Method for predicting fatigue life of high-pressure common rail pump oil supply element

Abstract

The present application relates to a high-pressure common rail pump oil supply element fatigue life prediction method, comprising the steps of: obtaining the structure composition, load condition, failure mode, stress analysis result of the oil supply element of the diesel engine high-pressure common rail pump; obtaining the key design parameters; carrying out single factor test design and response surface test design; establishing an oil supply element finite element model to obtain the stress distribution of the oil supply element under different stress analysis results; solving the stress field finite element model and recording the calculation results; obtaining the single factor regression analysis and response surface analysis of the key design parameters on the maximum equivalent stress of the key parts; the present application can greatly reduce the stress level of the oil supply element under the same working conditions; predict the fatigue life of the high-pressure pump oil supply element under different key parameter design values; predict the maximum equivalent stress of the oil supply element under different key parameter design value combinations; and greatly reduce the research and development cost.

Description

TECHNICAL FIELD

[0001] The present application relates to the technical field of diesel engine fuel supply, in particular to a fatigue life prediction method for a high-pressure common rail pump oil supply element. BACKGROUND

[0002] The high-pressure common rail pump of a diesel engine is one of the core components of the fuel supply system, and plays a crucial role in the working process of the common rail system and even the entire diesel engine. The working pressure of the high-pressure pump is one of its key performance parameters, which significantly affects the fuel atomization quality, combustion efficiency, and the power performance, emission performance, and reliability of the diesel engine.

[0003] In the prior art, the maximum working pressure of the high-pressure pump is usually not more than 180 MPa. If it is increased to more than 200 MPa, fatigue failure of the high-pressure pump oil supply element will occur after a certain period of operation, causing high-pressure seal failure, plunger assembly fatigue fracture, and other faults, which will significantly reduce the oil supply flow and efficiency, or even completely lose the working ability of the high-pressure pump.

[0004] To cope with the challenges of high-pressure working conditions, the prior art currently uses methods such as high-strength materials, changes in processing technology, or surface coating of components to improve the load-carrying capacity and durability of the high-pressure pump oil supply element. In addition to changing materials, optimization techniques are also widely used in high-pressure pump design work, but existing optimization techniques mostly focus on improving the overall performance of the high-pressure pump, and there is still a lack of specialized optimization design for key design parameters of the oil supply element, which limits the further improvement of the sealing performance and system reliability of the high-pressure pump.

[0005] The defects of the prior art are:

[0006] 1. Limited effect of changing materials: The degree of improvement of the working pressure and fatigue life of the high-pressure pump oil supply element by changing higher-strength materials is limited, and changing higher-strength materials will increase material costs and processing difficulty, and stress concentration problems are difficult to alleviate, thus failing to meet the actual application requirements of simultaneously improving the working pressure and durability of the high-pressure pump. At present, a domestic manufacturer has successfully increased the rated working pressure of a certain type of high-pressure pump from 160 MPa to 180 MPa by changing materials while ensuring the durability of the sealing structure, but as the rated pressure continues to increase, the durability of the sealing structure decreases significantly, so the effect of improving the working pressure and durability of the sealing structure by changing materials is limited;

[0007] 2. Lack of optimization design quantifiable standard: the method of optimizing the key design parameters of the oil supply element is generally more economical and comprehensive, but the optimization technology is mainly focused on the improvement of the overall performance of the high-pressure pump and even the common rail system, and the existing technology lacks optimization of the design parameters of the sealing structure. In addition, in order to guide the design and development of subsequent models, the influence of the key design parameters of the oil supply element on its durability needs to be obtained, and the existing technology lacks an evaluation method and process for the durability of the oil supply element of the high-pressure pump, and the influence of the key design on the durability of the oil supply element is not given.

[0008] Therefore, in order to improve the working pressure and durability of the existing high-pressure pump, and to provide a reference for the development of subsequent products, it is urgent to carry out the life evaluation and optimization of the oil supply element of the high-pressure common rail pump. SUMMARY

[0009] The present application is directed to the above problems, and provides a high-pressure common rail pump oil supply element fatigue life prediction method, which aims to greatly reduce the stress level of the oil supply element under the same working conditions; predict the fatigue life of the oil supply element of the high-pressure pump under different key parameter design values, and guide the improvement design of the case model; predict the maximum equivalent stress of the oil supply element under different combinations of key parameter design values, and guide the multi-parameter comprehensive improvement design of the case model; greatly reduce the research and development cost, and guide the design improvement.

[0010] To solve the above problems, the technical scheme provided by the present application is:

[0011] A high-pressure common rail pump oil supply element fatigue life prediction method, comprising the following steps:

[0012] S100. Obtain the structure, load condition, failure mode, and stress analysis result of the oil supply element of the high-pressure common rail pump of the diesel engine;

[0013] S200. According to the failure mode and stress analysis result of the oil supply element of the high-pressure common rail pump, the key design parameters of the oil supply element are screened out; and the maximum equivalent stress of the key part of the oil supply element is taken as the fatigue life evaluation index;

[0014] S300. For the oil supply element with the number of key design parameters less than 3, single factor test design is carried out; for the oil supply element with the number of key design parameters not less than 3, response surface test design is carried out, and a response surface test parameter matrix is obtained;

[0015] S400. According to the stress analysis result of the load condition in S100, a finite element model of the oil supply element is established according to the response surface test parameter matrix in S300, and the stress distribution of the oil supply element under different stress analysis results is obtained;

[0016] S500. Solve the stress field finite element model, and then record the calculation results of the maximum equivalent stress of the key parts of the oil supply element;

[0017] S600. Perform an influence analysis of the key design parameters of the oil supply element on fatigue life, and obtain a single-factor regression analysis and a response surface analysis of the key design parameters on the maximum equivalent stress of the key parts; wherein:

[0018] The single-factor regression analysis corresponds to a case where the number of key design parameters is less than 3;

[0019] The response surface analysis corresponds to a case where the number of key design parameters is not less than 3.

[0020] Preferably, in S100, the oil supply element of the high-pressure common rail pump comprises a plunger pair, a plunger sleeve-valve seat assembly, and a valve core assembly; wherein:

[0021] The plunger pair includes a plunger and a tappet;

[0022] The plunger sleeve-valve seat assembly includes a plunger sleeve and a valve seat;

[0023] The valve core assembly includes a valve core;

[0024] The load conditions of the high-pressure common rail pump include the fuel pressure on the surface of the plunger, the fuel pressure on the surface of the plunger sleeve, the fuel pressure on the surface of the valve seat, the fuel pressure on the surface of the valve core, the bolt pre-tightening force on the plunger sleeve, the spring force on the plunger, the spring force on the valve core, and the contact relationship between the surfaces of the plunger and the valve core in mutual contact;

[0025] The failure modes of the high-pressure common rail pump include the failure modes of the plunger pair, the failure modes of the plunger sleeve-valve seat assembly, and the failure modes of the valve core assembly; wherein:

[0026] The failure modes of the plunger pair include groove fillet fracture and plunger middle fatigue fracture;

[0027] The failure modes of the plunger sleeve-valve seat assembly include seal surface failure;

[0028] The failure modes of the valve core assembly include seal failure.

[0029] Preferably, in S200, the key design parameters of the oil supply element include plunger groove fillet radius, plunger bottom spherical surface diameter, plunger sleeve seal surface edge fillet radius, plunger sleeve seal surface edge taper angle, plunger sleeve seal surface boss thickness, plunger sleeve seal ring belt diameter, valve core taper angle, and valve core assembly taper angle difference.

[0030] Preferably, in S300, the number of the key design parameters of the plunger counterpart is 2, including groove fillet radius and bottom spherical surface diameter; the single-factor test design is carried out for the groove fillet radius and the bottom spherical surface diameter respectively;

[0031] The number of the key design parameters of the valve seat-plunger sleeve assembly is 4, including sealing surface edge fillet radius, plunger sleeve sealing surface edge taper angle, sealing surface boss thickness and sealing ring belt diameter; the single-factor test design is carried out for the sealing surface edge fillet radius, the plunger sleeve sealing surface edge taper angle, the sealing surface boss thickness and the sealing ring belt diameter respectively, and then the Box-Behnken test design is carried out for the sealing surface edge fillet radius, the plunger sleeve sealing surface edge taper angle, the sealing surface boss thickness and the sealing ring belt diameter.

[0032] Preferably, the value range of the groove fillet radius is 0.2mm-0.8mm, with a graduation value of 0.1mm;

[0033] The value range of the bottom spherical surface diameter is 440mm-800mm, with a graduation value of 40mm;

[0034] The value range of the sealing surface edge fillet radius is 0mm-0.4mm, with a graduation value of 0.05mm;

[0035] The value range of the plunger sleeve sealing surface edge taper angle is 90°-150°, with a graduation value of 15°;

[0036] The value range of the sealing surface boss thickness is 0.15mm-0.4mm, with a graduation value of 0.05mm;

[0037] The value range of the sealing ring belt diameter is 19.2mm-21.2mm, with a graduation value of 0.4mm.

[0038] Preferably, in S400, according to the response surface test parameter matrix in S300 and the stress analysis results of the load condition in S100, the finite element model of the oil supply element is established; then the stress field finite element model is established for each oil supply element and is simplified, and after the calculation grid is divided, the solving is carried out to obtain the stress distribution of the oil supply element under different stress analysis results under different key design parameters.

[0039] Preferably, in S500, after the solving of the stress field finite element model is completed, the calculation results of the maximum equivalent stress of the key parts of the plunger, the plunger sleeve and the valve core are recorded in the following manner:

[0040] The recorded results of the finite element model of the plunger are the maximum equivalent stresses at the plunger groove fillet and the center of the plunger under different key design parameters;

[0041] The recorded results of the finite element model of the plunger sleeve are the maximum equivalent stresses of the sealing surface of the plunger sleeve and the valve seat under different key design parameters;

[0042] The recorded results of the finite element model of the spool are the maximum equivalent stresses of the contact ring of the spool and the valve seat under different key design parameters.

[0043] Preferably, in S600, after the single-factor regression analysis and the response surface analysis of the maximum equivalent stresses of the key parts under the key design parameters are completed, the optimization design of the key design parameters of the oil supply element is finally completed according to the functional expression between the maximum equivalent stress and the key design parameters; specifically:

[0044] The single-factor regression analysis is achieved by fitting the maximum equivalent stresses of different key design parameters of the oil supply element by using a regression equation, and a functional expression of the maximum equivalent stress of the oil supply element with respect to a single design parameter is obtained;

[0045] The response surface analysis is achieved by fitting the calculated maximum equivalent stresses of the oil supply element under different key design parameters by using a regression equation, and a functional expression of the maximum equivalent stress of the oil supply element with respect to all the key design parameters is obtained;

[0046] The optimization design is achieved by calculating the optimal design parameter combination according to the obtained functional expression under a given optimization criterion.

[0047] Preferably, in the optimization design, the minimization of the maximum equivalent stress of the sealing surface of the valve seat-plunger sleeve assembly is taken as the given optimization criterion, and the optimal values of the four key design parameters, i.e., the plunger sleeve pre-tightening force, the cone angle, the boss thickness and the ring diameter, are obtained.

[0048] Compared with the prior art, the present application has the following advantages:

[0049] 1. The present application realizes a significant reduction in the stress level of the oil supply element under the same working conditions, as shown in FIG. 4 and FIG. 5. Figure 9 As can be seen from the comparison, the maximum equivalent stress after optimization is reduced by about 50%;

[0050] 2. The application provides the influence law of single key design parameter on the maximum equivalent stress of the oil supply element, and gives the numerical relationship between the design parameter and the maximum equivalent stress and fatigue life, thereby the fatigue life of the high-pressure pump oil supply element under different key parameter design values can be predicted, and the case model improvement design is guided;

[0051] 3. The application provides the response surface equation of the key design parameter on the maximum equivalent stress of the valve seat-plunger sleeve assembly sealing surface, thereby the maximum equivalent stress of the oil supply element under different key parameter design value combinations can be predicted, and the multi-parameter comprehensive improvement design of the case model is guided;

[0052] 4. Compared with the technical route of replacing materials and processes, the technical process provided by the application can greatly reduce the research and development cost in the high-load improvement development process of the high-pressure pump oil supply element, thereby guiding the design improvement. BRIEF DESCRIPTION OF DRAWINGS

[0053] Figure 1 The figure is a method flowchart of the embodiment of the application;

[0054] Figure 2a The figure is a force analysis schematic diagram of the oil outlet valve of the embodiment of the application;

[0055] Figure 2b The figure is a force analysis schematic diagram of the oil inlet valve of the embodiment of the application;

[0056] Figure 2c The figure is a fuel pressure analysis schematic diagram of the plunger pair of the embodiment of the application;

[0057] Figure 2d The figure is a force analysis schematic diagram of the bottom of the plunger of the embodiment of the application;

[0058] Figure 2e The figure is a force analysis schematic diagram of the plunger sleeve-valve seat assembly of the embodiment of the application;

[0059] Figure 3 The figure is a finite element model schematic diagram of the embodiment of the application;

[0060] Figure 4a The figure is an equivalent stress nephogram of the finite element calculation result of the plunger sleeve of the embodiment of the application;

[0061] Figure 4b The figure is an equivalent stress nephogram of the finite element calculation result of the valve seat-plunger sleeve sealing surface of the embodiment of the application;

[0062] Figure 5a The figure is a variation law curve schematic diagram of the maximum equivalent stress of the sealing outer ring with the outer circle diameter of the embodiment of the application;

[0063] Figure 5b The average equivalent stress of the sealing outer ring of the specific embodiment of the present application varies with the outer diameter, and the schematic diagram of the variation law curve is shown in the figure;

[0064] Figure 6 The fatigue life of the specific embodiment of the present application varies with the design parameter (single factor), and the schematic diagram of the variation law curve is shown in the figure;

[0065] Figure 7a The response surface of R2 to a and F in the response surface of the key design parameter to the maximum equivalent stress of the sealing surface of the valve seat-plunger sleeve is shown in the figure;

[0066] Figure 7b The response surface of R2 to h and F in the response surface of the key design parameter to the maximum equivalent stress of the sealing surface of the valve seat-plunger sleeve is shown in the figure;

[0067] Figure 7c The response surface of R2 to D and F in the response surface of the key design parameter to the maximum equivalent stress of the sealing surface of the valve seat-plunger sleeve is shown in the figure;

[0068] Figure 7d The response surface of R2 to a and H in the response surface of the key design parameter to the maximum equivalent stress of the sealing surface of the valve seat-plunger sleeve is shown in the figure;

[0069] Figure 7e The response surface of R2 to a and D in the response surface of the key design parameter to the maximum equivalent stress of the sealing surface of the valve seat-plunger sleeve is shown in the figure;

[0070] Figure 7f The response surface of R2 to D and H in the response surface of the key design parameter to the maximum equivalent stress of the sealing surface of the valve seat-plunger sleeve is shown in the figure;

[0071] Figure 8 The response surface equation expression and fitting parameter of the specific embodiment of the present application are shown in the figure;

[0072] Figure 9 The cloud diagram of the maximum equivalent stress of the sealing contact surface after optimization of the specific embodiment of the present application is shown in the figure. DETAILED DESCRIPTION

[0073] The present application will be further illustrated by the following specific embodiments, which are only used to illustrate the present application and not used to limit the scope of the present application, and the modification of various equivalent forms of the present application by those skilled in the art after reading the present application falls within the scope defined by the claims attached hereto.

[0074] As Figure 1As shown, a fatigue life prediction method for a high-pressure common rail pump oil supply component includes the following steps:

[0075] S100. Obtain the structural composition, load conditions, failure modes, and force analysis results of the fuel supply components of the diesel engine high-pressure common rail pump.

[0076] In this specific embodiment, in S100, the oil supply components of the high-pressure common rail pump include a plunger assembly, a plunger sleeve-valve seat assembly, and a valve core assembly; wherein:

[0077] The plunger pair includes a plunger and a tappet.

[0078] The plunger sleeve-valve seat assembly includes a plunger sleeve and a valve seat.

[0079] The valve core assembly includes a valve core.

[0080] like Figures 2a to 2e As shown, the load conditions of the high-pressure common rail pump include the fuel pressure on the surface of the plunger, the fuel pressure on the surface of the plunger sleeve, the fuel pressure on the surface of the valve seat, the fuel pressure on the surface of the valve core, the bolt preload on the plunger sleeve, the spring force on the plunger, the spring force on the valve core, and the contact relationship between the contact surfaces between the plunger and the valve core.

[0081] The failure modes of high-pressure common rail pumps include failure modes of the plunger pair, failure modes of the plunger sleeve-valve seat assembly, and failure modes of the valve core assembly.

[0082] It should be noted that the failure mode is the failure mode of the oil supply component and needs to be analyzed in conjunction with the product failure.

[0083] In this specific embodiment, specifically:

[0084] The failure modes of the plunger pair include groove fillet fracture and fatigue fracture in the middle of the plunger.

[0085] It should be noted that the failure mode of the plunger pair is related to the groove corner radius and the bottom spherical diameter.

[0086] Failure modes of the plunger sleeve-valve seat assembly include failure of the sealing surface.

[0087] It should be noted that the failure mode of the plunger sleeve-valve seat assembly is related to four sealing surface structural parameters: the radius of the sealing surface edge fillet, the cone angle of the plunger sleeve sealing surface edge, the thickness of the sealing surface boss, and the diameter of the sealing ring.

[0088] Failure modes of valve core components include seal failure.

[0089] It should be noted that the failure mode of the valve core assembly is related to two parameters: the valve core cone angle and the valve core assembly cone surface angle difference.

[0090] S200. According to the failure mode and stress analysis result of the oil supply element of the high-pressure common rail pump, key design parameters of the oil supply element are screened; and the maximum equivalent stress of the key part of the oil supply element is taken as the fatigue life evaluation index.

[0091] In the embodiment, in S200, the key design parameters of the oil supply element include the groove fillet radius of the plunger, the bottom spherical diameter of the plunger, the edge fillet radius of the plunger sleeve sealing surface, the edge taper angle of the plunger sleeve sealing surface, the boss thickness of the plunger sleeve sealing surface, the sealing ring belt diameter of the plunger sleeve, the valve core taper angle, and the taper angle difference of the valve core assembly.

[0092] S300. Single-factor test design is carried out on the oil supply element with the number of key design parameters less than 3; and response surface test design is carried out on the oil supply element with the number of key design parameters not less than 3, to obtain a response surface test parameter matrix.

[0093] It should be noted that the role of S300 is to provide an analysis basis for subsequent analysis of the influence of design parameters on the fatigue life of the oil supply element.

[0094] In the embodiment, the Box-Behnken response surface design method is adopted.

[0095] In the embodiment, in S300, the number of key design parameters of the plunger assembly is 2, including the groove fillet radius and the bottom spherical diameter; and single-factor test design is carried out on the groove fillet radius and the bottom spherical diameter, respectively.

[0096] The number of key design parameters of the valve seat-plunger sleeve assembly is 4, including the sealing surface edge fillet radius, the plunger sleeve sealing surface edge taper angle, the sealing surface boss thickness, and the sealing ring belt diameter; single-factor test design is first carried out on the sealing surface edge fillet radius, the plunger sleeve sealing surface edge taper angle, the sealing surface boss thickness, and the sealing ring belt diameter, respectively, and then Box-Behnken test design is carried out on the sealing surface edge fillet radius, the plunger sleeve sealing surface edge taper angle, the sealing surface boss thickness, and the sealing ring belt diameter.

[0097] In the embodiment, the groove fillet radius is in the range of 0.2mm-0.8mm, with a graduation value of 0.1mm.

[0098] The bottom spherical diameter is in the range of 440mm-800mm, with a graduation value of 40mm.

[0099] The sealing surface edge fillet radius is in the range of 0mm-0.4mm, with a graduation value of 0.05mm.

[0100] The plunger sleeve sealing surface edge taper angle is in the range of 90°-150°, with a graduation value of 15°.

[0101] The thickness of the sealing surface boss ranges from 0.15mm to 0.4mm, with 0.05mm as the graduation value.

[0102] The diameter of the sealing ring belt ranges from 19.2mm to 21.2mm, with a graduation value of 0.4mm.

[0103] In this specific embodiment, the obtained Box-Behnken test design table is shown in Table 1:

[0104] Table 1. Box-Behnken test design table for valve seat-plunger sleeve assembly

[0105]

[0106] It should be noted that, as can be seen from Table 1, for the diesel engine of this specific embodiment, the edge fillet radius has a relatively low impact on its life, while the preload force during assembly has a relatively large impact on its life. Therefore, in the Box-Behnken test design table, this specific embodiment uses the assembly preload force instead of the edge fillet radius.

[0107] like Figure 3 As shown, S400. According to the response surface test parameter matrix in S300 and the stress analysis results of the load condition in S100, a finite element model of the oil supply component is established to obtain the stress distribution of the oil supply component under different stress analysis results.

[0108] In this specific embodiment, in S400, a finite element model of the oil supply component is established according to the response surface test parameter matrix in S300 and the stress analysis results of the load condition in S100. Then, a stress field finite element model is established for each oil supply component and simplified. After dividing the calculation grid, the model is solved to obtain the stress distribution of the oil supply component under different stress analysis results under different key design parameters.

[0109] S500. Solve the stress field finite element model, and then record the calculation results of the maximum equivalent stress of the key parts of the oil supply component.

[0110] In this specific embodiment, the calculation results of S500 are shown in the sixth column of Table 1.

[0111] like Figures 4a to 4b As shown, it should be noted that the maximum equivalent stress in the finite element calculation results is directly related to the fatigue life of the oil supply component. Therefore, the recorded results are the maximum equivalent stress in the finite element model calculation results of different oil supply components.

[0112] As shown in FIG5 , in this specific embodiment, in S500 , after solving the stress field finite element model, the calculation results of the maximum equivalent stress of the key parts of the plunger, plunger sleeve, and valve core are recorded in the following manner:

[0113] The recorded results of the finite element model of the plunger are the maximum equivalent stress at the plunger groove corners and the center of the plunger under different key design parameters. The calculation result is the fatigue life evaluation index of the plunger.

[0114] The recorded results of the finite element model of the plunger sleeve are the maximum equivalent stress on the sealing surface of the plunger sleeve and the valve seat under different key design parameters. The calculation result is the fatigue life evaluation index of the plunger sleeve.

[0115] The recorded results of the finite element model of the valve core are the maximum equivalent stress of the contact ring between the valve core and the valve seat under different key design parameters. The calculation result is the fatigue life evaluation index of the valve core component.

[0116] like Figure 6 As shown, S600. Carry out analysis on the influence of key design parameters of oil supply components on fatigue life, and obtain single factor regression analysis and response surface analysis on the maximum equivalent stress of key parts of key design parameters;

[0117] in:

[0118] The univariate regression analysis corresponds to the case where the number of key design parameters is less than 3.

[0119] Response surface analysis corresponds to the case where the number of key design parameters is not less than 3.

[0120] In this specific embodiment, in S600, after completing the single factor regression analysis and response surface analysis of the key design parameters on the maximum equivalent stress of the key parts, the key design parameters of the oil supply component are finally optimized based on the functional expression between the maximum equivalent stress and the key design parameters; specifically:

[0121] Single factor regression analysis is achieved by using regression equations to fit the maximum equivalent stress under key design parameters of different oil supply components, and the functional expression of the maximum equivalent stress of the oil supply component with respect to a single design parameter is obtained.

[0122] Response surface analysis is performed by calculating the maximum equivalent stress of the oil supply component under different key design parameter combinations that have been calculated and designed in Table 1. Then, regression fitting is performed on the calculated results to obtain the functional expression of the maximum equivalent stress of the oil supply component with respect to all key design parameters.

[0123] Optimize the design by using the given optimization criteria and then calculating the optimal design parameter combination based on the obtained function expression.

[0124] In the embodiment, the minimum of the maximum equivalent stress of the sealing surface of the valve seat-plunger sleeve assembly is taken as the given optimization criterion, and the optimal values of the four key design parameters, i.e. the pre-tightening force of the plunger sleeve, the taper angle, the boss thickness and the ring belt diameter, are obtained.

[0125] In the embodiment, the single-factor fitting formula of the taper angle θ of the valve core assembly to the maximum equivalent stress of the valve core and the valve seat is expressed by formula (1):

[0126]

[0127] wherein σ θ,set is the maximum equivalent stress of the valve seat under different taper angles, with the unit of MPa; σ θ,valve is the maximum equivalent stress of the valve core under different taper angles, with the unit of MPa; and θ is the taper angle of the valve core assembly, with the unit of °.

[0128] In the embodiment, the single-factor fitting formula of the taper angle difference Δθ of the valve core assembly to the maximum equivalent stress σ Δθ of the valve core and the valve seat is expressed by formula (2):

[0129]

[0130] wherein σ Δθ,set is the maximum equivalent stress of the valve seat under different taper angle differences, with the unit of MPa; σ Δθ,valve is the maximum equivalent stress of the valve core under different taper angle differences, with the unit of MPa; and Δθ is the taper angle difference, with the unit of °.

[0131] In the embodiment, the single-factor fitting formula of the taper angle θ of the valve core assembly to the fatigue life t θ of the valve core and the valve seat is expressed by formula (3):

[0132]

[0133] wherein t θ,set is the fatigue cycle number of the valve seat under different taper angles; and t θ,valve is the fatigue cycle number of the valve core under different taper angles.

[0134] In the embodiment, the single-factor fitting formula of the taper angle difference Δθ of the valve core assembly to the fatigue life t θ of the valve core is expressed by formula (4):

[0135]

[0136] wherein t Δθ,valve is the fatigue cycle number of the valve core under different taper angle differences.

[0137] In this specific embodiment, the maximum equivalent stress σ on the sealing surface of the plunger sleeve is caused by the taper angle α of the sealing surface of the valve seat-plunger sleeve assembly. α The single factor fitting formula is expressed as formula (5):

[0138] σ α =-4.733×α+1193.884 (5)

[0139] Where: α is the maximum equivalent stress of the sealing surface under different plunger sleeve sealing surface edge cone angles, the unit is MPa; α is the valve seat-plunger sleeve assembly sealing surface edge cone angle, the unit is °.

[0140] In this specific embodiment, the maximum equivalent stress σ on the sealing surface of the plunger sleeve is caused by the thickness h of the boss of the valve seat-plunger sleeve assembly sealing surface. h The single factor fitting formula is expressed as formula (6):

[0141] σ h =-150.274×h+585.289 (6)

[0142] Where: h is the maximum equivalent stress on the sealing surface of the plunger sleeve under different boss thicknesses of the valve seat-plunger sleeve assembly, in MPa; h is the boss thickness of the sealing surface of the valve seat-plunger sleeve assembly, in mm.

[0143] In this specific embodiment, the maximum equivalent stress σ on the sealing surface of the plunger sleeve is caused by the diameter D of the sealing surface of the valve seat-plunger sleeve assembly. D The single factor fitting formula is expressed as formula (7):

[0144] σ D =-55.639×D+1650.746 (7)

[0145] Where: D It is the maximum equivalent stress of the plunger sleeve sealing surface under different valve seat-plunger sleeve assembly sealing surface annulus diameters, in MPa; D is the valve seat-plunger sleeve assembly sealing annulus diameter, in mm.

[0146] In this specific embodiment, the taper angle α of the sealing surface of the valve seat-plunger sleeve assembly has an effect on the fatigue life t of the plunger sleeve sealing surface. α The single factor fitting formula is expressed as formula (8):

[0147]

[0148] Where: t α is the number of fatigue cycles of the piston sleeve sealing surface under different sealing surface edge cone angles; α is the sealing surface edge cone angle, unit is °.

[0149] In this specific embodiment, the diameter D of the sealing surface of the valve seat-plunger sleeve assembly has an effect on the fatigue life t of the sealing surface of the plunger sleeve. D The single factor fitting formula is expressed as formula (9):

[0150]

[0151] Where: t D is the fatigue life cycle number of the piston sleeve sealing surface under different sealing ring diameters; D is the sealing ring diameter, in mm.

[0152] like Figures 7a to 7f As shown in Figure 2, it should be noted that the above single factor influence regularity models are all obtained through parameter fitting. In the fitting process, basic functions such as polynomial, exponential function, logarithmic function, power function, etc. are used for trial, and finally R is selected. 2 The fitting model with the highest value is used as the single-factor influence law model. This method can maximize the relevance and effectiveness of the model. The response surface equation of the maximum equivalent stress σ on the sealing surface of the valve seat-plunger sleeve assembly with the preload force F, the sealing surface edge cone angle α, the sealing surface boss thickness h, and the annular band outer diameter D is expressed as Equation (10):

[0153]

[0154] Where: σ is the maximum equivalent stress of the piston sleeve sealing surface, in MPa; F is the bolt preload, in N; α is the taper angle of the sealing surface edge, in degrees; h is the thickness of the sealing surface boss, in mm; D is the outer diameter of the sealing ring, in mm.

[0155] like Figure 8 As shown in the figure, it is important to note that the response surface model described above was obtained using the Box-Behnken response surface design (BBD) method because the Box-Behnken design offers significant advantages over other experimental design methods. Specifically, compared with the full factorial design, it significantly reduces the number of experimental points and is suitable for fitting quadratic polynomial models, whereas the full factorial design requires more points to obtain the same information. Compared with the central composite design (CCD), the BBD avoids star-point experiments outside the factor range, reducing experimental risk, and evenly distributes experimental points within the experimental space, resulting in better model fitting. Compared with the Latin square design, the BBD is more flexible, not only allowing for the screening of experimental factors but also allowing for the direct fitting of quadratic models. Therefore, the BBD design offers comprehensive advantages in terms of the number of experiments, safety, and model fitting ability, making it well-suited for multi-factor response optimization problems.

[0156] like Figure 9As shown in the figure, combined with the practical problems such as the wide variety of key design parameters of high-pressure common rail pumps, the cumbersome steps of modifying the finite element model and the high requirements for computer computing power, it can be seen that the use of BBD can provide a better solution and can obtain the interactive effects of multiple parameters on the response variables with a smaller number of test points.

[0157] In this specific embodiment, the response surface model, BBD, describes the relationship between each factor and the response variable through a quadratic polynomial model, and the relationship is expressed as follows:

[0158]

[0159] Where: Y is the response variable; β0 is the constant term; β i is the first-order effect coefficient; β ii is the second-order effect coefficient; β ij is the interaction effect coefficient; k is the number of factors.

[0160] Based on this model, the constant terms and coefficients of the response surface model of the maximum equivalent stress σ on the sealing surface of the valve seat-plunger sleeve assembly and key design parameters can be obtained by performing regression analysis on the test data.

[0161] Figure 4 and Figure 9 The following are the stress distribution cloud diagrams of the contact sealing surface before and after optimization. By comparison, it can be clearly seen that the maximum equivalent stress after optimization is reduced by about 50%.

[0162] In the foregoing detailed description, various features are grouped together in a single embodiment to simplify the disclosure. This method of disclosure should not be interpreted as reflecting an intention that embodiments of the claimed subject matter require more features than are expressly recited in each claim. On the contrary, as reflected in the appended claims, the invention comprises less than all the features of any individual disclosed embodiment. The appended claims are hereby expressly incorporated into the detailed description, with each claim standing on its own as a separate preferred embodiment of the invention.

[0163] The above description of the disclosed embodiments is intended to enable any person skilled in the art to implement or use the present invention. Various modifications to these embodiments will be readily apparent to those skilled in the art, and the general principles defined herein may be applied to other embodiments without departing from the spirit and scope of the present disclosure. Therefore, the present disclosure is not limited to the embodiments presented herein but is intended to be accorded the widest scope consistent with the principles and novel features disclosed herein.

[0164] The above description includes examples of one or more embodiments. Of course, not all possible combinations of components or methods described above will be employed to make or use the embodiments nor will all of

[0165] The above description is included to enable a person skilled in the art to make or use the application. The above description does not preclude various alternatives or variations. For instance, the above description includes examples of one or more embodiments. Of course, not all possible combinations of components or methods described above will be employed to make or use the embodiments nor will all of

[0165] The above description includes examples of one or more embodiments. Of course, not all possible combinations of components or methods described above will be employed to make or use the embodiments nor will all of

Claims

1. A method for predicting fatigue life of a high-pressure common rail pump fuel supply component, characterized by: The following steps are involved: S100 obtains the structural composition, load conditions, failure modes, and force analysis results of the fuel supply components of the diesel engine high-pressure common rail pump; S200. Based on the failure mode of the fuel supply component of the high-pressure common rail pump and the stress analysis results, screen and obtain key design parameters of the fuel supply component; and use the maximum equivalent stress of the key parts of the fuel supply component as a fatigue life evaluation indicator; S300. For the fuel supply component with the number of key design parameters less than 3, a single-factor experimental design is performed; for the fuel supply component with the number of key design parameters not less than 3, a response surface experimental design is performed to obtain a response surface experimental parameter matrix; S400. According to the response surface test parameter matrix in S300 and the force analysis results of the load condition in S100, a finite element model of the oil supply component is established to obtain the stress distribution of the oil supply component under different force analysis results; S500. Solve the stress field finite element model, and then record the calculation results of the maximum equivalent stress of the key parts of the oil supply element; S600. Carry out an analysis of the influence of the key design parameters of the oil supply element on fatigue life, and obtain a single factor regression analysis and response surface analysis of the key design parameters on the maximum equivalent stress of the key parts; wherein: The single factor regression analysis corresponds to the case where the number of the key design parameters is less than 3; The response surface analysis corresponds to the case where the number of the key design parameters is not less than 3.

2. The fatigue life prediction method of a high-pressure common rail pump oil supply component according to claim 1 is characterized in that: In S100, the oil supply element of the high-pressure common rail pump includes a plunger assembly, a plunger sleeve-valve seat assembly, and a valve core assembly; wherein: The plunger pair includes a plunger and a tappet; The plunger sleeve-valve seat assembly includes a plunger sleeve and a valve seat; The valve core assembly includes a valve core; The load condition of the high-pressure common rail pump includes the fuel pressure on the surface of the plunger, the fuel pressure on the surface of the plunger sleeve, the fuel pressure on the surface of the valve seat, the fuel pressure on the surface of the valve core, the bolt pre-tightening force on the plunger sleeve, the spring force on the plunger, the spring force on the valve core, and the contact relationship between the contacting surfaces of the plunger and the valve core; The failure modes of the high-pressure common rail pump include the failure modes of the plunger pair, the failure modes of the plunger sleeve-valve seat assembly, and the failure modes of the valve core assembly; wherein: The failure modes of the plunger pair include groove fillet fracture and plunger middle fatigue fracture; The failure mode of the plunger sleeve-valve seat assembly includes sealing surface failure; The failure mode of the valve core assembly includes sealing failure.

3. The fatigue life prediction method of a high-pressure common rail pump oil supply component according to claim 2 is characterized in that: In S200, the key design parameters of the oil supply element include the radius of the plunger groove fillet, the diameter of the plunger bottom spherical surface, the radius of the fillet of the plunger sleeve sealing surface edge, the cone angle of the plunger sleeve sealing surface edge, the thickness of the plunger sleeve sealing surface boss, the diameter of the plunger sleeve sealing ring, the valve core cone angle, and the cone angle difference of the valve core assembly.

4. The fatigue life prediction method of a high-pressure common rail pump oil supply component according to claim 3 is characterized in that: In S300, the number of the key design parameters of the plunger assembly is 2, including the groove fillet radius and the bottom spherical diameter; the single-factor experimental design is carried out for the groove fillet radius and the bottom spherical diameter respectively; The number of key design parameters of the valve seat-plunger sleeve assembly is 4, including the sealing surface edge fillet radius, the plunger sleeve sealing surface edge cone angle, the sealing surface boss thickness, and the sealing ring band diameter; the single-factor experimental design is first carried out for the sealing surface edge fillet radius, the plunger sleeve sealing surface edge cone angle, the sealing surface boss thickness, and the sealing ring band diameter, and then the Box-Behnken experimental design is carried out for the sealing surface edge fillet radius, the plunger sleeve sealing surface edge cone angle, the sealing surface boss thickness, and the sealing ring band diameter.

5. The fatigue life prediction method of a high-pressure common rail pump oil supply component according to claim 4 is characterized in that: The radius of the groove fillet is in the range of 0.2 mm to 0.8 mm, with 0.1 mm as the graduation value; The diameter of the bottom spherical surface ranges from 440 mm to 800 mm, with a graduation value of 40 mm; The range of the fillet radius of the sealing surface edge is 0mm to 0.4mm, with 0.05mm as the graduation value; The taper angle of the plunger sleeve sealing surface edge ranges from 90° to 150°, with 15° as the graduation value; The thickness of the sealing surface boss ranges from 0.15 mm to 0.4 mm, with a graduation value of 0.05 mm; The diameter of the sealing ring band ranges from 19.2 mm to 21.2 mm, with a graduation value of 0.4 mm.

6. The fatigue life prediction method of a high-pressure common rail pump oil supply component according to claim 5, characterized in that: In S400, according to the response surface test parameter matrix in S300 and the force analysis results of the load condition in S100, a finite element model of the oil supply component is established; then, a stress field finite element model is established for each of the oil supply components and simplified, and a calculation grid is divided and solved to obtain the stress distribution of the oil supply component under different force analysis results under different key design parameters.

7. The fatigue life prediction method of a high-pressure common rail pump oil supply component according to claim 6, characterized in that: In S500, after solving the stress field finite element model, the calculation results of the maximum equivalent stress of the key parts of the plunger, the plunger sleeve, and the valve core are recorded in the following manner: The recorded results of the finite element model of the plunger are the maximum equivalent stresses at the fillet of the plunger groove and the center of the plunger under different key design parameters; The recorded result of the finite element model of the plunger sleeve is the maximum equivalent stress of the sealing surface between the plunger sleeve and the valve seat under different key design parameters; The record result of the finite element model of the valve core is the maximum equivalent stress of the contact ring zone between the valve core and the valve seat under different key design parameters.

8. The fatigue life prediction method of a high-pressure common rail pump oil supply component according to claim 7, characterized in that: In S600, after completing the single factor regression analysis and response surface analysis of the key design parameters on the maximum equivalent stress of the key parts, the key design parameters of the oil supply component are finally optimized based on the functional expression between the maximum equivalent stress and the key design parameters; Specifically: The single factor regression analysis is achieved by fitting the maximum equivalent stress of the oil supply element under different key design parameters using a regression equation, and obtaining a functional expression of the maximum equivalent stress of the oil supply element with respect to a single design parameter; The response surface analysis is performed by calculating and designing the maximum equivalent stress of the oil supply component under different combinations of the key design parameters, and then performing regression fitting on the calculated results to obtain a functional expression of the maximum equivalent stress of the oil supply component with respect to all the key design parameters; The optimization design uses a given optimization criterion and then calculates the optimal design parameter combination based on the obtained function expression.

9. The fatigue life prediction method of a high-pressure common rail pump oil supply component according to claim 8, characterized in that: In the optimization design, the minimization of the maximum equivalent stress on the sealing surface of the valve seat-plunger sleeve assembly is used as the given optimization criterion to obtain the optimal values ​​of the four key design parameters, namely, plunger sleeve preload force, cone angle, boss thickness, and annular band diameter.

Images