Large-tonnage valve high-temperature fatigue wear test device

Abstract

The invention discloses a large-tonnage valve high-temperature fatigue wear test device, which comprises a rack, a loading platform, a lifting mechanism for driving the loading platform to perform lifting motion, and a heating mechanism for heating a valve seat ring, the invention relates to an air valve displacement control device which comprises an air valve, a rotation driving mechanism for driving the air valve to rotate, an air valve displacement control mechanism for driving the air valve to reset and a pressure applying mechanism. The pressure applying mechanism comprises a pressure applying driving assembly, a large disc spring assembly and a small disc spring assembly, the pressure driving assembly comprises at least two eccentric cams which are driven by the same rotary driving shaft and have phase difference, and at least two paths of power transmission units which are arranged in one-to-one correspondence with the at least two eccentric cams, and the eccentric distances of the at least two eccentric cams are different; therefore, at least two different working conditions of valve opening, seating and detonation pressure impact are simulated. The problem that multiple working conditions and large-tonnage impact of the valve cannot be simulated at the same time in the prior art is solved.

Description

Technical Field

[0001] This invention relates to the field of valve testing technology, and in particular to a high-temperature fatigue wear testing device for large-tonnage valves. Background Technology

[0002] The valve-seat contact pair ensures the sealing of the combustion chamber and is a crucial component of it, controlling the engine's "breathing." With the continuous improvement of engine technology in my country, there is a growing demand for higher power density—greater power and smaller size—leading to continuously increasing temperatures and pressures within the engine cylinders, further deteriorating the working environment for engine components. Simultaneously, increasing reliability requirements place higher demands on various engine components. Furthermore, the use of cleaner fuels such as natural gas in recent years has resulted in higher combustion temperatures and cleaner combustion products within the engine cylinders. While appropriate combustion products covering the contact surfaces can provide lubrication and protection for gasoline and diesel engines, this "protective" effect is significantly reduced for natural gas engines. During each engine cycle, the intake and exhaust valves undergo opening and closing, enduring frequent reciprocating seating impacts, direct scouring from high-temperature combustion gases, and corrosive effects from exhaust gases—creating a harsh working environment. For example, the actual temperature of the valve disc bottom surface in a high-performance diesel engine is approximately 700-800℃, and the combustion knock pressure has increased from below 20 bar to 25-30 bar, significantly increasing the impact load on the valves. A 150kN valve high-temperature impact wear device can more accurately simulate the high-temperature and high-pressure impact environment of valves in actual operation, thus more realistically reflecting the valve wear condition. Therefore, higher requirements are placed on valve performance.

[0003] Traditional methods for assessing high-temperature valve wear performance have significant limitations. While engine testing can reflect real-world operating conditions, it is time-consuming, costly, and complex, making it unsuitable for rapid R&D needs. Material-level friction and wear testing can only test basic material properties and cannot reproduce the dynamic behavior of valves during the engine's operating cycle, such as the coupling effect of rotational motion, seat impact, and detonation pressure impact. Therefore, it cannot accurately capture the valve "sinking" mechanism caused by wear in actual use. The test data from dedicated high-temperature valve wear testing devices differ significantly from those from material-level tests, highlighting the necessity of dedicated equipment. Existing devices mostly use mechanical or hydraulic drives. Mechanical drives are simple in structure but have low load loading accuracy and limited seat frequency, making it difficult to accurately simulate high-dynamic conditions. Hydraulic drives, while responsive and capable of high load loading, are prone to oil leaks that pollute the environment, and the purchase and maintenance costs are too high. Furthermore, most current devices lack the ability to simultaneously simulate valve rotation conditions and high-tonnage explosion impacts, especially failing to stably reproduce instantaneous impact forces at the 150kN level. This results in test results deviating from actual engine operating conditions, making it impossible to comprehensively assess the fatigue wear characteristics of valves under high-temperature, high-pressure combined stress. To address these issues, existing technologies urgently need improvement. Summary of the Invention

[0004] In view of the shortcomings of the existing technology, the present invention provides a solution.

[0005] This invention is achieved using the following technical solution: A high-tonnage valve high-temperature fatigue wear testing device, characterized in that it comprises: frame; A loading platform is installed on the frame. The loading platform is provided with a valve, a valve seat that cooperates with the valve head, a rotary drive mechanism for driving the valve to rotate, and a valve displacement control mechanism for driving the valve to reset. A lifting mechanism, mounted on the frame, is used to drive the loading platform to move vertically up and down; The heating mechanism, mounted on the frame, is used to heat the valve seat rings; The pressure-applying mechanism includes a pressure-applying drive assembly, a large disc spring assembly and a small disc spring assembly that sequentially transmit load. The pressure-applying drive assembly includes at least two eccentric cams driven by the same rotary drive shaft and having a phase difference, and at least two power transmission units corresponding one-to-one with the at least two eccentric cams. Each power transmission unit includes a transmission component, a hammer, and a fixed guide structure. The first end of the transmission component is connected to the corresponding eccentric cam, and the second end is hinged to the hammer. The hammer is slidably connected to the guide structure. Each eccentric cam is configured to drive its corresponding hammer through its respective transmission component under the drive of the rotary drive shaft. Each of the hammers is constrained by the guide structure to reciprocate along a set straight path. The large disc spring assembly is configured to receive the force output by the hammer and reciprocate linearly. The load of the large disc spring assembly is adjustable. The input end of the small disc spring assembly is connected to the output end of the large disc spring assembly. The output end of the small disc spring assembly is used to apply pressure to the valve, and the valve distance is adjustable. At least two eccentric cams have different eccentric distances, so that the hammers corresponding to them can apply different loads to the valve through the large disc spring assembly and the small disc spring assembly, thereby simulating at least two different working conditions in valve opening, seating, and burst pressure impact.

[0006] Furthermore, the at least two eccentric cams include a central eccentric cam and two side eccentric cams, and the at least two transmission units include a central hammer and two side hammers. The eccentric distance of the central eccentric cam is smaller than that of the side eccentric cams. The central eccentric cam drives the large disc spring assembly to generate a first impact through the central hammer to simulate the valve seating condition. The two side eccentric cams drive the large disc spring assembly to generate a second impact greater than the first impact through their respective side hammers to simulate the burst pressure impact condition.

[0007] Furthermore, the large disc spring assembly includes a first housing slidably connected to the frame, a large disc spring group housed inside the first housing, and a hammer head connected to the large disc spring group. The first housing is provided with a detachable maintenance panel. By opening the maintenance panel, the number of the large disc spring group can be increased, decreased, or replaced to adjust the load of the large disc spring assembly. The head of the hammer head is provided with a pressure sensor, and the hammer head constitutes the output end of the large disc spring assembly.

[0008] Furthermore, the first housing is provided with a disc spring cavity for accommodating the disc spring assembly. A disc spring cover is installed on the upper side of the disc spring cavity, and a mounting block and an adjusting block are provided on the lower side. The upper part of the mounting block forms a bearing surface for supporting the disc spring assembly. The adjusting block is located below the mounting block, and the two are connected by a relative inclined surface. The first housing is provided with an adjusting bolt. The shank of the adjusting bolt passes through the first housing and is threadedly connected to the adjusting block. Rotating the adjusting bolt can drive the adjusting block to move relative to the mounting block, thereby changing the axial position of the mounting block to adjust the preload of the disc spring assembly.

[0009] Furthermore, the disc spring assembly includes: The second housing is fixedly mounted on the frame; A displacement element is movably disposed within the second housing. The displacement element is connected to the output end of the large disc spring assembly and can be driven by it to move axially. The displacement element has a small disc spring cavity. The mounting base plate is movably mounted within the small disc spring cavity. The lower end of the small disc spring assembly is placed on the mounting base plate; A disc spring shaft has its lower end inserted into the small disc spring assembly and its lower end is restricted to move within the small disc spring cavity. A top head is installed on the upper end of the disc spring shaft that extends out of the second housing. The top head constitutes the output end of the small disc spring assembly. The distance between the top head and the valve can be adjusted by adjusting the number of small disc spring assemblies.

[0010] Furthermore, the small disc spring assembly also includes an adjusting nut. The displacement element has a threaded section, and the adjusting nut is screwed onto the threaded section. The end face of the adjusting nut has multiple pushing members. The pushing members pass through the through hole reserved in the displacement element and abut against the mounting base plate. By rotating the adjusting nut to change its axial position on the threaded section, the pushing members can be driven to push the mounting base plate to move within the small disc spring cavity, thereby adjusting the preload state of the small disc spring assembly.

[0011] Furthermore, the upper end of the disc spring shaft is provided with a clamping assembly for clamping the top head. The clamping assembly includes a connecting plate, a chuck seat, a collet, and a chuck lock nut. The bottom of the connecting plate is provided with a groove for fitting and installing with the upper end of the disc spring shaft. The chuck seat is fixed on the connecting plate. The collet is installed in the center hole of the chuck seat and the two are fixed together by a bevel engagement. The chuck lock nut is threaded to the chuck seat and is fixed together with the collet by a bevel engagement. The fixed end of the top head is clamped and fixed by the collet.

[0012] Furthermore, the lifting mechanism includes a lifting drive source, a transmission assembly, and two worm gear lifts for lifting the loading platform. Each worm gear lift includes a first worm wheel and a first worm that mesh with each other, and a lead screw driven by the first worm wheel and capable of vertical movement. The upper end of the lead screw movably extends through the mounting plate at the top of the frame, and the lower end of the lead screw is fixedly connected to the loading platform. The first worms of the two worm gear lifts are coaxially connected through a coupling to achieve synchronous input. An input gear is fixedly provided at the end of one of the first worms. The lifting drive source includes a handle rotatably mounted on the frame and a drive gear coaxially fixed to the handle's rotating shaft. The input gear and the drive gear are connected through a transmission assembly.

[0013] Furthermore, the rotary drive mechanism includes a gear reducer motor, a second worm wheel and a second worm gear that mesh with each other, and a connecting shaft. The second worm gear is driven to the output shaft of the gear reducer motor, and the second worm wheel is connected to the connecting shaft through a shaft hole fitting structure to drive the connecting shaft to rotate synchronously. The connecting shaft is used to drive the valve to rotate.

[0014] Furthermore, the connecting shaft is provided with an adjustable locking mechanism at one end for connecting the valve, for clamping and locking the valve tail. The adjustable locking mechanism includes an axial center hole on the end face of the connecting shaft and a locking element. The locking element can be operated to change the diameter of the axial center hole, thereby clamping or releasing the valve tail.

[0015] Furthermore, the valve displacement control mechanism includes a spring chamber and a return spring installed in the spring chamber. The tail of the valve extends into the spring chamber and is fixedly connected to one end of the linkage shaft. The end of the linkage shaft used to connect the valve is fitted with an upper thrust ball bearing and a lower thrust ball bearing. The return spring is fitted outside the linkage shaft, and the two ends of the return spring abut against the inner end face of the spring chamber and the upper thrust ball bearing, respectively.

[0016] Furthermore, the heating mechanism includes an induction heater and an electric heating coil connected to the induction heater. The electric heating coil is wound around the valve seat ring below the valve seat ring. The innermost diameter of the electric heating coil is larger than the outer diameter of the valve seat ring. An annular heat preservation cover is provided on the outside of the electric heating coil. The annular heat preservation cover is fixed on the loading platform. A cooling cylinder is installed on the loading platform. The interior of the cooling cylinder forms a circulating water cooling cavity and an axial channel. The circulating water cooling cavity is connected to an external water source. The valve stem is located in the axial channel.

[0017] Furthermore, it also includes a temperature measuring mechanism, which comprises a thermocouple and an infrared thermometer. The head of the thermocouple is configured to be aligned with the stem of the valve, and the optical probe of the infrared thermometer is configured to be aligned with the valve seat to measure its operating temperature in a non-contact manner.

[0018] Furthermore, the bottom of the cooling cylinder is provided with a valve clamp, the valve clamp is provided with a groove that matches the valve seat ring, the axial channel is provided with a guide for wrapping the valve stem, the upper end face of the valve clamp is provided with a through hole for the valve stem to pass through the guide, and the guide abuts against the upper end face of the valve clamp to constrain the guide to move downward.

[0019] Compared with the prior art, the beneficial effects of the present invention include at least the following: This invention, through the design of a pressure application mechanism, a large disc spring assembly, a small disc spring assembly, a rotary drive mechanism, a valve displacement control mechanism, a heating mechanism, and a lifting mechanism, better simulates the wear test of the valve-seat pair under different loads, speeds, and temperatures. The design of different phase differences and eccentric distances between the eccentric wheel cams allows for the simulation of actual valve opening, seating, and detonation impact conditions. The large disc spring assembly can control the load force by increasing or decreasing the number of large disc springs, better simulating the high-temperature wear conditions of modern high-performance engine valves. This avoids the disadvantages of long testing times, high costs, and complex processes associated with whole-engine testing, and enables the evaluation of high-temperature valve wear performance within a very short test cycle. The small disc spring assembly can adjust the distance from valve opening to seating by increasing or decreasing the number of small disc springs. The rotary drive mechanism can drive the valve to rotate, more closely reflecting the actual operating conditions of an engine and further improving the reliability of the test. Through the above design, the technical problems of existing technologies in simulating multiple valve operating conditions, large-tonnage impacts, and adjustment flexibility are effectively solved. This device can evaluate the high-temperature fatigue wear performance of valves in a way that is closer to the actual engine working environment within a very short test cycle. It provides a reliable test basis for the research and development and optimization of engine components, avoids the time-consuming and costly whole engine test, and overcomes the shortcomings of existing dedicated test devices. Attached Figure Description

[0020] Figure 1 This is one of the schematic diagrams of a high-tonnage valve high-temperature fatigue wear testing device according to an embodiment of the present invention; Figure 2 This is the second schematic diagram of the high-tonnage valve high-temperature fatigue wear testing device according to an embodiment of the present invention; Figure 3 This is the third schematic diagram of the high-tonnage valve high-temperature fatigue wear testing device according to an embodiment of the present invention; Figure 4This is the fourth schematic diagram of the high-tonnage valve high-temperature fatigue wear testing device according to an embodiment of the present invention; Figure 5 This is a cross-sectional view of the high-tonnage valve high-temperature fatigue wear testing device according to an embodiment of the present invention; Figure 6 A schematic diagram of the pressure application component according to an embodiment of the present invention; Figure 7 A partial structural cross-sectional view of the pressure-driving component according to an embodiment of the present invention; Figure 8 Assembly cross-sectional view of the large disc spring assembly and the small disc spring assembly according to an embodiment of the present invention; Figure 9 A cross-sectional view of the cooling cylinder according to an embodiment of the present invention; Figure 10 A schematic diagram of the linkage shaft according to an embodiment of the present invention; Figure 11 yes Figure 10 A sectional view of the structure; Figure 12 This is a schematic diagram of the intermediate eccentric cam according to an embodiment of the present invention; Figure 13 This is a schematic diagram of a side-eccentric cam according to an embodiment of the present invention; Figure 14 This is a displacement trajectory diagram of the top head according to an embodiment of the present invention; In the diagram: 1. Frame; 2. Loading platform; 31. First worm gear; 32. Lead screw; 33. Coupling; 34. Input gear; 35. Transmission assembly; 36. Handle; 41. Gear reduction motor; 42. Second worm wheel; 43. Second worm gear; 44. Connecting shaft; 441. Axial center hole; 442. Cutout; 443. Threaded hole; 51. Spring chamber; 52. Return spring; 53. Upper thrust ball bearing; 54. Lower thrust ball bearing; 61. Induction heater; 62. Heating coil; 63. Annular insulation cover; 64. Cooling cylinder; 65. Circulating water cooling cavity; 66. Conduit; 67. Valve clamp; 71. Thermocouple; 72. Infrared thermometer; 81. Pressure drive assembly; 810. Rotary drive shaft; 811. Central eccentric cam; 8 12. Side eccentric cam; 813. Transmission component; 814. Hammer; 815. Guide structure; 82. Large disc spring assembly; 821. First housing; 822. Large disc spring assembly; 823. Hammer head; 824. Maintenance panel; 825. Pressure sensor; 826. Large disc spring cover; 827. Mounting block; 828. Adjusting block; 829. Adjusting bolt; 83. Small disc spring assembly; 831. Second housing; 832. Displacement element; 833. Mounting base plate; 834. Small disc spring assembly; 835. Disc spring shaft; 836. Top head; 837. Adjusting nut; 838. Pushing component; 839. Clamping assembly; 8391. Connecting plate; 8392. Chuck seat; 8393. Collet; 8394. Chuck lock nut; 91. Valve; 92. Valve seat ring. Detailed Implementation

[0021] Exemplary embodiments will now be described more fully with reference to the accompanying drawings. However, these exemplary embodiments can be implemented in many forms and should not be construed as limited to the embodiments set forth herein; rather, they are provided to make the invention more comprehensive and complete, and to fully convey the concept of the exemplary embodiments to those skilled in the art. The same reference numerals in the drawings denote the same or similar structures, and therefore repeated descriptions of them will be omitted.

[0022] The terms used to express position and direction in this invention are illustrated with reference to the accompanying drawings, but changes can be made as needed, and all such changes are included within the scope of protection of this invention.

[0023] like Figures 1 to 14 As shown, the present invention provides a high-tonnage valve high-temperature fatigue wear testing device, comprising: Rack 1; Loading platform 2 is installed on the frame 1. The loading platform 2 is provided with valve 91, valve seat ring 92 that cooperates with the head of valve 91, rotary drive mechanism for driving valve 91 to rotate, and valve displacement control mechanism for driving valve 91 to reset. A lifting mechanism, mounted on the frame 1, is used to drive the loading platform 2 to move vertically up and down; A heating mechanism, mounted on the frame 1, is used to heat the valve seat 92; The pressure-applying mechanism includes a pressure-applying drive assembly 81, a large disc spring assembly 82 and a small disc spring assembly 83 that sequentially transmit loads. The pressure-applying drive assembly 81 includes at least two eccentric cams driven by the same rotary drive shaft 810 and having a phase difference, and at least two power transmission units corresponding to the at least two eccentric cams. Each power transmission unit includes a transmission component 813, a hammer 814, and a fixed guide structure 815. The first end of the transmission component 813 is connected to the corresponding eccentric cam, and the second end is hinged to the hammer 814. The hammer 814 is slidably connected to the guide structure 815. Each eccentric cam is configured to drive the corresponding hammer through its respective transmission component 813 under the drive of the rotary drive shaft 810. 814, causing each of the hammers 814 to reciprocate linearly along a set straight path under the constraint of the guide structure 815. The large disc spring assembly 82 is configured to receive the force output by the hammers 814 and reciprocate linearly, and the load of the large disc spring assembly 82 is adjustable. The input end of the small disc spring assembly 83 is connected to the output end of the large disc spring assembly 82. The output end of the small disc spring assembly 83 is used to apply pressure to the valve 91, and the distance of the valve 91 is adjustable. At least two eccentric cams have different eccentric distances, so that the hammers 814 corresponding to them can apply different loads to the valve 91 through the large disc spring assembly 82 and the small disc spring assembly 83, thereby simulating at least two different working conditions of valve 91 opening, sitting, and burst pressure impact.

[0024] In this embodiment, the experimental principle is as follows: First, the valve 91 and valve seat 92 to be tested are mounted on the loading platform 2, and the tail of the valve 91 is clamped and fixed by the rotary drive mechanism so that it can rotate. The return spring 52 in the valve displacement control mechanism keeps the valve 91 in a naturally drooping open state, that is, the head of the valve 91 is separated from the valve seat 92.

[0025] The lifting mechanism is activated, driving the loading platform 2 to move vertically to ensure that the valve 91 can precisely engage with the valve seat 92 when seated, and to adjust the initial distance between the top head 836 of the small disc spring assembly 83 and the head of the valve 91. The heating mechanism is activated to induction heat the valve seat 92, gradually increasing its temperature and stabilizing it at the preset test temperature.

[0026] The pressure application mechanism begins operation. The rotary drive shaft 810 in the pressure application drive assembly 81 begins to rotate. Three eccentric cams are mounted on this rotary drive shaft 810: a central eccentric cam 811 and two side eccentric cams 812. The eccentric distance D1 of the central eccentric cam 811 is designed to be a smaller value, for example, 6 mm, while the eccentric distance D2 of the two side eccentric cams 812 is designed to be a larger value, for example, 20 mm. There is a preset phase difference between these eccentric cams; for example, the side eccentric cams 812 have a phase difference of 160 degrees relative to the central eccentric cam 811.

[0027] When the rotary drive shaft 810 rotates, the central eccentric cam 811 first pushes the large disc spring assembly 82 through its corresponding power transmission unit (including the transmission component 813 and the hammer 814). At this time, the large disc spring assembly 82 receives the force of the hammer 814, generating an impact load that simulates the valve 91 sitting. Since the eccentric distance of the central eccentric cam 811 is relatively small, it drives the hammer 814 to produce a relatively small displacement. This displacement is transmitted through the large disc spring assembly 82 and the small disc spring assembly 83, causing the top 836 of the small disc spring assembly 83 to contact the valve head of the valve 91 and smoothly push the valve 91 towards the valve seat 92, thus realizing the valve 91 sitting.

[0028] Following this, due to the phase difference between the eccentric cams, when the central eccentric cam 811 completes its push stroke and the valve 91 is in the seated state, the two side eccentric cams 812 begin to push the disc spring assembly 82 through their respective power transmission units. Because the side eccentric cams 812 have a larger eccentric distance, they drive the hammer 814 to produce a larger displacement. This larger displacement generates an instantaneous explosive pressure impact load much greater than the seating impact. This process simulates the impact of internal combustion chamber explosion pressure on the valve 91.

[0029] After the side eccentric cam 812 completes its explosive pressure impact stroke, as the rotary drive shaft 810 continues to rotate, the side eccentric cam 812 and the middle eccentric cam 811 sequentially disengage from the power transmission unit, and the compressive force of the large disc spring assembly 82 and the small disc spring assembly 83 is gradually released. The return spring 52 in the valve displacement control mechanism pushes the valve 91 away from the valve seat ring 92, returning it to its initial open state, completing a full test cycle. The entire process is conducted continuously in a high-temperature environment to simulate the high-temperature fatigue wear of the valve 91 in an actual engine. By adjusting the number of disc springs or the preload of the large disc spring assembly 82, the load magnitude of the seat impact and explosive pressure impact can be changed; by adjusting the number of disc springs or the installation position of the small disc spring assembly 83, the seat stroke of the valve 91 and the distance between the valve head 836 and the valve 91 can be adjusted.

[0030] Compared to traditional valve 91 wear testing devices, both mechanical and hydraulic drives have limitations in simulating actual engine operating conditions. For example, some mechanical drive devices may only be able to simulate a single impact load, or may be unable to precisely control the valve 91's seating stroke and detonation impact force, leading to deviations between test results and actual conditions. While hydraulic drive devices offer improvements in accuracy and load capacity, their inherent oil leakage and high cost limit their widespread application. This embodiment introduces at least two eccentric cams with phase differences and different eccentric distances, combined with adjustable load large disc spring assembly 82 and small disc spring assembly 83, to achieve accurate simulation of various complex operating conditions such as valve 91 opening, seating, and detonation impact. Specifically, the central eccentric cam 811 simulates valve 91 seating, while the side eccentric cam 812 simulates detonation impact. Through a sophisticated phase difference design, the two work collaboratively under the drive of the same rotating drive shaft 810, forming a continuous and realistic cyclic process. This multi-condition integrated simulation capability is difficult to achieve with existing single-drive or simple mechanical structures.

[0031] Furthermore, the adjustable load characteristic of the large disc spring assembly 82 allows test personnel to flexibly adjust the magnitude of the impact load according to different engine models and operating conditions, thereby covering a wider range of tests. The adjustable valve 91 distance of the small disc spring assembly 83 further enhances the precise control over the valve 91 seating stroke and impact energy. Compared to existing devices with fixed parameters or inconvenient adjustments, the adjustment flexibility of this embodiment significantly enhances the adaptability and accuracy of the test.

[0032] In summary, this embodiment, through its innovative pressure application mechanism design, effectively solves the technical problems of existing technologies in simulating multiple operating conditions, large-tonnage impacts, and adjustment flexibility of valve 91. This device can evaluate the high-temperature fatigue wear performance of valve 91 in a manner closer to the actual engine operating environment within a very short test cycle, providing a reliable test basis for the research and optimization of engine components. It avoids time-consuming and costly whole-engine testing and overcomes the shortcomings of existing dedicated testing devices.

[0033] In a preferred embodiment, at least two eccentric cams include a central eccentric cam 811 and two side eccentric cams 812, and at least two transmission units include a central hammer 814 and two side hammers 814. The eccentric distance of the central eccentric cam 811 is smaller than the eccentric distance of the side eccentric cams 812. The central eccentric cam 811 drives the large disc spring assembly 82 to generate a first impact through the central hammer 814 to simulate the valve 91 seating condition. The two side eccentric cams 812 respectively drive the large disc spring assembly 82 through the corresponding side hammers 814 to generate a second impact greater than the first impact to simulate the burst pressure impact condition.

[0034] In this embodiment, at least two eccentric cams are configured to include a central eccentric cam 811 and two side eccentric cams 812. This configuration allows the pressure applying mechanism to simultaneously handle impact loads of different types and intensities, providing a basis for simulating the force on the valve 91 under different operating conditions. The eccentric distance of the central eccentric cam 811 is smaller than that of the side eccentric cams 812, and this difference is key to achieving different impact force magnitudes. For example, the eccentric distance of the central eccentric cam 811 can be designed to be smaller to produce smaller displacement, while the eccentric distance of the side eccentric cams 812 can be designed to be larger to produce larger displacement. Specifically, the solution of this application, through ingenious mechanical structure design, enables the pressure applying mechanism to accurately simulate two key operating conditions of the valve 91 in actual operation: seat impact and burst pressure impact. Specifically, the rotary drive shaft 810 in the pressure applying mechanism drives the eccentric cam group with different eccentric distances and specific phase differences. When the rotary drive shaft 810 is driven to rotate (which can be driven by a motor), firstly, the middle eccentric cam 811 with a smaller eccentric distance drives the large disc spring assembly 82 to produce a relatively small displacement through its corresponding middle hammer 814, thereby applying a small first impact force to the valve 91. This first impact force is precisely used to simulate the working condition of the valve 91 contacting the valve seat ring 92 when it is normally seated. Subsequently, due to the preset phase difference between the eccentric cam assemblies, the two side eccentric cams 812 with larger eccentric distances begin to drive the large disc spring assembly 82 through their respective side hammers 814, thereby driving the large disc spring assembly 82 to produce a relatively large displacement, and thus applying a second impact force to the valve 91 that is much greater than the first impact. This second impact force is used to simulate the high-intensity impact condition that the valve 91 is subjected to during engine combustion detonation. In this way, the pressure applying mechanism can simulate the valve 91's seating and detonation impact in a continuous cycle, according to precise timing and different load magnitudes, thus more realistically reproducing the force situation of the valve 91 under actual engine operating conditions. This design, combined with the adjustable load and distance characteristics of the large disc spring assembly 82 and small disc spring assembly 83 in the aforementioned pressure applying mechanism, allows the entire test device to flexibly adjust the magnitude of the impact force and the valve 91's position stroke, further improving the accuracy and adaptability of the simulation. Through the above technical solution, this application can accurately distinguish and simulate the differences in the magnitude and timing of the impact force between the valve 91 seating condition and the detonation impact condition, solving the problem of insufficient simulation accuracy in traditional solutions. By configuring eccentric cams with different eccentric distances and corresponding hammers 814, precise control of applying different loads to the valve 91 is achieved, enabling the seating impact and detonation impact to be simulated realistically and independently. This tiered impact simulation capability significantly improves the accuracy and reliability of the high-temperature fatigue wear test of valve 91, providing more realistic data support for evaluating the performance of valve 91 under complex operating conditions.

[0035] As one specific implementation method, refer to Figure 14 The two side eccentric cams 812 have a 160-degree phase difference relative to the central eccentric cam 811. The eccentric distance D1 of the central eccentric cam 811 can be designed to be 6mm, while the eccentric distance D2 of the two side eccentric cams 812 can be designed to be 20mm. Each eccentric cam is connected to the corresponding hammer 814 through an eccentric bushing, and the hammer 814 reciprocates under the constraint of a linear guide rail. Specifically, the side hammers 814 and the central hammer 814 perform different functions: the central hammer 814, due to the smaller eccentric distance of the central eccentric cam 811 and the fact that there is only one hammer 814, means that it can push the displacement of the mandrel 836 to be small. Its function is to push the mandrel 836 to contact the valve head 91 and seat the valve 91 on the valve seat ring 92 (i.e., Figure 14 The red dot indicates that valve 91 is seated. Note that the middle hammer 814 is now at its highest position (the limit to which the mandrel 836 can be pushed: valve 91 is seated). If the length of valve 91 is different, the position of the loading platform 2 can be adjusted to achieve the same seating effect. The two hammers 814, due to the longer eccentricity of the eccentric cams on both sides, can push the mandrel 836 further. Since both hammers 814 are under force simultaneously, this means that the two hammers 814 are responsible for delivering the explosive pressure impact to valve 91. The two hammers 814 begin working after the middle hammer 814 reaches its highest displacement. Due to the high engine speed, the time it takes for the two hammers 814 to replace the middle hammer 814 is negligible. At this point, valve 91 is seated, and the two hammers 814 move upwards, replacing the middle hammer 814. Because the position of valve 91 is restricted, the disc spring compresses rapidly in a short time, generating explosive pressure. The limit displacement of the two hammers 814 is shown as the blue dot in the diagram. Note that the displacement difference between the blue and red dot positions represents the ultimate compression height of the disc spring. This can also be adjusted by changing the loading platform 2 or the disc spring stiffness to change the burst pressure value. From Figure 14 It can be clearly seen that the two side hammers 814 and the middle hammer 814 (note that the displacement of the two side hammers 814 is necessarily related to the two side eccentric cams, and the same applies to the middle hammer 814; the displacement of the hammer 814 is determined by the eccentric distance of the eccentric cam) form a cycle due to the phase difference and the difference in the eccentric distance of the eccentric cam: the initial state of valve 91 is a naturally drooping state controlled by the return spring 52 (disengaged from valve seat 92) - the displacement of the middle hammer 814 (dotted line in the figure) displaces valve 91 through the top head 836 - the two side hammers 814 then move upward, and due to the restriction of valve 91 position, the disc spring begins to compress rapidly to generate explosive pressure impact - the two side hammers 814 move downward, and the middle hammer 814 takes over, and valve 91 returns to the naturally drooping state due to the action of the return spring - the middle hammer 814... Figure 14The composite displacement, i.e., the displacement of the top 836, represents the entire cyclic working process of the valve 91 fatigue testing machine, and can also be understood as the displacement process of the large disc spring cavity. Note: The position of the vertical axis 0 in the figure is where the hammer 814 is located just touching the bottom force plate of the large disc spring assembly 82.

[0036] In a preferred embodiment, the large disc spring assembly 82 includes a first housing 821 slidably connected to the frame 1, a large disc spring group 822 housed inside the first housing 821, and a hammer head 823 connected to the large disc spring group 822. The first housing 821 is provided with a detachable maintenance panel 824. By opening the maintenance panel 824, the number of large disc spring groups 822 can be increased, decreased, or replaced to adjust the load of the large disc spring assembly 82. A pressure sensor 825 is provided at the head of the hammer head 823, and the hammer head 823 constitutes the output end of the large disc spring assembly 82.

[0037] In this embodiment, the first housing 821 is slidably connected to the frame 1. This slidable connection is typically achieved through mechanisms such as guide rails, sliders, or rollers. As one implementation, the first housing 821 may have a slider on its side, which engages with a corresponding guide rail on the frame 1, allowing the first housing 821 to slide vertically. To address the inconvenience of traditional load adjustment, the first housing 821 is cleverly designed with a detachable maintenance panel 824. When adjusting the load applied to the valve 91, the operator does not need to extensively disassemble the entire pressure-applying mechanism; they can simply open the maintenance panel 824 to directly operate the internal disc spring assembly 822, such as increasing or decreasing the number of disc springs or replacing them with different specifications. In this way, the overall load of the disc spring assembly 82 can be precisely adjusted, thereby achieving precise adjustment of the load applied to the valve 91. This adjustment method is intuitive and reliable, and can adapt to the load requirements of different test conditions.

[0038] A pressure sensor 825 is installed at the head of the hammer 823. This pressure sensor 825 is a device that converts the applied pressure into an electrical signal output. Its placement at the head of the hammer 823 allows it to directly measure the instantaneous pressure or impact force applied by the hammer 823 to the valve 91. This provides real-time force feedback data for the testing process, which is crucial for precise control and monitoring of the test load, ensuring the accuracy and repeatability of the test, and making load control during the test process more refined and visualized. The hammer 823 constitutes the output end of the large disc spring assembly 82, clarifying its role in the entire pressure-applying mechanism—it is the final link in the transmission of force from the large disc spring assembly 822. This means that the impact force generated by the large disc spring assembly 822 ultimately acts on the small disc spring assembly 83 through the hammer 823, and the movement and force state of the hammer 823 directly reflect the working output of the large disc spring assembly 82. Through the above structural design, the problems of complex operation, low efficiency, and possible test interruption when adjusting the load applied to the valve 91 in traditional large-tonnage valve 91 high-temperature fatigue wear test devices are effectively solved. This allows the large-tonnage valve 91 high-temperature fatigue wear test device to more efficiently and accurately adjust and control the impact load applied to the valve 91 when simulating different working conditions such as valve 91 opening, seating, and detonation impact, significantly improving the flexibility, convenience of operation, and reliability of data in the test.

[0039] In a preferred embodiment, the first housing 821 has a large disc spring cavity for accommodating the large disc spring assembly 822. A large disc spring cover 826 is installed on the upper side of the large disc spring cavity, and a mounting block 827 and an adjusting block 828 are provided on the lower side. The upper part of the mounting block 827 forms a bearing surface for supporting the large disc spring assembly 822. The adjusting block 828 is located below the mounting block 827, and the two are connected by a relative inclined surface. The first housing 821 is provided with an adjusting bolt 829. The shank of the adjusting bolt 829 passes through the first housing 821 and is threadedly connected to the adjusting block 828. Rotating the adjusting bolt 829 can drive the adjusting block 828 to move relative to the mounting block 827, thereby changing the axial position of the mounting block 827 to adjust the preload of the large disc spring assembly 822.

[0040] In this embodiment, the large disc spring cavity inside the first housing 821 provides a stable accommodating space for the large disc spring assembly 822, and the large disc spring cover 826 installed on the upper side of the large disc spring cavity provides upper limit and protection for the large disc spring assembly 822. At the lower part of the large disc spring cavity, the bearing surface formed on the upper part of the mounting block 827 directly supports the large disc spring assembly 822, and the adjusting block 828 is located below the mounting block 827; the two are connected by a relative inclined surface fit. When the adjusting bolt 829 is rotated externally, the shank of the adjusting bolt 829 passes through the first housing 821 and is threadedly connected to the adjusting block 828, so that the rotational movement of the adjusting bolt 829 is converted into the horizontal movement of the adjusting block 828. Due to the inclined surface fit between the adjusting block 828 and the mounting block 827, the horizontal movement of the adjusting block 828 drives the mounting block 827 to move vertically. The change in the axial position of the mounting block 827 directly changes the compression of the large disc spring assembly 822, thereby precisely adjusting the preload of the large disc spring assembly 822. Through the above structural design, the process of adjusting the preload of the large disc spring assembly 822 becomes more convenient and precise. Operators no longer need to perform complex disassembly and assembly; they can achieve stepless adjustment of the preload of the large disc spring assembly 822 simply by rotating the adjusting bolt 829, significantly improving the efficiency and accuracy of load adjustment. This adjustment method effectively avoids the frequent disassembly and maintenance problems that may exist in traditional adjustment methods, reducing maintenance costs and time, and extending the service life of the equipment. At the same time, the precise preload adjustment capability allows the large disc spring assembly 82 to more accurately provide the required load, thereby improving the realism and reliability of the entire high-tonnage valve 91 high-temperature fatigue wear testing device in simulating different working conditions.

[0041] In a preferred embodiment, the small disc spring assembly 83 includes: The second housing 831 is fixedly mounted on the frame 1; The displacement element 832 is movably disposed within the second housing 831. The displacement element 832 is connected to the output end of the large disc spring assembly 82 and can be driven by it to move axially. The displacement element 832 is provided with a small disc spring cavity. The mounting base plate 833 is movably installed in the small disc spring cavity, and the lower end of the small disc spring assembly 834 is placed on the mounting base plate 833. The disc spring shaft 835 has its lower end inserted into the small disc spring assembly 834 and its lower end is restricted to move within the small disc spring cavity. The disc spring shaft 835 extends out of the upper end of the second housing 831 and is equipped with the top head 836. The top head 836 constitutes the output end of the small disc spring assembly 83. The distance between the top head 836 and the valve 91 can be adjusted by changing the number of the small disc spring assemblies 834.

[0042] In this embodiment, the displacement element 832 is a movable component inside the small disc spring assembly 83. Its main function is to receive the impact force from the large disc spring assembly 82 and transmit it to the top head 836. The mounting base plate 833 is a movable platform inside the small disc spring cavity, used to support the small disc spring group 834. Its movable mounting means that it can move axially within the small disc spring cavity, thereby changing the effective length or preload state of the small disc spring group 834. The small disc spring group 834 can be composed of one or more stacked disc springs. The lower end of the small disc spring group 834 is placed directly on the mounting base plate 833, and its compression state is changed by moving the mounting base plate 833. The disc spring shaft 835 is a rod-shaped structure that passes through the center of the small disc spring group 834. Its main function is to guide the axial compression and rebound of the small disc spring group 834 and to serve as the central axis for force transmission. The lower end of the disc spring shaft 835 passes through the central hole of the small disc spring group 834, ensuring that the disc spring group can be uniformly compressed axially when under force, avoiding radial off-center loading. The lower end of the disc spring shaft 835 is restricted to move within the small disc spring cavity. This restriction means it cannot freely detach from the cavity, but it can move axially within it. The top end 836 is the component that directly contacts the valve 91. It constitutes the output end of the small disc spring assembly 83 and serves as the interface for physical contact between the small disc spring assembly 83 and the valve 91, directly applying the force generated by the small disc spring assembly 83 to the valve 91. Adjusting the number of small disc spring assemblies 834 allows for adjustment of the distance between the top end 836 and the valve 91, which is one of the core innovations of this design. By increasing or decreasing the number of disc springs in the small disc spring assembly 834, the overall height and stiffness of the small disc spring assembly 834 can be directly changed. When the number of disc springs increases, the initial height of the small disc spring assembly 834 increases, thereby reducing the distance between the valve head 836 and the valve 91 when unloaded. Conversely, when the number of disc springs decreases, the distance between the valve head 836 and the valve 91 increases. This adjustment method is simple in structure and intuitive in operation, requiring no complex mechanical transmission mechanism, and can achieve precise adjustment of the distance between the valve head 836 and the valve 91 to adapt to different valve 91 sizes or simulate different seat strokes.

[0043] The small disc spring assembly 83 of this application, through its ingenious structural design, achieves the function of applying pressure to the valve 91 and flexibly adjusting the distance between the mandrel 836 and the valve 91. When the large disc spring assembly 82 in the pressure applying mechanism generates force, its output end directly drives the displacement element 832 to move axially. The displacement element 832 then drives the disc spring shaft 835 to move axially, thereby allowing the mandrel 836 on the disc spring shaft 835 to directly contact the valve 91, applying impact force to the valve 91, thus simulating the valve 91's seating and detonation impact. The ingenuity of this solution lies in the fact that by adjusting the number of small disc spring assemblies 834, the overall height and stiffness of the small disc spring assemblies 834 can be directly changed. When the number of small disc spring assemblies 834 changes, the relative position of the mounting base plate 833 within the small disc spring cavity changes accordingly, thus affecting the distance between the mandrel 836 and the valve 91 in the initial state. This mechanism allows the test apparatus to quickly and easily adapt to valves 91 of different lengths, or to simulate the working state of valves 91 under different seating strokes, greatly improving the flexibility and adaptability of the test. It effectively solves the problems of inconvenient adjustment of the distance between the mandrel 836 and the valve 91, and the inability to quickly adapt to different valve 91 sizes or testing requirements in traditional apparatuses. Furthermore, the small disc spring assembly 83 not only receives and transmits the impact force of the large disc spring assembly 82, but also, through its adjustable distance characteristics, makes the simulation of the valve 91 seating stroke more refined and realistic, thus more accurately simulating the complex force state and movement trajectory of the valve 91 in actual engine operation. This combination not only ensures the pressure and frequency applied, but also, through the unique design of the small disc spring assembly 83, achieves precise control of the valve 91 seating distance, making the test results closer to actual working conditions and improving the authenticity and reliability of the test.

[0044] In one specific implementation, the small disc spring assembly 83 further includes an adjusting nut 837. The displacement element 832 has a threaded section, and the adjusting nut 837 is screwed onto the threaded section. The end face of the adjusting nut 837 is provided with a plurality of pushers 838. The pushers 838 pass through the through hole reserved in the displacement element 832 and abut against the mounting base plate 833. By rotating the adjusting nut 837 to change its axial position on the threaded section, the pushers 838 can be driven to push the mounting base plate 833 to move within the small disc spring cavity, thereby adjusting the preload state of the small disc spring assembly 834.

[0045] In this embodiment, when the operator rotates the adjusting nut 837, it engages with the threaded section on the displacement element 832, causing the adjusting nut 837 to move axially along the threaded section. Multiple pushers 838 on the end face of the adjusting nut 837 pass through the pre-drilled holes in the displacement element 832 and directly abut against the mounting base plate 833 as the adjusting nut 837 moves axially. With continued rotation of the adjusting nut 837, the pushers 838 push the mounting base plate 833 to move axially within the small disc spring cavity. The movement of the mounting base plate 833 directly changes the initial compression of the small disc spring assembly 834, thus adjusting its preload. In this way, the initial spring force of the small disc spring assembly 834 can be precisely set, thereby finely adjusting the distance between the mandrel 836 and the valve 91, as well as the initial impact force when the valve 91 sits, effectively solving the inconvenience and inaccuracy problems that may exist in traditional adjustment methods. This adjustment mechanism works in conjunction with the large disc spring assembly 82 and the small disc spring assembly 83 in the pressure application mechanism, enabling the entire pressure application mechanism to more flexibly and accurately simulate loads under different operating conditions such as valve 91 opening, seating, and burst pressure impact. This adjustment method provides significant convenience and accuracy, especially when precise control of the valve 91 seating impact force or valve 91 stroke is required. The design of multiple pushers 838 ensures uniform force distribution on the mounting base plate 833, preventing tilting and thus guaranteeing the stability and consistency of the small disc spring assembly 834 compression. This significantly improves the realism and reliability of the test device in simulating actual operating conditions, making the test results more valuable and providing more accurate and repeatable test conditions for evaluating the high-temperature fatigue wear performance of large-tonnage valves 91.

[0046] In a preferred embodiment, the upper end of the disc spring shaft 835 is provided with a clamping assembly 839 for clamping the top head 836. The clamping assembly 839 includes a connecting plate 8391, a chuck seat 8392, a collet 8393, and a chuck lock nut 8394. The bottom of the connecting plate 8391 is provided with a groove for fitting and mounting with the upper end of the disc spring shaft 835. The chuck seat 8392 is fixed on the connecting plate 8391. The collet 8393 is installed in the center hole of the chuck seat 8392 and the two are fixed together by a bevel engagement. The chuck lock nut 8394 is threadedly connected to the chuck seat 8392, and the chuck lock nut 8394 and the collet 8393 are fixed together by a bevel engagement. The fixed end of the top head 836 is clamped and fixed by the collet 8393.

[0047] In this embodiment, the connecting plate 8391 serves as the base of the clamping assembly 839, with a groove at its bottom for fitting into the upper end of the disc spring shaft 835. This fitting connection effectively prevents relative rotation and axial movement of the connecting plate 8391 relative to the disc spring shaft 835, thereby ensuring the stability and reliability of the force transmission path between the entire clamping assembly 839 and the disc spring shaft 835. The collet seat 8392 is securely fixed to this connecting plate 8391, providing solid support for the subsequent clamping mechanism. The collet 8393 is precisely installed in the center hole of the collet seat 8392, and the two are engaged by a bevel. This bevel engagement is one of the core mechanisms for achieving the clamping function, enabling the collet 8393 to effectively convert axial force into radial clamping force. To further enhance the reliability and adjustability of the clamping, the collet lock nut 8394 is threadedly connected to the collet seat 8392. When the chuck lock nut 8394 is rotated, it moves axially on the chuck seat 8392. Simultaneously, the chuck lock nut 8394 and the collet 8393 are engaged via a bevel. Therefore, the axial movement of the chuck lock nut 8394 acts on the collet 8393 through the bevel, forcing the inner hole of the collet 8393 to contract radially, thereby tightly gripping the fixed end of the mandrel 836. Through this multi-component synergistic structure, the mandrel 836 can be firmly and precisely clamped at the output end of the disc spring shaft 835. The engagement of the connecting plate 8391 and the disc spring shaft 835 ensures overall stability, the chuck seat 8392 provides support, and the collet 8393 and the chuck lock nut 8394, through a double bevel engagement and threaded adjustment, achieve a controllable and powerful clamping force on the mandrel 836. In addition, the design of threaded connection and bevel fit makes the installation and disassembly of the 836 mandrel quick and convenient, greatly simplifies maintenance operations, shortens test interruption time, and improves the overall operating efficiency of the test device.

[0048] In a preferred embodiment, the lifting mechanism includes a drive assembly, a transmission assembly 35, and two worm gear lifts for lifting the loading platform 2. Each worm gear lift includes a first worm wheel and a first worm 31 that mesh with each other, and a lead screw 32 driven by the first worm wheel and capable of vertical movement. The upper end of the lead screw 32 extends movably through the mounting plate at the top of the frame 1, and the lower end of the lead screw 32 is fixedly connected to the loading platform 2. The first worms 31 of the two worm gear lifts are coaxially connected through a coupling 33 to achieve synchronous input. An input gear 34 is fixedly provided at the end of one of the first worms 31. The drive assembly includes a handle 36 rotatably mounted on the frame 1 and a drive gear coaxially fixed on the shaft of the handle 36. The input gear 34 and the drive gear are connected through the transmission assembly 35.

[0049] In this embodiment, when the operator rotates the handle 36 in the drive assembly, the drive gear coaxially fixed on the shaft of the handle 36 rotates accordingly. This drive gear transmits rotational power to the input gear 34 fixed to the end of the first worm 31 of one of the worm gear jacks via the transmission assembly 35. Since the first worms 31 of the two worm gear jacks are coaxially connected via a coupling 33, when one first worm 31 receives input power and begins to rotate, the other first worm 31 is also forced to rotate synchronously by the coupling 33, rotating at the exact same speed and direction. The two synchronously rotating first worms 31 mesh with their corresponding first worm wheels, efficiently converting the rotational motion into the rotation of the first worm wheels. Subsequently, the lead screw 32, driven by the first worm wheels, converts the rotational motion of the first worm wheels into its own vertical linear motion through the action of the threaded pair. Because the upper ends of the two lead screws 32 extend through the mounting plate at the top of the frame 1 and their lower ends are fixedly connected to the loading platform 2, and the movement of the two lead screws 32 is completely synchronized, the loading platform 2 can move vertically smoothly and precisely under the combined action of the two worm gear lifters. This design cleverly utilizes the synchronous transmission characteristics of the coupling 33 and the self-locking and high transmission ratio characteristics of the worm gear mechanism, ensuring that the loading platform 2 remains horizontal during lifting, effectively avoiding the platform tilting or valve 91 position deviation problems caused by asynchronous lifting on both sides in traditional solutions. In this way, the lifting mechanism of this application can provide a highly stable and accurate loading platform 2 for the high-temperature fatigue wear test of valve 91, thereby significantly improving the accuracy of valve 91 seating impact simulation and the reliability of test results.

[0050] In a preferred embodiment, the rotary drive mechanism includes a gear reducer motor 41, a second worm wheel 42 and a second worm 43 that mesh with each other, and a connecting shaft 44. The second worm 43 is driven to the output shaft of the gear reducer motor 41. The second worm wheel 42 is connected to the connecting shaft 44 through a shaft hole fitting structure to drive the connecting shaft 44 to rotate synchronously. The connecting shaft 44 is used to drive the valve 91 to rotate.

[0051] In this embodiment, the geared motor 41 is a power device integrating a motor and a gear reducer. Its main function is to provide stable and controllable rotational power, and to reduce speed and increase torque through the reducer. This motor can be a combination of an AC servo motor and a planetary gear reducer, or a combination of a DC brushless motor and a helical gear reducer, to meet different control precision and load requirements. Using the geared motor 41 as a power source, it provides stable and controllable torque output, ensuring a smooth and efficient driving process. The meshing second worm gear 42 and second worm 43 utilize the characteristics of worm gear transmission to achieve a high reduction ratio and self-locking function, enhancing the accuracy and anti-interference capability of rotational motion. The second worm 43 is driven by the output shaft of the geared motor 41, simplifying the transmission chain, reducing intermediate links, and improving overall efficiency. The second worm gear 42 is connected to the connecting shaft 44 through a shaft-hole mating structure, ensuring synchronous rotation, eliminating the risk of slippage, and guaranteeing motion consistency. The connecting shaft 44 directly drives the valve 91 to rotate, achieving efficient power transmission, enabling the valve 91 to accurately simulate the rotational behavior under actual working conditions during high-temperature wear tests. This structured transmission path ensures that valve 91 can achieve stable, precise and controllable rotational motion in the high-temperature fatigue wear test device, thereby effectively solving the problems of unstable drive, low efficiency and insufficient control precision in traditional solutions.

[0052] In a preferred embodiment, the connecting shaft 44 is provided with an adjustable locking mechanism at one end for connecting the valve 91, for clamping and locking the tail of the valve 91. The adjustable locking mechanism includes an axial center hole 441 provided on the end face of the connecting shaft 44 and a locking element. The locking element can be operated to change the diameter of the axial center hole 441, thereby clamping or releasing the tail of the valve 91.

[0053] In this embodiment, the locking element can be operated to change the diameter of the axial center hole 441, enabling dynamic adjustment of the effective size of the axial center hole 441 to achieve compatible clamping of valve 91 tails with different diameters. For example, radial contraction or expansion can be achieved by screwing in and out, by tapered engagement, or by lever mechanism driving the jaws to open and close. Ultimately, the locking element can grip or release the tail of the valve 91. When the hole diameter is reduced to less than the diameter of the valve 91 tail, the locking element generates radial pressure on the valve 91 tail, achieving gripping; when the hole diameter is increased, the radial pressure is released, achieving release, thereby ensuring that the valve 91 does not loosen or fall off under high-speed rotation and impact loads, while facilitating quick replacement of the valve 91 during testing. By setting an adjustable locking mechanism at one end of the connecting shaft 44 used to connect the valve 91, the adaptability problem of clamping the tail of the valve 91 is cleverly solved. Specifically, when installing valve 91, the tail of valve 91 is first inserted into the axial center hole 441 on the end face of the connecting shaft 44. This hole provides initial positioning and guidance for valve 91. Subsequently, the locking element is operated to change the effective diameter of the axial center hole 441. For example, when the locking element is operated to decrease the diameter, it applies radial pressure to the tail of valve 91 inserted therein, thereby clamping and locking the tail of valve 91. Conversely, when replacing valve 91, the locking element is operated to increase the diameter, which loosens the tail of valve 91, facilitating its removal. This adjustable design allows the linkage shaft 44 to flexibly adapt to the tail of valves 91 of different sizes, ensuring that the valve 91 maintains a stable and reliable connection when rotating under the drive of the rotary drive mechanism (including the gear reduction motor 41, the second worm gear 42 and the second worm 43 and the linkage shaft 44). This avoids slippage or damage caused by insecure clamping, thereby ensuring that the rotary drive mechanism can accurately and effectively transmit the rotational motion to the valve 91, improving the accuracy and reliability of the test.

[0054] Specifically, the locking element described in the figure can be a locking screw. The wall of the axial center hole 441 is provided with at least one axially extending cut 442, thereby forming at least one elastic clamping part. The connecting shaft 44 is provided with a threaded hole 443 that communicates with the axial center hole 441 and is screwed into by the locking screw. By tightening the locking screw, its end can push against the elastic clamping part to cause it to elastically deform and shrink the diameter of the axial center hole 441, thereby achieving the clamping or releasing of the tail of the valve 91.

[0055] In a preferred embodiment, the valve displacement control mechanism includes a spring chamber 51 and a return spring 52 installed in the spring chamber 51. The tail of the valve 91 extends into the spring chamber 51 and is fixedly connected to one end of the connecting shaft 44. The connecting shaft 44 is fitted with an upper thrust ball bearing 53 and a lower thrust ball bearing 54 at one end for connecting the valve 91. The return spring 52 is fitted outside the connecting shaft 44, and the two ends of the return spring 52 abut against the inner end face of the spring chamber 51 and the upper thrust ball bearing 53, respectively.

[0056] In this embodiment, when the pressure applying mechanism applies pressure to the valve 91 to cause it to sit or be impacted, the valve 91 will drive the connecting shaft 44 fixedly connected to it to move upward. At this time, the return spring 52 sleeved on the outside of the connecting shaft 44 is compressed. One end of the return spring 52 abuts against the inner end face of the spring chamber 51, and the other end abuts against the upper thrust ball bearing 53 on the connecting shaft 44. This structure ensures that the return spring 52 is subjected to uniform force during compression, avoiding spring skewing or instability. When the pressure of the pressure applying mechanism is released, the compressed return spring 52 will release its stored elastic potential energy, generating an upward restoring force. This restoring force acts on the connecting shaft 44 through the upper thrust ball bearing 53, thereby driving the valve 91 fixedly connected to the connecting shaft 44 to move upward, so that it accurately returns from the seated state to the initial open position. During this process, the upper thrust ball bearing 53 and lower thrust ball bearing 54 sleeved on the connecting shaft 44 significantly reduce the frictional resistance of the connecting shaft 44 during axial reciprocating motion, ensuring the smoothness and stability of the valve 91's movement. Simultaneously, as the output end of the rotary drive mechanism, the connecting shaft 44 not only controls the axial displacement of the valve 91 but also drives the valve 91 to rotate, allowing the valve 91 to maintain its rotational state during the reset process, more realistically simulating the working conditions of an actual engine. This integrated design enables the valve 91 to maintain high stability and precision in its reset action while withstanding axial impact and rotational motion, effectively solving the problem of valve 91 position deviation caused by unstable reset action in traditional solutions.

[0057] In a preferred embodiment, the heating mechanism includes an induction heater 61 and an electric heating coil 62 connected to the induction heater 61. Specifically, the induction heater 61 is preferably a high-frequency induction heater 61, whose output power is adjustable to adapt to different heating requirements. The electric heating coil 62 is coiled below the valve seat ring 92. The innermost diameter of the electric heating coil 62 is larger than the outer diameter of the valve seat ring 92. An annular heat insulation cover 63 is provided on the outside of the electric heating coil 62. The annular heat insulation cover 63 is fixed on the loading platform 2. A cooling cylinder 64 is installed on the loading platform 2. The interior of the cooling cylinder 64 forms a circulating water cooling cavity 65 and an axial channel. The circulating water cooling cavity 65 is connected to an external water source. The stem of the valve 91 is located in the axial channel.

[0058] In this embodiment, an induction heater 61 drives an electric heating coil 62 to perform non-contact heating of the valve seat 92. The electric heating coil 62 is coiled below the valve seat 92, ensuring that heat acts directly and efficiently on the valve seat 92, avoiding heat loss during the transfer process. Simultaneously, the innermost diameter of the electric heating coil 62 is designed to be larger than the outer diameter of the valve seat 92. This not only prevents overheating or damage that might occur from direct contact between the coil and the seat, but more importantly, it ensures a uniform distribution of the magnetic field in the valve seat 92 region. This allows the valve seat 92 to obtain a uniform temperature field, avoiding the generation of local hot or cold spots, and improving the accuracy of heating and the realism of the simulation. To further reduce heat loss to the environment, an annular insulation cover 63 is placed outside the electric heating coil 62 and firmly fixed to the loading platform 2, forming an effective heat insulation barrier. This concentrates heat around the valve seat 92, thereby maintaining the temperature stability of the valve seat 92 region and improving heating efficiency. Meanwhile, to address the potential overheating issue of the valve stem in high-temperature environments, a cooling cylinder 64 is installed on the loading platform 2. This cooling cylinder 64 features a cleverly designed circulating water cooling cavity 65 and an axial channel. The circulating water cooling cavity 65 is connected to an external water source, allowing the cooling medium (e.g., cooling water) to continuously circulate within the cavity, efficiently absorbing and carrying away heat. The valve stem is precisely located within the axial channel of the cooling cylinder 64, ensuring direct or indirect contact with the cooling area formed by the cooling medium. This design allows the valve stem to receive effective and targeted cooling while the valve seat 92 is heated at high temperatures, preventing deformation or performance degradation due to overheating. Through this ingenious design combining heating and cooling, this device can accurately simulate the high-temperature environment of the engine valve 91 during actual operation, while effectively controlling the temperature of each component of the valve 91, ensuring the accuracy of the test and the long-term reliability of the device. It effectively solves the technical problems of uneven heating, severe heat loss, and the lack of an effective cooling mechanism for the valve stem in traditional heating methods.

[0059] As a preferred implementation, the device also includes a temperature measuring mechanism comprising a thermocouple 71 and an infrared thermometer 72, wherein the head of the thermocouple 71 is configured to be aligned with the stem of the valve 91, and the optical probe of the infrared thermometer 72 is configured to be aligned with the valve seat 92 to measure its operating temperature in a non-contact manner.

[0060] In this embodiment, with the heating mechanism simulating a high-temperature environment by heating the valve seat 92, the addition of a temperature measuring mechanism enables the test device not only to simulate high temperatures but also to acquire real-time and accurate data on the actual operating temperature of key components under high temperatures. Specifically, the head of the thermocouple 71 is aligned with the stem of the valve 91, allowing direct contact measurement of the stem temperature and providing high-precision point temperature data; the optical probe of the infrared thermometer 72 is aligned with the valve seat 92, using a non-contact method to measure the seat's operating temperature, avoiding physical interference or damage to the test components. This combination of thermocouple 71 and infrared thermometer 72 ensures comprehensive and real-time acquisition of temperature information for the valve 91 stem and valve seat 92 during the simulation of a high-temperature environment, providing accurate data support for the test. This precise temperature monitoring capability allows for more refined temperature control during the test, making the test results more convincing and more realistically reflecting the high-temperature fatigue wear of the valve 91 during actual operation.

[0061] In a preferred embodiment, the bottom of the cooling cylinder 64 is provided with a valve clamp 67, the valve clamp 67 is provided with a groove that is adapted to the valve seat 92, the axial channel is provided with a guide tube 66 for wrapping the stem of the valve 91, the upper end face of the valve clamp 67 is provided with a through hole for the stem of the valve 91 to pass through the guide tube 66, and the guide tube 66 abuts against the upper end face of the valve clamp 67 to constrain the guide tube 66 to move downward.

[0062] In this embodiment, a valve clamp 67 is provided at the bottom of the cooling cylinder 64, forming a precise fit and support with the valve seat 92 and the guide 66, thereby ensuring the stable positioning of the valve seat 92 and the axial stability of the guide 66. Specifically, the specially designed groove on the valve clamp 67 has a shape and size that precisely matches the valve seat 92, allowing the valve seat 92 to be firmly embedded and positioned, effectively preventing radial or axial displacement under high temperature and high frequency impact. At the same time, the guide 66 provided in the axial channel is used to guide and protect the reciprocating motion of the valve 91 stem, ensuring the straightness of the valve 91 movement. To prevent the guide 66 from shifting downward under the impact of the valve 91, a through hole is provided on the upper end face of the valve clamp 67. This through hole not only allows the valve 91 stem to pass smoothly through and enter the guide 66, but more importantly, the lower end face of the guide 66 directly abuts against the upper end face of the valve clamp 67. This contact relationship allows the valve clamp 67 to provide reliable axial support for the guide 66, transferring the axial load of the guide 66 to the stable clamp and cylinder block structure. Through this structural combination, the positions of the valve seat 92 and the guide 66 are precisely and stably controlled, ensuring the accuracy and repeatability of the valve 91 during seating and detonation impact simulations. This design, combined with the basic cooling cylinder 64 structure, not only achieves effective cooling of the valve stem of 91 but also solves the stability problem of the valve seat 92 and the guide 66 under impact conditions, significantly improving the reliability of the entire test setup.

[0063] The device of this invention features a simple design. Through the design of eccentric cam assemblies with different phase differences, large and small disc spring assemblies 83, a rotary drive mechanism, a valve displacement control mechanism, a heating mechanism, and a lifting mechanism, it better simulates the wear test of the valve 91-seat pair under different loads, valve 91 speeds, and temperatures. The design of the phase difference and eccentric distance between the eccentric cam assemblies allows for the simulation of actual working conditions such as valve 91 opening, seating, and pressure surge. The large disc spring assembly 82 can control the load force by increasing or decreasing the number of large disc spring assemblies 822, better simulating the high-temperature wear of valve 91 in modern high-performance engines. This avoids the disadvantages of long testing times, high costs, and complex processes associated with whole-engine testing, and enables the evaluation of valve 91 high-temperature wear performance within a very short test cycle, providing a reliable basis for actual production. The small disc spring assembly 83 can adjust the distance from valve 91 opening to seating by increasing or decreasing the number of small disc spring assemblies 834. The rotary drive mechanism can drive the valve 91 to rotate, which more closely matches the actual engine operating conditions and further improves the reliability of the test.

[0064] Although embodiments of the present invention have been shown and described above, it is understood that the above embodiments are exemplary and should not be construed as limiting the present invention. Those skilled in the art can make changes, modifications, substitutions and variations to the above embodiments within the scope of the invention without departing from the principles and spirit of the invention, and all such changes should fall within the protection scope of the claims of the present invention.

Claims

1. A high-tonnage valve high-temperature fatigue wear testing device, characterized in that, include: frame; A loading platform is installed on the frame. The loading platform is provided with a valve, a valve seat that cooperates with the valve head, a rotary drive mechanism for driving the valve to rotate, and a valve displacement control mechanism for driving the valve to reset. A lifting mechanism, mounted on the frame, is used to drive the loading platform to move vertically up and down; The heating mechanism, mounted on the frame, is used to heat the valve seat rings; The pressure-applying mechanism includes a pressure-applying drive assembly, a large disc spring assembly and a small disc spring assembly that sequentially transmit load. The pressure-applying drive assembly includes at least two eccentric cams driven by the same rotary drive shaft and having a phase difference, and at least two power transmission units corresponding one-to-one with the at least two eccentric cams. Each power transmission unit includes a transmission component, a hammer, and a fixed guide structure. The first end of the transmission component is connected to the corresponding eccentric cam, and the second end is hinged to the hammer. The hammer is slidably connected to the guide structure. Each eccentric cam is configured to drive its corresponding hammer through its respective transmission component under the drive of the rotary drive shaft. Each of the hammers is constrained by the guide structure to reciprocate along a set straight path. The large disc spring assembly is configured to receive the force output by the hammer and reciprocate linearly. The load of the large disc spring assembly is adjustable. The input end of the small disc spring assembly is connected to the output end of the large disc spring assembly. The output end of the small disc spring assembly is used to apply pressure to the valve, and the valve distance is adjustable. At least two eccentric cams have different eccentric distances, so that the hammers corresponding to them can apply different loads to the valve through the large disc spring assembly and the small disc spring assembly, thereby simulating at least two different working conditions in valve opening, seating, and burst pressure impact.

2. The high-tonnage valve high-temperature fatigue wear testing device according to claim 1, characterized in that, The at least two eccentric cams include a central eccentric cam and two side eccentric cams, and the at least two transmission units include a central hammer and two side hammers. The eccentric distance of the central eccentric cam is smaller than the eccentric distance of the side eccentric cams. The central eccentric cam drives the disc spring assembly to generate a first impact through the central hammer to simulate the valve seating condition. Two eccentric cams on the sides drive the disc spring assembly through corresponding side hammers to generate a second impact greater than the first impact, in order to simulate the explosive pressure impact condition.

3. The high-tonnage valve high-temperature fatigue wear testing device according to claim 1, characterized in that, The large disc spring assembly includes a first housing slidably connected to the frame, a large disc spring assembly housed inside the first housing, and a hammer head connected to the large disc spring assembly. The first housing is provided with a detachable maintenance panel. By opening the maintenance panel, the number of large disc spring assemblies can be increased, decreased, or replaced to adjust the load of the large disc spring assembly. The head of the hammer head is provided with a pressure sensor, and the hammer head constitutes the output end of the large disc spring assembly.

4. The high-tonnage valve high-temperature fatigue wear testing device according to claim 3, characterized in that, The first housing has a large disc spring cavity for accommodating the large disc spring assembly. A large disc spring cover is installed on the upper side of the large disc spring cavity, and a mounting block and an adjusting block are provided on the lower side. The upper part of the mounting block forms a bearing surface for supporting the large disc spring assembly. The adjusting block is located below the mounting block, and the two are connected by a relative inclined surface. The first housing has an adjusting bolt. The shank of the adjusting bolt passes through the first housing and is threadedly connected to the adjusting block. Rotating the adjusting bolt can drive the adjusting block to move relative to the mounting block, thereby changing the axial position of the mounting block to adjust the preload of the large disc spring assembly.

5. The high-tonnage valve high-temperature fatigue wear testing device according to claim 1, characterized in that, The disc spring assembly includes: The second housing is fixedly mounted on the frame; A displacement element is movably disposed within the second housing. The displacement element is connected to the output end of the large disc spring assembly and can be driven by it to move axially. The displacement element has a small disc spring cavity. The mounting base plate is movably mounted within the small disc spring cavity. The lower end of the small disc spring assembly is placed on the mounting base plate; A disc spring shaft has its lower end inserted into the small disc spring assembly and its lower end is restricted to move within the small disc spring cavity. A top head is installed on the upper end of the disc spring shaft that extends out of the second housing. The top head constitutes the output end of the small disc spring assembly. The distance between the top head and the valve can be adjusted by adjusting the number of small disc spring assemblies.

6. The high-tonnage valve high-temperature fatigue wear testing device according to claim 5, characterized in that, The small disc spring assembly also includes an adjusting nut. The displacement element has a threaded section, and the adjusting nut is screwed onto the threaded section. The end face of the adjusting nut has multiple pushing members. The pushing members pass through the through hole reserved in the displacement element and abut against the mounting base plate. By rotating the adjusting nut to change its axial position on the threaded section, the pushing members can be driven to push the mounting base plate to move within the small disc spring cavity, thereby adjusting the preload state of the small disc spring assembly.

7. The high-temperature fatigue wear testing device for large-tonnage valves according to claim 5, characterized in that, The upper end of the disc spring shaft is provided with a clamping assembly for clamping the top head. The clamping assembly includes a connecting plate, a chuck seat, a collet, and a chuck lock nut. The bottom of the connecting plate is provided with a groove for fitting and installing with the upper end of the disc spring shaft. The chuck seat is fixed on the connecting plate. The collet is installed in the center hole of the chuck seat and the two are fixed together by a bevel engagement. The chuck lock nut is threaded to the chuck seat and the chuck lock nut is fixed together with the collet by a bevel engagement. The fixed end of the top head is clamped and fixed by the collet.

8. The high-tonnage valve high-temperature fatigue wear testing device according to claim 1, characterized in that, The lifting mechanism includes a lifting drive source, a transmission assembly, and two worm gear lifts for lifting the loading platform. Each worm gear lift includes a first worm wheel and a first worm that mesh with each other, and a lead screw driven by the first worm wheel and capable of vertical movement. The upper end of the lead screw movably extends through the mounting plate at the top of the frame, and the lower end of the lead screw is fixedly connected to the loading platform. The first worms of the two worm gear lifts are coaxially connected by a coupling to achieve synchronous input. An input gear is fixedly provided at the end of one of the first worms. The lifting drive source includes a handle rotatably mounted on the frame and a drive gear coaxially fixed to the handle's rotating shaft. The input gear and the drive gear are connected by a transmission assembly.

9. The high-tonnage valve high-temperature fatigue wear testing device according to claim 1, characterized in that, The rotary drive mechanism includes a gear reducer motor, a second worm wheel and a second worm gear that mesh with each other, and a connecting shaft. The second worm gear is driven to the output shaft of the gear reducer motor, and the second worm wheel is connected to the connecting shaft through a shaft hole fitting structure to drive the connecting shaft to rotate synchronously. The connecting shaft is used to drive the valve to rotate.

10. The high-tonnage valve high-temperature fatigue wear testing device according to claim 9, characterized in that, The connecting shaft is provided with an adjustable locking mechanism at one end for connecting the valve, which is used to clamp and lock the valve tail. The adjustable locking mechanism includes an axial center hole on the end face of the connecting shaft and a locking element. The locking element can be operated to change the diameter of the axial center hole, thereby clamping or releasing the valve tail.

11. The high-tonnage valve high-temperature fatigue wear testing device according to claim 9 or 10, characterized in that, The valve displacement control mechanism includes a spring chamber and a return spring installed in the spring chamber. The tail of the valve extends into the spring chamber and is fixedly connected to one end of the linkage shaft. The end of the linkage shaft used to connect the valve is fitted with an upper thrust ball bearing and a lower thrust ball bearing. The return spring is fitted outside the linkage shaft, and the two ends of the return spring abut against the inner end face of the spring chamber and the upper thrust ball bearing, respectively.

12. The high-tonnage valve high-temperature fatigue wear testing device according to claim 1, characterized in that, The heating mechanism includes an induction heater and an electric heating coil connected to the induction heater. The electric heating coil is coiled below the valve seat. The innermost diameter of the electric heating coil is larger than the outer diameter of the valve seat. An annular heat insulation cover is provided on the outside of the electric heating coil. The annular heat insulation cover is fixed on the loading platform. A cooling cylinder is installed on the loading platform. The interior of the cooling cylinder forms a circulating water cooling cavity and an axial channel. The circulating water cooling cavity is connected to an external water source. The valve stem is located in the axial channel.

13. The high-tonnage valve high-temperature fatigue wear testing device according to claim 1, characterized in that, It also includes a temperature measuring mechanism, which comprises a thermocouple and an infrared thermometer. The head of the thermocouple is configured to be aligned with the stem of the valve, and the optical probe of the infrared thermometer is configured to be aligned with the valve seat to measure its operating temperature in a non-contact manner.

14. The high-tonnage valve high-temperature fatigue wear testing device according to claim 12, characterized in that, The bottom of the cooling cylinder is provided with a valve clamp, which has a groove that matches the valve seat. The axial channel is provided with a guide tube for wrapping the valve stem. The upper end face of the valve clamp is provided with a through hole for the valve stem to pass through the guide tube. The guide tube abuts against the upper end face of the valve clamp to constrain the guide tube from moving downward.

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