A method and device for predicting the sealing performance of a main drive of a shield machine under actual complex service conditions

Abstract

The application discloses a shield machine main drive sealing performance prediction method and device under actual complex service conditions, and relates to the field of mixed lubrication. The method takes a rotary lip seal as a research object, is based on a two-dimensional axisymmetric finite element model of a shield machine main drive rotary lip seal, and combines solid mechanics, lubricating oil film fluid mechanics, sealing lip surface micro-convex body contact mechanics, sealing lip surface normal and tangential elastic deformation, and thermodynamic analysis to quantitatively analyze data representing sealing performance of the shield machine main drive seal under actual complex service conditions, such as viscous friction and friction torque under different pressure differentials. The application can predict the sealing performance of the shield machine main drive seal under actual complex service conditions.

Description

Technical Field

[0001] This application relates to the field of mixed lubrication, and in particular to a method and device for predicting the sealing performance of the main drive of a tunnel boring machine under actual complex service conditions. Background Technology

[0002] The main drive seal provides lubrication and sealing during tunnel boring machine (TBM) operation, and is crucial for ensuring the safe and reliable operation of the TBM. Failure of the main drive seal during TBM service can lead to serious engineering accidents and significant economic losses. Developing software to predict the performance of the TBM main drive rotary lip seal holds promise for solving this problem.

[0003] Numerous sealing technologies have been reported in the existing literature. For example, in 2004, Salant and Rocke first introduced the average Reynolds equation with a flow factor into the rotary lip seal model and used statistical methods to describe the surface geometry of the rotary lip seal. In 2005, Rocke and Salant established a rotary lip seal model by combining the elastic deformation of the rotary lip and the hydrodynamics of the lubricating oil film. In 2006, Maoui and Hajjam studied the influence of the local coupling thermal effect between the rotating shaft and the lubricating oil film on the sealing performance. In 2011, Jia and Salant et al. predicted the effect of shaft surface roughness on pumping rate. In 2015, Gadari and Hajjam et al. established a rotary lip seal model considering different sealing lip surface roughness and shaft surface roughness.

[0004] The sealing technologies described above are all based on elastohydrodynamic lubrication, while tunnel boring machines typically operate under mixed lubrication conditions. In 2007, Shen and Salant considered that when the rotating shaft operates at low speeds, the micro-protrusions on the sealing lip surface come into contact with the shaft, resulting in mixed lubrication. Therefore, they established a mixed elastohydrodynamic lubrication model for rotating lips. In 2013, Guo et al. established a mixed lubrication model based on the average Reynolds equation, combining coupled analysis of contact mechanics, fluid mechanics, and deformation mechanics. In 2020, a mixed thermo-elastohydrodynamic lubrication model for Gaussian surface rotating lip seals was first established, combining thermodynamics, fluid mechanics, deformation mechanics, and contact mechanics. In 2021, a mixed thermo-elastohydrodynamic lubrication model for non-Gaussian surfaces was further established. In 2022, a two-dimensional rotating lip elastohydrodynamic lubrication model was established, focusing on analyzing the influence of shaft surface texture on sealing performance. In 2024, an improved mixed lubrication model was proposed, incorporating the true morphology of the sealing lip and rotating shaft surfaces.

[0005] Although sealing technology is becoming increasingly sophisticated, there is currently no software available to predict the sealing performance of the main drive seals of tunnel boring machines under complex actual service conditions. Summary of the Invention

[0006] The purpose of this application is to provide a method and device for predicting the sealing performance of the main drive seal of a tunnel boring machine under actual complex service conditions, which can predict the sealing performance of the main drive seal of the tunnel boring machine under actual complex service conditions.

[0007] To achieve the above objectives, this application provides the following solution:

[0008] Firstly, this application provides a method for predicting the sealing performance of the main drive seal of a tunnel boring machine under actual complex service conditions, comprising: obtaining the static contact pressure and contact length under the current pressure difference using finite element simulation based on a two-dimensional axisymmetric finite element model of the rotating lip seal of the main drive of the tunnel boring machine; wherein the pressure difference is the pressure difference between the oil side and the silt side; setting an initial value for the oil film thickness; determining the flow factor based on the initial value of the oil film thickness; solving the discrete Reynolds equation using an alternating hermit algorithm based on the flow factor and the contact length under the current pressure difference to obtain the fluid pressure, back pumping rate, and viscous friction force; solving the GW contact model using micro-protrusion contact mechanics to obtain the micro-protrusion contact pressure and micro-protrusion contact friction force; and obtaining the micro-protrusion contact pressure and micro-protrusion contact friction force using the linear contact elastic deformation equation of the elastic half-space based on the fluid pressure, micro-protrusion contact pressure, static contact pressure under the current pressure difference, and contact length under the current pressure difference. Elastic deformation; based on micro-elastic deformation, fluid pressure, micro-protrusion contact pressure, and static contact pressure under the current pressure difference, a new oil film thickness is obtained; the new oil film thickness is compared with the initial oil film thickness value to determine whether the oil film thickness converges; if it does not converge, the new oil film thickness replaces the initial oil film thickness value, and the process returns to the step "determine the flow factor based on the initial oil film thickness value"; if it converges, the sealing contact area temperature is obtained using thermodynamic analysis based on the micro-protrusion contact pressure; when the sealing contact area temperature does not converge, the new oil film thickness replaces the initial oil film thickness value, and the process returns to the step "determine the flow factor based on the initial oil film thickness value"; when the sealing contact area temperature converges, the friction torque is determined based on the micro-protrusion contact friction force and viscous friction force, and the friction torque and anti-pumping rate are output; the friction torque and the anti-pumping rate characterize the sealing performance of the tunnel boring machine's main drive rotary lip seal.

[0009] Secondly, this application provides a device for predicting the sealing performance of the main drive of a tunnel boring machine under actual complex service conditions, including: a simulation module, an initialization module, a flow factor determination module, a Reynolds equation solving module, a contact model solving module, a micro-elastic deformation determination module, an oil film thickness updating module, a comparison module, a replacement module, a temperature calculation module, a circulation module, and a sealing performance determination module.

[0010] The simulation module is used to obtain the static contact pressure and contact length under the current pressure difference based on the two-dimensional axisymmetric finite element model of the rotating lip seal of the tunnel boring machine's main drive using finite element simulation; the pressure difference is the pressure difference between the oil side and the silt side. The initialization module is used to set the initial value of the oil film thickness. The flow factor determination module is used to determine the flow factor based on the initial value of the oil film thickness. The Reynolds equation solving module is used to solve the discrete Reynolds equation using the alternating hermit algorithm based on the flow factor and the contact length under the current pressure difference, obtaining the fluid pressure, pump-back rate, and viscous friction. The contact model solving module is used to solve the GW contact model using micro-forehead contact mechanics, obtaining the micro-forehead contact pressure and micro-forehead contact friction. The micro-elastic deformation determination module is used to obtain the micro-elastic deformation based on the fluid pressure, micro-forehead contact pressure, static contact pressure under the current pressure difference, and contact length under the current pressure difference, using the line contact elastic deformation equation of elastic half-space. The oil film thickness update module is used to obtain a new oil film thickness based on microelastic deformation, fluid pressure, micro-protrusion contact pressure, and static contact pressure under the current pressure difference. The comparison module compares the new oil film thickness with the initial oil film thickness value to determine if the oil film thickness has converged. The replacement module replaces the initial oil film thickness value with the new oil film thickness if convergence does not occur, and calls the flow factor determination module. The temperature calculation module, if convergence occurs, obtains the sealing contact area temperature using thermodynamic analysis based on the micro-protrusion contact pressure. The circulation module replaces the initial oil film thickness value with the new oil film thickness if the sealing contact area temperature does not converge, and calls the flow factor determination module. The sealing performance determination module determines the friction torque based on the micro-protrusion contact friction force and viscous friction force when the sealing contact area temperature converges, and outputs the friction torque and back-pumping rate; the friction torque and the back-pumping rate characterize the sealing performance of the tunnel boring machine's main drive rotary lip seal.

[0011] According to the specific embodiments provided in this application, this application has the following technical effects:

[0012] This application provides a method and device for predicting the sealing performance of the main drive seal of a tunnel boring machine (TBM) under actual complex service conditions. Based on a two-dimensional axisymmetric finite element model of the TBM main drive rotary lip seal, the static contact pressure is obtained through solid mechanics analysis, the fluid pressure is obtained through lubricating oil film hydrodynamics, the micro-protrusion contact pressure and micro-protrusion contact friction are obtained through micro-protrusion contact mechanics, the micro-elastic deformation is calculated using the line contact elastic deformation equation of the elastic half-space, and the temperature of the sealing contact area is obtained through thermodynamic analysis. Combining solid mechanics, lubricating oil film hydrodynamics, micro-protrusion contact mechanics of the sealing lip surface, normal and tangential elastic deformation of the sealing lip surface, and thermodynamic analysis, the method quantitatively analyzes the data characterizing the sealing performance of the TBM main drive seal under actual complex service conditions, such as the back pumping rate and friction torque under different pressure differentials, thus realizing the prediction of the sealing performance of the TBM main drive seal under actual complex service conditions. Attached Figure Description

[0013] To more clearly illustrate the technical solutions in the embodiments of this application or related technologies, the drawings used in the embodiments will be briefly introduced below. Obviously, the drawings described below are only some embodiments of this application. For those skilled in the art, other drawings can be obtained based on these drawings without creative effort.

[0014] Figure 1 A flowchart illustrating a method for predicting the sealing performance of the main drive of a tunnel boring machine under actual complex service conditions, provided as an embodiment of this application;

[0015] Figure 2 A schematic diagram of the research object provided in one embodiment of this application;

[0016] Figure 3 A more detailed flowchart illustrating a method for predicting the sealing performance of the main drive of a tunnel boring machine under actual complex service conditions, as provided in one embodiment;

[0017] Figure 4 A finite element model, equivalent stress distribution, and static contact pressure diagram of a rotary lip seal provided in an embodiment of this application;

[0018] Figure 5 A schematic diagram of the normal and tangential unit stiffness matrix of the main drive rotary lip seal of a tunnel boring machine under actual complex service conditions, provided as an embodiment of this application;

[0019] Figure 6 This is a schematic diagram illustrating the variation of temperature, lubricating oil viscosity, fluid pressure, micro-protrusion contact pressure, and lubricating oil film thickness distribution at different speeds, provided in an embodiment of this application.

[0020] Figure 7A schematic diagram illustrating the variation of total friction force, circumferential deformation, reverse pumping rate, frictional torque, micro-protrusion contact friction force, and viscous friction force at different speeds, provided for an embodiment of this application.

[0021] Figure 8 A schematic diagram showing the fluid pressure distribution, micro-protrusion contact pressure distribution, lubricating oil film thickness distribution, and circumferential deformation variation under different roughnesses in an embodiment of this application;

[0022] Figure 9 This is a schematic diagram illustrating the distribution of viscous friction force, micro-protrusion contact friction force, reverse pumping rate, and friction torque variation under different roughnesses, provided for an embodiment of this application.

[0023] Figure 10 This is a schematic diagram illustrating the variation of viscous friction, micro-protrusion contact friction, total friction, temperature, lubricating oil viscosity, and friction torque under different pressure differentials, provided in an embodiment of this application.

[0024] Figure 11 This is a schematic diagram illustrating the fluid pressure distribution, micro-protrusion contact pressure distribution, lubricating oil film thickness, circumferential deformation, and reverse pumping rate variation under different pressure differentials, provided in an embodiment of this application. Detailed Implementation

[0025] The technical solutions of the embodiments of this application will be clearly and completely described below with reference to the accompanying drawings. Obviously, the described embodiments are only some embodiments of this application, and not all embodiments. Based on the embodiments of this application, all other embodiments obtained by those skilled in the art without creative effort are within the scope of protection of this application.

[0026] To make the above-mentioned objectives, features and advantages of this application more apparent and understandable, the application will be further described in detail below with reference to the accompanying drawings and specific embodiments.

[0027] Regarding the issue of main drive seal failure in tunnel boring machines, in one exemplary embodiment, such as... Figure 1 As shown, a method for predicting the sealing performance of the main drive of a tunnel boring machine under actual complex service conditions is provided, including the following steps 101 to 111. Wherein:

[0028] Step 101: Based on the two-dimensional axisymmetric finite element model of the main drive rotating lip seal of the tunnel boring machine, the static contact pressure and contact length under the current pressure difference are obtained by finite element simulation; the pressure difference is the pressure difference between the oil side and the silt side.

[0029] Step 102: Set the initial value for oil film thickness.

[0030] Step 103: Determine the flow factor based on the initial value of the oil film thickness.

[0031] Step 104: Based on the flow factor and the contact length under the current pressure difference, solve the discrete Reynolds equation using the alternating hermit algorithm to obtain the fluid pressure, pump-back ratio, and viscous friction.

[0032] Step 105: Solve the GW contact model using micro-protrusion contact mechanics to obtain the micro-protrusion contact pressure and micro-protrusion contact friction.

[0033] Step 106: Based on the fluid pressure, micro-protrusion contact pressure, static contact pressure under the current pressure difference, and contact length under the current pressure difference, obtain the micro-elastic deformation using the line contact elastic deformation equation of the elastic half-space.

[0034] Step 107: Obtain the new oil film thickness based on microelastic deformation, fluid pressure, micro-protrusion contact pressure, and static contact pressure under the current pressure difference.

[0035] Step 108: Compare the new oil film thickness with the initial oil film thickness to determine whether the oil film thickness has converged.

[0036] Step 101: If convergence fails, replace the initial value of the oil film thickness with the new oil film thickness and return to step "Determine the flow factor based on the initial value of the oil film thickness".

[0037] Step 109: If convergence is achieved, the temperature of the sealed contact area is obtained by thermodynamic analysis based on the contact pressure of the micro-protrusion.

[0038] Step 110: When the temperature in the sealing contact area does not converge, replace the initial value of the oil film thickness with the new oil film thickness, and return to step "Determine the flow factor based on the initial value of the oil film thickness".

[0039] Step 111: When the temperature of the sealing contact area converges, determine the friction torque based on the contact friction force and viscous friction force of the micro-protrusions, and output the friction torque and the back pumping rate; the friction torque and the back pumping rate characterize the sealing performance of the main drive rotary lip seal of the tunnel boring machine.

[0040] The research object of this application is a rotary lip seal for the main drive of a tunnel boring machine, which consists of a single lip seal ring and a rotating shaft. The single lip seal ring is made of nitrile rubber, and the rotating shaft is made of structural steel. During the operation of the rotary lip seal, the pressure of the medium it experiences is axisymmetric, and to improve computational efficiency, it is simplified into a two-dimensional axisymmetric model.

[0041] The theoretical basis of this application is that during the operation of a rotary lip seal, there is a complex coupling effect between the solid mechanics of the interference fit of the seal, the fluid mechanics of the lubricating oil film, the contact mechanics of the micro-protrusions on the surface of the single lip seal, the micro-elastic deformation of the surface of the single lip seal, and the thermodynamic analysis of the sealing contact area.

[0042] Steps 101 to 111 above are implemented, taking the main drive rotary lip seal of a tunnel boring machine (TBM) as the research object, and considering the coupling effects between multiple physical fields such as solid mechanics, fluid mechanics, micro-protrusion contact mechanics, micro-elastic deformation, and thermodynamics. Combining theoretical foundations and finite element techniques, a method for predicting the sealing performance of the TBM main drive seal under actual complex service conditions is established for the first time. Data characterizing the sealing performance, such as the back pumping rate and frictional torque, are quantitatively analyzed under different operating speeds, surface roughnesses, and large pressure differentials. This application provides a numerical simulation software that avoids expensive experiments and labor costs, offering an effective method for predicting the sealing performance of the TBM main drive seal.

[0043] In another exemplary embodiment of this application, solid mechanics employs finite element method analysis, simplifying the rotating lip seal into a two-dimensional axisymmetric finite element model, and simplifying the rotational motion into translational motion. The Yeoh-3rdorder model is used to fit the uniaxial tensile test data of nitrile rubber material. A multi-region method is used to divide the mesh, where the rotating shaft size is 0.5 mm, the sealing lip size is 0.5 mm, and the single-lip seal ring lip size is 0.1 mm. For ease of calculation, the mesh near the contact point between the sealing ring lip and the rotating shaft is refined, and the mesh size at the contact point is set to 0.01 mm. The mesh is divided into 10649 elements and 32841 nodes. The interference fit between the sealing lip and the rotating shaft is 12 mm, the pressure difference between the mud / sand side and the oil side is 0.1 MPa, the mud / sand side pressure is 1.1 MPa, and the fluid side pressure is 1.0 MPa. The static contact pressure P is obtained through finite element simulation calculation. sc and contact length L y The contact length refers to the region from the oil side to the silt side where the static contact pressure is not zero.

[0044] refer to Figure 2 The schematic diagram of the rotary lip seal shown in part (a) consists of a single-lip sealing ring and a rotating shaft. (See reference...) Figure 2 The schematic diagram of the sealing contact area shown in part (b) assumes that the surface of the rotating shaft is smooth and the surface of the sealing lip is distributed with micro-protrusions. Figure 2 Part (c) shows the forces acting on the sealing contact area. Here, y represents the axial direction, and x represents the circumferential direction, i.e., the direction of rotation of the rotating shaft. The medium on one side of the seal is mud and sand, and the medium on the other side is lubricating oil.

[0045] Therefore, step 101 above can be replaced by the following steps 201 to 204:

[0046] Step 201: Simplify the main drive rotary lip seal of the tunnel boring machine into a two-dimensional axisymmetric finite element model; the main drive rotary lip seal of the tunnel boring machine includes a single lip seal ring and a rotating shaft, the single lip seal ring is in contact with the rotating shaft, and the material of the single lip seal ring is nitrile rubber.

[0047] Step 202: Mesh the two-dimensional axisymmetric finite element model using a multi-region method.

[0048] Step 203: Fit the uniaxial tensile test data of nitrile rubber using the Yeoh-3rdorder model.

[0049] Step 204: Based on the two-dimensional axisymmetric finite element model after meshing and the fitted uniaxial tensile test data of nitrile rubber, the static contact pressure and contact length under the current pressure difference are obtained by finite element simulation.

[0050] In another exemplary embodiment of this application, the formula for determining the flow factor in step 103 above is:

[0051]

[0052] In the formula, This represents the circumferential pressure-flow factor. denoted by axial pressure-flow factor, H represents dimensionless oil film thickness, D′ and r represent fitting constants, and γ represents surface morphology parameters; This represents the shear flow factor at any angle in the x-direction. Let θ represent the shear flow factor at any angle in the y-direction, and let θ represent the angle between the micro-protrusions on the sealing lip surface and the circumferential coordinate system. V represents the shear flow factor. r1 and V r2 γ represents the variance ratio, γ1 represents the surface topography parameter of the rotating shaft, γ2 represents the surface topography parameter of the sealing lip, and Φ represents the variance ratio. s The functions represent γ1, γ2, and H, σ1 and σ2 represent the surface roughness of the rotating shaft and the sealing lip, respectively, σ represents the root mean square roughness, and A2, A3, α4, α5, and α6 all represent fitting constants.

[0053] In another exemplary embodiment of this application, step 104 described above can be replaced by steps 301 to 303:

[0054] Step 301: Introduce the flow factor and cavitation factor into the Reynolds equation to obtain the Reynolds equation after the introduction: In the formula, H represents the dimensionless oil film thickness, h represents the lubricating oil film thickness, σ represents the root mean square roughness; F represents the cavitation factor; Φ represents the general variable. X, Y, ξ represents intermediate parameters, x represents the circumferential direction, and y represents the axial direction. Indicates axial dimensionless, L x L represents the circumferential length. y Indicates the contact length. The average clearance is represented by erf, the error function by μ, the lubricating oil viscosity by U, and the relative velocity between the rotating shaft and the sealing lip by P. ref P represents the reference pressure. cav Indicates cavitation pressure; and These represent the pressure-flow factor at any angle. and These represent the shear flow factors at arbitrary angles. Density flux factor at any angle; θ represents the angle between the micro-protrusions on the sealing lip surface and the circumferential coordinate system. This represents the circumferential pressure-flow factor. Indicates the axial pressure-flow factor. denoted by density flow factor, and a2 and b2 both represent fitting constants;

[0055] Step 302: Discretize the introduced Reynolds equation using the finite volume method to obtain the discrete Reynolds equation.

[0056] Step 303: Based on the flow factor and the contact length under the current pressure difference, the discrete Reynolds equation is solved using the alternating hermit algorithm to obtain the fluid pressure, pump-back ratio, and viscous friction.

[0057] In another exemplary embodiment of this application, steps 103 and 104 above relate to lubricating oil film hydrodynamics. The lubricating oil film hydrodynamic analysis is described by the average Reynolds equation, taking into account the surface roughness of the sealing lip and the cavitation effect of the fluid. The average flow factor and cavitation factor are introduced into the Reynolds equation. The Reynolds equation is discretized into a two-dimensional finite equation system using the finite volume method, and the fluid pressure P is obtained by solving the system using the alternating hermit algorithm. f .

[0058] The Reynolds equation is:

[0059]

[0060] In the formula, and Represents the pressure-flow factor at any angle. and Represents the shear flow factor at any angle. The density flux factor at any angle is shown below:

[0061]

[0062] In the formula, and Indicates the pressure-flow factor. Indicates the shear flow factor. θ represents the density flow factor, and θ represents the angle between the micro-protrusions on the sealing lip surface and the circumferential coordinate system.

[0063]

[0064] Wherein, γ is the surface morphology parameter, and D′ and r are shown in Table 1.

[0065] Table 1 D′ and r

[0066] γ D′ r H 1 / 9 1.48 0.42 H>1 1 / 6 1.38 0.42 H>1 1 / 3 1.18 0.42 H>0.75 1 0.90 0.56 H>0.5 3 0.225 1.5 H>0.5 6 0.520 1.5 H>0.5 9 0.870 1.5 H>0.5

[0067]

[0068] σ1 and σ2 are the surface roughness of the rotating shaft and the sealing lip, respectively, and A2, A3, α4, α5, and α6 are shown in Table 2.

[0069] Table 2 A2, A3, α4, α5, α6

[0070] γ <![CDATA[A2]]> <![CDATA[α4]]> <![CDATA[α5]]> <![CDATA[α6]]> <![CDATA[A3]]> 1 / 9 2.046 1.12 0.78 0.03 1.856 1 / 6 1.962 1.08 0.77 0.03 1.754 1 / 3 1.858 1.01 0.76 0.03 1.561 1 1.899 0.98 0.92 0.05 1.126 3 1.560 0.85 1.13 0.08 0.556 6 1.290 0.62 1.09 0.08 0.388 9 1.011 0.54 1.07 0.08 0.295

[0071]

[0072] In the formula, The values ​​of a2 and b2 are shown in Table 3.

[0073] Table 3 Values ​​of a2 and b2

[0074] γ <![CDATA[a2]]> <![CDATA[b2]]> 1 / 9 0.926 0.005 1 / 3 0.686 0.019 1 0.657 0.081 3 0.730 0.503 9 0.927 3.704

[0075] Reference Figure 3 The Reynolds equations are discretized into a two-dimensional finite equation system using the finite volume method and written in nodal coordinate form as follows:

[0076]

[0077] make:

[0078]

[0079] In the formula: H represents the dimensionless oil film thickness, P i,j =F i,j Ф i,j δX and δY represent the fluid pressure, respectively; δX and δY represent the distance between two adjacent nodes in the X and Y directions; and ΔX and ΔY represent the distance between two adjacent interfaces.

[0080] The above node coordinates can then be simplified as follows:

[0081] A i,j φ i,j =B i,j Φ i+1,j +C i,j Φ i-1,j +D i,j Φ i,j+1 +E i,j Φ i,j-1 +F i,j ;

[0082] The tridiagonal matrix algorithm (TDMA) is used to solve the problem along the axis, let B i,j Φ i+1,j +C i,j Φ i-1,j +F i,j =M i,j ;

[0083] but:

[0084] A i,j Φ i,j =D i,j Φ i,j+1 +E i,j Φ i,j-1 +M i,j ;

[0085]

[0086] Let Φ i,j-1 =P i,j-1 Φ i,j +Q i,j-1 Solve the following two equations simultaneously:

[0087]

[0088] but:

[0089] A i,j Φ i,j =D i,j Φ i,j+1 +E i,j P i,j-1 Φ i,j +E i,j Q i,j-1 +M i,j ;

[0090]

[0091] Therefore, we can conclude that:

[0092]

[0093] because:

[0094]

[0095] Therefore, P can be obtained by iterating from j=1 to j=NY-2. i,j and Q i,j .

[0096] Φ i,NY-2 =P i,NY-2 Φ i,NY-1 +Q i,NY-2 ;

[0097] Φ i,NY-1 These are boundary conditions, and Φ can be solved. i,NY-2 According to Φ i,j =P i,j Φ i,j+1 +Q i,j By iterating from j = NY-2 to j = 1, each Φ can be calculated. i,j .

[0098] Solve using the Circular Tridiagonal Matrix Algorithm (CTDMA) in the circumferential direction:

[0099] A i,j Φ i,j =B i,j Φ i+1,j +C i,j Φ i-1,j +D i,j Φ i,j+1 +E i,j Φ i,j-1 +F i,j ;

[0100] Let N i,j =D i,j Φ i,j+1 +E i,j Φ i,j-1 +F i,j ,but:

[0101] A i,j Φ i,j =B i,j Φ i+1,j +C i,j Φ i-1,j +N i,j ;

[0102] Similar to the Y direction, in the X direction, we assume: Φ i,j =P i,j Φ i+1,j +R i,j Φ N-1,j +Q i,j ;

[0103]

[0104] but:

[0105]

[0106]

[0107] Therefore, any P can be calculated. i,j R i,j and Q i,j Then, solve Φ. N-1,j :

[0108] Let i = N-1, according to A i,j Φ i,j =B i,j Φ i+1,j +C i,j Φ i-1,j +N i,j ,have:

[0109] A N-1,j Φ N-1,j =B N-1,j Φ 0,j +C N-1,j Φ N-2,j +N N-1,j (1)

[0110] Using this formula: Φ i,j =P i,j Φ i+1,j +R i,j Φ N-1,j +Q i,j We can obtain:

[0111] Φ N-2,j =P N-2, jΦ N-1,j +R N-2,j Φ N-1,j +Q N-2,j Substitute (1)

[0112] (A N-1,j -C N-1,j P N-2,j -C N-1,j R N-2,j )Φ N-1,j =B N-1,jΦ 0,j +C N-1,j Q N-2,j +N N-1,j

[0113] Let:

[0114] S 0,j = A N-1,j -C N-1,j P N-2,j -C N-1,j R N-2,j ;

[0115] T 0,j = B N-1,j ;

[0116] U 0,j = D N-1,j φ N-1,j+1 +E N-1,j Φ N-1,j-1 +F N-1,j +Cx -1,j Q N-2,j ;

[0117] Then:

[0118]

[0119] S i-1,j Φ N-1,j = T i-1,j (P i-1,j Φ i,j +R i-1,j Φ N-1,j +Q i-1,j )+U i-1,j ;

[0120] (S i-1,j -T i-1,j R i-1,j )Φ N-1,j = T i-1,j P i-1,j Φ i,j +T i-1,j Q i-1,j +u i-1,j ;

[0121] Therefore:

[0122] S i,j = S i-1,j -T i-1,j *R i-1,j ;

[0123] T i,j = T i-1,j *P i-1,j ;

[0124] U i,j =U i-1,j +T i-1,j *Q i-1,j ;

[0125] Let i = N in (2),

[0126] S N-1,j Φ N-1,j =T N-1,j Φ N-1,j +U N-1,j ;

[0127] but:

[0128]

[0129] Then substitute Φ i,j =P i,j Φ i+1,j +R i,j Φ N-1,j +Q i,j Calculate Φ sequentially N-2,j ……Φ 0,j .

[0130] Solving P using the Alternating Implicit Algorithm (ADI) ij (fluid pressure P) f (discrete form).

[0131]

[0132] The reverse pumping rate (reverse pumping ratio) can be expressed as:

[0133] In the formula,

[0134] Furthermore, viscous shear stress can be expressed as:

[0135]

[0136] In the formula, The term is the average value of the sliding velocity component of the shear stress. and It is the shear stress factor.

[0137]

[0138] In the formula, z = H / 3.

[0139]

[0140] In the formula, the values ​​of D and s are shown in Table 4.

[0141] Table 4 shows the values ​​of D and s.

[0142] γ D s 1 / 9 1.51 0.52 1 / 3 1.47 0.58 1 1.40 0.66 3 0.98 0.79 9 0.73 0.91

[0143]

[0144] Φ fs =0 H≥7;

[0145]

[0146] Viscous friction force F l for:

[0147]

[0148] In the formula, A is the solution domain.

[0149] In another exemplary embodiment of this application, the contact mechanics of the micro-protrusions on the surface of the single-lip seal ring is described by the GW contact model, and the micro-protrusion contact pressure is P. c Therefore, step 105 above can be replaced by steps 401 to 403:

[0150] Step 401: Define the GW contact model as follows: In the formula, P c The contact pressure of the micro-protrusion is represented by E′, the equivalent elastic modulus is E′, the equivalent density of the micro-protrusion is η, R is the radius of the rotation axis, and H represents the oil film thickness; z = H / 3, where z represents one-third of the oil film thickness.

[0151] Step 402: Solve the GW contact model using micro-protrusion contact mechanics to obtain the micro-protrusion contact pressure.

[0152] Step 403: Based on the contact pressure of the micro-protrusion, use the formula F c =∫∫ A fP c dxdy, calculates the contact friction force of the micro-convex body; where F c Let f be the contact friction force of the micro-protrusion, f be the friction coefficient between the rotating shaft and the sealing lip, and A be the solution domain.

[0153] In another exemplary embodiment of this application, the micro-elastic deformation analysis includes normal deformation analysis. The normal micro-elastic deformation H of the sealing element in the sealing area during the rotation of the shaft is measured. def Originating from fluid pressure P f Contact pressure P with micro-protrusion c The supporting force and static contact pressure P sc The pressure difference ΔP causes normal deformation of the sealing element. Microelastic deformation analysis also includes tangential deformation. The tangential microelastic deformation τ of the sealing element in the sealing area during shaft rotation is also considered. def Originating from the shear stress τ of the lubricating oil filmavg Frictional shear stress τ between the micro-protrusions on the sealing lip surface c The combined effect of these factors means that step 106 can be replaced by steps 501 to 503:

[0154] Step 501: Based on the fluid pressure, the micro-protrusion contact pressure, and the static contact pressure under the current pressure difference, use the formula ΔP = P f +P c -P sc Calculate the pressure difference between the supporting force and the static contact pressure; the supporting force consists of fluid pressure and micro-protrusion contact pressure; where ΔP represents the pressure difference between the supporting force and the static contact pressure, P f P represents fluid pressure. c P represents the contact pressure of the micro-convexity. sc This indicates static contact pressure.

[0155] Step 502: Based on the pressure difference between the supporting force and the static contact pressure, use the formula... Calculate the normal deformation; where H def Denotes normal deformation, E′ is the equivalent elastic modulus, and L y Let y represent the contact length, y represent the axial direction, and s represent the integral variable.

[0156] Step 503: Based on the contact length under the current pressure difference, use the formula... Calculate the tangential deformation; where τ def Indicates tangential deformation, L x τ represents the circumferential length. avg τ represents viscous shear stress. c Let L represent the contact stress of the micro-convex body, and x represent the circumferential direction. At this point, L... x =L y .

[0157] Step 504: Combine the normal deformation and tangential deformation to form micro-elastic deformation.

[0158] In another exemplary embodiment of this application, the thermodynamic analysis is calculated using a global thermal method. The sealing contact area generates a large amount of heat due to friction, causing the temperature of the sealing contact area to rise. The thermodynamic equation is expressed as:

[0159]

[0160] q y : Rotary lip seal reverse pumping rate, ρ: fluid density, C oil Specific heat capacity of lubricating oil, L; Length of rotating shaft, h exchange: Thermal conductivity of the rotating shaft. The heat in the sealing area mainly originates from the combined effects of fluid viscous stress and micro-protrusion contact shear stress, and the generated heat Φ1 is expressed as:

[0161] Φ1=∫∫ A (τ avg +τ c Udxdy;

[0162] In the formula, A is the solution domain and U is the relative sliding velocity.

[0163] Therefore, step 109 above can be replaced by the following steps 601 to 602:

[0164] Step 601: Based on the contact pressure of the micro-convex body, use the formula τ c =fP c Calculate the contact stress of the micro-convex body; where τ c P represents the contact stress of the micro-protrusion, f is the coefficient of friction between the rotating shaft and the sealing lip, and P is the coefficient of friction between the rotating shaft and the sealing lip. c This indicates the contact pressure of the micro-protrusion.

[0165] Step 602: Based on the contact stress of the micro-protrusion, obtain the temperature of the sealed contact area through thermodynamic equations.

[0166] In another exemplary embodiment of this application, the formula for determining the friction torque in step 111 is:

[0167] T n =F all ·R;

[0168] F all =F l +F c ;

[0169] In the formula, T n F represents frictional torque. all F represents the total frictional force, R is the radius of the axis of rotation, and F is the total frictional force. l F represents viscous friction. c This represents the contact friction force of micro-protrusions.

[0170] In another exemplary embodiment of this application, the temperature increase mainly affects the viscosity of the lubricating oil, and the change in lubricating oil viscosity causes changes in fluid pressure and oil film thickness. Therefore, iterative calculations are required until the temperature converges. After step 109 above, the method may further include: based on the temperature of the sealing contact area, using the Reynolds viscosity-temperature equation μ(T)=μ0exp[-α(TT)] ref]] Calculate the viscosity of the lubricating oil at the temperature of the sealing contact area. Where μ(T) represents the viscosity of the lubricating oil at temperature T in the sealing contact area, μ0 represents the viscosity of the lubricating oil at temperature T0 in the sealing contact area, α represents the rate of change of the lubricating oil viscosity with temperature, and T ref Indicates ambient temperature.

[0171] The following is combined with Figures 2-11 The method described in this application will be further elaborated.

[0172] 1. Research object: Rotary lip seal of main drive of tunnel boring machine.

[0173] refer to Figure 2 The schematic diagram of the rotary lip seal shown in (a) consists of a single-lip sealing ring and a rotating shaft. (See reference...) Figure 2 The schematic diagram of the sealing contact area shown in (b) assumes that the surface of the rotating shaft is smooth and the surface of the sealing lip is composed of micro-protrusions. Here, y represents the axial direction and x represents the circumferential direction, i.e., the direction of rotation of the rotating shaft. The medium on one side of the seal is mud and sand, and the medium on the other side is lubricating oil, with a pressure difference of 0.1 MPa between the two media.

[0174] 2. Calculation process as follows Figure 3 As shown.

[0175] 1) Input basic parameters, through Figure 4 The finite element model shown in section (a) calculates the static contact pressure distribution P under different pressure differences. sc and contact length L y , Figure 4 Part (b) in the figure is the equivalent stress diagram of the sealing lip under actual service conditions. It can be seen that the equivalent stress at the contact point between the lip and the rotating shaft and at the tooth root position is the largest -1.8MPa. Figure 4 Part (c) shows the pressure distribution at the sealing lip contact. Figure 4 Part (d) in the figure represents the pressure distribution of the sealing lip contact under different pressure differentials;

[0176] 2) Assuming the initial oil film thickness distribution H, solve for the flow factor. and

[0177] 3) The Reynolds equations are discretized into a system of linear two-equations using the Finite Volume Method (FVM). The Tridiagonal Matrix Method (TDMA) is used to solve the discretized Reynolds equations in the axial direction, and the Cyclic Tridiagonal Matrix Method (CTDMA) is used to solve the discretized Reynolds equations in the circumferential direction. The Alternating Obscure Algorithm (ADI) is used to repeat the calculation until Φ reaches the convergence criterion (1e-13), and the fluid pressure P is output. f ;

[0178] 4) Solve for the contact pressure P of the micro-assurances based on the contact mechanics of micro-assurances.c ;

[0179] 5) Based on such Figure 5 The normal stiffness matrix shown in part (a) is... Figure 5 The tangential stiffness matrix shown in part (b) is used to calculate the normal deformation H. def and tangential deformation τ def ;

[0180] 6) Based on the micro-protrusion contact pressure P c and fluid pressure P f With static contact pressure P sc Calculate the new oil film thickness;

[0181] 7) Compare the thickness of the new and old oil films to determine whether the convergence standard (0.01 micrometers) has been met;

[0182] 8) If the oil film thickness does not converge, calculate the microelastic normal and tangential deformation based on the elastic half-space line contact. The normal and tangential deformations cause changes in the oil film thickness and micro-protrusion orientation, and the changes in micro-protrusion orientation cause changes in the flow factor.

[0183] 9) Solve for the flow factor and repeat steps 2)-9) until the oil film thickness reaches the convergence criterion;

[0184] 10) Calculate the fluid viscous pressure τ avg Contact pressure τ with micro-protrusions c Frictional force F caused by the combined action all ;

[0185] 11) Calculate the temperature T of the sealed contact area based on thermodynamic analysis, and determine whether the temperature has reached the convergence standard (1℃);

[0186] 12) If the temperature does not converge, repeat steps 2)-11) until the temperature reaches the convergence criterion;

[0187] 13) Output reverse pumping rate R V and friction torque T n The entire calculation process is now complete.

[0188] Table 5 shows the actual complex service conditions, material performance parameters, and lubricating oil performance parameters of the main drive seal of the tunnel boring machine.

[0189] Table 5 Default parameters used in the fluid-structure-thermal coupling simulation model

[0190] parameter value parameter value Oil side fluid pressure 1.0MPa Sediment-side fluid pressure 1.1MPa Reference pressure 1.0MPa cavitation pressure 0MPa Shaft rotation speed 3m / s coefficient of friction 0.2 Elastic modulus of sealing material 6.93MPa Poisson's ratio of sealing material 0.49 Shaft material elastic modulus 1.0 GPa Shaft material Poisson's ratio 0.29 fluid viscosity 0.08 Pa*s fluid density 890kg / L Fluid viscosity-temperature coefficient 0.024 Fluid specific heat 1880 J / (kg·K) Surface morphology parameters of seal 1 / 3 Shaft radius 6m Root mean square surface roughness of seal 1μm Root mean square surface roughness of shaft 0μm Surface roughness orientation of seal 90° Fluid surface heat transfer coefficient <![CDATA[200W / (m 2 ·K)]]> temperature 25℃ － －

[0191] 3. The effect of speed on sealing performance.

[0192] In this application, Figure 6Part (a) shows the variation of temperature and lubricating oil viscosity at different speeds. Figure 6 Part (b) shows the variation of fluid pressure distribution at different velocities. Figure 6 Part (c) shows the variation of the contact pressure distribution of the micro-protrusions at different speeds. Figure 6 Part (d) shows the variation of lubricating oil film thickness distribution at different speeds. For example... Figure 6 As shown in (a)-(d), with the increase of speed, the rotational speed of the rotating shaft increases from 0 m / s to 6 m / s. The temperature in the sealing contact area increases linearly, reaching a maximum of 71℃, while the lubricating oil viscosity decreases non-linearly, from 0.08 Pas to 0.025 Pas. In the sealing contact area, the fluid pressure and lubricating oil film thickness continuously increase, but the increasing trend gradually decreases. The contact pressure of the micro-protrusion continuously decreases, and the slowing trend gradually weakens, indicating that the fluid pumping effect gradually strengthens. Figure 7 The variation of total friction force at different speeds is shown in part (a). As the speed increases, the total friction force in the sealed contact area continuously decreases, as shown in part (a). Figure 7 The variation of circumferential deformation at different speeds is shown in part (b) of the diagram, and the corresponding circumferential deformation continuously decreases. For example... Figure 7 The variation of the reverse pumping rate at different speeds is shown in section (c). When the main drive seal of the tunnel boring machine is stationary, the reverse pumping rate is negative (-8.5 mm). 3 At speeds of 6 m / s, sediment will enter the lubricating oil side. As the speed increases, the reverse pumping rate gradually increases; for example, when the speed increases to 6 m / s, the reverse pumping rate is 220 mm. 3 The speed of friction torque is approximately 0.5 m / s, mainly due to the axial position of the maximum circumferential deformation being closer to the sediment side. This aligns with the reverse pumping mechanism proposed by Qian and KM et al., which states that the closer the axial position of the maximum circumferential deformation is to the sediment side, the better the pumping capacity. The variation of friction torque with speed is shown in the figure below. Figure 7 The variation of friction torque at different speeds is shown in section (d). The friction torque first increases sharply, then gradually decreases, and finally remains constant. For example, at a speed of 3 m / s, the friction torque is 79304 Nm. Friction torque mainly includes viscous friction torque and micro-protrusion contact friction torque. Figure 7 Part (e) shows the variation of the contact friction force of the micro-protrusions at different speeds. Figure 7 Part (f) shows the variation of viscous friction at different speeds. For example... Figure 7 As shown in (e)-(f), at low speeds, the micro-protrusion contact friction dominates, causing a momentary increase in frictional torque. As speed increases, viscous friction gradually increases, while the micro-protrusion contact friction decreases rapidly, resulting in a decrease in total friction. When the speed increases to 2 m / s, the total friction remains essentially constant, and the corresponding frictional torque also remains constant.

[0193] 4. The effect of surface roughness on sealing performance.

[0194] In this application, Figure 8 Part (a) shows the variation of fluid pressure distribution under different roughness conditions. Figure 8 Part (b) shows the variation of the contact pressure distribution of the micro-protrusions under different roughness conditions. Figure 8 Part (c) shows the variation of lubricating oil film thickness distribution under different roughness conditions. Figure 8 Part (d) in the figure shows the variation of circumferential deformation under different roughnesses. For example... Figure 8 As shown in (a)-(d), keeping other parameters constant, with increasing roughness, the fluid pressure and lubricating oil film thickness gradually decrease, while the micro-protrusion contact pressure and circumferential deformation gradually increase, indicating that the reverse pumping effect gradually weakens. Furthermore, the axial position of the maximum circumferential deformation is closer to the silt side and gradually moves away from the silt side with increasing roughness, further illustrating the decrease in the reverse pumping rate according to the reverse pumping theory. However, as... Figure 9 As shown in section (c), the back pumping rate increases non-linearly with increasing surface roughness. This is mainly because the axial position of the maximum circumferential deformation is closer to the mud and sand side, and the sealing system always has back pumping capability. The increase in flow rate caused by the increase in surface roughness leads to an increase in the back pumping rate. For example, the back pumping rate increases from 74 mm... 3 / s (0.8 micrometers) increased to 248mm 3 / s (2 micrometers). Figure 9 Part (a) shows the variation of viscous friction force distribution under different roughness conditions. Figure 9 Part (b) shows the variation of the contact friction force distribution of micro-protrusions under different roughness conditions. Figure 9 Part (c) shows the variation of reverse pumping rate under different roughness conditions. Figure 9 Part (d) in the figure shows the variation of friction torque under different roughnesses. For example... Figure 9 As shown in (a)-(b), the viscous friction gradually decreases, while the contact friction of the micro-protrusions continuously increases. Figure 9 As shown in section (d), the friction torque exhibits a non-linear increasing trend with increasing surface roughness, mainly due to the increased contact pressure of the micro-protrusions caused by the increased roughness. For example, the friction torque increases from 73212 Nm (0.8 μm) to 98171 Nm (2 μm).

[0195] 5. The effect of pressure difference on sealing performance.

[0196] As the pressure difference between the oil side and the silt side increases (0.1MPa-0.2MPa), the contact length increases significantly (1.68mm-4.9mm), and the maximum static contact pressure gradually decreases (4.95MPa-3.66MPa).

[0197] Figure 10 Part (a) shows the variation of viscous friction and micro-protrusion contact friction under different pressure differentials. Figure 10 Part (b) shows the variation of total frictional force under different pressure differences. Figure 10 Part (c) shows the viscosity variation of lubricating oil under the same pressure difference. Figure 10 Part (d) in the diagram shows the variation of frictional torque under different pressure differentials. For example... Figure 10 As shown in (a)-(d), with the increase of pressure difference, the viscous friction and micro-protrusion contact friction in the sealing contact area gradually increase, corresponding to an increase in total friction. The increase in total friction causes the temperature in the sealing contact area to rise (47.6℃-60.4℃), which in turn causes the viscosity of the lubricating oil to decrease (0.047Pas-0.034Pas) and the friction torque to increase (79304Nm-124307Nm).

[0198] Figure 11 Part (a) shows the variation of fluid pressure distribution and micro-protrusion contact pressure distribution under different pressure differentials. Figure 11 Part (b) shows the variation of lubricating oil film thickness under different pressure differentials. Figure 11 Part (c) shows the variation of tangential deformation under different pressure differentials. Figure 11 Part (d) shows the variation of the reverse pumping rate under different pressure differentials. For example... Figure 11 As shown in (a)-(d), with the increase of pressure difference and contact length, the fluid pressure, micro-protrusion contact pressure, lubricating oil film thickness, and circumferential deformation change significantly. For example, as the fluid pressure amplitude gradually decreases, the micro-protrusion contact pressure amplitude first increases, then decreases, and then increases again; the minimum lubricating oil film thickness first decreases, then increases, and then decreases again; the circumferential deformation gradually increases; and the reverse pumping first increases, then decreases, and then increases again (0.1MPa-123mm). 3 / s, 0.12MPa-173mm 3 / s, 0.14MPa-126mm 3 / s, 0.16MPa-87mm 3 / s, 0.18MPa-53mm 3 / s, 0.2MPa-62mm 3 / s), with a pressure difference of 0.12MPa, the maximum back pumping rate is -173mm. 3 / s. Furthermore, increased pressure differential leads to increased frictional torque, exacerbating wear and causing changes in the contact length and static contact pressure of the sealing contact area, thus affecting sealing performance. Therefore, controlling the pressure differential within a certain range can effectively promote a sealing effect.

[0199] This application focuses on rotary lip seals and establishes a fluid-solid-thermal coupled hybrid lubrication numerical simulation model by combining solid mechanics, lubricating oil film fluid mechanics, contact mechanics of micro-protrusions on the sealing lip surface, normal and tangential elastic deformation of the sealing lip surface, and thermodynamic analysis. Quantitative analysis is performed on the sealing performance of the main drive seal of a tunnel boring machine under actual complex service conditions, such as different operating speeds, surface roughness, and large pressure differentials, including backflow rate and frictional torque. This application provides an effective way to predict the sealing performance of the main drive seal of a tunnel boring machine in advance, potentially avoiding seal failure and saving economic costs.

[0200] Based on the same inventive concept, this application also provides a device for predicting the main drive sealing performance of a tunnel boring machine under actual complex service conditions, used to implement the aforementioned method for predicting the main drive sealing performance under actual complex service conditions. The solution provided by this device is similar to the solution described in the above method. Therefore, the specific limitations of one or more embodiments of the device for predicting the main drive sealing performance of a tunnel boring machine under actual complex service conditions provided below can be found in the limitations of the method for predicting the main drive sealing performance of a tunnel boring machine under actual complex service conditions described above, and will not be repeated here.

[0201] In an exemplary embodiment, a device for predicting the sealing performance of the main drive of a tunnel boring machine under actual complex service conditions is provided, comprising: a simulation module, an initialization module, a flow factor determination module, a Reynolds equation solving module, a contact model solving module, a micro-elastic deformation determination module, an oil film thickness updating module, a comparison module, a replacement module, a temperature calculation module, a circulation module, and a sealing performance determination module.

[0202] The simulation module is used to obtain the static contact pressure and contact length under the current pressure difference based on the two-dimensional axisymmetric finite element model of the rotating lip seal of the tunnel boring machine's main drive using finite element simulation; the pressure difference is the pressure difference between the oil side and the silt side. The initialization module is used to set the initial value of the oil film thickness. The flow factor determination module is used to determine the flow factor based on the initial value of the oil film thickness. The Reynolds equation solving module is used to solve the discrete Reynolds equation using the alternating hermit algorithm based on the flow factor and the contact length under the current pressure difference, obtaining the fluid pressure, pump-back rate, and viscous friction. The contact model solving module is used to solve the GW contact model using micro-forehead contact mechanics, obtaining the micro-forehead contact pressure and micro-forehead contact friction. The micro-elastic deformation determination module is used to obtain the micro-elastic deformation based on the fluid pressure, micro-forehead contact pressure, static contact pressure under the current pressure difference, and contact length under the current pressure difference, using the line contact elastic deformation equation of elastic half-space. The oil film thickness update module is used to obtain a new oil film thickness based on microelastic deformation, fluid pressure, micro-protrusion contact pressure, and static contact pressure under the current pressure difference. The comparison module compares the new oil film thickness with the initial oil film thickness value to determine if the oil film thickness has converged. The replacement module replaces the initial oil film thickness value with the new oil film thickness if convergence does not occur, and calls the flow factor determination module. The temperature calculation module, if convergence occurs, obtains the sealing contact area temperature using thermodynamic analysis based on the micro-protrusion contact pressure. The circulation module replaces the initial oil film thickness value with the new oil film thickness if the sealing contact area temperature does not converge, and calls the flow factor determination module. The sealing performance determination module determines the friction torque based on the micro-protrusion contact friction force and viscous friction force when the sealing contact area temperature converges, and outputs the friction torque and back-pumping rate; the friction torque and the back-pumping rate characterize the sealing performance of the tunnel boring machine's main drive rotary lip seal.

[0203] The technical features of the above embodiments can be combined in any way. For the sake of brevity, not all possible combinations of the technical features in the above embodiments are described. However, as long as there is no contradiction in the combination of these technical features, they should be considered to be within the scope of this specification.

[0204] This document uses specific examples to illustrate the principles and implementation methods of this application. The descriptions of the above embodiments are only for the purpose of helping to understand the methods and core ideas of this application. Furthermore, those skilled in the art will recognize that, based on the ideas of this application, there will be changes in the specific implementation methods and application scope. Therefore, the content of this specification should not be construed as a limitation of this application.

Claims

1. A method for predicting the sealing performance of the main drive of a tunnel boring machine under actual complex service conditions, characterized in that, include: Based on the two-dimensional axisymmetric finite element model of the main drive rotary lip seal of the tunnel boring machine, the static contact pressure and contact length under the current pressure difference are obtained by finite element simulation; the pressure difference is the pressure difference between the oil side and the silt side. Set the initial value for oil film thickness; The flow factor is determined based on the initial value of the oil film thickness; Based on the flow factor and the contact length under the current pressure difference, the discrete Reynolds equation is solved using the alternating hermit algorithm to obtain the fluid pressure, pump-back ratio, and viscous friction. The GW contact model is solved using micro-assurance contact mechanics to obtain the micro-assurance contact pressure and micro-assurance contact friction. Based on the fluid pressure, micro-protrusion contact pressure, static contact pressure under the current pressure difference, and contact length under the current pressure difference, the micro-elastic deformation is obtained using the line contact elastic deformation equation of the elastic half-space. The new oil film thickness is obtained based on microelastic deformation, fluid pressure, micro-protrusion contact pressure, and static contact pressure under the current pressure difference. Compare the new oil film thickness with the initial oil film thickness to determine whether the oil film thickness has converged. If convergence fails, replace the initial value of the oil film thickness with the new oil film thickness and return to the step "Determine the flow factor based on the initial value of the oil film thickness"; If convergence occurs, the temperature of the sealed contact area can be obtained using thermodynamic analysis based on the contact pressure of the micro-protrusions. When the temperature in the sealing contact area does not converge, replace the initial value of the oil film thickness with the new oil film thickness and return to the step "Determine the flow factor based on the initial value of the oil film thickness"; When the temperature of the sealing contact area converges, the friction torque is determined based on the contact friction force and viscous friction force of the micro-protrusions, and the friction torque and the back pumping rate are output; the friction torque and the back pumping rate characterize the sealing performance of the main drive rotary lip seal of the tunnel boring machine.

2. The method for predicting the sealing performance of the main drive of a tunnel boring machine under actual complex service conditions as described in claim 1, characterized in that, Based on the two-dimensional axisymmetric finite element model of the main drive rotating lip seal of the tunnel boring machine, finite element simulation was used to obtain the static contact pressure and contact length under the current pressure difference, specifically including: The main drive rotary lip seal of the tunnel boring machine is simplified into a two-dimensional axisymmetric finite element model; the main drive rotary lip seal of the tunnel boring machine includes a single lip seal ring and a rotating shaft, the single lip seal ring is in contact with the rotating shaft, and the material of the single lip seal ring is nitrile rubber; A multi-region method was used to mesh the two-dimensional axisymmetric finite element model; The uniaxial tensile test data of nitrile rubber were fitted using the Yeoh-3rd order model; Based on the two-dimensional axisymmetric finite element model after meshing and the fitted uniaxial tensile test data of nitrile rubber, the static contact pressure and contact length under the current pressure difference are obtained by finite element simulation.

3. The method for predicting the sealing performance of the main drive of a tunnel boring machine under actual complex service conditions as described in claim 1, characterized in that, The formula for determining the flow factor is: In the formula, This represents the circumferential pressure-flow factor. denoted by axial pressure-flow factor, H represents dimensionless oil film thickness, D′ and r represent fitting constants, and γ represents surface morphology parameters; This represents the shear flow factor at any angle in the x-direction. Let θ represent the shear flow factor at any angle in the y-direction, and let θ represent the angle between the micro-protrusions on the sealing lip surface and the circumferential coordinate system. V represents the shear flow factor. r1 and V r2 γ represents the variance ratio, γ1 represents the surface topography parameter of the rotating shaft, γ2 represents the surface topography parameter of the sealing lip, and Φ represents the variance ratio. s The functions represent γ1, γ2, and H, σ1 and σ2 represent the surface roughness of the rotating shaft and the sealing lip, respectively, σ represents the root mean square roughness, and A2, A3, α4, α5, and α6 all represent fitting constants.

4. The method for predicting the sealing performance of the main drive of a tunnel boring machine under actual complex service conditions as described in claim 1, characterized in that, Based on the flow factor and the contact length under the current pressure difference, the discrete Reynolds equation is solved using the alternating hermit algorithm to obtain the fluid pressure, pump-back ratio, and viscous friction force, specifically including: By introducing the flow factor and cavitation factor into the Reynolds equation, we obtain the Reynolds equation after the introduction: In the formula, H represents the dimensionless oil film thickness, h represents the oil film thickness, σ represents the root mean square roughness; F represents the cavitation factor; Φ represents the general variable. X, Y, ξ represents intermediate parameters, x represents the circumferential direction, and y represents the axial direction. Indicates axial dimensionless, L x L represents the circumferential length. y Indicates the contact length. The average clearance is represented by erf, the error function by μ, the lubricating oil viscosity by U, and the relative velocity between the rotating shaft and the sealing lip by P. ref P represents the reference pressure. cav Indicates cavitation pressure; and These represent the pressure-flow factor at any angle. and These represent the shear flow factors at arbitrary angles. Density flux factor at any angle; θ represents the angle between the micro-protrusions on the sealing lip surface and the circumferential coordinate system. This represents the circumferential pressure-flow factor. Indicates the axial pressure-flow factor. denoted by density flow factor, and a2 and b2 both represent fitting constants; Discretize the introduced Reynolds equation using the finite volume method to obtain the discrete Reynolds equation; Based on the flow factor and the contact length under the current pressure difference, the discrete Reynolds equation is solved using the alternating hermit algorithm to obtain the fluid pressure, pump-back ratio, and viscous friction.

5. The method for predicting the sealing performance of the main drive of a tunnel boring machine under actual complex service conditions as described in claim 1, characterized in that, The GW contact model is solved using micro-assurance contact mechanics to obtain the micro-assurance contact pressure and micro-assurance contact friction, specifically including: Define the GW contact model as follows: In the formula, P c The contact pressure of the micro-protrusion is represented by E′, the equivalent elastic modulus is E′, the equivalent density of the micro-protrusion is η, R is the radius of the rotation axis, and H represents the oil film thickness; z = H / 3, where z represents one-third of the oil film thickness. The GW contact model is solved using micro-protrusion contact mechanics to obtain the micro-protrusion contact pressure. Based on the contact pressure of the micro-protrusions, using the formula F c =∫∫ A fP c dxdy, calculates the contact friction force of the micro-convex body; where F c Let f be the contact friction force of the micro-protrusion, f be the friction coefficient between the rotating shaft and the sealing lip, and A be the solution domain.

6. The method for predicting the sealing performance of the main drive of a tunnel boring machine under actual complex service conditions as described in claim 1, characterized in that, Based on the fluid pressure, micro-convexity contact pressure, static contact pressure under the current pressure difference, and contact length under the current pressure difference, the micro-elastic deformation is obtained using the line contact elastic deformation equation of the elastic half-space, specifically including: Based on the fluid pressure, the micro-protrusion contact pressure, and the static contact pressure under the current pressure difference, the formula ΔP = P is used. f +P c -P sc Calculate the pressure difference between the supporting force and the static contact pressure; the supporting force consists of fluid pressure and micro-protrusion contact pressure; where ΔP represents the pressure difference between the supporting force and the static contact pressure, P f P represents fluid pressure. c P represents the contact pressure of the micro-convexity. sc Indicates static contact pressure; Based on the pressure difference between the supporting force and the static contact pressure, the formula is used. Calculate the normal deformation; where H def Denotes normal deformation, E′ is the equivalent elastic modulus, and L y represents the contact length, y represents the axial direction, and s represents the integral variable; Based on the contact length under the current pressure difference, use the formula Calculate the tangential deformation; where τ def Indicates tangential deformation, L x τ represents the circumferential length. avg τ represents viscous shear stress. c This represents the contact stress of the micro-protrusion, where x represents the circumferential direction; The normal deformation and tangential deformation together constitute microelastic deformation.

7. The method for predicting the sealing performance of the main drive of a tunnel boring machine under actual complex service conditions as described in claim 1, characterized in that, Based on the contact pressure of the micro-protrusions, the temperature of the sealing contact area is obtained using thermodynamic analysis, specifically including: Based on the contact pressure of the micro-protrusion, using the formula τ c =fP c Calculate the contact stress of the micro-convex body; where τ c P represents the contact stress of the micro-protrusion, f is the coefficient of friction between the rotating shaft and the sealing lip, and P is the coefficient of friction between the rotating shaft and the sealing lip. c Indicates the contact pressure of the micro-protrusion; Based on the contact stress of the micro-protrusions, the temperature of the sealed contact area is obtained through thermodynamic equations.

8. The method for predicting the sealing performance of the main drive of a tunnel boring machine under actual complex service conditions according to claim 1, characterized in that, The formula for determining the friction torque is: T n =F all ·R; F all =F l +F c ; In the formula, T n F represents frictional torque. all F represents the total frictional force, R is the radius of the axis of rotation, and F is the total frictional force. l F represents viscous friction. c This represents the contact friction force of micro-protrusions.

9. The method for predicting the sealing performance of the main drive of a tunnel boring machine under actual complex service conditions according to claim 1, characterized in that, If convergence occurs, the temperature of the sealed contact area is obtained using thermodynamic analysis based on the contact pressure of the micro-protrusions. This process then includes: Based on the temperature of the sealed contact area, the Reynolds viscosity-temperature equation μ(T)=μ0exp[-α(TT) is used. ref )] Calculate the viscosity of the lubricating oil at the temperature of the sealed contact area; In the formula, μ(T) represents the viscosity of the lubricating oil at a temperature of T in the sealing contact area, μ0 represents the viscosity of the lubricating oil at a temperature of T0 in the sealing contact area, α represents the rate of change of the lubricating oil viscosity with temperature, and T ref Indicates ambient temperature.

10. A device for predicting the sealing performance of the main drive of a tunnel boring machine under actual complex service conditions, characterized in that, The device for predicting the sealing performance of the main drive of the tunnel boring machine under actual complex service conditions includes: The simulation module is used to obtain the static contact pressure and contact length under the current pressure difference based on the two-dimensional axisymmetric finite element model of the rotating lip seal of the main drive of the tunnel boring machine; the pressure difference is the pressure difference between the oil side and the silt side. The initialization module is used to set the initial value of the oil film thickness; A flow factor determination module is used to determine the flow factor based on the initial value of the oil film thickness; The Reynolds equation solving module is used to solve the discrete Reynolds equation using the alternating hermit algorithm based on the flow factor and the contact length under the current pressure difference, to obtain the fluid pressure, pump-back ratio and viscous friction. The contact model solution module is used to solve the GW contact model using micro-protrusion contact mechanics to obtain the micro-protrusion contact pressure and micro-protrusion contact friction. The micro-elastic deformation determination module is used to obtain the micro-elastic deformation based on the fluid pressure, micro-protrusion contact pressure, static contact pressure under the current pressure difference, and contact length under the current pressure difference, using the line contact elastic deformation equation of the elastic half-space. The oil film thickness update module is used to obtain a new oil film thickness based on micro-elastic deformation, fluid pressure, micro-protrusion contact pressure and static contact pressure under the current pressure difference. The comparison module is used to compare the new oil film thickness with the initial oil film thickness to determine whether the oil film thickness has converged. The replacement module is used to replace the initial value of the oil film thickness with a new oil film thickness and call the flow factor determination module if convergence fails. The temperature calculation module is used to obtain the temperature of the sealed contact area based on the contact pressure of the micro-protrusions using thermodynamic analysis if convergence is achieved. The circulation module is used to replace the initial value of the oil film thickness with a new oil film thickness and call the flow factor determination module when the temperature of the sealing contact area does not converge. The sealing performance determination module is used to determine the friction torque based on the contact friction force and viscous friction force of the micro-protrusions when the temperature of the sealing contact area converges, and outputs the friction torque and the back pumping rate; the friction torque and the back pumping rate characterize the sealing performance of the main drive rotary lip seal of the tunnel boring machine.

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