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

Abstract

The invention relates to a method for predicting the fatigue life of an oil supply element of a high-pressure common-rail pump. The method comprises the following steps: acquiring the structural composition, load condition, failure form and stress analysis result of the oil supply element of the high-pressure common-rail pump of a diesel engine; key design parameters are obtained; carrying out single-factor test design and response surface test design; establishing a finite element model of the oil supply element, and obtaining stress distribution conditions of the oil supply element under different stress analysis results; solving the stress field finite element model, and recording a calculation result; obtaining single-factor regression analysis and response surface analysis of the key design parameters on the maximum equivalent stress of the key parts; under the same working condition, the stress level of the oil supply element is greatly reduced; predicting the fatigue life of the oil supply element of the high-pressure pump under different key parameter design values; predicting the maximum equivalent stress of the oil supply element under different key parameter design value combinations; and the research and development cost is greatly reduced.

Description

Technical Field

[0001] The invention relates to the technical field of diesel engine fuel supply, and in particular to a fatigue life prediction method for a high-pressure common rail pump fuel supply component. Background Art

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

[0003] In the prior art, the maximum working pressure of a high-pressure pump usually does not exceed 180MPa. If it is increased to above 200MPa, after running for a certain period of time, fatigue failure of the oil supply components of the high-pressure pump is likely to occur, leading to failure of the high-pressure seal, fatigue fracture of the plunger assembly and other faults. In the best case, the oil supply flow and efficiency are significantly reduced, and in the worst case, the high-pressure pump completely loses its working ability.

[0004] In order to cope with the challenges of high-pressure working conditions, existing technologies currently use high-strength materials, replacement of processing technology or coating of component surfaces to improve the load-bearing capacity and durability of high-pressure pump oil supply components. In addition to replacing materials, optimization technology is also widely used in high-pressure pump design, but existing optimization technologies are mostly focused on improving the overall performance of high-pressure pumps, while the specialized optimization design for key design parameters of oil supply components is still insufficient, which limits the further improvement of high-pressure pump sealing performance and system reliability.

[0005] The defects of the prior art are:

[0006] 1. The effect of replacing materials is limited: The degree to which the working pressure and fatigue life of the oil supply components of the high-pressure pump can be improved by replacing higher-strength materials is limited, and replacing higher-strength materials will increase material costs and processing difficulties, and the stress concentration problem is 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 160MPa to 180MPa by replacing materials while ensuring the durability of the sealing structure. However, as the rated pressure continues to increase, the durability of its sealing structure decreases significantly. Therefore, the effect of increasing the working pressure and durability of the sealing structure by replacing materials is relatively limited;

[0007] 2. Lack of quantifiable standards for optimized design: Optimizing the key design parameters of the fuel supply components is usually more economical and comprehensive, but the optimization technology is mostly focused on improving the overall performance of the high-pressure pump and even the common rail system. 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, it is also necessary to obtain the influence of the key design parameters of the fuel supply components on their durability. The existing technology lacks an evaluation method and process for the durability of the high-pressure pump fuel supply components, and does not give the influence of the key design on the durability of the fuel supply components.

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

[0009] In view of the above problems, the present invention provides a fatigue life prediction method for a fuel supply component of a high-pressure common rail pump, the purpose of which is to achieve a significant reduction in the stress level of the fuel supply component under the same working conditions; predict the fatigue life of the fuel supply component of the high-pressure pump under different key parameter design values, and guide the improved design of the case model; predict the maximum equivalent stress of the fuel supply component under different key parameter design value combinations, and guide the multi-parameter comprehensive improved design of the case model; significantly reduce the R&D cost and guide the design improvement.

[0010] To solve the above problems, the technical solution provided by the present invention is:

[0011] A method for predicting fatigue life of a high-pressure common rail pump fuel supply component comprises the following steps:

[0012] S100. Obtaining the structural composition, load condition, failure mode, and force analysis results of the fuel supply component of the diesel engine high-pressure common rail pump;

[0013] S200. According to the failure mode and the stress analysis result of the fuel supply component of the high-pressure common rail pump, the key design parameters of the fuel supply component are screened and obtained; and the maximum equivalent stress of the key parts of the fuel supply component is used as the fatigue life evaluation index;

[0014] S300. For the fuel supply component whose number of key design parameters is less than 3, a single factor experimental design is carried out; for the fuel supply component whose number of key design parameters is not less than 3, a response surface experimental design is carried out to obtain a response surface experimental parameter matrix;

[0015] S400. According to the force analysis results in S300 and the force analysis results in the load form 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;

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

[0017] S600. Carry out the influence analysis of the key design parameters of the oil supply element on the fatigue life, and obtain the single factor regression analysis and 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 the case where the number of the key design parameters is less than 3;

[0019] The response surface analysis corresponds to the key design parameters where the number of the key design parameters is less than 3.

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

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

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

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

[0024] 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 preload 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 that contact each other;

[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 mode of the plunger sleeve-valve seat assembly includes sealing surface failure;

[0028] The failure mode of the valve core component includes sealing failure.

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

[0030] Preferably, in S300, the number of the key design parameters of the plunger pair 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;

[0031] The number of key design parameters of the valve seat-plunger sleeve assembly is 4, including the sealing surface edge fillet radius, the contact 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 contact 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 contact surface edge cone angle, the sealing surface boss thickness, and the sealing ring band diameter.

[0032] Preferably, 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;

[0033] The diameter of the bottom spherical surface ranges from 440 mm to 800 mm, with 40 mm as the graduation value;

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

[0035] The taper angle of the contact surface edge ranges from 90° to 150°, with 15° as the graduation value;

[0036] The thickness of the sealing surface boss ranges from 0.15 mm to 0.4 mm, with 0.05 mm as the graduation value;

[0037] The diameter of the sealing ring belt ranges from 19.2 mm to 21.2 mm, with 0.4 mm as the graduation value.

[0038] Preferably, in S400, according to the force analysis results in S300 and the force analysis results in the load form in S100, a finite element model of the oil supply element is established; then a stress field finite element model is established for each of the oil supply elements and simplified, and the model is solved after dividing the calculation grid to obtain the stress distribution of the oil supply element under different force analysis results under different key design parameters.

[0039] Preferably, 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:

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

[0041] The record 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;

[0042] 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.

[0043] Preferably, 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 element are optimized according to the functional relationship between the maximum equivalent stress and the key design parameters; specifically:

[0044] The single factor influence 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 to a single design parameter;

[0045] The response surface analysis is achieved by calculating and designing the maximum equivalent stress of the oil supply element under different combinations of the key design parameters, and then performing regression fitting on the calculation results to obtain a functional expression of the maximum equivalent stress of the oil supply element with respect to all the key design parameters;

[0046] The optimization design uses a given optimization criterion and then calculates the optimal design parameter combination based on the obtained functional relationship.

[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 used as a given optimization criterion to obtain the optimal values ​​of the four key design parameters, namely, the plunger sleeve preload force, cone angle, boss thickness, and annular band diameter, so as to achieve the maximum equivalent stress of the sealing surface of the valve seat-plunger sleeve assembly.

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

[0049] 1. The present invention realizes a significant reduction in the stress level of the oil supply component under the same working conditions. Fig. 9 It is not difficult to see from the comparison that the maximum equivalent stress after optimization is reduced by about 50%;

[0050] 2. Since the present invention provides the influence law of a single key design parameter on the maximum equivalent stress of the oil supply component, and gives the numerical relationship between the design parameter and the maximum equivalent stress and fatigue life, it can be used to predict the fatigue life of the oil supply component of the high-pressure pump under different key parameter design values, and guide the improved design of the case model;

[0051] 3. Since the present invention provides a response surface equation of the key design parameters to the maximum equivalent stress of the valve seat-plunger sleeve assembly sealing surface, it can be used to predict the maximum equivalent stress of the oil supply component under different key parameter design value combinations, and guide the multi-parameter comprehensive improvement design of the case model;

[0052] 4. Since the present invention is in the process of improving the high-load capacity of the oil supply component of the high-pressure pump, compared with the technical route of replacing materials, processes, etc., the proposed technical process can greatly reduce the research and development costs and guide the design improvement. BRIEF DESCRIPTION OF THE DRAWINGS

[0053] Figure 1 A schematic diagram of a method flow chart of a specific embodiment of the present invention;

[0054] Figure 2a It is a schematic diagram of force analysis of an oil outlet valve according to a specific embodiment of the present invention;

[0055] Figure 2b It is a schematic diagram of force analysis of an oil inlet valve according to a specific embodiment of the present invention;

[0056] Figure 2c It is a schematic diagram of analyzing the fuel pressure of the plunger pair according to a specific embodiment of the present invention;

[0057] Figure 2d It is a schematic diagram of the force analysis of the bottom of the plunger in a specific embodiment of the present invention;

[0058] Figure 2e A schematic diagram of force analysis of a plunger sleeve-valve seat assembly according to a specific embodiment of the present invention;

[0059] Figure 3 It is a schematic diagram of a finite element model of a specific embodiment of the present invention;

[0060] Figure 4a It is an equivalent stress cloud diagram of the finite element calculation result of the plunger sleeve of a specific embodiment of the present invention;

[0061] Figure 4b It is an equivalent stress cloud diagram of the finite element calculation result of the valve seat-plunger sleeve sealing surface of a specific embodiment of the present invention;

[0062] Figure 5a It is a schematic diagram of a curve showing the variation of the maximum equivalent stress of the sealing outer ring with the outer diameter in a specific embodiment of the present invention;

[0063] Figure 5b It is a schematic diagram of a curve showing the variation of the average equivalent stress of the sealing outer ring with the outer diameter according to a specific embodiment of the present invention;

[0064] Figure 6 It is a schematic diagram of a curve showing the variation of fatigue life with design parameters (single factor) in a specific embodiment of the present invention;

[0065] Figure 7a It is a schematic diagram of the response surface of R2 to α and F in the maximum equivalent stress response surface of the valve seat-plunger sleeve sealing surface of the key design parameters of a specific embodiment of the present invention;

[0066] Figure 7b It is a schematic diagram of the response surface of R2 to h and F in the maximum equivalent stress response surface of the valve seat-plunger sleeve sealing surface of the key design parameters of a specific embodiment of the present invention;

[0067] Figure 7c It is a schematic diagram of the response surface of R2 to D and F in the maximum equivalent stress response surface of the valve seat-plunger sleeve sealing surface of the key design parameters of a specific embodiment of the present invention;

[0068] Figure 7d It is a schematic diagram of the response surface of R2 to α and H in the maximum equivalent stress response surface of the valve seat-plunger sleeve sealing surface of the key design parameters of a specific embodiment of the present invention;

[0069] Figure 7e It is a schematic diagram of the response surface of R2 to α and D in the maximum equivalent stress response surface of the valve seat-plunger sleeve sealing surface of the key design parameters of a specific embodiment of the present invention;

[0070] Figure 7f It is a schematic diagram of the response surface of R2 to D and H in the maximum equivalent stress response surface of the valve seat-plunger sleeve sealing surface of the key design parameters of a specific embodiment of the present invention;

[0071] Figure 8 The response surface equation expression and fitting parameter schematic diagram of a specific embodiment of the present invention;

[0072] Fig. 9 This is a cloud diagram of the maximum equivalent stress on the sealing contact surface after optimization of a specific embodiment of the present invention. DETAILED DESCRIPTION

[0073] The present invention is further explained below in conjunction with specific embodiments. It should be understood that these embodiments are only used to illustrate the present invention and are not used to limit the scope of the present invention. After reading the present invention, various equivalent forms of modifications to the present invention by those skilled in the art all fall within the scope defined by the claims attached to this application.

[0074] like Figure 1As shown, a method for predicting fatigue life of a high-pressure common rail pump fuel supply component comprises the following steps:

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

[0076] In this specific embodiment, in S100, the oil supply element of the high-pressure common rail pump includes a plunger pair, 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 Figure 2a to Figure 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 contacting surfaces of the plunger and the valve core.

[0081] 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.

[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-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 sealing surface edge fillet radius, the contact surface edge cone angle, the sealing surface boss thickness, and the sealing ring band diameter.

[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 angle difference.

[0090] S200. According to the failure mode and stress analysis results of the fuel supply component of the high-pressure common rail pump, the key design parameters of the fuel supply component are screened and the maximum equivalent stress of the key parts of the fuel supply component is used as the fatigue life evaluation index.

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

[0092] S300. For the fuel supply components whose number of key design parameters is less than 3, a single factor experimental design is carried out; for the fuel supply components whose number of key design parameters is not less than 3, a response surface experimental design is carried out to obtain a response surface experimental parameter matrix;

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

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

[0095] In this specific embodiment, in S300, the number of key design parameters of the plunger pair is 2, including the groove fillet radius and the bottom spherical diameter; single-factor experimental design is carried out for 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 contact surface edge cone angle, the sealing surface boss thickness, and the sealing ring diameter; first, single-factor experimental design is carried out for the sealing surface edge fillet radius, the contact surface edge cone angle, the sealing surface boss thickness, and the sealing ring diameter, and then Box-Behnken experimental design is carried out for the sealing surface edge fillet radius, the contact surface edge cone angle, the sealing surface boss thickness, and the sealing ring diameter.

[0097] In this specific embodiment, the radius of the groove fillet ranges from 0.2 mm to 0.8 mm, with 0.1 mm as the graduation value.

[0098] The diameter of the bottom spherical surface ranges from 440 mm to 800 mm, with a graduation value of 40 mm.

[0099] The radius of the sealing surface edge fillet ranges from 0 mm to 0.4 mm, with 0.05 mm as the graduation value.

[0100] The taper angle of the contact surface edge ranges from 90° to 150°, with 15° as the graduation value.

[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, from Table 1, it can be seen that 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 to replace the edge fillet radius.

[0107] like Figure 3 As shown, S400. According to the force analysis results in S300 and the force analysis results in the load form 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.

[0108] In this specific embodiment, in S400, a finite element model of the oil supply component is established according to the force analysis results in S300 and the force analysis results in the load form in S100; then a stress field finite element model is established for each oil supply component and simplified, and the 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.

[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 result of S500 is shown in the sixth column of Table 1.

[0111] like Figure 4a to Figure 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 FIG. 5 , 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 of the sealing surface between 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 the analysis of the influence of the key design parameters of the oil supply component on the fatigue life, and obtain the single factor regression analysis and response surface analysis of the key design parameters on the maximum equivalent stress of the key parts;

[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 critical design parameters is 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 optimized according to the functional relationship between the maximum equivalent stress and the key design parameters; specifically:

[0121] The single factor influence analysis is realized by fitting the maximum equivalent stress under the key design parameters of different oil supply components through the regression equation, and the functional expression of the maximum equivalent stress of the oil supply component to a single design parameter is obtained.

[0122] The response surface analysis is carried out by calculating and designing the maximum equivalent stress of the oil supply component under different key design parameter combinations in Table 1, and then performing regression fitting on the calculation results to obtain the functional expression of the maximum equivalent stress of the oil supply component for all key design parameters.

[0123] Optimize the design, use the given optimization criteria, and then calculate the optimal design parameter combination based on the obtained functional relationship.

[0124] In this specific embodiment, 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 a given optimization criterion to obtain the optimal values ​​of the four key design parameters of the plunger sleeve preload, cone angle, boss thickness, and annular band diameter, thereby obtaining the maximum equivalent stress of the sealing surface of the valve seat-plunger sleeve assembly.

[0125] In this specific embodiment, the single factor fitting formula of the maximum equivalent stress of the valve core and the valve seat by the cone angle θ of the valve core assembly is expressed as follows:

[0126]

[0127] Where: θ,set is the maximum equivalent stress of the valve seat under different cone angles, in MPa; σ θ,valve is the maximum equivalent stress of the valve core under different cone angles, in MPa; θ is the cone angle of the valve core assembly, in degrees.

[0128] In this specific embodiment, the maximum equivalent stress σ of the valve core and valve seat caused by the angle difference Δθ of the valve core component cone Δθ The single factor fitting formula is expressed as follows:

[0129]

[0130] Where: Δθ,set is the maximum equivalent stress of the valve seat under different cone angle differences, in MPa; σ Δθ,valve is the maximum equivalent stress of the valve core under different cone angle differences, in MPa; Δθ is the cone angle difference, in degrees.

[0131] In this specific embodiment, the valve core component taper angle Δθ has an effect on the fatigue life t of the valve core and the valve seat. θ The single factor fitting formula is expressed as follows:

[0132]

[0133] Where: t θ,set is the number of fatigue cycles of the valve seat under different cone angles; t θ,valve is the number of fatigue cycles of the valve core under different cone angles.

[0134] In this specific embodiment, the valve core component cone angle difference Δθ has an effect on the fatigue life t of the valve core. θ The single factor fitting formula is expressed as follows:

[0135]

[0136] Where: t Δθ,valve It is the number of fatigue cycles of the valve core under different cone 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 follows:

[0138] σ a =-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, in MPa; α is the valve seat-plunger sleeve assembly sealing surface edge cone angle, in degrees.

[0140] In this specific embodiment, the maximum equivalent stress σ of the boss thickness h on the sealing surface of the plunger sleeve is h The single factor fitting formula is expressed as follows:

[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 sealing surface, in MPa; h is the boss thickness of the valve seat-plunger sleeve assembly sealing surface, in mm.

[0143] In this specific embodiment, the maximum equivalent stress σ of the sealing surface of the valve seat-plunger sleeve assembly ring zone diameter D on the sealing surface of the plunger sleeve D The single factor fitting formula is expressed as follows:

[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 annular zone diameters, in MPa; D is the valve seat-plunger sleeve assembly sealing annular zone 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 sealing surface of the plunger sleeve. α The single factor fitting formula is expressed as follows:

[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 follows:

[0150]

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

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

[0153]

[0154] Among them: 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 °; 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 should be noted that the above response surface model is obtained by the Box-Behnken response surface experimental design (BBD) method, because compared with other experimental design methods, the Box-Behnken design has significant advantages; specifically: compared with the full factorial design, it significantly reduces the number of experimental points and is suitable for fitting quadratic polynomial models, while the full factorial design requires more points to obtain the same information; compared with the central composite design (CCD), BBD avoids star point experiments outside the factor range, reduces experimental risks, and evenly distributes experimental points in the experimental space, and the model fitting effect is better; compared with the Latin square design, BBD is more flexible, not only can the experimental factors be screened, but also the quadratic model can be directly fitted. Therefore, the BBD design has comprehensive advantages in terms of the number of experiments, safety, and model fitting ability, and is very suitable for multi-factor response optimization problems.

[0156] like Fig. 9As shown in the figure, combined with the practical problems such as the wide variety of key design parameters of the high-pressure common rail pump, the cumbersome steps to modify 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 through 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 term and various coefficients in the response surface model of the maximum equivalent stress σ of the valve seat-plunger sleeve assembly sealing surface with key design parameters can be obtained by regression analysis of the test data.

[0161] Figure 4 and Fig. 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 above 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 the embodiments of the claimed subject matter require more features than are clearly stated in each claim. On the contrary, as reflected in the appended claims, the invention is in a state of having less than all the features of the disclosed individual embodiments. Therefore, 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 disclosed embodiments are described above to enable any person skilled in the art to implement or use the present invention. Various modifications of these embodiments are obvious to those skilled in the art, and the general principles defined herein may also 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 given herein, but is consistent with the broadest scope of the principles and novel features disclosed in this application.

[0164] The above description includes examples of one or more embodiments. Of course, it is not possible to describe all possible combinations of components or methods for the purpose of describing the above embodiments, but it should be recognized by those skilled in the art that the various embodiments may be further combined and arranged. Therefore, the embodiments described herein are intended to cover all such changes, modifications and variations that fall within the scope of protection of the appended claims. In addition, with respect to the term "comprising" used in the specification or claims, the word is covered in a manner similar to the term "including", just as "including," is explained as a transitional word in the claims. In addition, any term "or" used in the specification of the claims is intended to mean "non-exclusive or".

[0165] The specific implementation methods described above further illustrate the objectives, technical solutions and beneficial effects of the present invention in detail. It should be understood that the above description is only a specific implementation method of the present invention and is not intended to limit the scope of protection of the present invention. Any modifications, equivalent substitutions, improvements, etc. made within the spirit and principles of the present invention should be included in the scope of protection of the present invention.

Claims

1. A method for predicting fatigue life of a high-pressure common rail pump fuel supply component, characterized in that: The following steps are involved: S100. Obtaining the structural composition, load condition, failure mode, and force analysis results of the fuel supply component of the diesel engine high-pressure common rail pump; S200. According to the failure mode and the stress analysis result of the fuel supply component of the high-pressure common rail pump, the key design parameters of the fuel supply component are screened and obtained; and the maximum equivalent stress of the key parts of the fuel supply component is used as the fatigue life evaluation index; S300. For the fuel supply component whose number of key design parameters is less than 3, a single factor experimental design is carried out; for the fuel supply component whose number of key design parameters is not less than 3, a response surface experimental design is carried out to obtain a response surface experimental parameter matrix; S400. According to the force analysis results in S300 and the force analysis results in the load form 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. Solving the stress field finite element model, and then recording the calculation results of the maximum equivalent stress of the key parts of the oil supply element; S600. Carry out the influence analysis of the key design parameters of the oil supply element on the fatigue life, and obtain the 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 key design parameters where the number of the key design parameters is less than 3.

2. The fatigue life prediction method of the 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 pair, a plunger sleeve-valve seat assembly, and a valve core assembly; wherein: The plunger pair comprises a plunger and a tappet; The plunger sleeve-valve seat assembly comprises 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 preload 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 that contact each other; 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 component includes sealing failure.

3. The fatigue life prediction method of the 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 plunger sleeve sealing surface edge fillet, the taper 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 taper angle, and the taper angle difference of the valve core assembly.

4. The fatigue life prediction method of the 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 pair 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 contact 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 contact 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 contact surface edge cone angle, the sealing surface boss thickness, and the sealing ring band diameter.

5. The fatigue life prediction method of the 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 40 mm as the graduation value; The value range of the fillet radius of the edge of the sealing surface is 0mm to 0.4mm, with 0.05mm as the graduation value; The taper angle of the contact 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 0.05 mm as the graduation value; The diameter of the sealing ring belt ranges from 19.2 mm to 21.2 mm, with 0.4 mm as the graduation value.

6. The fatigue life prediction method of the high-pressure common rail pump oil supply component according to claim 5 is characterized in that: In S400, according to the force analysis results in S300 and the force analysis results in the load form 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 the high-pressure common rail pump oil supply component according to claim 6 is 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 stress at the fillet of the plunger groove and at the center of the plunger under different key design parameters; The record 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 the high-pressure common rail pump oil supply component according to claim 7 is 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 element are optimized according to the functional relationship between the maximum equivalent stress and the key design parameters. Specifically: The single factor influence 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 to a single design parameter; The response surface analysis is achieved by calculating and designing the maximum equivalent stress of the oil supply element under different combinations of the key design parameters, and then performing regression fitting on the calculation results to obtain a functional expression of the maximum equivalent stress of the oil supply element 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 functional relationship.

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

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