the overall vibration peak of the trailer and increasing the adhesion of the wheel body 13 when turning to avoid derailment.

[0072] This application does not specify the specific form of the elastic structure, but only requires that it be a structural member with a certain degree of elastic deformation. The arrangement of the elastic structure between the axle 14 and the mounting seat 12 is also not specifically limited, and the elastic structure can be arranged in a lattice, closed, stepped, spiral, or gradient shape, etc. As a preferred embodiment, the elastic structure can include one or more elastic bodies 17 spaced apart along the axial direction of the axle 14, with the elastic bodies 17 being configured as a closed structure extending circumferentially around the axle 14. Those skilled in the art will appreciate that by constructing the elastomer 17 as a closed structure extending circumferentially around the axle 14, on the one hand, the elastomer 17 is integrally sleeved on the outer circumferential surface of the axle 14, facilitating installation and radial positioning; on the other hand, the load borne by the axle 14 in each radial direction can be stored and consumed through the elastomer 17, thereby improving the elastic structure's ability to respond to forces in all directions when distributing forces when the trailer passes through complex road sections, thereby improving the force balance of the wheel assembly and the adhesion of the wheel body 13 to the track 2, so that the trailer has a better vibration reduction effect and vibration peak shaving effect when dealing with complex road sections. It should be noted that when the elastomer 17 is in a closed shape as a whole, in order to improve its adaptability to the outer periphery of the axle 14, it is preferred that the elastomer 17 is adapted to the cross-sectional shape of the axle 14. For example, when the elastomer 17 is sleeved on the shaft section of the axle 14 with a circular cross-section, the elastomer 17 can be constructed as an annular closed structure. For another example, when the elastomer 17 is sleeved on the shaft section of the axle 14 with a rectangular cross-section, the elastomer 17 can be constructed as a rectangular closed structure.

[0073] Further preferably, the elastomer 17 is compressed between the mounting seat 12 and the wheel axle 14, and the compression rate λ of the elastomer 17 is ≤ 30%. It will be understood by those skilled in the art that the compression rate of the elastomer 17 after being compressed between the mounting seat 12 and the wheel axle 14 is a key factor in balancing the rigidity and vibration reduction performance of the trailer structure. If the compression rate of the elastomer 17 is too large, the rigidity of the elastomer 17 itself and even the trailer structure will be too large, while the vibration reduction performance will be weakened. If the compression rate of the elastomer 17 is too small, the vibration reduction performance of the trailer will be improved while losing the strength environment that is most favorable for the operation of the bearing 15. Therefore, the present invention obtains the optimal solution for the compression rate after comprehensively considering various factors, that is, the compression rate λ of the elastomer 17 is ≤ 30%, which satisfies the excellent vibration reduction performance of the trailer and improves the noise reduction effect while ensuring that the bearing 15 operates in the most favorable strength environment, thereby effectively extending the service life of the bearing 15. It is understandable that the compression rate of the elastic body 17 when the trailer is unloaded and the compression rate of the elastic body 17 when loaded are necessarily different. The compression rate of the elastic body 17 at different circumferential positions will also vary. In particular, under the same operating conditions, the portion of the elastic body 17 located above the wheel axle 14 and the portion located below the wheel axle 14 will exhibit different compression rates due to the difference in the loads they are subjected to. The limiting condition of the elastic body 17 compression rate λ≤30% in this application applies to both the trailer's unloaded and loaded states, and also to the compression rate at all circumferential positions of the elastic body 17. That is, the compression rate λ≤30% must be maintained at any position of the elastic body 17 under any trailer operating condition. Later in this specification, in the trailer vibration reduction method, the optimal solution for the compression rate λ≤30% of the elastic body 17 will be demonstrated by way of examples, which will not be elaborated here.

[0074] In a preferred embodiment, the elastic body 17 is configured as a rubber ring with a circular, polygonal or irregular cross-section. Figure 2 As shown, the elastic body 17 is configured as a rubber ring with a circular cross section, that is, the elastic body 17 is an O-ring. The number and material of the O-ring can be adjusted according to the load, and it is easy to replace after damage. The clearance of the mounting hole 121 is adjusted according to the load to achieve the purpose of vibration reduction and noise reduction; Figure 12 The elastic body 17 is shown as a rubber ring with a conical cross-section. This type of rubber ring can adjust the elastic modulus and thickness of the material according to the load, has good shock absorption capabilities, is simple to position, and is convenient to fix. It ensures that the radial and axial stiffness and vibration of the trailer are well balanced during operation. The conical rubber ring can also absorb some axial vibration and reduce rigid collision. Of course, in actual production, if there are special requirements for the structure of the wheel axle 14 and the mounting seat 12, or for the installation of the elastic body 17, the cross-section of the elastic body 17 can be set to an irregular shape to improve adaptability.

[0075] In a preferred embodiment, the elastomer 17 is configured as a single material, which includes an elastic material that can undergo elastic compression. For example, the elastomer 17 can be made of only a single material such as rubber, which can utilize the viscosity and elasticity of the material itself for vibration reduction and noise reduction. In the actual implementation process, the number of elastomers 17 can be increased or decreased according to the load of the conveyor to meet the requirements of bearing stiffness and vibration reduction effect. In another preferred embodiment, the elastomer 17 can also be configured as a composite material, which includes a support skeleton and an elastic material that can undergo elastic compression surrounding the outside of the support skeleton. For example, the support skeleton can be made of a metal material to form a stable internal support base for the elastomer 17, and the elastic material is coated on the outside of the support skeleton by many suitable optional methods such as pasting, hot melting, and socketing. Among them, the elastic material can be a polymer material such as rubber.

[0076] In a preferred embodiment where the elastomer 17 is configured as an O-ring with a circular cross-section, as Figure 2 shown, an annular groove 141 adapted to a single O-ring can also be provided on the outer peripheral surface of the wheel axle 14. The cross-sectional area S2 of the annular groove 141 and the cross-sectional area S1 of the O-ring satisfy: 0.6 ≤ S1 / S2 ≤ 1.2. Among them, the cross-sectional area S1 is the standard area of the O-ring in the initial uncompressed state. The depth h of the annular groove 141 in the radial direction of the mounting hole 121 or the wheel axle 14 and the cross-sectional diameter d of the O-ring satisfy: 0.5 < h / d < 1. The cooperation of the O-ring and the annular groove 141, on the one hand, can use the side walls on both sides of the annular groove 141 to form axial limits for the O-ring to prevent the O-ring from shifting axially along the wheel axle 14 and affecting the working performance. On the other hand, it can also store and consume the load received by the wheel axle 14 in the axial direction through the abutment of the elastomer 17 and the side walls on both sides of the annular groove 141, forming vibration reduction in the axial direction of the wheel assembly. During specific assembly, the elastomer 17 can be first sleeved on the wheel axle 14 and pre-tightened in the annular groove 141, and then the wheel axle 14 can be fitted with the mounting hole 121 of the mounting seat 12. In addition, the reasonable matching of the sizes of the O-ring and the annular groove 141 is an important guarantee for improving the vibration reduction effect and extending the service life of the O-ring. In this application, by making the cross-sectional area S? of the annular groove 141 and the cross-sectional area S1 of the O-ring satisfy: 0.6 ≤ S1 / S2 ≤ 1.2, and making the depth h of the annular groove 141 in the radial direction of the mounting hole 121 or the wheel axle 14 and the cross-sectional diameter d of the O-ring satisfy: 0.5 < h / d < 1. Taking the cross-sectional area S2 of the annular groove 141 as 1 cm 2 and the depth h as 1 cm as an example, the range of the cross-sectional area S1 of the O-ring can be 0.6 - 1.2 cm 2The cross-sectional diameter d of the O-ring can range from 1 to 2 cm, excluding the range boundaries of 1 cm and 2 cm. This rationally optimizes the size matching between the O-ring and the annular groove 141, effectively preventing the O-ring from being too large or too small. Macroscopically, this size design ensures that the inner and outer sides of the O-ring are in stable and reliable contact with the axle 14 and the mounting seat 12, respectively, and has ample deformation space and stable deformation capacity, ensuring long-term and effective vibration reduction. Microscopically, this size design helps further rationally optimize important vibration-related factors, including the pre-compression rate of the O-ring when unloaded, the compression rate after loading, stiffness, quantity, and resonance. It should be noted that Figure 2 The embodiment in which five O-rings are arranged between the mounting seat 12 and the axle 14 is only an illustrative embodiment and does not limit the present application. Other appropriate numbers of O-rings can be reasonably selected according to relevant vibration reduction requirements.

[0077] The present application does not limit the structure of the annular groove 141. As a preferred embodiment, Figure 2 As shown, the annular groove 141 is configured as a structure with a rectangular cross section. As another preferred embodiment, the annular groove can also be configured as a structure with a conical cross section.

[0078] The present application does not specifically limit the structure of the mounting base 12, and includes but is not limited to the following embodiments:

[0079] Example 3: Figures 1 to 3 As shown, the mounting base 12 is constructed as an integrally formed structure, with mounting holes 121 extending through both ends of the mounting base 12. Specifically, to ensure structural strength, the mounting base 12 can be constructed from a one-piece metal block manufactured through processes such as stamping and cutting. Of course, other suitable materials may also be used. Because the mounting hole 121 extends through both ends of the mounting base 12, the axle 14 can pass through one end of the mounting hole 121 and exit through the other end. The exiting end can be used to position the mounting base 12.

[0080] Example 4: Figure 11As shown, the mounting seat 12 includes a first mounting body 122 and a second mounting body 123 arranged at intervals, the mounting hole 121 includes a first through hole 1211 passing through the first mounting body 122 and a second through hole 1212 passing through the second mounting body 123, the first mounting body 122 is close to the wheel body 13, and the second mounting body 123 is located on the side of the first mounting body 122 away from the wheel body 13, and the elastic structure is distributed between the first mounting body 122 and the wheel axle 14 and / or between the second mounting body 123 and the wheel axle 14. Those skilled in the art will appreciate that by constructing the mounting base 12 as a split structure consisting of a first mounting body 122 and a second mounting body 123, it is convenient for the first mounting body 122 and the second mounting body 123 to be independently designed and processed into forms that respectively play different roles on the wheel axle 14. For example, the first through hole 1211 can be constructed as a circular through hole that matches the circular cross-section of the wheel axle 14, and the second through hole 1212 can be constructed as a long strip hole or an irregular hole with a limiting surface, and a mating surface that fits the limiting surface is provided on the outer surface of the wheel axle 14. In this way, after the wheel axle 14 passes through the second mounting body 123, the cooperation between the limiting surface and the mating surface can limit the rotation of the wheel axle 14 relative to the mounting base 12. Specifically, considering that the axle 14 in the first mounting body 122 is subjected to the greatest force during the operation of the trailer, an elastic structure can be provided only in the first through hole 1211, that is, the elastic structure is distributed between the first mounting body 122 and the axle 14, and no elastic structure is provided between the second mounting body 123 and the axle 14. As a preferred embodiment, Figure 11 As shown, a number of elastic bodies 17 can be arranged between the first mounting body 122 and the axle 14 and between the second mounting body 123 and the axle 14, so that energy can be dissipated and vibration can be reduced between the axle 14 and the first mounting body 122 and the second mounting body 123 respectively through the elastic bodies 17.

[0081] Regarding the axial limitation method of the wheel axle 14 after the wheel axle 14 is matched with the mounting seat 12, in a preferred embodiment, as shown in FIG. Figure 11 As shown, positioning baffles 18 can be provided at both ends of the mounting base 12. Positioning baffles 18 are used to position the wheel axle 14. Positioning baffles 18 can be connected to the wheel axle 14 via locking members 19. For example, locking members 19 can be locking screws, thereby preventing the wheel axle 14 from moving left or right along the axial direction. Elastic shock-absorbing pads 20 can be provided between the positioning baffles 18 and the mounting base 12 to provide axial cushioning, absorbing and cushioning the impact of axial loads, and thus achieving vibration and noise reduction.

[0082] Regarding the fixing method of the mounting base 12 on the frame 11, this application does not specifically limit it, and includes but is not limited to the following embodiments:

[0083] Example 5: Welding the mounting base 12 to the vehicle frame 11. If the mounting base 12 is integrally formed, it can be welded to the vehicle frame 11 as a whole. If the mounting base 12 is a split structure consisting of a first mounting body 122 and a second mounting body 123, the first mounting body 122 and the second mounting body 123 can be welded to the vehicle frame 11 separately.

[0084] Example 6: The mounting base 12 is connected to the vehicle frame 11 via threaded fasteners. If the mounting base 12 is integrally formed, it can be fixed to the vehicle frame 11 as a whole using threaded fasteners such as screws and bolts. If the mounting base 12 is a separate structure consisting of a first mounting body 122 and a second mounting body 123, the first mounting body 122 and the second mounting body 123 can be fixed to the vehicle frame 11 separately using threaded fasteners such as screws and bolts.

[0085] Example 7: The mounting base 12 is riveted to the frame 11. If the mounting base 12 is integrally formed, it can be riveted to the frame 11 as a whole. If the mounting base 12 is a split structure consisting of a first mounting body 122 and a second mounting body 123, the first mounting body 122 and the second mounting body 123 can be riveted to the frame 11 separately.

[0086] As a preferred embodiment of the present application, Figure 2 and Figure 13 As shown, the mounting base 12 is provided with a tapered guide slope 124 at the end of the mounting hole 121 near the wheel body 13. The tapered guide slope 124 is used to guide the elastic structure to be installed in the mounting hole 121. Specifically, taking the embodiment in which the mounting base 12 is composed of the first mounting body 122 and the second mounting body 123 as an example, in order to facilitate the installation of the elastic body 17 between the wheel axle 14 and the first mounting body 122 and on the wheel axle 14, the tapered guide slope 124 is provided on the side of the first mounting body 122 near the wheel body 13. Those skilled in the art will understand that in order to ensure that the elastic structure has a vibration reduction effect between the mounting seat 12 and the axle 14, the outer contour of the elastic structure must be larger than the inner contour of the mounting hole 121 provided in the mounting seat 12. Therefore, the end face of the mounting seat 12 will form a stop for the elastic structure and make it difficult to assemble the elastic structure into the mounting hole 121. Therefore, by providing the mounting seat 12 with a conical guide slope 124 at the end edge of the mounting hole 121, the conical guide slope 124 can be used to squeeze the elastic structure during the assembly process of the elastic structure, causing it to deform and guide the elastic structure into the mounting hole 121.

[0087] Compared to the aforementioned embodiment in which the elastic body 17 is installed in the annular groove 141 provided on the axle 14, as another preferred embodiment of the present application, the mounting base 12 can also be provided with a mounting groove on the inner wall of the mounting hole 121, and the elastic structure can be installed in the mounting groove. During assembly, the elastic structure can be first installed in the mounting hole 121 and pre-tightened in the mounting groove, and then the axle 14 is matched with the mounting hole 121. Specifically, the structure of the mounting groove should be compatible with the elastic structure. For example, if the elastic structure is an annular elastic body, the mounting groove can be provided as an annular mounting groove provided along the circumference of the inner wall of the mounting hole 121.

[0088] The present application provides a trailer vibration reduction method, based on the aforementioned embodiment in which the elastomer 17 is compressed between the mounting seat 12 and the wheel axle 14 with a compression rate λ≤30%, and the elastomer 17 is an O-ring that cooperates with the annular groove 141. Based on the size limitations of the O-ring and the annular groove 141, the O-ring is in a pre-tightened and compressed state in the annular groove 141. In this pre-tightened and compressed state, the O-ring belongs to an extrusion elastic vibration reduction method, which relies on being pre-extruded in the annular groove 141 and elastically deformed to generate a pre-tightening force. At the same time, the pressure of the material medium also squeezes the O-ring during operation, causing it to generate a self-tightening force. Therefore, when selecting the optimal decompression rate of the O-ring, many factors and parameters such as stiffness, resonance frequency, contact pressure, avoidance of permanent deformation, and service life should be comprehensively considered to ensure that the O-ring has excellent load-bearing performance, reduces the risk of resonance, and increases its service life.

[0089] The following demonstrates that the compression rate of the O-ring λ≤30% is the optimal solution through the vibration reduction method:

[0090] 1. Resonance control strategy of trailer oscillation system

[0091] The natural frequency ω0 of the trailer system satisfies:

[0092]

[0093] Where: m1 is the total mass borne by the O-ring, in kg; K is the combined stiffness, in N / m.

[0094]

[0095] Where: K1 is the stiffness coefficient of the O-ring, and K2 is the stiffness coefficient of the bearing.

[0096] because Therefore, increasing K1 will increase the natural frequency and increase the risk of resonance. In engineering, the general approach is to increase mass or reduce stiffness to avoid resonance. Since K2>>K1, K≈K1.

[0097] The main frequency of wheel-rail rolling vibration of wheel-rail belt conveyor is above 50Hz. According to engineering experience, the natural frequency of belt conveyor generally does not resonate below 20Hz. Therefore, if 20Hz is used as the standard, the resonance frequency f should meet the following requirements:

[0098]

[0099] If 50Hz is used as the standard, the resonant frequency f (unit: Hz) should meet the following requirements:

[0100]

[0101] If the resonant frequency standard is expressed as S, the resonant frequency f should satisfy:

[0102]

[0103] Further:

[0104] K1<4π 2 S 2 m1

[0105] Table 1 below is the stiffness test data table of the O-ring. According to the stiffness test data trend in Table 1 below, the stiffness prediction curve of the O-ring is obtained as follows:

[0106] y=e bx

[0107] Where b is the transfer coefficient between λ and k1, 26≤b≤30.

[0108] Therefore, the relationship between the compression rate and stiffness of the O-ring is:

[0109] k1=e bλ ×10 3 (N / m)

[0110] In other words, the stiffness of the O-ring is set to k1, k1 = e bλ N / mm, 26≤b≤30.

[0111] Table 1

[0112]

[0113] Furthermore, the number of O-rings is set to n. From the above, the relationship between the compression rate λ and the stiffness of n rubber rings is:

[0114] K1=nk1=ne bλ ×10 3 (N / m)

[0115] From the above, it can be concluded that in order to effectively prevent resonance in the trailer vibration reduction system, the stiffness k1 of the O-ring and the resonance frequency f, the total mass m1 borne by the O-ring and the number n of O-rings must meet the following requirements:

[0116] Among them, 0 <f≤50Hz;

[0117] In other words, bλ ×10 3 <4π 2 f 2 m1; take the standard, that is: ne bλ ×10 3 <4π 2 S 2 m1.

[0118] Specifically, in, is the original dynamic load on the O-ring, m1 is the total mass of the O-ring, g is the acceleration due to gravity, and a is the acceleration of the trailer. Since a has nonlinear characteristics, the peak acceleration a can be used. max Instead, d is the cross-sectional diameter of the O-ring and h is the depth of the annular groove.

[0119] When b=28, in order to prevent the trailer system from resonating, then:

[0120] ne 28λ ×10 3 <4π 2 S 2 m1

[0121] Further:

[0122]

[0123] When m1=300kg, n=1, S=20Hz, λ should be <30%. If S=50Hz, λ should be <37%.

[0124] When m1 = 300kg, n = 2, S = 20Hz, λ should be < 28%. If S = 50Hz, λ should be < 34%.

[0125] When m1 = 300kg, n = 4, S = 20Hz, λ should be < 25%. If S = 50Hz, λ should be < 32%;

[0126] When m1 = 300kg, n = 6, S = 20Hz, λ should be < 24%; if S = 50Hz, λ should be < 30%;

[0127] When m1=300kg, n=8, S=20Hz, λ should be <23%; if S=50Hz, λ should be <29%.

[0128] 2. Prediction method of bearing service life

[0129] The system vibration attenuation can be expressed by the damping ratio ζ:

[0130]

[0131] Where: c1 is the damping coefficient.

[0132] The trailer system is an underdamped system (0<ζ<1), and the natural frequency ω0 and the damping ratio ζ have the following relationship:

[0133]

[0134] Further:

[0135]

[0136] Further:

[0137]

[0138] Therefore, in the trailer vibration system, increasing the stiffness of the O-ring will reduce the damping ratio, causing the damping ratio to be far away from the critical damping (ζ=1), which is not conducive to vibration attenuation.

[0139] Let the service life of the bearing be L 10 , according to the bearing life calculation formula:

[0140]

[0141] Among them, C r is the rated load of the bearing, K1 is the stiffness coefficient of the O-ring, and K2 is the stiffness coefficient of the bearing.

[0142] If the trailer system does not have an O-ring, there are:

[0143]

[0144] Where: X2 is the bearing displacement.

[0145] Assuming the damping ratio ζ is equal to the decay rate c of the bearing displacement X2, after adding the O-ring, the bearing load Q2 becomes:

[0146] Q2=K2X2-cK2X2=K2X2-ζK2X2

[0147] Therefore, the larger ζ is, the higher the vibration attenuation is, and the above formula can be written as:

[0148]

[0149] Q2 can be written as:

[0150]

[0151] Further:

[0152]

[0153] The relationship between bearing life and compression rate is:

[0154]

[0155] or:

[0156]

[0157] λ and L 10 It is negatively correlated; further, through life experiments, the relationship between the rubber ring compression rate λ and the bearing vibration acceleration a can be obtained, and then the equivalent load value can be directly obtained by Q2=mg+ma to calculate the bearing life.

[0158] 3. Prediction method of O-ring service life

[0159] O-rings are polymer materials. The fatigue life of polymer materials is directly related to strain, and strain is directly related to compression rate. The following table 1 is the O-ring strain slope difference table, λ i Indicates that the compression amount is i%. It can be seen that the strain change is the largest when the compression rate of the O-ring is 30%, reflecting that the fatigue life of the O-ring decreases fastest when the compression amount exceeds 30%.

[0160] Table 2

[0161]

[0162] Furthermore, the relationship between the compression rate λ of the O-ring and the service life L0 of the O-ring can be obtained based on the experimental data of the service life of the O-ring.

[0163] From the above analysis, it can be seen that as the compression rate λ of the O-ring increases, the stiffness k1 of the O-ring increases and the load-bearing performance is enhanced. When the compression rate λ exceeds 23%-37%, the resonance risk increases and the life L0 of the O-ring decreases. When the compression rate λ exceeds 30%, the strain increases most significantly, and the bearing life L 10 Therefore, the determination of the compression rate λ of the O-ring needs to consider k1, f, L0, L 10 Factors such as λ should be guaranteed to be ≤ 30%.

[0164] During the operation of the trolley, the wheel body 13 rolls on the track 2. The material load and wheel-rail force experience energy storage and energy consumption when passing through the O-ring, so that the overall vibration of the trolley is attenuated. The closer the damping ratio ζ is to 1, the better the effect. Among them, the force on the O-ring is divided into two stages: in the first stage, the cross-sectional shape of the O-ring is improved by pre-tightening the mounting seat 12, thereby improving the load-bearing capacity or stiffness of the O-ring. The above stiffness experiment proves that the O-ring is compressed in a limited space, and its stiffness shows an obvious nonlinear change with the compression rate; in the second stage, an axial load is applied to the O-ring, and the O-ring is uniaxially compressed. Since the stiffness of the O-ring in the first stage has been greatly improved, the displacement of the O-ring in the second stage is small, which can be approximately regarded as a linear deformation.

[0165] To sum up, the trailer vibration reduction method of the present application, based on the method of matching the size of the annular groove 141 with the size of the O-ring, further integrates the prediction of O-ring stiffness with compression rate, calculation of the number of O-rings, trailer oscillation system resonance prevention strategy, and bearing service life with compression rate prediction method, so that without large-scale modification of the overall structure of the trailer, a good vibration reduction effect can be achieved, the risk of trailer resonance is greatly reduced, and the service life of O-rings and bearing wearing parts is increased as much as possible. Moreover, the present method can improve the vibration reduction effect by changing the compression rate, quantity, and material of the O-ring according to actual requirements, and the O-ring is easy to maintain and replace, the overall structure is simple, the stability is good, and it has strong practicality.

[0166] Anything not described in this application can be achieved by adopting or drawing on existing technologies.

[0167] The various embodiments in this specification are described in a progressive manner, and the same or similar parts between the various embodiments can be referred to each other. Each embodiment focuses on the differences from other embodiments.

[0168] The foregoing is merely an embodiment of the present application and is not intended to limit the present application. For those skilled in the art, the present application may have various changes and variations. Any modifications, equivalent replacements, improvements, etc. made within the spirit and principles of the present application should all be included within the scope of the claims of the present application.

Claims

1. A trolley, used for a wheel-rail type belt conveyor, for supporting and pulling the conveyor belt on the track of the wheel-rail type belt conveyor and traveling along the conveyor belt, characterized in that: The trailer includes a frame and a plurality of wheel assemblies distributed on both sides of the frame, the frame is provided with mounting seats corresponding to the wheel assemblies one by one, the wheel assembly includes a wheel body, a wheel axle and a bearing, the wheel axle is transmission-connected to the wheel body through the bearing, at least a portion of the wheel axle is passed through a mounting hole provided on the mounting seat and is gap-fitted with the mounting hole, and an elastic structure is provided in the mounting hole between the mounting seat and the wheel axle.

2. The trailer according to claim 1, characterized in that: The elastic structure includes one or a plurality of elastic bodies arranged at intervals along the axial direction of the wheel axle, and the elastic body is configured as a closed structure extending around the circumference of the wheel axle.

3. The trailer according to claim 2, characterized in that: The elastic body is compressed between the mounting seat and the wheel axle, and a compression rate λ of the elastic body is ≤30%.

4. The trailer according to claim 3, characterized in that: The elastic body is configured as an O-ring with a circular, polygonal or irregular cross-section.

5. The trailer according to claim 3, characterized in that: The elastic body is configured as a single material, which includes an elastic material that can be elastically compressed; Alternatively, the elastomer is configured as a composite material, which includes a support skeleton and an elastic material that is elastically compressible and surrounds the support skeleton.

6. The trailer according to claim 4, characterized in that: The elastomer is configured as an O-ring with a circular cross-section, and an annular groove adapted to a single O-ring is provided on the outer peripheral surface of the wheel shaft. The cross-sectional area S2 of the annular groove and the cross-sectional area S1 of the O-ring satisfy: 0.6≤S1 / S2≤1.2, wherein the cross-sectional area S1 is the standard area of ​​the O-ring in the initial uncompressed state, and the depth h of the annular groove along the radial direction of the mounting hole or the wheel shaft and the cross-sectional diameter d of the O-ring satisfy: 0.5<h / d<1.

7. The trailer according to claim 6, characterized in that: The annular groove is configured as a structure with a rectangular or conical cross section.

8. The trailer according to claim 1, characterized in that: The mounting seat is constructed as an integrally formed structure, and the mounting holes pass through both ends of the mounting seat; Alternatively, the mounting seat includes a first mounting body and a second mounting body arranged at intervals, the mounting hole includes a first through hole passing through the first mounting body and a second through hole passing through the second mounting body, the first mounting body is close to the wheel body, the second mounting body is located on the side of the first mounting body away from the wheel body, and the elastic structure is distributed between the first mounting body and the wheel axle and / or between the second mounting body and the wheel axle.

9. The trailer according to claim 8, characterized in that: The mounting seat is welded to the vehicle frame; Alternatively, the mounting base is connected to the vehicle frame via threaded fasteners; Alternatively, the mounting seat and the vehicle frame are riveted.

10. The trailer according to claim 1, characterized in that: The mounting seat is provided with a conical guiding slope at the end of the mounting hole close to one end of the wheel body, and the conical guiding slope is used to guide the elastic structure to be installed into the mounting hole; And / or, the mounting seat is provided with a mounting groove on the inner wall of the mounting hole, and the elastic structure is mounted in the mounting groove.

11. A method for reducing vibration of a trailer, applied to the trailer according to claim 6, characterized in that: The vibration reduction method includes: the stiffness of the O-ring is set to k1, The prediction method of O-ring stiffness with compression rate is as follows: O-ring stiffness k1 = e bλ N / mm, where b is the transmission coefficient between λ and k1, 26≤b≤30.

12. The method for reducing vibration of a trailer according to claim 11, characterized in that: The method for determining the number of O-rings is also included, which is specifically, the number of O-rings in, is the original dynamic load on the O-ring, m1 is the total mass borne by the O-ring, g is the acceleration due to gravity, a is the acceleration of the trailer, d is the cross-sectional diameter of the O-ring, and h is the depth of the annular groove.

13. The trailer vibration reduction method according to claim 12, characterized in that: The stiffness k1 of the O-ring and the resonant frequency f, the total mass m1 borne by the O-ring and the number n of O-rings satisfy: Among them, 0 <f≤50Hz。 14. The trailer vibration reduction method according to claim 12, characterized in that: The compression rate λ and resonance frequency f of the O-ring, the total mass m1 borne by the O-ring and the number n of O-rings satisfy: Among them, 0 <f≤50Hz。

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