A pressure relay and hydraulic control system for vehicle overload protection

Abstract

The application relates to the technical field of relays, in particular to a pressure relay and a hydraulic control system for vehicle overload protection, wherein the pressure relay comprises a spring assembly which comprises coaxially nested first and second springs, the axial two ends of the first and second springs are respectively in contact with a first pressure bearing surface and a second pressure bearing surface, and the first and second pressure bearing surfaces are oppositely arranged; the first and second springs have equal effective turns and pitches, so that the first and second springs are synchronously axially compressed and elastically deformed between the first and second pressure bearing surfaces. The problem of trigger precision reduction caused by temperature drift of a traditional pressure relay is solved.

Description

Technical Field

[0001] This invention relates to the field of relay technology, and in particular to a pressure relay and a hydraulic control system for vehicle overload protection. Background Technology

[0002] Currently, vehicles play a crucial role in airport ground support, primarily used to lift supplies such as food and beverages to hatches. To ensure safety during the lifting process, vehicles typically have pressure relays installed in their hydraulic systems to monitor hydraulic circuit pressure. When the load exceeds a preset threshold, the increased hydraulic pressure triggers the pressure relay, outputting an electrical signal to the controller to cut off lifting power, thus achieving overload protection.

[0003] However, existing pressure relays have limitations in adaptability to specific environments. Firstly, vehicle operating environments experience extreme temperature variations, such as loading food in cold storage and unloading it outdoors, typically switching between -40°C and 80°C. The internal spring shear modulus of ordinary pressure relays changes drastically with temperature, causing overload trigger pressure drift and affecting detection accuracy and safety. Secondly, at the moment of container start-up or due to pulse impacts caused by uneven ground, the hydraulic system pressure generates instantaneous spikes. Existing technology is prone to false alarms and frequent system shutdowns, impacting operational efficiency. Therefore, there is an urgent need for a pressure relay and control system that can adapt to the vehicle's operating environment and offers better reliability. Summary of the Invention

[0004] In this section, as well as in the abstract and title of this application, some simplifications or omissions may be made to avoid obscuring the purpose of this section, the abstract, and the title of this application, and such simplifications or omissions shall not be used to limit the scope of the invention.

[0005] To address the problem of poor adaptability of pressure relays in the prior art to the vehicle's operating environment, one objective of this invention is to provide a pressure relay comprising a spring assembly, which includes a first spring and a second spring coaxially nested together. The axial ends of the first spring and the second spring respectively abut against a first pressure-bearing surface and a second pressure-bearing surface, and the first pressure-bearing surface and the second pressure-bearing surface are arranged facing each other. The first spring and the second spring have equal effective coils and pitch, so that the first spring and the second spring generate synchronous axial compressive elastic deformation between the first pressure-bearing surface and the second pressure-bearing surface.

[0006] In a preferred embodiment of the pressure relay of the present invention, the first spring and the second spring have opposite helical directions, and the helix angles of the first spring and the second spring range from 3° to 6°; the stiffness of the first spring... Stiffness of the second spring The inner diameter D1 of the first spring and the outer diameter D2 of the second spring satisfy the following condition: 0.95≤(K1·D1) / (K2·D2)≤1.15.

[0007] In a preferred embodiment of the pressure relay of the present invention, the first spring is a spring steel with a negative elastic modulus temperature coefficient, whose elastic modulus decreases as the temperature increases; the second spring is a constant elastic alloy with a positive elastic modulus temperature coefficient, whose elastic modulus increases as the temperature increases; the thermoelastic coefficients of the first spring and the second spring have opposite signs to compensate for the combined stiffness of the spring assembly under different ambient temperatures. Furthermore, the equivalent thermoelastic coefficient of the spring assembly satisfy: , , in, These are the stiffnesses of the first spring and the second spring, respectively. These are the thermoelastic coefficients of the first spring and the second spring, respectively.

[0008] In a preferred embodiment of the pressure relay described in this invention, the stiffness of the first spring is: Stiffness of the second spring satisfy: .

[0009] In a preferred embodiment of the pressure relay of the present invention, the inner diameter D1 of the first spring and the outer diameter D2 of the second spring satisfy the condition: 1.02≤D1 / D2≤1.04.

[0010] As a preferred embodiment of the pressure relay of the present invention, it further includes: a housing having an inner cavity; a pressure adjusting knob disposed on the housing; a plunger slidably mounted in the inner cavity; a switch contact disposed in the inner cavity and located on the sliding path of the plunger, the plunger being adapted to contact the switch contact; the opposing sides of the pressure adjusting knob and the plunger are respectively the first pressure-bearing surface and the second pressure-bearing surface; the two ends of the first spring and the second spring respectively abut against the pressure adjusting knob and the plunger.

[0011] As a preferred embodiment of the pressure relay of the present invention, it further includes an oil inlet connected to the inner cavity, the oil inlet being used to deliver hydraulic oil into the inner cavity.

[0012] Another objective of this invention is to provide a hydraulic control system for vehicle overload protection, comprising a hydraulic pump station, a control unit, a hydraulic cylinder for driving the lifting and lowering of the vehicle body, and the aforementioned pressure relay; the rodless chamber of the hydraulic cylinder is connected to the output end of the hydraulic pump station via a hydraulic pipeline; the pressure relay is connected to the oil inlet path of the rodless chamber; the switch contact of the pressure relay is electrically connected to the control unit, and the control unit controls the hydraulic cylinder to perform lifting locking or alarm actions according to the signal state of the switch contact.

[0013] As a preferred embodiment of the hydraulic control system for vehicle overload protection described in this invention, it further includes a lifting control valve group disposed between the hydraulic pump station and the hydraulic cylinder, wherein the control unit is electrically connected to the electromagnetic control terminal of the lifting control valve group; when the hydraulic pressure sensed by the pressure relay reaches a preset overload threshold, the control unit controls the lifting control valve group to switch to the neutral lock-up state, so that the hydraulic cylinder stops rising while maintaining the current height.

[0014] As a preferred embodiment of the hydraulic control system for vehicle overload protection described in this invention, the control unit is equipped with a filtering and timing module; the control unit determines a true overload and triggers protection logic only when the duration of the overload electrical signal received from the pressure relay exceeds the anti-shake duration threshold Δt; wherein the anti-shake duration threshold Δt is set to 0.5s~2s to filter the instantaneous pulse pressure generated by the vehicle body at the moment of startup or under bumpy conditions.

[0015] The beneficial effects of this invention are as follows: This invention achieves synchronous compression deformation of the two springs by matching the effective coils and pitch of the coaxially nested first and second springs with equal values. Combined with a design of thermoelastic coefficients with opposite signs, the temperature drift of the first spring's stiffness is compensated by the temperature characteristics of the second spring's elastic modulus, thus controlling the equivalent thermoelastic coefficient of the spring assembly within a certain range. Within the full temperature range of -40℃ to 80℃, the temperature drift rate of the combined stiffness of the spring assembly is less than 1.5%, solving the problem of decreased triggering accuracy caused by temperature drift in traditional pressure relays and ensuring the stability of the overload protection threshold under extreme temperature difference environments. Attached Figure Description

[0016] To more clearly illustrate the technical solutions of the embodiments of the present invention, the accompanying drawings used in the description of the embodiments will be briefly introduced below. Obviously, the accompanying drawings described below are only some embodiments of the present invention. For those skilled in the art, other drawings can be obtained based on these drawings without creative effort.

[0017] Figure 1This is a schematic diagram of the spring assembly of the present invention.

[0018] Figure 2 This is a schematic diagram of the pressure relay of the present invention.

[0019] Figure 3 This is a cross-sectional schematic diagram of the pressure relay of the present invention.

[0020] Figure 4 This is a schematic diagram of the hydraulic control system for vehicle overload protection according to the present invention.

[0021] Figure 5 This is a graph showing the numerical relationship between the helix angle and the cancellation rate of the present invention.

[0022] Figure 6 This is a graph showing the numerical relationship between the helix angle and the absolute value of the linearity error of the pressure in this invention.

[0023] Figure 7 This is a graph showing the numerical relationship between the helix angle, radial sway, and minimum gap between the rings under ultimate compression in this invention.

[0024] Figure 8 This is a graph showing the numerical relationship between the helix angle and the number of fatigue life cycles of the present invention.

[0025] In the diagram: 101, spring assembly; 101a, first spring; 101b, second spring; S1, first pressure bearing surface; S2, second pressure bearing surface; 102, housing; 102a, inner cavity; 103, pressure adjustment knob; 104, plunger; 105, switch contact; 106, oil inlet; 200, hydraulic pump station; 300, control unit; 301, filter timing module; 400, hydraulic cylinder; 401, rodless chamber; 500, lifting control valve assembly; 501, electromagnetic control terminal. Detailed Implementation

[0026] To make the objectives, features, and advantages of the present invention more apparent and understandable, the specific embodiments of the present invention will be described in detail below with reference to the accompanying drawings.

[0027] Many specific details are set forth in the following description in order to provide a full understanding of the invention. However, the invention may also be practiced in other ways different from those described herein, and those skilled in the art can make similar extensions without departing from the spirit of the invention. Therefore, the invention is not limited to the specific embodiments disclosed below.

[0028] Secondly, the term "one embodiment" or "embodiment" as used herein refers to a specific feature, structure, or characteristic that may be included in at least one implementation of the present invention. The phrase "in one embodiment" appearing in different places in this specification does not necessarily refer to the same embodiment, nor is it a single or selective embodiment that is mutually exclusive with other embodiments.

[0029] Example 1, see Figure 1 This is the first embodiment of the present invention. This embodiment provides a pressure relay, which includes a spring assembly 101.

[0030] Specifically, the spring assembly 101 includes a first spring 101a and a second spring 101b coaxially nested together. The first spring 101a is the outer coil spring, and the second spring 101b is the inner coil spring, with their central axes coinciding. Both ends of the first spring 101a and the second spring 101b abut against the first bearing surface S1 and the second bearing surface S2, respectively. The first bearing surface S1 and the second bearing surface S2 are two parallel and facing planes. The central axes of the first bearing surface S1 and the second bearing surface S2 coincide with the central axes of the first spring 101a and the second spring 101b, ensuring that the first spring 101a and the second spring 101b experience uniform force and no radial off-center load during compression.

[0031] The first spring 101a and the second spring 101b have equal effective coils n and pitch t, ​​that is, the axial center distance between two adjacent effective coils n of the first spring 101a is equal to the axial center distance between two adjacent effective coils n of the second spring 101b. It is worth noting that a spring has effective coils n and support coils. The effective coils n refer to the coils that expand, deform, and exert elastic force when subjected to force. The support coils refer to the coils that are ground flat, tightly attached, and do not deform at both ends of the spring. They only serve to support, level, and bear pressure for positioning and are not counted as effective coils n. The support coils are generally 1 to 2 coils. This design ensures that when the first spring 101a and the second spring 101b move towards each other on the first bearing surface S1 and the second bearing surface S2, they generate synchronous axial compressive elastic deformation. The first spring 101a and the second spring 101b always maintain synchronous extension and contraction throughout the entire compression stroke, without relative axial displacement. This ensures that the load distribution between the first spring 101a and the second spring 101b in the dual-spring design always maintains the preset ratio, and there will be no overload of a single spring in the first spring 101a and the second spring 101b.

[0032] By using a coaxial nested, equal-pitch double spring design with a first spring 101a and a second spring 101b, the first spring 101a and the second spring 101b are compressed and deformed synchronously, ensuring the stability of the load distribution of the spring assembly 101. This provides a structural basis for subsequent temperature compensation and torsional cancellation designs. At the same time, the parallel double spring structure improves the load-bearing capacity and fatigue life of the spring assembly 101. Compared with a single spring structure, the stress per coil under the same load is lower, making it more suitable for the high-frequency reciprocating operation requirements of vehicles.

[0033] Example 2, see Figures 1 to 8 This is the second embodiment of the present invention, which is based on embodiment 1.

[0034] Specifically, the spiral directions of the first spring 101a and the second spring 101b are opposite. In this embodiment, the coil of the first spring 101a is a right-hand spiral, and the coil of the second spring 101b is a left-hand spiral. Simultaneously, the helix angle of both the first spring 101a and the second spring 101b ranges from 3° to 6°. In this embodiment, 4.5° is preferred. Under axial compression, the circumferential torsional torque generated by the first spring 101a and the circumferential torsional torque generated by the second spring 101b are opposite in direction and basically cancel each other out. According to the mechanical principle of cylindrical helical springs, the circumferential torsional torque M generated by the cylindrical helical spring under axial compression is positively correlated with the spring's stiffness K, axial compression displacement Δx, fitting diameter D (for the first spring 101a, it refers to its inner diameter D1; for the second spring 101b, it refers to its outer diameter D2), and the angle function value related to the helix angle α. Since the first spring 101a and the second spring 101b are coaxially nested, their axial ends respectively abut against the same set of opposing first bearing surfaces S1 and second bearing surfaces S2, and their effective number of turns n and pitch t are equal, therefore... The axial compression displacement Δx is the same; and the helix angles of the first spring 101a and the second spring 101b are the same. Therefore, the ratio of the circumferential torsional torque generated by the first spring 101a and the second spring 101b is equivalent to the ratio of the product of their stiffness and the mating diameter. If the circumferential torsional torque M1 generated by the first spring 101a and the circumferential torsional torque M2 generated by the second spring 101b are to be approximately equal in magnitude and basically cancel each other out, i.e., M1≈M2, then it is necessary to satisfy K1·D1≈K2·D2, i.e. (K1·D1) / (K2·D2)≈1, where K1 is the stiffness of the first spring 101a, D1 is the inner diameter of the first spring 101a, K2 is the stiffness of the second spring 101b, and D2 is the outer diameter of the second spring 101b.

[0035] In this embodiment, the stiffness K1 and inner diameter D1 of the first spring 101a and the stiffness K2 and outer diameter D2 of the second spring 101b satisfy the following condition: 0.95 ≤ (K1·D1) / (K2·D2) ≤ 1.15. When the above condition is met, the circumferential torsional torque cancellation rate of the dual springs can reach more than 95%, which can effectively eliminate the circumferential torsional stress generated when the spring assembly 101 is axially compressed, and avoid problems such as coil interference, jamming, and radial runout.

[0036] The experiments were conducted as follows: Baseline operating conditions: ambient temperature 25℃, axial rated compressive load 1000N (corresponding to the vehicle's rated load), compression stroke 10.5mm; Extreme operating conditions: axial overload load 1200N (120% of the rated load), compression stroke 12.6mm; Wide temperature range operating conditions: -40℃ cold storage low temperature, 80℃ apron high temperature full temperature range coverage; Vibration operating conditions: airport ground service environment 10g acceleration, 10~500Hz random vibration, simulating a strong vibration working environment. The test indicators are as follows: Circumferential torsional torque cancellation rate: the closer the value is to 100%, the better the torsional cancellation effect of the dual springs and the more thorough the torsional elimination; Pressure linearity error: corresponds to the pressure monitoring accuracy. The smaller the value, the better the linear correspondence between the spring compression and the load, and the higher the triggering accuracy of the pressure relay; Radial runout under rated load: corresponds to anti-runout and anti-interference performance. The smaller the value, the better the running stability of the dual springs and the less likely the coils will collide or jam; Minimum coil gap under ultimate compression: corresponds to the anti-coil-by-coil capability. The larger the value, the less likely the coils will collide or wear during the compression process of the dual springs; Estimated fatigue life cycles: corresponds to the operational reliability under strong vibration conditions. The larger the value, the longer the fatigue life of the dual springs. Table 1 is a comparison table of the core performance of spring assembly 101 under different helix angles.

[0037] Table 1:

[0038] The experimental conclusions are as follows: Circumferential torsional moment cancellation rate: Increases first and then decreases with increasing helix angle, reaching 99.7% at 4.5°, close to the theoretically complete cancellation state; after deviating from 4.5°, the cancellation rate drops rapidly, falling below 97% below 4.0° or above 5.0°, failing to achieve a stable torsional cancellation effect. Pressure linearity error: At 4.5°, the error is only ±0.32%, the smallest among the 12 data sets, indicating optimal pressure monitoring linearity and highest accuracy; excessively small or large helix angles will cause nonlinear distortion in the compression of spring assembly 101, significantly increasing the error. Radial runout: At 4.5°, the radial runout is only 0.041mm, providing the best anti-runout and anti-interference effect; after deviating from the optimal value, the runout gradually increases, easily causing coil collisions and jamming. Minimum coil clearance at ultimate compression: At 4.5°, the clearance reaches 0.52mm, the largest in the group, ensuring no spring coil merging or wear even under 120% overload limit conditions, demonstrating optimal anti-merging performance. Fatigue life estimated cycles: At 4.5°, the fatigue life reaches 2.12 million cycles, the highest value among the 12 sets of data; due to the complete elimination of torsional additional stress, the spring only bears pure shear stress, resulting in the lowest fatigue damage rate and the longest service life. 4.4°~4.6° is the optimal helix angle range for spring assembly 101, with 4.5° being the core preferred value; when the helix angle is too small (<4.0°), the pitch t is too small, the circumferential torsional displacement distortion of the coil is aggravated, and the torsional torque matching degree decreases; when the helix angle is too large (>5.0°), the axial component of the spring coil accounts for too high a proportion, the torsional torque is nonlinearly distorted, and the torques with opposite directions of rotation cannot be matched and canceled; 4.5° can simultaneously meet the requirements of torsional cancellation, higher monitoring accuracy, better operational stability, and longer fatigue life, and is suitable for the wide temperature range, strong vibration, and high reliability operating conditions of vehicles. It is worth noting that an excessively small helix angle will cause the spring assembly 101 to generate additional nonlinear circumferential displacement during compression, exacerbating the torsional torque mismatch; conversely, an excessively large helix angle will increase the proportion of the axial force, leading to a higher sensitivity of the torsional torque to manufacturing errors, thus also making precise matching impossible. Experiments revealed an optimal solution within the 3°~6° range, especially around 4.5°, where the combined impact of the two non-ideal effects is minimized, thereby achieving near-cancellation of torque.

[0039] Preferably, the thermoelastic coefficient of the first spring 101a and the thermoelastic coefficient of the second spring 101b have opposite signs (opposite signs mean that one thermoelastic coefficient is positive and the other is negative); wherein the first spring 101a is made of spring steel with a negative temperature coefficient of elastic modulus, and in this embodiment, it is preferably 60Si2MnA spring steel, whose elastic modulus decreases with increasing temperature and whose thermoelastic coefficient is negative. The second spring 101b is made of a constant elastic alloy with a positive elastic modulus and a constant temperature coefficient. In this embodiment, it is preferably a Ni42CrTi constant elastic alloy, whose elastic modulus increases with increasing temperature and whose thermoelastic coefficient is... And the thermoelastic coefficient The selectable range is This is used to compensate for the combined stiffness of the spring assembly 101 under different ambient temperatures. Furthermore, the equivalent thermoelastic coefficient of the spring assembly 101... satisfy: The formula for calculating the equivalent thermoelastic coefficient is as follows: In the formula, The stiffnesses of the first spring 101a and the second spring 101b are respectively. The thermoelastic coefficients of the first spring 101a and the second spring 101b are respectively. In this embodiment, through the above parameter matching, the combined stiffness temperature drift rate of the spring assembly 101 is less than 1.5% within the ambient temperature range of -40℃ to 80℃. By matching the opposite thermoelastic coefficients of the first spring 101a and the second spring 101b, stiffness compensation of the spring assembly 101 is achieved across the entire temperature range (-40℃ to 80℃), keeping the equivalent thermoelastic coefficient within a low range. This solves the problem of pressure trigger threshold deviation caused by spring stiffness drift under extreme temperature differences, ensuring that the overload protection accuracy remains within aviation-grade standards when the vehicle frequently switches between cold storage and apron operations.

[0040] Preferably, the stiffness of the first spring 101a Stiffness of the second spring 101b satisfy: In this embodiment, the first spring 101a is the main spring that bears the rated preload force. The stiffness of this embodiment is... The value is 52 N / mm. The second spring 101b is a temperature-compensated auxiliary spring. The stiffness of this embodiment is... The stiffness of the first spring 101a is 48 N / mm. Stiffness of the second spring 101b The stiffness ratio is 1.08:1. This specific stiffness ratio ensures that the first spring 101a, acting as the main spring, bears the majority of the load, achieving stable calibration of the pressure threshold. Simultaneously, it ensures that the temperature compensation effect of the second spring 101b, acting as the auxiliary spring, can cover the stiffness temperature drift of the first spring 101a, while avoiding the influence of excessive stiffness of the second spring 101b on the pressure calibration of the first spring 101a. This achieves an optimal balance between load-bearing capacity and temperature compensation effect. Table 2 shows the comparison table of core design parameters for the dual-spring system.

[0041] Table 2:

[0042] The experimental environment was as follows: Temperature coverage: -40℃ (low temperature in cold storage), 0℃, 25℃ (baseline ambient temperature), 50℃, 80℃ (high temperature on the apron), with a total temperature difference of 120℃, covering the entire operating temperature range of the equipment; Load conditions: rated compressive load of 1000N (corresponding to the rated load of the vehicle), and 1200N overload limit condition; Blank control group: existing general-purpose single-spring structure (with the same parameters and materials as the main spring, i.e., the first spring 101a, without a temperature-compensated auxiliary spring). Test indicators were as follows: Temperature drift rate of the combined stiffness across the entire temperature range of -40℃ to 80℃: must meet the requirement of <1.5%, as shown in Table 3. Table 3 shows the different... A table comparing the temperature drift of the stiffness of spring assembly 101 across the entire temperature range.

[0043] Table 3:

[0044] Experimental data proves that The optimal value for overall performance is achieved, with the best temperature drift control: the maximum stiffness temperature drift change rate across the entire temperature range is only 0.97%, far superior to <1.5%. Compared with the existing single-spring structure, the temperature drift is reduced by 67.9%, solving the pressure drift problem under extreme temperature differences.

[0045] Preferably, the inner diameter D1 of the first spring 101a and the outer diameter D2 of the second spring 101b satisfy the condition: 1.02 ≤ D1 / D2 ≤ 1.04. This ensures that there is no radial interference between the first spring 101a and the second spring 101b during compression, and also avoids radial runout caused by excessive gap. Through a specific inner and outer diameter ratio design, within the full temperature range of thermal expansion and contraction and compression stroke, there will be no interference or jamming caused by high-temperature expansion, nor radial runout or wobbling caused by low-temperature contraction, further improving the operational stability of the spring assembly 101.

[0046] Furthermore, the pressure relay also includes a housing 102, a pressure adjusting knob 103, a plunger 104, and a switch contact 105. The housing 102 serves as the mounting base for the pressure relay and contains a sealed inner cavity 102a. The pressure adjusting knob 103 is threaded onto the right end of the housing 102 and can be rotated to adjust the axial feed. The plunger 104 is slidably mounted in the inner cavity 102a via a guide engagement and can slide left and right axially. The switch contact 105 is fixedly mounted on the left end of the inner cavity 102a and is located on the sliding path of the plunger 104. When the plunger 104 slides to the left to a preset stroke, it contacts the switch contact 105. When the plunger 104 is subjected to hydraulic pressure and moves to the right, it disengages from the switch contact 105. The left end face of the pressure adjustment knob 103 is the first pressure bearing surface S1, and the right end face of the plunger 104 is the second pressure bearing surface S2. The two ends of the first spring 101a and the second spring 101b respectively abut against the left end face of the pressure adjustment knob 103 and the right end face of the plunger 104 to achieve axial positioning and load transmission. The pressure adjustment knob 103 is used to adjust the preload of the first spring 101a and the second spring 101b, i.e., to calibrate the overload trigger threshold. The plunger 104 converts the hydraulic pressure into axial thrust, realizing the conversion of hydraulic pressure into the compression of the first spring 101a and the second spring 101b. Finally, the switch contact 105 outputs an electrical signal, forming a complete pressure sensing and triggering circuit.

[0047] Furthermore, the pressure relay also includes an oil inlet 106, which is located at the lower end of the housing 102 and communicates with the inner cavity 102a. The oil inlet 106 is used to connect to the hydraulic pipeline and deliver hydraulic oil into the inner cavity 102a, so that the hydraulic pressure acts directly on the pressure-bearing surface of the plunger component 104. The oil inlet enables the real-time introduction of hydraulic system pressure, allowing the plunger component 104 to directly sense the real-time pressure of the hydraulic circuit.

[0048] Example 3, see Figures 1 to 8 This embodiment corresponds to a hydraulic control system for vehicle overload protection. Specifically, it includes a hydraulic pump station 200, a control unit 300, a hydraulic cylinder 400 for driving the vehicle body to rise and fall, and a pressure relay as described in Embodiments 1 and 2. In this embodiment, the hydraulic cylinder 400 is a single-acting plunger hydraulic cylinder, which is located at the bottom of the vehicle body and is used to drive the body to rise and fall.

[0049] Specifically, the rodless chamber 401 of the hydraulic cylinder 400 is connected to the output end of the hydraulic pump station 200 via a hydraulic pipeline. The hydraulic pump station 200 provides high-pressure hydraulic oil to the lifting system, driving the hydraulic cylinder 400 to extend and raise the housing. The oil inlet 106 of the pressure relay is connected to the oil inlet path of the rodless chamber 401 of the hydraulic cylinder 400 via a pressure measuring pipeline, monitoring the hydraulic pressure of the rodless chamber of the hydraulic cylinder 400 in real time. This hydraulic pressure is positively correlated with the housing load. The switch contact 105 of the pressure relay is electrically connected to the control unit 300 via a control cable. The control unit 300 is the vehicle's PLC controller, configured to control the hydraulic cylinder 400 to perform a lifting and locking action or trigger an audible and visual alarm based on the signal status of the switch contact 105. By monitoring the hydraulic pressure corresponding to the housing load in real time through the pressure relay, accurate identification of overload conditions is achieved. At the same time, the control unit 300 implements protective actions to form overload protection, avoiding the safety risks of overloaded operation of the housing.

[0050] Preferably, the hydraulic control system further includes a lifting control valve assembly 500. The lifting control valve assembly 500 is installed on the hydraulic pipeline between the hydraulic pump station 200 and the hydraulic cylinder 400. In this embodiment, the lifting control valve assembly 500 is a three-position four-way solenoid directional valve with an O-type neutral position function. Its P port is connected to the output end of the hydraulic pump station 200, its A port is connected to the rodless chamber 401 of the hydraulic cylinder 400, and its T port is connected to the oil tank. The control unit 300 is electrically connected to the solenoid control terminal 501 of the lifting control valve assembly 500 to control the position switching of the directional valve. When the hydraulic pressure sensed by the pressure relay reaches the preset overload threshold, the switch contact 105 switches its state and sends an overload electrical signal to the control unit 300. The control unit 300 immediately outputs a control signal to de-energize the lifting control valve assembly 500 and switch it to the neutral position lock-up state. The O-type neutral position function can completely close the rodless chamber 401 of the hydraulic cylinder 400, so that the hydraulic cylinder 400 maintains its current height and stops its upward movement, thereby achieving overload lock-up protection. The lifting control valve group with O-type center position function realizes center position locking protection in case of overload. Compared with the traditional unloading protection logic, it not only stops the overload operation of the box from continuing to rise, but also prevents the box from falling, which meets the usage requirements of ground equipment.

[0051] Furthermore, the control unit 300 integrates a filtering timing module 301, which is a timer function module within the PLC. The control unit 300 is configured to determine a true overload and trigger the protection logic only when the duration of the overload electrical signal received from the pressure relay exceeds a preset anti-shake duration threshold Δt. The anti-shake duration threshold Δt ranges from 0.5s to 2s, and in this embodiment, 1s is preferred. This duration effectively filters instantaneous pulse pressure generated at the moment of vehicle body startup and under bumpy conditions on airport roads, preventing false triggering. Through the anti-shake design of the filtering timing module 301, the continuous pressure signal of a true overload is effectively distinguished from the instantaneous pressure spikes generated by bumps and impacts, solving the problems of frequent false alarms and abnormal shutdowns in traditional systems. While ensuring the effectiveness of the overload protection function, this improves the continuity of vehicle operations and the efficiency of flight support.

[0052] To facilitate understanding of the technical solution of this invention, its working principle is explained in detail below: By rotating the pressure adjustment knob 103 axially, the initial preload of the spring assembly 101 is adjusted, so that the plunger 104 moves to the right and triggers the switching state of the switch contact 105. The pressure is consistent with the preset overload threshold. After calibration, the pressure adjustment knob 103 is locked to complete the threshold setting. Under normal circumstances, the plunger 104 is in contact with the switch contact 105. However, under the action of hydraulic pressure, the plunger 104 will be released from the switch contact 105, causing the switch contact 105 to switch states.

[0053] During normal operation, the vehicle body load is within the rated range, the hydraulic pressure of the rodless chamber 401 of the hydraulic cylinder 400 is lower than the preset overload threshold, the hydraulic thrust cannot overcome the preload of the spring assembly 101, the plunger 104 remains in the initial position, the switch contact 105 is in the initial state, the control unit 300 does not receive an overload signal, the lifting control valve group 500 can switch positions normally, and the body can be lifted and lowered normally.

[0054] During the overload protection phase, when the load on the housing exceeds the rated value, the hydraulic pressure in the rodless chamber 401 of the hydraulic cylinder 400 reaches the preset overload threshold. The hydraulic thrust pushes the plunger 104 to slide to the right, compressing the spring assembly 101 and causing the switch contact 105 to switch states, sending an overload electrical signal to the control unit 300. The filter timing module 301 verifies the signal for anti-shake. When the duration of the signal exceeds the anti-shake duration threshold Δt, the control unit 300 determines it as a real overload, controls the lifting control valve group 500 to switch to the neutral lock state, keeps the hydraulic cylinder 400 at the current height and stops rising, and triggers an audible and visual alarm, completing the overload protection.

[0055] During the reset phase, when the overload is removed, the hydraulic pressure in the rodless chamber of the hydraulic cylinder 400 drops below the reset pressure. The spring assembly 101 pushes the plunger 104 to reset to the left, the switch contact 105 returns to its initial state, and after the control unit 300 receives the reset signal, it releases the lockout protection, and the lifting system resumes normal operation.

[0056] It should be noted that the above embodiments are only used to illustrate the technical solutions of the present invention and are not intended to limit it. Although the present invention has been described in detail with reference to preferred embodiments, those skilled in the art should understand that modifications or equivalent substitutions can be made to the technical solutions of the present invention without departing from the spirit and scope of the technical solutions of the present invention, and all such modifications or substitutions should be covered within the scope of the claims of the present invention.

Claims

1. A pressure relay characterized by: include, The spring assembly (101) includes a first spring (101a) and a second spring (101b) coaxially nested together. The axial ends of the first spring (101a) and the second spring (101b) respectively abut against a first bearing surface (S1) and a second bearing surface (S2). The first bearing surface (S1) and the second bearing surface (S2) are arranged facing each other. The first spring (101a) and the second spring (101b) have equal effective coils (n) and pitch (t) so that the first spring (101a) and the second spring (101b) generate synchronous axial compressive elastic deformation between the first bearing surface (S1) and the second bearing surface (S2); The helical line of the first spring (101a) and the helical line of the second spring (101b) are opposite in rotation direction, the helical pitch of the first spring (101a) and the helical pitch of the second spring (101b) are in the range of 3°-6°, the stiffness of the first spring (101a) is opposite to the stiffness of the second spring (101b) , the inner diameter D1 of the first spring (101a) and the outer diameter D2 of the second spring (101b) satisfy: 0.95≤(K1·D1) / (K2·D2)≤1.

15. The first spring (101a) is made of spring steel with a negative elastic modulus temperature coefficient, and its elastic modulus decreases as the temperature increases; the second spring (101b) is made of a constant elastic alloy with a positive elastic modulus temperature coefficient, and its elastic modulus increases as the temperature increases. The thermoelastic coefficient of the first spring (101a) and the thermoelastic coefficient of the second spring (101b) have opposite signs to compensate for the combined stiffness of the spring assembly (101) under different ambient temperatures. Furthermore, the equivalent thermoelastic coefficient of the spring assembly (101) satisfy: ， ， in, , These are the stiffnesses of the first spring (101a) and the second spring (101b), respectively. , The thermoelastic coefficients are the first spring (101a) and the second spring (101b), respectively.

2. The pressure relay as described in claim 1, characterized in that: Stiffness of the first spring (101a) Stiffness of the second spring (101b) Satisfy: 0.95≤ / ≤1.

1.

3. The pressure relay as described in claim 1, characterized in that: The inner diameter D1 of the first spring (101a) and the outer diameter D2 of the second spring (101b) satisfy the following condition: 1.02≤D1 / D2≤1.

04.

4. The pressure relay as described in any one of claims 1 to 3, characterized in that: It also includes, The housing (102) has an inner cavity (102a) inside it; A pressure adjustment knob (103) is provided on the housing (102); A plunger (104) is slidably mounted in the inner cavity (102a); A switch contact (105) is disposed in the inner cavity (102a) and located on the sliding path of the plunger (104), the plunger (104) being adapted to contact the switch contact (105); The opposing sides of the pressure adjustment knob (103) and the plunger (104) are the first pressure bearing surface (S1) and the second pressure bearing surface (S2), respectively. The two ends of the first spring (101a) and the second spring (101b) respectively abut against the pressure adjustment knob (103) and the plunger (104).

5. The pressure relay as described in claim 4, characterized in that: It also includes an oil inlet (106) which is connected to the inner cavity (102a) and is used to deliver hydraulic oil into the inner cavity (102a).

6. A hydraulic control system for vehicle overload protection, characterized in that: include, The hydraulic pump station (200), the control unit (300), the hydraulic cylinder (400) for driving the lifting and lowering of the vehicle body, and the pressure relay as described in claim 1; The rodless chamber (401) of the hydraulic cylinder (400) is connected to the output end of the hydraulic pump station (200) through a hydraulic pipeline; The pressure relay is connected to the oil inlet path of the rodless chamber (401); The pressure relay's switch contact (105) is electrically connected to the control unit (300), and the control unit (300) controls the hydraulic cylinder (400) to perform lifting lock or alarm actions according to the signal status of the switch contact (105).

7. The hydraulic control system for vehicle overload protection as described in claim 6, characterized in that: It also includes a lifting control valve group (500) disposed between the hydraulic pump station (200) and the hydraulic cylinder (400), wherein the control unit (300) is electrically connected to the electromagnetic control terminal (501) of the lifting control valve group (500); When the hydraulic pressure sensed by the pressure relay reaches the preset overload threshold, the control unit (300) controls the lifting control valve group (500) to switch to the neutral lock state so that the hydraulic cylinder (400) stops rising while maintaining the current height.

8. The hydraulic control system for vehicle overload protection as described in claim 7, characterized in that: The control unit (300) is equipped with a filter timing module (301). The control unit (300) determines that there is a real overload and triggers the protection logic only when the duration of the overload electrical signal received from the pressure relay exceeds the anti-shake duration threshold Δt; wherein, the anti-shake duration threshold Δt is 0.5s to 2s, and is used to filter the instantaneous pulse pressure generated by the vehicle body at the moment of start-up or under bumpy conditions.

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