A method for optimizing a bogie of a rail vehicle

Abstract

The application discloses an optimization method of a rail vehicle bogie, which optimizes the rail vehicle bogie by optimizing a suspension device and a suspension component of the rail vehicle bogie. The suspension device is optimized by strain design of a conical spring rubber layer of a primary conical spring in a limited space, anti-falling performance optimization of a lateral stopper and torsional stiffness design of a support ball hinge in an anti-rolling torsion bar device. The suspension component is optimized by anisotropic design of horizontal and vertical stiffness of a center pin sleeve in a traction device and optimization of radial bearing and deflection capacity of a traction rod node. The application can further optimize various performances of the suspension device and the suspension component of the rail vehicle bogie, and meets more and more severe performance requirements of users on the rail vehicle bogie.

Description

Technical Field

[0001] This invention relates to a method for optimizing the performance of a railway vehicle bogie, specifically a method for optimizing a locomotive bogie by optimizing the suspension device and suspension components of the railway vehicle bogie, belonging to the field of locomotive bogie technology. Background Technology

[0002] The bogie of a rail vehicle is the most critical system for ensuring the safety of rail vehicle operation and improving its speed and comfort. During normal operation of a rail vehicle, the bogie plays a vital role in load bearing, guidance, traction, and braking.

[0003] The bogie suspension system is a crucial component of the bogie, and its characteristics directly determine the magnitude of the wheel-rail interaction force, thus influencing the operational stability of the rail vehicle. The bogie suspension system also directly impacts the ride smoothness of the rail vehicle; its parameters directly determine the vehicle's running quality. Modern rail vehicle bogie suspension systems are generally divided into primary and secondary suspension systems. The primary suspension system mainly includes primary conical springs, vertical hydraulic dampers, and axle boxes between the wheelsets and the frame. The secondary suspension system mainly includes traction devices, air spring systems, lateral stops, lateral hydraulic dampers, vertical hydraulic dampers, and anti-roll torsion bar systems between the bogie and the car body. Due to track irregularities and other factors, the wheel-rail interaction force constantly changes when the rail vehicle is running on the track, thus exciting vibrations in the rail vehicle system. To attenuate these vibrations, ensure passenger comfort, and improve the dynamic quality of the rail vehicle, most rail vehicle bogies use a parallel connection of axle box steel springs and vertical hydraulic dampers in their primary suspension to initially attenuate high-frequency vibrations from the wheelsets. The secondary suspension connects the car body and the bogie, separating the car body and the bogie from each other, thereby attenuating the high-frequency vibrations transmitted from the bogie to the car body.

[0004] The traction device of a railway vehicle bogie consists of a center pin seat, center pin, traction seat, and traction rod. It is a crucial component connecting the bogie to the car body, transmitting the traction force generated by the interaction between the wheel and rail, as well as the braking force generated by the braking system. The traction device must ensure that it does not interfere with the relative movement of the bogie and car body during normal operation of the railway vehicle, while also accommodating the angle of rotation of the car body relative to the bogie when the railway vehicle passes through a curve.

[0005] With technological advancements, the requirements for railway vehicle bogies are becoming increasingly stringent, such as high integration and applicability to diverse track conditions worldwide. Consequently, the demands on bogie suspension and traction systems are also rising, including stringent space requirements to achieve performance within a minimal footprint while maintaining high reliability. Components must be designed for lightweighting, modularity, and adjustability to ensure versatility across different platforms and projects. Key load-bearing components, such as primary conical springs and stoppers, often require non-linear characteristics to meet the end-user's demands for high comfort and large load capacity. Therefore, further optimization of the suspension system and its components in railway vehicle bogies is necessary.

[0006] Chinese invention patent CN109858120B, with an authorization announcement date of April 7, 2023, discloses a method for optimizing the dynamic parameters of a bogie suspension device in a high-speed train. The method includes: establishing a multibody dynamics model of the vehicle system; obtaining the actual vibration modes of the target vehicle; calculating the vibration modes of the multibody dynamics model of the vehicle system based on the root locus method; determining whether the calculated vibration modes of the multibody dynamics model of the vehicle system are consistent with the actual vibration modes of the target vehicle; if not, correcting the multibody dynamics model of the vehicle system and returning to the step of calculating the vibration modes of the multibody dynamics model of the vehicle system based on the root locus method; if yes, determining the dynamic parameters of the suspension device in the target vehicle that need optimization; and optimizing the dynamic parameters of the suspension device in the target vehicle based on the multibody dynamics model of the vehicle system.

[0007] Chinese invention patent CN109918793B, with an authorization announcement date of June 7, 2022, discloses a rapid optimization method for bogie suspension parameters. The steps are as follows: First, establish a vehicle dynamics simulation model based on the structural parameters of the high-speed train; Second, perform suspension parameter optimization analysis based on the constructed vehicle dynamics simulation model, determining the required input variables, value range, and output evaluation indicators. The input variables include the primary suspension longitudinal stiffness, primary suspension lateral stiffness, anti-hunting damper damping, secondary suspension lateral damper damping, anti-hunting damper joint stiffness, and secondary suspension lateral stiffness, which affect the lateral stability of the train. The six key suspension parameters for shock absorber joint stiffness are analyzed. The input variables are the suspension parameters to be optimized. Evaluation indices include lateral stability, equivalent cone robustness, and suspension parameter robustness. The output evaluation index is the optimization objective, with its value range determined based on existing suspension parameter values. Third, a joint simulation module is established between the vehicle dynamics model simulation and MATLAB using the SIMPACK scripting language. The NSGA-II multi-objective optimization algorithm is employed, setting the population size n, crossover probability Pc, mutation probability Pm, and iteration count N to update the design parameters. The NSGA-II algorithm is then evaluated based on the quantity and distribution of optimization results. The parameters of the I algorithm are adjusted, with n set to 5000 and N to 100 times. Fourth, based on the optimization objective, eigenvalue simulation calculations are performed on the vehicle dynamics model, and the optimized objective values ​​obtained from the simulation are extracted through the SIM PACK and MATLAB joint simulation interface. Fifth, the output evaluation index values ​​that meet the evaluation index range for the first time after simulation optimization are used as the Pareto front data. For subsequent output evaluation index values ​​that meet the evaluation index range, it is necessary to determine whether they are better than the Pareto front data. If not, proceed to step three and use NSGA-I. Algorithm I performs crossover and mutation operations on the design parameters based on the crossover probability Pc and mutation probability Pm to obtain a new set of design parameters, and assigns the updated design parameter values ​​to the dynamic model. If so, the design parameters replace the Pareto front data. Sixth, when Pareto front data is replaced, it is determined whether the current iteration number i is not less than N. If not, it returns to step three to continue updating the design parameter values ​​until the current iteration number i is greater than or equal to N, at which point the optimization design ends. Seventh, finally, the optimized data needs to be processed using the K-Means clustering method to obtain K representative suspension parameter sets. From these, the category that meets the design requirements is selected, and its corresponding parameter set is the bogie suspension parameter for the optimized design.

[0008] Although the aforementioned patent documents all optimize bogies or bogie suspension devices, their disclosed solutions differ from those in this application.

[0009] In summary, designing an optimization method for railway vehicle bogies to further enhance the performance of suspension devices and key suspension components, and to meet the increasingly stringent performance requirements of users for railway vehicle bogies, is an urgent technical problem to be solved. Summary of the Invention

[0010] The technical problem to be solved by the present invention is to provide an optimization method for rail vehicle bogies, which addresses the deficiencies in the existing technology. This method can further optimize the performance of the suspension device and its components in the rail vehicle bogie, and meet the increasingly demanding performance requirements of users for rail vehicle bogies.

[0011] To solve the above-mentioned technical problems, the technical solution adopted by the present invention is as follows: an optimization method for a railway vehicle bogie, which optimizes the bogie by optimizing the suspension device and suspension components of the railway vehicle bogie. The optimization of the suspension device is achieved by designing the primary conical spring rubber layer with equal strain within a limited space, optimizing the anti-fall-off performance of the lateral stop, and increasing the torsional stiffness of the support ball joint in the anti-roll torsion bar device. The optimization of the suspension components is achieved by designing the center pin sleeve in the traction device with anisotropic lateral and longitudinal stiffness and optimizing the traction tie rod nodes to improve radial load-bearing and deflection capabilities.

[0012] Preferably, the tapered spring includes an outer sleeve, a mandrel, and multiple spacers located between the outer sleeve and the mandrel. The outer sleeve, the mandrel, and the multiple spacers are bonded together by multiple rubber bodies through vulcanization. The inner circumferential surface of the outer sleeve and the spacers are both tapered. An end screw hole is also opened on one end of the outer sleeve.

[0013] The method of performing constant strain design of the tapered spring rubber layer in a limited space involves extending the radial width of the annular end toward the central axis of the tapered spring while keeping the outer diameter of the outer sleeve constant, to ensure the machining of the end screw hole, and then increasing the gap between the inner circumferential surface of the outer sleeve and the outermost spacer to achieve constant strain design of multiple rubber bodies located between the outer sleeve, the mandrel and multiple spacers after vulcanization.

[0014] Preferably, in the axial section of the conical spring, the angle between the inclined line of the conical spacer located on one side of the central axis of the conical spring and the central axis of the conical spring is set as θ1, so that θ1 is equal for each spacer, and the angle between the inclined line of the inner circumferential surface of the outer sleeve located on one side of the central axis of the conical spring and the central axis of the conical spring is set as θ2.

[0015] The specific steps of the method for achieving constant strain design of the rubber layer of a conical spring within a limited space are as follows: First, the inclined line at the position of the inner circumferential surface of the outer jacket located on one side of the central axis of the conical spring is set to be parallel to the inclined line at the position of the conical spacer located on one side of the central axis of the conical spring. While ensuring that the outer diameter of the outer jacket remains unchanged, the inclined line at the position of the inner circumferential surface of the outer jacket is moved parallel to the central axis of the conical spring to the inclined line P, so that the radial width of the annular end of the outer jacket is greater than the diameter of the end screw hole. The inclined line P intersects the horizontal line extending horizontally along the annular end of the upper end of the outer jacket at point S. Then, with point S as the center, the inclined line P is rotated by a certain angle away from the central axis of the conical spring to obtain the inclined line Q, so that the inclined line at the position of the inner circumferential surface of the outer jacket is moved to the inclined line Q. Then, θ2 > θ1, so as to increase the gap between the inner circumferential surface of the outer jacket and the outermost spacer. Through the above steps, the constant strain design of multiple rubber bodies located between the outer jacket, the mandrel and multiple spacers after vulcanization is achieved.

[0016] Preferably, the optimization of the anti-detachment performance of the transverse stop involves designing the transverse stop as a base plate, a rubber body, a wear plate, and a cover plate. The base plate, rubber body, and wear plate are vulcanized into a whole, and the cover plate is assembled at the center of the bottom end of the vulcanized whole. The wear plate is frustum-shaped, and the rubber body is also frustum-shaped. The center of the frustum of the rubber body has a through hole along the axial direction, and the wear plate is located in the through hole of the rubber body. The center of the base plate also has a through hole, and the cover plate is installed at the through hole of the base plate. The edge of the base plate near its through hole also has a stepped hole, which allows rubber to flow into the stepped hole to form an anti-detachment buckle, thereby preventing the rubber body from falling off.

[0017] Preferably, the supporting ball joint includes an outer sleeve, a hollow mandrel disposed inside the outer sleeve, and an annular rubber located between the outer sleeve and the hollow mandrel, wherein the outer sleeve and the hollow mandrel are bonded together by vulcanization of the annular rubber;

[0018] The design to increase the torsional stiffness of the supporting ball joint in the anti-roll torsion bar device is to wrap at least one layer of fabric in the annular rubber and along the circumference of the annular rubber as a torsional stiffness reinforcing material to increase the torsional stiffness of the supporting ball joint.

[0019] When at least one layer of fabric is wound around the annular rubber, a layered structure of rubber and fabric is formed inside the annular rubber.

[0020] Before the ball joint is vulcanized, strips of fabric and strips of rubber are prefabricated. The strips of fabric and rubber are then stacked together and rolled into a ring. The ring-shaped fabric and rubber are then placed between the outer jacket and the hollow mandrel for product vulcanization, resulting in a layered structure of rubber and fabric inside the ring-shaped rubber.

[0021] Preferably, the supporting ball joint includes an outer sleeve, a hollow mandrel disposed inside the outer sleeve, and an annular rubber located between the outer sleeve and the hollow mandrel, wherein the outer sleeve and the hollow mandrel are bonded together by vulcanization of the annular rubber;

[0022] The design to increase the torsional stiffness of the supporting ball joint in the anti-roll torsion bar device involves setting a limiting structure one on the outer sleeve and a limiting structure two on the hollow mandrel. The limiting structure one and the limiting structure two have a circumferential gap L. When torsion occurs, the annular rubber deforms under the action of external force, causing the outer sleeve and the hollow mandrel to rotate relative to each other in the circumferential direction, and the gap L decreases. The limiting structure one and the limiting structure two then rotate relative to each other in the circumferential direction. When the gap is eliminated, the limiting structure one and the limiting structure two come into contact to form a circumferential limiting structure.

[0023] Preferably, the limiting structure is an opening on the outer sleeve and a pin on the hollow mandrel. The pin extends radially into the opening on the outer sleeve, and the gap L is the space between one side of the pin and one inner side of the opening along the circumferential direction. After the product is vulcanized, the annular rubber fills the gap L.

[0024] During torsion, the pin rotates relative to the opening, the gap L decreases, the rubber in the gap L is compressed, and normal torsional deformation occurs in the early stage. When the gap L disappears, one side of the pin contacts one inner side of the opening to form a circumferential limiting structure.

[0025] Preferably, the center pin sleeve includes a center pin and a pin seat;

[0026] The design of the center pin sleeve with anisotropic lateral and longitudinal stiffness involves adding an inner sleeve inside the pin seat. The inner circumferential surface of the inner sleeve is elliptical. The center pin is inserted into the inner sleeve and connects with it. A free gap is formed between the inner sleeve and the center pin along the major axis of the inner circumferential surface of the elliptical inner sleeve, and a clearance fit is formed between the inner sleeve and the center pin along the minor axis of the inner circumferential surface of the elliptical inner sleeve. The major axis of the inner circumferential surface of the elliptical inner sleeve is consistent with the lateral direction of the rail vehicle, and the minor axis of the inner circumferential surface of the elliptical inner sleeve is consistent with the longitudinal direction of the rail vehicle.

[0027] When the rail vehicle is traveling in a straight line, the longitudinal stiffness is formed by the contact between the center pin and the inner sleeve to transmit traction or braking force. When the rail vehicle is turning a large curve or a gentle curve, the center pin will move in the free clearance and will not contact the inner sleeve in the transverse direction. When the rail vehicle is turning a small curve or a sharp curve, the center pin will first move in the free clearance, and after eliminating the free clearance, it will contact the inner sleeve to form transverse stiffness.

[0028] A rubber layer is provided in the inner sleeve. By adopting a non-uniform thickness design for the rubber layer, the lateral stiffness formed when the center pin contacts the inner sleeve is greater than the longitudinal stiffness formed when the center pin contacts the inner sleeve.

[0029] Preferably, the cross-section of the center pin is circular. When the center pin is in the center position, along the minor axis direction, the diameter of the center pin matches the minor axis of the inner circumferential surface of the elliptical inner sleeve, so that a clearance fit is formed between the inner sleeve and the center pin. Along the major axis direction, the diameter of the center pin is less than the major axis of the inner circumferential surface of the elliptical inner sleeve, so that a free gap is left between the inner circumferential surface of the inner sleeve and the center pin.

[0030] Preferably, the traction rod node includes a node mandrel, a node rubber body, and a node outer sleeve, all vulcanized as a single unit.

[0031] The optimization of the traction rod node to improve radial load-bearing and deflection capacity involves providing cavities at both ends of the node rubber body in the radial direction, and providing rubber stops protruding towards the node outer sleeve in the cavities, so that the cavities form two V-shaped shapes between the rubber stops and the node rubber body on both sides of the circumference. The node mandrel located at the rubber stops is provided with a metal stop protruding towards the node outer sleeve.

[0032] The beneficial effects of this invention are as follows: This invention can further optimize the performance of various suspension devices and suspension components in railway vehicle bogies, meeting the increasingly stringent performance requirements of users for railway vehicle bogies. By optimizing the series of conical springs, within a limited space, the conical springs can ensure that sufficient space is reserved at the annular end of their outer sleeve for machining end screw holes, and that each rubber body achieves a state of equal strain under pressure. This prevents defects such as frequent cracking and damage to the rubber body during use, avoiding premature failure of the rubber body, thereby improving the product's service life and performance. By optimizing the lateral stop, specifically by setting the wear plate to a frustum-shaped structure, it will not fall off the rubber body after vulcanization; furthermore, by setting stepped holes in the base plate, the rubber body forms anti-detachment buckles at the stepped holes, preventing the rubber body from falling off and ensuring the stability of the composite lateral stop. By wrapping at least one layer of fabric around the annular rubber of the supporting ball joint as a torsional stiffness reinforcing material, the fabric is used to tighten the annular rubber under load, reducing the degree of deformation. This increases the torsional stiffness of the supporting ball joint, reduces the deformation in the torsional direction, prevents excessive displacement of the torsion bar from colliding with other components, and reduces noise and abnormal sounds. Through the above improvements, the designed supporting ball joint can generate normal torsional deformation in the early stage. After torsion at a certain angle, it is torsionally limited, achieving nonlinear torsional stiffness, thus preventing excessive displacement of the torsion bar from colliding with other components and reducing noise and abnormal sounds. By adding an inner sleeve with an elliptical inner circumference between the center pin and the pin seat, and by setting a rubber layer in the inner sleeve and designing the rubber layer with non-uniform thickness, this invention can ensure the anisotropic lateral and longitudinal stiffness of the center pin sleeve, and also ensure that the pin seat can withstand greater lateral impact force when the rail vehicle is turning sharp bends. This prevents the pin seat from being damaged and failing during operation, thus ensuring the normal operation of the rail vehicle. This invention reduces the contact area of ​​the rubber by creating cavities at both radial ends of the rubber body, and then placing rubber stops protruding towards the outer sheath within the cavities, allowing the joint to deflect even after deformation under high stiffness. Furthermore, by placing metal stops at the rubber stops, the joint can still provide high radial stiffness. Attached Figure Description

[0033] Figure 1 This is a schematic diagram of the axial cross-sectional structure of a conical spring in the prior art;

[0034] Figure 2 for Figure 1 A partial structural diagram of the area between the outermost layer and the innermost layer;

[0035] Figure 3 This is a schematic diagram illustrating the principle of the design method for the conical spring in Embodiment 1 of the present invention;

[0036] Figure 4 This is a schematic diagram of the axial cross-sectional structure of the conical spring in Embodiment 1 of the present invention;

[0037] Figure 5 This is a schematic diagram of the axial cross-sectional structure of the spacer sleeve of the conical spring in Embodiment 1 of the present invention;

[0038] Figure 6 for Figure 4 Enlarged structural diagram of section A in the middle;

[0039] Figure 7 for Figure 4 Enlarged structural diagram of section B in the middle;

[0040] Figure 8 This is a schematic diagram of the overall structure of the transverse stop in Embodiment 1 of the present invention;

[0041] Figure 9 for Figure 8 A cross-sectional view;

[0042] Figure 10 for Figure 8 A diagram showing the view from below;

[0043] Figure 11 This is a bottom view of the base plate of the lateral stop in Embodiment 1 of the present invention;

[0044] Figure 12 This is a schematic diagram of the cover plate structure of the transverse stop in Embodiment 1 of the present invention;

[0045] Figure 13 This is a schematic diagram of the rubber structure of the lateral stop in Embodiment 1 of the present invention;

[0046] Figure 14 for Figure 13 A cross-sectional view;

[0047] Figure 15 This is a schematic diagram of the wear plate structure of the transverse stop in Embodiment 1 of the present invention;

[0048] Figure 16 This is a schematic diagram of the axial cross-sectional structure of the supporting ball joint in Embodiment 1 of the present invention;

[0049] Figure 17 This is a radial cross-sectional view of the supporting ball joint in Embodiment 1 of the present invention;

[0050] Figure 18 This is a three-dimensional structural diagram of the prefabricated rubber and fabric being stacked when making a supporting ball joint according to Embodiment 1 of the present invention.

[0051] Figure 19This is a schematic diagram of the radial cross-sectional structure after the pre-made rubber and fabric are rolled up when making the supporting ball joint according to Embodiment 1 of the present invention.

[0052] Figure 20 for Figure 16 Enlarged structural diagram of section C;

[0053] Figure 21 This is a schematic diagram of the axial cross-sectional structure of the supporting ball joint in Embodiment 2 of the present invention;

[0054] Figure 22 For along Figure 21 Schematic diagram of the cross-sectional structure along line AA;

[0055] Figure 23 for Figure 21 A schematic diagram of a partial axial cross-section of a structure located at an opening.

[0056] Figure 24 When the supporting ball joint is not twisted, it is located Figure 22 Enlarged structural diagram of section D in the middle;

[0057] Figure 25 When the supporting ball joint twists and forms a circumferential limiting structure, it is located at Figure 22 Enlarged structural diagram of section D in the middle;

[0058] Figure 26 This is a schematic diagram of the axial cross-sectional structure of the center pin sleeve according to Embodiment 1 of the present invention;

[0059] Figure 27 for Figure 26 A schematic diagram of the cross-sectional structure along line AA.

[0060] Figure 28 This is a top view of the inner sleeve of the center pin sleeve according to Embodiment 1 of the present invention;

[0061] Figure 29 This is a schematic axial cross-sectional view of the inner sleeve of the center pin sleeve according to Embodiment 1 of the present invention.

[0062] Figure 30 This is a schematic diagram of the overall structure of the traction rod node in Embodiment 1 of the present invention;

[0063] Figure 31 for Figure 30 Side view diagram;

[0064] Figure 32 for Figure 31 Schematic sectional view (without section lines drawn on the mandrel);

[0065] Figure 33 for Figure 32 Enlarged view of a portion;

[0066] Figure 34 for Figure 31 Another sectional view (without section lines drawn on the mandrel);

[0067] Figure 35 This is a schematic diagram of the rubber body structure of the traction rod node in Embodiment 1 of the present invention;

[0068] Figure 36 for Figure 35 Schematic diagram from the end side view;

[0069] Figure 37 for Figure 36 Enlarged view of a portion;

[0070] Figure 38 for Figure 36 Cross-sectional view;

[0071] Figure 39 This is a schematic diagram of the overall structure of the mandrel of the traction rod node in Embodiment 1 of the present invention;

[0072] Figure 40 for Figure 39 Front view diagram;

[0073] Figure 41 for Figure 40 Cross-sectional view;

[0074] Figure 42 for Figure 41 Cross-sectional view;

[0075] Figure 43 This is a cross-sectional view of the mandrel of the traction rod node in Embodiment 3 of the present invention (wherein the inner mandrel is not shown with cross-sectional lines);

[0076] Figure 44 This is a cross-sectional view of the mandrel of the traction rod node in Embodiment 3 of the present invention from another angle (wherein the inner mandrel is not shown with cross-sectional lines);

[0077] Figure 45 for Figure 44 Enlarged view of a portion;

[0078] Figure 46 This is a cross-sectional view of the traction rod node in Embodiment 4 of the present invention (the inner mandrel is not shown with cross-sectional lines);

[0079] Figure 47 This is a cross-sectional view of the traction rod node in Embodiment 4 of the present invention from another angle (where the inner mandrel is not shown with cross-sectional lines);

[0080] Figure 48 and Figure 49These are schematic diagrams showing cross-sectional views of the existing traction rod nodes from two different directions.

[0081] In the diagram: 1. Outer sleeve, 111. Annular end, 112. Stepped section, 2. Mandrel, 3. Spacer, 4. Rubber body, 411. Arc-shaped surface, 412. Sloping surface one, 413. Sloping surface two, 414. Rubber body body, 5. End screw hole, 6. Lower conical spacer, 7. Annular spacer, 8. Cover, 811. Gas through hole, 9. Damping adjustment space, 10. Sealing ring, 11. Base plate, 12. Wear plate, 121. Positioning hole, 122. Circular stop, 123. Wear plate body, 13. Cover plate, 14. Mounting hole, 15. Through hole, 16. Annular boss, 17. Annular step, 18. Stepped hole, 19. 20. Anti-detachment buckle, 21. Retaining ring, 22. Stop hole, 23. Ring rubber, 24. Torsion bar, 25. Fabric, 26. Rubber belt, 27. Opening, 28. Pin body, 29. Space 1, 30. Pin seat, 30. Inner sleeve, 301. Rubber layer, 302. Outer sleeve, 303. Inner sleeve body, 304. Bottom cover, 31. Center pin, 32. Bolt hole, 33. Drain hole, 34. Node outer sleeve, 35. Node rubber body, 351. Node rubber body, 36. Node mandrel, 361. Inner node mandrel, 362. Outer node mandrel, 37. Cavity, 38. Rubber stop, 39. Protective layer, 40. Metal stop. Detailed Implementation

[0082] The technical solution of the present invention will be further described in detail below with reference to the accompanying drawings and specific embodiments.

[0083] This application mainly optimizes the primary conical spring in the suspension system, the lateral stop and the support ball joint in the anti-roll torsion bar device, as well as the center pin sleeve and traction rod node in the traction device. These are described in detail below:

[0084] Example 1:

[0085] I. Optimization of a series of conical springs:

[0086] A conical spring is a rubber elastic positioning and support element suitable for various subway and light rail vehicles both domestically and internationally. Primarily used in vehicle suspension systems, it provides elastic positioning, shock absorption, and cushioning, while also eliminating high-frequency vibrations and improving vibration performance, thus ensuring vehicle safety and comfort. Conical springs are often installed as primary suspension components between the frame and bogie, mainly serving as elastic supports and transmitting loads. They are a core component of rail vehicle suspension systems, and their performance directly affects vehicle stability.

[0087] like Figure 1As shown, an existing conical spring includes an outer sleeve 1, a spindle 2, and multiple spacers 3 located between the outer sleeve 1 and the spindle 2. The outer sleeve 1, spindle 2, and multiple spacers 3 are made of metal and are bonded together by multiple rubber bodies 4 through vulcanization. For ease of subsequent description, the large end of the conical spring is defined as the upper end, and the small end of the conical spring is defined as the lower end. The spacers 3 are set as conical spacers. In the axial section of the conical spring, the angle between the inclined line of the conical spacer 3 located on one side of the central axis L of the conical spring and the central axis L of the conical spring is set as θ1. Then, θ1 is equal for each spacer 3. The inner circumferential surface of the outer sleeve 1 is also set as conical. The angle between the inclined line of the inner circumferential surface of the outer sleeve 1 located on one side of the central axis L of the conical spring and the central axis L of the conical spring is set as θ2. Then, θ1 = θ2. At the large end of the conical spring, an end screw hole 5 is also provided on one end of the outer sleeve 1, which is an annular end 111.

[0088] One of the problems with the existing technology is that after a period of operation, the rubber body always develops defects such as cracks and damage, causing rubber failure, which in turn affects the service life and performance of the product.

[0089] To this end, the applicant conducted research and found that, Figure 2As shown, during the design process, it is necessary to consider the equal strain design of each rubber layer. The equal strain design of rubber layers refers to using the equal strain principle to consider the strain and stress distribution of the rubber material during the design of rubber products, in order to ensure the performance and lifespan of the product. Therefore, under the premise of equal strain design of the rubber layers, theoretically, the tilt angles of the spacer 3 and the outer sleeve 1 should be designed according to the concept of θ1 = θ2, so that each layer of rubber 4 is in an equal strain state. In this case, theoretically, the tilt line M at the position of the inner circumference surface of the outer sleeve 1 should be located outside the tilt line N at the actual position of the inner circumference surface of the outer sleeve 1 (i.e., away from the central axis L of the conical spring). Let the distance between the tilt line M and the outermost spacer 3 be H1. However, since the outer diameter Φ of the outer sleeve 1 is fixed and cannot be changed, the design space is limited. At the same time, an end screw hole 5 needs to be opened on the annular end 111 of the outer sleeve 1. At this time, since the reserved size of the radial width D of the annular end 111 is insufficient, the inclined line M at the theoretical position of the inner circumferential surface of the outer sleeve 1 will interfere with the end screw hole 5. Therefore, it is impossible to process the end screw hole 5 on the annular end 111. In order to leave enough radial width D to process the end screw hole 5, the inclined line M at the theoretical position of the inner circumferential surface of the outer sleeve 1 must be shifted inward (i.e., closer to the central axis L of the conical spring) to the position of the inclined line N. Let the distance between the inclined line N and the outermost spacer 3 be H2. It can be seen that H2 < H1. In fact, H1 and H2 represent the thickness of the rubber body 4 between the outer jacket 1 and the outermost spacer 3. When H2 < H1, it means that the actual thickness of the rubber body 4 between the outer jacket 1 and the outermost spacer 3 is less than the theoretical thickness of the rubber body 4 between the outer jacket 1 and the outermost spacer 3. This will cause the rubber bodies to not be in a state of equal strain when the product is under pressure, and the rubber bodies will always have defects such as cracks and damage, resulting in rubber failure, which in turn affects the service life and performance of the product.

[0090] Therefore, such as Figure 3 and Figure 4As shown, in the design, the applicant sets the inner circumferential surface of the outer sleeve 1 into a conical shape. First, the inclined line of the inner circumferential surface of the outer sleeve 1, located on one side of the central axis L of the conical spring, is set parallel to the inclined line of the conical spacer 3, also located on one side of the central axis L of the conical spring. With the outer diameter Φ of the outer sleeve 1 fixed, the inclined line of the inner circumferential surface of the outer sleeve 1 is moved parallel to the central axis L of the conical spring to the inclined line P. This makes the radial width D of the annular end 111 of the outer sleeve 1 greater than the diameter of the end screw hole 5, thus reserving sufficient space for machining the end screw hole 5. Because it is a parallel movement... Therefore, the angle between the inclined line P and the central axis L of the conical spring is equal to θ1. The inclined line P intersects the horizontal line extending horizontally along the annular end 111 at the upper end of the outer sleeve 1 at point S. Then, with point S as the center, the inclined line P is rotated by a certain angle away from the central axis L of the conical spring to obtain the inclined line Q, so that the inclined line at the position of the inner circumferential surface of the outer sleeve 1 is located on the inclined line Q. At this time, the angle θ2 between the inclined line at the position of the inner circumferential surface of the outer sleeve 1 on the side of the central axis L of the conical spring and the central axis L of the conical spring is greater than the angle θ1 between the inclined line at the position of the conical spacer 3 on the side of the central axis L of the conical spring and the central axis L of the conical spring. In this embodiment, since the inclined line P is rotated by a certain angle away from the central axis L of the conical spring, i.e., the angle θ2 > the angle θ1, from Figure 3 As can be seen, the gap H3 between the inclined line Q and the outermost spacer 3 is actually greater than the gap H4 between the inclined line P and the outermost spacer 3. This design, while ensuring sufficient space is reserved at the annular end 111 of the outer jacket 1 for machining the end screw hole 5, increases the thickness of the rubber body 4 between the outer jacket 1 and the outermost spacer 3. This ensures that the rubber bodies achieve a state of equal strain when the product is under pressure. In summary, through the above design method, within a limited space, the conical spring can ensure that sufficient space is reserved at the annular end of its outer jacket for machining the end screw hole, and also ensure that the rubber bodies achieve a state of equal strain when the product is under pressure. This prevents the conical spring from frequently developing defects such as cracks and damage to the rubber body during use, avoiding premature failure of the rubber body, thereby improving the service life and performance of the product.

[0091] Conical springs, typically used as primary suspension components, allow for significant vertical displacement of the axle box relative to the frame during vehicle operation. Lateral and longitudinal relative displacements between the wheelset and the frame are achieved through the deformation of the primary conical springs. Therefore, specific requirements are placed on the elastic stiffness of the primary conical springs in the vertical, lateral, and longitudinal directions. Furthermore, under different vertical loads, the vertical stiffness of the primary conical springs is required to increase non-linearly with increasing load. Figure 5As shown, the applicant also provides a lower conical spacer 6 at the lower end of the conical spacer 3. The conical spacer 3 and the lower conical spacer 6 can be integrated into one structure. The angle between the inclined line of the inner circumferential surface of the lower conical spacer 6 located on one side of the central axis L of the conical spring and the central axis L of the conical spring is set to θ3, then θ3 > θ1. In effect, an outwardly extending section is provided at the lower end of the conical spacer 3. When the product is subjected to vertical pressure, the rubber body at the lower end of the conical spacer deforms and bulges out to contact the lower conical spacer 6, realizing the characteristic that the vertical stiffness of the primary conical spring increases non-linearly with the increase of load. Through experiments, θ1 ≤ 7° can be set, which improves the lateral stiffness of the primary conical spring and achieves the requirement of a lateral to vertical stiffness ratio of 7.6 or higher. Setting θ3≥28° achieves the requirement of more than doubling the vertical nonlinear stiffness increment of the primary conical spring, reduces the vertical creep of the primary conical spring, and improves the fatigue life of the primary conical spring.

[0092] An annular spacer 7 is also provided at the upper end of the conical spacer 3. The conical spacer 3, the lower conical spacer 6, and the annular spacer 7 can also be configured as an integral structure. By providing the annular spacer 7, the upper end of the conical spacer 3 extends vertically upward, which facilitates the demolding of the conical spring after vulcanization.

[0093] like Figure 6 The lower end face of the rubber body 4 is set as an inverted face, which includes an arc-shaped face 411, a first inclined face 412 set at the upper end of the arc-shaped face 411, and a second inclined face 413 set at the lower end of the arc-shaped face 411. This reduces the amount of rubber and wrinkles in the non-working area, increases the strain space of the rubber during operation, and ensures the design of uniform strain in a single layer.

[0094] like Figure 4 As shown, a cover 8 is also provided at the upper end of the outer sleeve 1 of the conical spring. The cover 8 and the upper end of the conical spring are used to form a damping adjustment space 9. When a larger stiffness of the conical spring is required, gas, such as compressed air or inert gas, can be filled into the damping adjustment space 9. At this time, the stiffness of the product is actually equal to the stiffness of the conical spring plus the stiffness of the gas, which can meet the larger stiffness requirement. When a larger stiffness is not required, the gas in the damping adjustment space 9 can be discharged.

[0095] A gas passage 811 is provided on the cover 8, which communicates with the damping adjustment space 9. When gas is introduced into the damping adjustment space 9, it is injected into the damping adjustment space 9 through the gas passage 811. After injection, a sealing plug (not shown in the figure) is inserted into the gas passage 811 for sealing. When it is necessary to release the gas, simply remove the sealing plug, and when the conical spring is compressed, the gas in the damping adjustment space 9 will automatically be released through the gas passage 811.

[0096] like Figure 7 As shown, a full-circle stepped portion 112 is provided on the annular end 111 at the upper end of the outer jacket 1. A sealing ring 10 is also provided in the stepped portion 112. The cover 8 is press-fitted into the stepped portion 112, which can improve the sealing performance of the damping adjustment space 9.

[0097] As can be seen from the above technical solutions, by optimizing the series of conical springs, within a limited space, the conical springs can ensure that sufficient space is reserved at the annular end of the outer sleeve for machining the end screw holes, and that each rubber body achieves a state of uniform strain under pressure. This prevents defects such as cracks and damage to the rubber body during use, avoiding premature failure and thus improving the product's service life and performance. By setting a lower conical spacer at the lower end of the conical spacer, the vertical stiffness of the series of conical springs increases non-linearly with increasing load. By setting the lower end face of the rubber body as an inverted surface, the amount of rubber in the non-working area and wrinkles are reduced, increasing the strain space of the rubber during operation and ensuring a single-layer uniform strain design. By setting a damping adjustment space, the stiffness of the product can be adjusted according to the actual working requirements.

[0098] II. Optimization of lateral stops:

[0099] Example 1:

[0100] Lateral stops are commonly used rubber-metal composite vibration damping and limiting components. Structurally, they typically consist of a vulcanized base plate, a rubber body, and a wear plate. They are widely used in various vibration damping and limiting applications, especially in the bogies of rail vehicles, primarily bearing lateral compressive loads and vertical and longitudinal shear loads. During use, to withstand lateral loads, lateral stops must meet large variable stiffness requirements. Furthermore, during the operation of rail vehicles, situations often arise where lateral compression and vertical and longitudinal shear occur simultaneously, leading to the wear plate detaching from the rubber body, or even the rubber body detaching from the base plate. Therefore, it is necessary to improve the stability and reliability of existing composite lateral stops.

[0101] The applicant provides a variable stiffness composite lateral stop with protective functions, such as Figures 8 to 10 As shown, the assembly includes a base plate 11, a rubber body 4, a wear plate 12, and a cover plate 13. The base plate 11, rubber body 4, and wear plate 12 are vulcanized and bonded together, and the cover plate 13 is assembled at the center of the base plate 11. To prevent the wear plate 12 from separating from the rubber body 4 during use, the wear plate 12 is designed with a frustum shape, so that the wear plate 12 is surrounded within the conical space of the rubber body 4. Even if the bond between the wear plate 12 and the rubber body 4 fails, the wear plate 12 will not fall out.

[0102] like Figure 11 As shown, the base plate 11 has mounting holes 14 at its edge to mount the entire transverse stop onto the vehicle, and a through hole 15 at its center. An annular boss 16 is located at the bottom edge of the through hole 15. Correspondingly, as... Figure 12 As shown, an annular step 17 is provided on the outer periphery of the cover plate 13. After the cover plate 13 is assembled into the through hole 15 of the bottom plate, the annular step 17 cooperates with the annular boss 16 to limit the position of the cover plate 13.

[0103] Still Figure 11 As shown, a stepped hole 18 is provided at the outer edge of the annular boss 16 of the base plate 11. In this embodiment, four stepped holes 18 are evenly arranged. The end of the stepped hole 18 closer to the rubber body 4 is a large hole, and the end closer to the rubber body 4 is a small hole. The diameter ratio between the small hole and the large hole is 1:2. During the vulcanization process, as... Figure 13 As shown, the rubber of the rubber body 4 enters the stepped hole 18 to form an anti-detachment buckle 19, and as... Figure 14 As shown, the ratio of the rubber diameter d at the small hole of the anti-detachment buckle 19 to the outer diameter D at the bottom of the rubber body 414 is d:D = 0.1 to 0.2. The ratio of the distance S between the edge of the anti-detachment buckle 19 and the rubber body 414 in contact with the outer side of the bottom of the rubber body 414 to the outer diameter D at the bottom of the rubber body 414 is s:D = 1:10, so as to ensure that the anti-detachment buckle 19 can achieve the effect of preventing the rubber body 4 from falling off.

[0104] like Figure 9 and Figure 15 As shown, a positioning hole 121 is provided at the center of the top of the wear plate 12 to position the wear plate 12 during vulcanization. A circular stop 122 extending downwards is provided at the center of the bottom plate of the wear plate 12, and the diameter of the bottom end of the circular stop 122 is smaller than the diameter at the end connected to the wear plate body 123. The included angle b between the two sides of the stop 122 on the same radial direction is 10° to 15° to facilitate demolding after vulcanization. Accordingly, as Figure 14As shown, a stepped retaining ring 20 is provided at the center of the rubber body 4. The retaining ring 20 has a flared structure with a smaller upper end and a larger lower end. The included angle α between the two sides on the same radial direction on its inner side is 6° to 10°, which also facilitates the mold release after vulcanization. The ratio of the height A of the retaining ring 20 to the height B of the circular stop 122 is A:B = 1.3 to 1.5, so that the variable stiffness of the retaining ring 20 can be activated before the lateral stop reaches its hard stop function. The ratio of the inner diameter E at the lower opening of the retaining ring 20 to the outer diameter F at the bottom of the circular stop 122 is E:F = 1.3 to 1.5, so as to prevent the retaining ring 20 from being sheared by the circular stop 122 and the cover plate 13, causing it to crack before its service life. At the same time, the above structure separates the bottom of the circular stop 122 from the bottom of the retaining ring 20, preventing the rubber body 4 from blocking the path of the hard stop of the wear plate 12 during deformation, and also preventing the rubber from tearing during deformation. Moreover, as Figure 12 As shown, a stop hole 21 is provided at the center of the cover plate 13. When the lateral stop undergoes large deformation, the circular stop 122 falls into the stop hole 21, realizing a three-way limiting function. The distance H between the bottom of the circular stop 122 and the lower end face of the base plate 11 is the hard stop clearance of the lateral stop, which can be selected according to the actual operating conditions of the vehicle.

[0105] like Figure 14 As shown, the rubber at the top of the rubber body 4 has a gentle slope, while the rubber on the sides has a steeper slope. This allows for sufficient deformation space during small rubber displacements, achieving a relatively low axial stiffness to meet vehicle operation requirements. When the axial deformation reaches a certain level, the retaining ring 20 contacts the cover plate 13, and the wear plate 12 and cover plate 13 compress this portion of the rubber. Simultaneously, because the circular stop 122 obstructs the rubber's deformation space, the stiffness of the lateral stop increases rapidly, achieving high stiffness under large deformation. As the axial displacement further increases, the circular stop 122 extends into the stop hole 21 and contacts the steering frame beam, achieving a hard stop function. Moreover, the ratio of the inner diameter G of the stop hole 21 to the outer diameter F of the bottom of the circular stop 122 is G:F = 1.2. Chamfers are designed at the top of the stop hole 21 and the bottom of the circular stop 122, which, combined with the tapered design of the circular stop 122, further eliminates the influence of radial misalignment.

[0106] Meanwhile, by setting a cover plate 13 for subsequent assembly at the center of the base plate 11, the retaining ring 20 can be easily demolded after being formed during the vulcanization process, and the cover plate 13 can be glued to the base plate 11 with metal glue. When the lateral stop is installed, the bogie and the base plate 11 limit the position of the cover plate 13.

[0107] As can be seen from the above technical solutions, by optimizing the lateral stop—specifically by setting the wear plate to a frustum-shaped structure, which, after being vulcanized integrally with the rubber body, prevents it from falling off—and by setting stepped holes in the base plate, the rubber body forms anti-detachment buckles at the stepped holes, preventing the rubber body from falling off and ensuring the stability of the composite lateral stop, the following measures are taken. By setting a circular stop at the bottom of the wear plate and a stop hole in the center of the cover plate, when the vertical load is large and the vertical displacement increases, the circular stop falls into the stop hole, thus achieving the vehicle's vertical, lateral, and longitudinal three-way limiting function. The bottom cover plate is designed as an assembly structure, allowing for free replacement. By replacing the cover plate with different sizes and specifications, the variable stiffness performance and the three-way limiting range can be freely adjusted. Increasing the thickness of the base plate allows the rubber to contact the base plate earlier, thus achieving variable stiffness earlier. Furthermore, by increasing or decreasing the diameter of the stop hole in the cover plate, the vertical and longitudinal limiting ranges of the lateral stop can be adjusted. This greatly expands the applicability of the lateral stop.

[0108] III. Optimization of the support ball joint in the anti-roll torsion bar device:

[0109] An anti-roll torsion bar is a vehicle safety component installed between the car body and the bogie to provide anti-roll performance. It offers excellent anti-roll capabilities, ensuring the vehicle does not exceed permissible clearances under dynamic conditions and improving ride comfort. The anti-roll torsion bar works on the principle of leverage, converting the vertical movement of the suspension into torque. This torque is transmitted to the vehicle chassis through the torsion bar, reducing the risk of rollover. When the vehicle traverses a curve, the body tilts outward due to centripetal force. The anti-roll torsion bar generates a torque opposite to this tilt, restoring the body to balance and ensuring safe driving. The support ball joint is a crucial component of the anti-roll torsion bar. The torsion bar passes through and engages with the support ball joint, supporting the torsion bar on the bogie or car body. A problem with existing technologies is that some support ball joints have insufficient torsional stiffness, resulting in significant deformation in the torsional direction. This leads to excessive displacement of the connected torsion bar, causing it to collide with other components during operation, resulting in noise and abnormal sounds.

[0110] Therefore, the applicant adopted two optimization methods:

[0111] (I) First, the optimization method for the first type of supporting ball joint will be explained:

[0112] Example 1:

[0113] like Figure 16As shown, the supporting ball joint includes an outer sleeve 1, a hollow spindle 2 disposed inside the outer sleeve 1, and an annular rubber 22 located between the outer sleeve 1 and the hollow spindle 2. The outer sleeve 1 and the hollow spindle 2 are bonded together by vulcanization of the annular rubber 22. A torsion bar 23 passes through the hollow spindle 2 and is connected to it. When torsion occurs, it is actually a relative circumferential rotation between the outer sleeve 1 and the hollow spindle 2. The applicant has found that the torsional stiffness of the supporting ball joint is low, mainly because the annular rubber 22 is disposed between the outer sleeve 1 and the hollow spindle 2. The annular rubber 22 is a highly elastic polymer compound at room temperature with large molecular chain flexibility. Under external force, it can undergo large deformation, thus causing excessive relative circumferential rotation between the outer sleeve 1 and the hollow spindle 2. Therefore, the applicant has wound at least one layer of fabric 24 around the annular rubber 22 and along its circumference as a torsional stiffness reinforcing material to increase the torsional stiffness of the supporting ball joint. In this embodiment, the fabric is a curtain fabric. During operation, when the outer sleeve 1 and the hollow spindle 2 rotate relative to each other in the circumferential direction, the annular rubber 22 is deformed under load. At this time, the annular rubber 22 is pulled by the full circle of cloth 24, which reduces the degree of deformation, thereby increasing the torsional stiffness of the supporting ball joint, reducing the degree of deformation in the torsional direction, avoiding excessive displacement of the torsion bar and collision with other parts, and reducing noise and abnormal sounds.

[0114] After at least one layer of fabric 24 is wrapped, it is best to form a structure in which the rubber and fabric are distributed in layers inside the annular rubber 22, which can further increase the torsional stiffness of the supporting ball joint.

[0115] To form a layered structure of rubber and fabric, the vulcanization steps for supporting the ball joint are as follows: First, as... Figure 18 and Figure 19 As shown, a strip of fabric 24 and a strip of rubber 25 are prefabricated. The strip of fabric 24 and the rubber 25 are then stacked together and rolled into a ring. The ring of fabric 24 and rubber 25 is then placed between the outer jacket 1 and the hollow mandrel 2 for product vulcanization. During the vulcanization process, no additional rubber is needed. Instead, the ring of rubber 25 is used to vulcanize and bond the outer jacket 1 and the hollow mandrel 2 together.

[0116] Before stacking, the fabric 24 and rubber strip 25 need to be thinned to control their thickness. This helps ensure good contact between the fabric 24 and rubber strip 25 and prevents misalignment of the fabric 24 during vulcanization from affecting product performance. Through testing, the preferred ranges for the thickness K of the thinned fabric 24 and the thickness H of the thinned rubber strip 25 are: 0.8mm ≤ K ≤ 1mm for the thickness K of the thinned fabric 24, and 1mm ≤ H ≤ 1.5mm for the thickness H of the thinned rubber strip 25.

[0117] The number of fabric and rubber layers formed after rolling can be adjusted according to actual needs. When the product requires greater torsional stiffness, more turns can be made to increase the number of fabric and rubber layers formed after rolling. Conversely, fewer turns can be made to reduce the number of fabric and rubber layers formed after rolling.

[0118] like Figure 17 As shown, the addition of fabric 24 to the annular rubber 22 results in a larger outer diameter of the annular rubber 22 compared to when fabric 24 is not added. Therefore, the outer jacket 1 adopts a multi-lobed design; in this embodiment, it is a three-lobed design. After the product is vulcanized, no pre-compression is applied, and an opening gap A is left between adjacent lobe outer jackets. When the product is installed, the supporting ball joint is pressed into the mounting hole. During installation, the lobe outer jackets contract under force, and the opening gap A disappears, achieving product pre-compression and significantly improving the product's various stiffness properties. In this embodiment, the opening gap A is 1.2 mm.

[0119] like Figure 18 and Figure 20 As shown, the dimensions of the prefabricated rubber strip 25 are greater than the dimensions of the strip fabric 24. This allows the rubber strip 25 to completely enclose the fabric 24 after vulcanization. In particular, the width of the prefabricated rubber strip 25 is greater than the width of the strip fabric 24, ensuring that both ends of the fabric 24 are concealed within the rubber strip 25 and not exposed. A gap C is maintained between the end faces of the rubber strip 25 and the end faces of the fabric 24. This prevents cracks from easily forming at the exposed ends of the fabric during operation, which could lead to failure.

[0120] like Figure 18 As shown, the applicant also studied the order in which the strip of fabric 24 and the rubber strip 25 are stacked together. The stacking order is divided into two categories: placing the fabric 24 on top of the rubber strip 25 or placing the fabric 24 on the bottom of the rubber strip 25. Experiments revealed that when the fabric 24 is placed on top of the rubber strip 25 and rolled into a ring for vulcanization, the torsional stiffness of the product is greater than when the fabric 24 is placed on the bottom of the rubber strip 25 and rolled into a ring for vulcanization. Therefore, it is preferable to stack the fabric 24 on top of the rubber strip 25 during manufacturing.

[0121] Through the above improvements, at least one layer of fabric is wrapped around the annular rubber of the supporting ball joint as a torsional stiffness reinforcing material. When the annular rubber deforms under load, the entire loop of fabric pulls on the annular rubber, reducing the degree of deformation and thus increasing the torsional stiffness of the supporting ball joint. This reduces the degree of deformation in the torsional direction, prevents excessive displacement of the torsion bar from colliding with other components, and reduces noise and abnormal sounds. By designing specific vulcanization steps, a layered structure of rubber and fabric is formed inside the annular rubber, further increasing the torsional stiffness of the supporting ball joint. By designing the dimensions of the fabric and rubber when stacked, the ends of the fabric are concealed within the rubber, preventing cracks that could easily occur during operation and lead to failure. (II) Next, the second optimization method for the supporting ball joint is described:

[0122] Example 2:

[0123] like Figure 21 and Figure 22 As shown, the supporting ball joint includes an outer sleeve 1, a hollow spindle 2 disposed inside the outer sleeve 1, and an annular rubber 22 located between the outer sleeve 1 and the hollow spindle 2. The outer sleeve 1 and the hollow spindle 2 are bonded together by vulcanization of the annular rubber 22. A torsion bar 23 passes through the hollow spindle 2 and is connected to the hollow spindle 2. A first limiting structure is provided on the outer sleeve 1, and a second limiting structure is provided on the hollow spindle 2. The first limiting structure and the second limiting structure have a circumferential gap L. When torsion occurs, the annular rubber 22 deforms under the action of external force in the early stage, causing the outer sleeve 1 and the hollow spindle 2 to rotate relative to each other in the circumferential direction, and the gap L becomes smaller. The first limiting structure and the second limiting structure rotate relative to each other in the circumferential direction accordingly. When the rotation reaches a certain angle, the gap between the first limiting structure and the second limiting structure is eliminated, and the first limiting structure and the second limiting structure come into contact to form a circumferential limiting structure, which prevents further relative circumferential rotation between the outer sleeve 1 and the hollow spindle 2, thereby realizing a nonlinear change in torsional stiffness. The circumferential gap between the first and second limiting structures is designed to allow for normal torsional deformation in the early stages of the supporting ball joint under torsional force. When the gap is eliminated, the first and second limiting structures come into contact to form a circumferential limiting structure, thereby achieving a non-linear change in torsional stiffness. Through this design method, the designed supporting ball joint can generate normal torsional deformation in the early stages. After torsion at a certain angle, torsional limiting is performed, achieving non-linear torsional stiffness. This avoids excessive displacement of the torsion bar and collision with other components, reducing noise and abnormal sounds.

[0124] like Figures 23 to 25As shown, in this embodiment, the limiting structure consists of two parts: an opening 26 on the outer sleeve 1 and a pin 27 on the hollow mandrel 2. The pin 27 extends radially into the opening 26 on the outer sleeve 1, with a circumferential gap L between one side of the pin 27 and one inner side of the opening 26. After vulcanization, annular rubber 22 fills the gap L. During torsion, the pin 27 rotates relative to the opening 26, reducing the gap L and compressing the annular rubber 22, resulting in normal torsional deformation in the early stage. When the gap L disappears, one side of the pin 27 contacts one inner side of the opening 26, forming a circumferential limiting structure that prevents further relative circumferential rotation between the outer sleeve 1 and the hollow mandrel 2, thus achieving a nonlinear change in torsional stiffness. In this embodiment, the timing of the nonlinear inflection point of the torsional stiffness curve is determined by controlling the size of the gap L. When it is necessary for the nonlinear inflection point of the torsional stiffness curve to occur earlier, the gap L can be reduced; when it is necessary for the nonlinear inflection point of the torsional stiffness curve to occur later, the gap L can be increased.

[0125] like Figure 24 As shown, during manufacturing, the pin 27 can be first installed onto the hollow mandrel 2 by interference fit or welding, and then the outer sleeve 1 and the hollow mandrel 2 can be vulcanized. Alternatively, the outer sleeve 1 and the hollow mandrel 2 can be vulcanized, and then the pin 27 can be interference fitted onto the hollow mandrel 2. This eliminates the need for additional mold making; processing can be performed only after vulcanization, thus saving manufacturing costs.

[0126] The pin 27 is vulcanized in the annular rubber 22. The top of the pin 27 is lower than the top of the opening 26. This ensures that the outer diameter of the support ball joint does not change, thus not changing the original installation method of the support ball joint.

[0127] In this embodiment, a space 28 is left between the top surface of the annular rubber 22 in the opening 26 and the top of the opening 26. Rubber is not vulcanized within space 28, which meets the design requirements when the product has low radial stiffness. Alternatively, rubber can be vulcanized within space 7, which meets the design requirements when the product has high radial stiffness.

[0128] like Figure 21 and Figure 22As shown, two sets of pins 27 are provided, symmetrically distributed on the hollow mandrel 2 about the central axis of the hollow mandrel 2. Each set of pins 27 corresponds to one opening 26, therefore, there are two openings 26. The two openings 26 are symmetrically distributed on the outer sleeve 1 about the central axis of the outer sleeve 1. This arrangement can further ensure the formation of the circumferential limiting structure and realize the nonlinear change of torsional stiffness. Each set of pins 27 can contain one or more pins 27. In this embodiment, each set of pins 27 contains two pins 27.

[0129] Through the above improvements, the designed support ball joint can generate normal torsional deformation in the early stage. After tortuosity reaches a certain angle, it is torsionally limited, achieving nonlinear torsional stiffness. This avoids excessive displacement of the torsion bar and collision with other components, reducing noise and abnormal sounds. The nonlinear torsional stiffness is achieved by designing a specific torsional limiting structure. By controlling the size of the gap L, the timing of the nonlinear inflection point of the torsional stiffness curve can be controlled. By adjusting whether the space contains vulcanized rubber, the product can meet the radial stiffness requirements of different products.

[0130] IV. Optimization of the center pin sleeve in the traction device:

[0131] The car body and bogie frame of a rail vehicle are generally connected by a central traction device, which is a central pin sleeve. It includes a central pin and a pin seat. The central pin is installed on the bottom of the car body, and the pin seat is installed on the bogie frame. When the car is unloaded, the central pin is inserted into the pin seat to connect, so that the car body and the bogie frame are connected.

[0132] In rail vehicles, the direction of the vehicle's forward movement is generally referred to as the longitudinal direction, and the width direction of the vehicle body is referred to as the lateral direction. When the rail vehicle is traveling normally in a straight line, the longitudinal stiffness is mainly transmitted by the contact between the center pin and the pin sleeve. When the rail vehicle turns a curve, the center pin sleeve will bear a greater lateral force. Therefore, a free travel section needs to be designed in the lateral direction. That is, when the vehicle is traveling normally, the floating traction pin contacts the traction pin seat in the longitudinal direction to form longitudinal stiffness and transmit traction or braking force. When the vehicle turns a large curve (passing through a large curve radius), the floating traction pin will move in the lateral free play and will not contact the traction pin seat. At this time, the lateral stiffness is zero. When the vehicle turns a small curve (passing through a small curve radius), the floating traction pin will first move in the lateral free play, and after eliminating the lateral free play, it will contact the traction pin seat to form lateral stiffness.

[0133] The problem with the existing technology is:

[0134] 1. The towing pin seat is a cylindrical cast steel part, thinner in the transverse direction and thicker in the longitudinal direction. Therefore, when the floating towing pin contacts the towing pin seat, the resulting transverse stiffness is small, while the longitudinal stiffness is large; their transverse and longitudinal stiffness are anisotropic. Consequently, when the vehicle body makes sharp turns, the towing pin seat will bear a greater lateral impact force. Because the towing pin seat is thinner in the transverse direction, after a period of operation, it will be damaged and fail under the lateral impact force, thus compromising the normal operation of the vehicle body.

[0135] Second, the outermost layer of the floating traction pin is fitted with a rubber bushing. During operation, the rubber bushing of the floating traction pin frequently comes into contact with the cast steel traction pin seat, which leads to greater wear on the rubber bushing and reduces the service life of the floating traction pin. In addition, the contact between the rubber bushing and the traction pin seat generates a lot of noise, which has an impact on the environment.

[0136] To solve the first technical problem mentioned above, the applicant's improvement plan is as follows:

[0137] Example 1:

[0138] like Figures 26 to 28 As shown, a design method for the anisotropic lateral and longitudinal stiffness of a center pin sleeve is described. An inner sleeve 30 is press-fitted into the inside of the pin seat 29. The outer circumferential surface of the inner sleeve 30 is circular, matching the inner circumferential surface of the pin seat 29 for easy press-fitting. The inner circumferential surface of the inner sleeve 30 is elliptical. When the center pin sleeve is installed between the car body and bogie of a rail vehicle, the minor axis X of the elliptical inner circumferential surface is aligned with the longitudinal direction of the car body, and the major axis Y of the elliptical inner circumferential surface is aligned with the transverse direction of the car body. A center pin 31 is inserted into the inner sleeve 30 and engages with it. The cross-section of the center pin 31 is circular. When the center pin 31 is in the center position, that is, when the center point of the center pin 31 coincides with the center point of the elliptical inner sleeve 30, the diameter Φ of the center pin 31 along the minor axis direction X matches the minor axis of the inner circumference of the elliptical inner sleeve, thus forming a clearance fit between the center pin 31 and the inner sleeve 30. Along the major axis direction Y, the diameter Φ of the center pin 31 is less than the major axis of the inner circumference of the elliptical inner sleeve, thereby leaving a free gap L between the center pin 31 and the inner circumference of the inner sleeve 30. The free gap L is the lateral free gap. In this embodiment, the free gap L is set to 15mm.

[0139] A rubber layer 301 is provided inside the inner sleeve 30. The rubber layer 301 adopts a non-uniform thickness design, that is, the thickness D1 of the rubber layer 301 along the minor axis X is greater than the thickness D2 of the rubber layer 301 along the major axis Y. This results in the longitudinal stiffness being less than the transverse stiffness when the center pin 31 contacts the inner sleeve 30. After testing, the longitudinal stiffness is 120kN / mm and the transverse stiffness is (≥300kN / mm).

[0140] When the rail vehicle is traveling in a straight line, longitudinally, the center pin 31 contacts the inner sleeve 30 to form longitudinal stiffness and transmit traction or braking force. When the rail vehicle makes a large curve or gentle curve (passing through a large curve radius), laterally, the center pin 31 moves within the free clearance L and does not contact the inner sleeve 30; at this time, the lateral stiffness is zero. When the vehicle body makes a small curve or sharp curve, the center pin 31 first moves within the free clearance L, and after eliminating the free clearance L, it contacts the inner sleeve 30 to form lateral stiffness. At this time, since the lateral stiffness is much greater than the longitudinal stiffness, when the vehicle body makes a small curve or sharp curve (passing through a small curve radius), the pin seat and the inner sleeve 30 can withstand a greater lateral impact force together. This embodiment ensures both the anisotropy of the lateral and longitudinal stiffness of the center pin sleeve and the ability of the pin seat to withstand a greater lateral impact force when the rail vehicle makes small curves or sharp curves, preventing the pin seat from being damaged and failing during operation, thus ensuring the normal operation of the rail vehicle.

[0141] To solve the second technical problem mentioned above, such as Figure 27 and Figure 28 As shown, the inner sleeve 30 adopts a three-layer structure design, including an outer sleeve 302 and an inner sleeve 303. The rubber layer 301 is located between the outer sleeve 302 and the inner sleeve 303, and the outer sleeve 302 and the inner sleeve 303 are bonded together by vulcanization through the rubber layer 301. The outer sleeve 302 is made of metal and is circular in shape. The outer sleeve 302 matches the inner circumferential surface of the pin seat 29, which facilitates the press-fitting of the inner sleeve 30. The inner sleeve 303 is elliptical and is made of non-metallic wear-resistant material, such as nylon PA66 or PTFE. The center pin 31 is a metal cylinder, and its outer circumferential surface is not covered with a rubber bushing. In this way, when the center pin 31 contacts the inner sleeve 303, due to the advantages of non-metallic wear-resistant material such as good wear resistance, low friction, and low friction noise, excessive wear can be avoided, the service life of the center pin sleeve can be improved, friction noise can be reduced, and the environmental impact can be reduced.

[0142] It should be noted that in this embodiment, an inner sleeve is added between the pin seat and the center pin, and the inner sleeve is installed into the pin seat in a detachable manner. Even if the inner sleeve is damaged after working for a long time, it can be easily removed and replaced with a new inner sleeve during maintenance, thereby further improving the service life of the center pin sleeve.

[0143] like Figure 26 and Figure 29 As shown, during operation, when there are many switches on the track where the rail vehicle travels, the vibration is significant. The center pin 31 frequently impacts the inner sleeve 30, and over time, the inner sleeve 30 will rotate, causing product failure. To address this, the applicant has implemented two solutions: firstly, by press-fitting the outer sleeve 302 of the inner sleeve 30 into the pin seat 29, the interference fit prevents rotation of the inner sleeve; secondly, the applicant has designed bolt holes 32 at the bottom of the inner sleeve 30, using bolts (not shown in the figure) to pass through the pin seat 29 and bolt holes 32 to lock the inner sleeve 30 and pin seat 29 together, thereby further preventing the inner sleeve from rotating and preventing product failure.

[0144] from Figure 29 It can be seen that the inner sleeve 30 also includes a bottom cover 304. The bottom of the outer sleeve 302 is longer than the bottom of the inner sleeve 303 by a certain distance, so that the bottom of the outer sleeve 302 extends out of the bottom of the inner sleeve 303. The bottom cover 304 is connected to the inner circumferential surface of the bottom of the outer sleeve 302 by interference fit. Multiple bolt holes 32 are provided on the bottom cover 304.

[0145] A drain hole 33 is also provided on the bottom cover 304. During vehicle operation, the friction between the center pin and the inner sleeve may cause debris to fall into the bottom cover. In addition, during vehicle washing, water may enter the bottom cover. To prevent water, debris and other impurities from accumulating inside the bottom cover, a drain hole is designed so that sewage and friction debris can be discharged in a timely manner from the drain hole 33.

[0146] As can be seen from the above technical solution, this embodiment, by adding an inner sleeve with an elliptical inner circumferential surface between the center pin and the pin seat, and by designing a rubber layer in the inner sleeve and a non-uniform thickness design for the rubber layer, ensures that the transverse and longitudinal stiffness of the center pin sleeve is anisotropic, and that the pin seat can withstand greater lateral impact forces when the rail vehicle makes sharp turns, preventing damage and failure of the pin seat during operation, thus ensuring the normal operation of the rail vehicle. By designing the inner sleeve as a layered structure and using a non-metallic wear-resistant material for the innermost part, the center pin directly contacts the non-metallic wear-resistant inner sleeve during operation, thus avoiding excessive wear, increasing the service life of the center pin sleeve, reducing friction noise, and minimizing environmental impact. By installing the inner sleeve into the pin seat in a detachable manner, even if the inner sleeve is damaged after a long period of operation, it can be easily removed and replaced with a new inner sleeve during maintenance, further improving the service life of the center pin sleeve. By press-fitting the inner sleeve into the pin seat and locking the inner sleeve and pin seat together with bolts, the problem of rotation of the inner sleeve is further avoided, thus preventing product failure.

[0147] V. Optimization of the traction rod nodes in the traction device:

[0148] As a key component of the secondary traction system, the traction rod joint needs to transmit the vehicle's traction and braking forces, exhibiting large radial deformation. It also needs to withstand the deflection and torsional loads generated during vehicle operation. Therefore, it must simultaneously accommodate large radial loads and large deflection angles. However, large radial stiffness and low deflection stiffness are contradictory and cannot be simultaneously satisfied. Compared to unidirectional loading, this multi-directional composite loading condition is exponentially more severe. Therefore, compared to other linkage joints, the traction rod joint experiences harsher and more complex loading conditions, placing high demands on product reliability. Furthermore, due to space constraints, the traction rod joint has a relatively small external dimension, resulting in a low safety margin. Currently, domestic and international end-users and OEMs are increasingly demanding extended maintenance cycles for rail vehicle suspension components. Therefore, there is an urgent need to design traction rod joints with high radial stiffness and low deflection stiffness to extend their service life.

[0149] The applicant has made the following improvements:

[0150] Example 1:

[0151] A traction rod joint with high radial load capacity and high deflection capacity, such as Figures 30-34 As shown, from the outside in, it includes a node outer sleeve 34, a node rubber body 35, and a node mandrel 36, wherein the node rubber body 35 vulcanizes the node outer sleeve 34 and the node mandrel 36 into a whole. Figures 35-38As shown, the node rubber body 35 has cavities 37 at both ends in the radial direction, and the two symmetrically arranged cavities 37 symmetrically separate the two node rubber body bodies 351. A rubber stop 38 protruding towards the node outer sleeve 34 is provided between the cavity 37 and the node mandrel 36, and the rubber stop 38 is vulcanized on the node mandrel 36; a protective layer 39 is provided between the cavity 37 and the node outer sleeve 34, and the protective layer 39 is vulcanized on the node outer sleeve 34. As... Figure 38 As shown, the thickness h1 of the protective layer 39 rubber is between 0.5 and 1 mm, and its main function is to protect the inner wall of the node outer sleeve 34. Since it is inconvenient to perform other protective processing on the metal surface of this part of the node outer sleeve 34, the protective layer 39 prevents it from being exposed and corroded. The rubber thickness h2 of the rubber stop 38 is between 1 and 1.5 mm. In addition to protecting the node mandrel 36, the rubber here will also deform due to load. The distance between the rubber stop 38 and the protective layer 39 in the cavity 37, as well as the deformation of the rubber stop 38 and the protective layer 39 after load, constitute the displacement after load, which can be set differently according to different vehicle usage conditions.

[0152] Accordingly, such as Figures 39 to 42 As shown, a metal stop 40 protruding towards the node outer sleeve 34 is provided at the position where the node mandrel 36 is vulcanized and bonded to the rubber stop 38, so as to match the shape of the rubber stop 38 so that the rubber thickness is the same at all parts of the rubber stop 38.

[0153] like Figure 38 As shown in the sectional view, the shape of the outer surface of the rubber stop 38 along the axial direction is as follows: the middle part is a straight segment S1, and both ends of the straight segment S1 are connected to an outwardly convex arc segment S2. The other ends of the two arc segments S2 are connected to the two ends of the traction rod node, respectively. Figure 32 As shown, the center c of arc segment S2 lies on the straight line N, which is the line containing the center of straight segment S1 and the center of the traction rod node. Figure 33 As shown, the distance D between the endpoint d at both ends of the arc segment S2 and the plane where the end of the node sleeve 34 is located is ≤2mm. It can be located either outside or inside the end of the node sleeve 34.

[0154] Accordingly, such as Figure 40 and Figure 42 As shown, the shape of the outer surface of the metal stop 40 along the axial direction is also a straight segment in the middle and an outwardly convex arc segment at both ends of the straight segment, so as to better match the shape of the rubber stop 38.

[0155] like Figure 31 and Figure 36As shown, the rubber stop 38 surrounds the metal stop 40, and the cavity 37 is located on both sides of the rubber stop 38 along the circumferential direction, forming a V-shaped shape between the rubber stop 38 and the node rubber body 351. To ensure the overall performance of the traction rod node, such as... Figure 36 and Figure 38 As shown, the area corresponding to the surface formed by the straight segment S1 of the rubber stop 38 should satisfy the following condition: F / (L*W) ≤ mandrel yield strength, where F is the load, W is the width of the outer surface of the rubber stop 38 along the circumference, and L is the length of the straight segment S1. Simultaneously, to ensure the large radial stiffness requirement of the traction rod node, such as... Figure 36 As shown, the central angle α corresponding to the circumferential outer surface width W of the rubber stop 38 is ≤ π / 3. Therefore, in the design, the radial stiffness requirement should be considered first, and a larger width W should be selected to minimize the length L. In this way, during use, when the radial stiffness is not too large, the rubber stop 38 and the protective layer 39 will not come into close contact, and the traction rod node can easily deflect. Only after deflecting to a certain angle will the rubber stop 38 and the protective layer 39 come into contact. Moreover, the smaller L is, the later the rubber stop 38 and the protective layer 39 come into contact, that is, the smaller the deflection stiffness. When the radial stiffness is large, after the rubber stop 38 and the protective layer 39 come into contact, there is still space between the arc segment S2 of the rubber stop 38 and the protective layer for the traction rod node to deflect. Moreover, the smaller L is, the easier it is to complete the deflection.

[0156] Example 2:

[0157] This embodiment is a variation based on Embodiment 1. In this embodiment, the shape of the arc segment S2 of the rubber stop 38 can also be a diagonal segment. Correspondingly, the arc segment shape of the metal stop 40 also needs to be changed to a diagonal segment. Moreover, in the following embodiments, although the accompanying drawings are all shown as arc segments, diagonal segments can also be used, which will not be explained in detail later.

[0158] Example 3:

[0159] This embodiment is a modification of the node mandrel 36 based on the above embodiment, such as... Figure 43 and Figure 44As shown, the inner portion of the node mandrel 36 connected to its two ends (the flat sections) and the remaining outer portion are respectively machined into an inner node mandrel 361 and an outer node mandrel 362. This allows the inner node mandrel 361 to be machined from a smaller diameter material, while the outer node mandrel 362 can be machined from tubular material. Only the outer node mandrel 362 is vulcanized and bonded to the node rubber body 35, while the inner node mandrel 361 and outer node mandrel 362 are interference-fitted. The inner node mandrel 361 can also be recycled. Compared to a monolithic mandrel, this significantly reduces waste generated during processing. Furthermore, for traction ball joints with different flat section dimensions but identical dimensions in the middle, this structure improves the production efficiency of traction ball joints.

[0160] At the same time, such as Figure 45 As shown, the distance E between the outermost edge of the flat section and the maximum outer diameter of the inner node mandrel 361 satisfies: E≥1mm, so that the inner node mandrel 361 and the outer node mandrel 362 will not touch the flat section during press fitting, thus reducing costs while ensuring that there will be no interference marks in the exposed area of ​​the node mandrel 36.

[0161] Example 4:

[0162] This embodiment is a modification of the node outer cover 34 based on the above embodiment, such as... Figure 46 and Figure 47 As shown, the node sleeve 34 is designed to be thick in the middle and thin at both ends, so that the inner surface of the node sleeve 34 is a ring formed by a straight line segment T1 in the middle, and the two ends of the straight line segment T1 are respectively connected to a ring formed by an outwardly convex arc segment or oblique line segment T2. Moreover, the length of the straight line segment T1 here is equal to the length of the straight line segment S1 of the rubber stop 38, further reducing the deflection stiffness of the traction rod node.

[0163] The table below shows the traction rod node (referred to as structure A in the table) of the above embodiment 1 with the same installation dimensions, and the existing ones. Figure 48 and Figure 49 The deflection stiffness values ​​of the tie rod node (referred to as structure B in the table) under different load conditions are shown below:

[0164]

[0165]

[0166] As can be seen from the table above, under the same load conditions, compared with the structure of Example 1, the existing traction rod node structure has a larger increase in deflection stiffness, making it more difficult to deflect.

[0167] As can be seen from the above technical solution, it has the following characteristics:

[0168] 1. This invention reduces the contact area of ​​the rubber by creating cavities at both radial ends of the rubber body, and then placing rubber stops protruding towards the outer sheath within the cavities, allowing the joint to still deflect after deformation under high stiffness. Furthermore, by placing metal stops at the rubber stops, the joint can still provide high radial stiffness.

[0169] 2. This invention combines straight segments with oblique segments or outwardly convex arc segments on the axial outer surface of the rubber stop. When subjected to radial loads, the rubber surface first contacts the straight segments, and voids remain at both ends after contact, allowing the joint to still deflect after deformation. Simultaneously, while meeting stiffness requirements, the length of the straight segments is minimized to reduce the contact area of ​​the rubber at these points, ensuring deflection.

[0170] 3. By setting the surface shape of the metal stop to be the same as that of the rubber stop, the thickness of the rubber stop is basically the same everywhere, ensuring uniform deformation of the rubber. Furthermore, by controlling the rubber thickness to 1 to 1.5 mm, the invention can achieve anti-corrosion function, reduce the stress on the rubber during deformation, improve the service life of the rubber, and ensure the stability of rigidity.

[0171] 4. This invention improves service life by providing a protective layer on the outer casing, preventing the inner side of the casing from being exposed and corroded or rusted.

[0172] 5. By setting the outer jacket and protective layer into a surface axially formed by straight segments and oblique segments or circular arc segments, the present invention further reduces the contact area between the upper and lower surfaces of the cavity after node deformation, thereby facilitating node deflection.

[0173] In summary, this invention further optimizes the performance of suspension devices and components in railway vehicle bogies, meeting increasingly stringent performance requirements from users. By optimizing the primary conical spring, sufficient space is provided at the annular end of the outer sleeve for machining end bolt holes within a limited space. Simultaneously, the rubber components maintain a constant strain state under pressure, preventing cracks and damage during use and avoiding premature failure, thus improving product lifespan and performance. Furthermore, by optimizing the lateral stop by designing a frustum-shaped wear plate that integrates with the rubber body after vulcanization, it prevents detachment. Additionally, stepped holes in the base plate create anti-detachment buckles at these holes, further preventing detachment and ensuring the stability of the composite lateral stop. By wrapping at least one layer of fabric around the annular rubber of the supporting ball joint as a torsional stiffness reinforcing material, the fabric is used to tighten the annular rubber under load, reducing the degree of deformation. This increases the torsional stiffness of the supporting ball joint, reduces the deformation in the torsional direction, prevents excessive displacement of the torsion bar from colliding with other components, and reduces noise and abnormal sounds. Through the above improvements, the designed supporting ball joint can generate normal torsional deformation in the early stage. After torsion at a certain angle, it is torsionally limited, achieving nonlinear torsional stiffness, thus preventing excessive displacement of the torsion bar from colliding with other components and reducing noise and abnormal sounds. By adding an inner sleeve with an elliptical inner circumference between the center pin and the pin seat, and by setting a rubber layer in the inner sleeve and designing the rubber layer with non-uniform thickness, this invention can ensure the anisotropic lateral and longitudinal stiffness of the center pin sleeve, and also ensure that the pin seat can withstand greater lateral impact force when the rail vehicle is turning sharp bends. This prevents the pin seat from being damaged and failing during operation, thus ensuring the normal operation of the rail vehicle. This invention reduces the contact area of ​​the rubber by creating cavities at both radial ends of the rubber body, and then placing rubber stops protruding towards the outer sheath within the cavities, allowing the joint to deflect even after deformation under high stiffness. Furthermore, by placing metal stops at the rubber stops, the joint can still provide high radial stiffness.

[0174] In this embodiment, "multiple" refers to "two or more". The above embodiments are for illustrative purposes only and are not intended to limit the invention. Those skilled in the art can make various changes or modifications without departing from the spirit and scope of the invention. Therefore, all equivalent technical solutions should also fall within the protection scope of this invention, which is defined by the claims.

Claims

1. An optimization method for a railway vehicle bogie, characterized in that: It optimizes the locomotive bogie by optimizing the suspension device and suspension components of the rail vehicle bogie. Specifically, the suspension device is optimized by designing the primary conical spring rubber layer with equal strain within a limited space, optimizing the anti-fall-off performance of the lateral stop, and increasing the torsional stiffness of the support ball joint in the anti-roll torsion bar device. The suspension components are optimized by designing the center pin sleeve in the traction device with anisotropic lateral and longitudinal stiffness and optimizing the radial load-bearing and deflection capacity of the traction tie rod nodes. The center pin sleeve includes a center pin and a pin seat; The design of the center pin sleeve with anisotropic lateral and longitudinal stiffness involves adding an inner sleeve inside the pin seat. The inner circumferential surface of the inner sleeve is elliptical. The center pin is inserted into the inner sleeve and connects with it. A free gap is formed between the inner sleeve and the center pin along the major axis of the inner circumferential surface of the elliptical inner sleeve, and a clearance fit is formed between the inner sleeve and the center pin along the minor axis of the inner circumferential surface of the elliptical inner sleeve. The major axis of the inner circumferential surface of the elliptical inner sleeve is consistent with the lateral direction of the rail vehicle, and the minor axis of the inner circumferential surface of the elliptical inner sleeve is consistent with the longitudinal direction of the rail vehicle. When the rail vehicle is traveling in a straight line, the longitudinal stiffness is formed by the contact between the center pin and the inner sleeve to transmit traction or braking force. When the rail vehicle is turning a large curve or a gentle curve, the center pin will move in the free clearance and will not contact the inner sleeve in the transverse direction. When the rail vehicle is turning a small curve or a sharp curve, the center pin will first move in the free clearance, and after eliminating the free clearance, it will contact the inner sleeve to form transverse stiffness. A rubber layer is provided in the inner sleeve. By adopting a non-uniform thickness design for the rubber layer, the lateral stiffness formed by the contact between the center pin and the inner sleeve is greater than the longitudinal stiffness formed by the contact between the center pin and the inner sleeve. The cross-section of the center pin is circular. When the center pin is in the center position, along the minor axis direction, the diameter of the center pin matches the minor axis of the inner circumferential surface of the elliptical inner sleeve, so that a clearance fit is formed between the inner sleeve and the center pin. Along the major axis direction, the diameter of the center pin is less than the major axis of the inner circumferential surface of the elliptical inner sleeve, so that a free gap is left between the inner circumferential surface of the inner sleeve and the center pin.

2. The optimization method according to claim 1, characterized in that: The tapered spring includes an outer sleeve, a mandrel, and multiple spacers located between the outer sleeve and the mandrel. The outer sleeve, mandrel, and multiple spacers are bonded together by multiple rubber bodies through vulcanization. The inner circumferential surface of the outer sleeve and the spacers are both tapered. An end screw hole is also opened on one end of the outer sleeve. The method of performing constant strain design of the tapered spring rubber layer in a limited space involves extending the radial width of the annular end toward the central axis of the tapered spring while keeping the outer diameter of the outer sleeve constant, to ensure the machining of the end screw hole, and then increasing the gap between the inner circumferential surface of the outer sleeve and the outermost spacer to achieve constant strain design of multiple rubber bodies located between the outer sleeve, the mandrel and multiple spacers after vulcanization.

3. The optimization method according to claim 2, characterized in that: In the axial section of the conical spring, the angle between the inclined line of the conical spacer located on one side of the central axis of the conical spring and the central axis of the conical spring is set as θ1. Then the θ1 of each spacer is equal. The angle between the inclined line of the inner circumferential surface of the outer sleeve located on one side of the central axis of the conical spring and the central axis of the conical spring is set as θ2. The specific steps of the method for achieving constant strain design of the rubber layer of a conical spring within a limited space are as follows: First, the inclined line at the position of the inner circumferential surface of the outer jacket located on one side of the central axis of the conical spring is set to be parallel to the inclined line at the position of the conical spacer located on one side of the central axis of the conical spring. While ensuring that the outer diameter of the outer jacket remains unchanged, the inclined line at the position of the inner circumferential surface of the outer jacket is moved parallel to the central axis of the conical spring to the inclined line P, so that the radial width of the annular end of the outer jacket is greater than the diameter of the end screw hole. The inclined line P intersects the horizontal line extending horizontally along the annular end of the upper end of the outer jacket at point S. Then, with point S as the center, the inclined line P is rotated by a certain angle away from the central axis of the conical spring to obtain the inclined line Q, so that the inclined line at the position of the inner circumferential surface of the outer jacket is moved to the inclined line Q. Then, θ2 > θ1, so as to increase the gap between the inner circumferential surface of the outer jacket and the outermost spacer. Through the above steps, the constant strain design of multiple rubber bodies located between the outer jacket, the mandrel and multiple spacers after vulcanization is achieved.

4. The optimization method according to claim 1, characterized in that: The optimization of the anti-detachment performance of the transverse stop involves designing the transverse stop as a base plate, a rubber body, a wear plate, and a cover plate. The base plate, rubber body, and wear plate are vulcanized into a whole, and the cover plate is assembled at the center of the bottom end of the vulcanized whole. The wear plate is shaped like a frustum, and the rubber body is also shaped like a frustum. The center of the frustum of the rubber body has a through hole along the axial direction, and the wear plate is located in the through hole of the rubber body. The center of the base plate also has a through hole, and the cover plate is installed at the through hole of the base plate. The edge of the base plate near its through hole also has a stepped hole, which allows rubber to flow into the stepped hole to form an anti-detachment buckle, thereby preventing the rubber body from falling off.

5. The optimization method according to claim 1, characterized in that: The supporting ball joint includes an outer sleeve, a hollow mandrel disposed inside the outer sleeve, and an annular rubber located between the outer sleeve and the hollow mandrel. The outer sleeve and the hollow mandrel are bonded together by vulcanization of the annular rubber. The design to increase the torsional stiffness of the supporting ball joint in the anti-roll torsion bar device is to wrap at least one layer of fabric in the annular rubber and along the circumference of the annular rubber as a torsional stiffness reinforcing material to increase the torsional stiffness of the supporting ball joint. When at least one layer of fabric is wound around the annular rubber, a layered structure of rubber and fabric is formed inside the annular rubber. Before the ball joint is vulcanized, strips of fabric and strips of rubber are prefabricated. The strips of fabric and rubber are then stacked together and rolled into a ring. The ring-shaped fabric and rubber are then placed between the outer jacket and the hollow mandrel for product vulcanization, resulting in a layered structure of rubber and fabric inside the ring-shaped rubber.

6. The optimization method according to claim 1, characterized in that: The supporting ball joint includes an outer sleeve, a hollow mandrel disposed inside the outer sleeve, and an annular rubber located between the outer sleeve and the hollow mandrel. The outer sleeve and the hollow mandrel are bonded together by vulcanization of the annular rubber. The design to increase the torsional stiffness of the supporting ball joint in the anti-roll torsion bar device involves setting a limiting structure one on the outer sleeve and a limiting structure two on the hollow mandrel. The limiting structure one and the limiting structure two have a circumferential gap L. When torsion occurs, the annular rubber deforms under the action of external force, causing the outer sleeve and the hollow mandrel to rotate relative to each other in the circumferential direction, and the gap L decreases. The limiting structure one and the limiting structure two then rotate relative to each other in the circumferential direction. When the gap is eliminated, the limiting structure one and the limiting structure two come into contact to form a circumferential limiting structure.

7. The optimization method according to claim 6, characterized in that: The limiting structure is firstly an opening on the outer sleeve, and secondly a pin on the hollow mandrel. The pin extends radially into the opening on the outer sleeve, and the gap L is the space between one side of the pin and one inner side of the opening along the circumferential direction. After the product is vulcanized, the annular rubber fills the gap L. During torsion, the pin rotates relative to the opening, the gap L decreases, the rubber in the gap L is compressed, and normal torsional deformation occurs in the early stage. When the gap L disappears, one side of the pin contacts one inner side of the opening to form a circumferential limiting structure.

8. The optimization method according to claim 1, characterized in that: The traction rod node includes a node mandrel vulcanized as a whole, a node rubber body, and a node outer sleeve; The optimization of the traction rod node to improve radial load-bearing and deflection capacity involves providing cavities at both ends of the node rubber body in the radial direction, and providing rubber stops protruding towards the node outer sleeve in the cavities, so that the cavities form two V-shaped shapes between the rubber stops and the node rubber body on both sides of the circumference. The node mandrel located at the rubber stops is provided with a metal stop protruding towards the node outer sleeve.

Images