Double-eccentric C-shaped wear-resistant ball valve and sealing pair structure optimization method thereof

Abstract

The invention discloses a double-eccentric C-shaped wear-resistant ball valve and a sealing pair structure optimization method.The double-eccentric C-shaped wear-resistant ball valve comprises a valve body, a C-shaped ball body arranged in the valve body, a valve rod connected with the ball body and a valve seat assembly arranged in the valve.The double-eccentric C-shaped wear-resistant ball valve is matched with a fixed ball and a floating valve seat through double-eccentric kinematics, so that a sealing pair is separated more easily in the opening stage, and the service life of the sealing pair is prolonged. And the relative friction of the metal sealing surface in the whole stroke is reduced. The belleville spring arranged behind the valve seat provides stable pre-tightening and forms a superposition sealing effect with medium pressure, continuity of sealing contact can be kept easily when the pressure changes, meanwhile, the graphite ring and the graphite pressing ring are arranged at the position of a sealing pair, abrasion-resistant surface treatment is adopted, and the adaptability of the sealing pair in the particle-containing medium and high-temperature environment is improved. By carrying out thermosetting coupling analysis and parameter optimization on key structure sizes of the valve seat and the ball body, the contact state can be improved under different temperature and pressure combinations, so that the structural design is more targeted and interpretable.

Description

Technical Field

[0001] This invention relates to the field of industrial valve technology, specifically to a double-eccentric C-type wear-resistant ball valve and a method for optimizing the structure of its sealing pair. Background Technology

[0002] In coal chemical, papermaking, slurry, and other slurry transportation applications, the media often exhibit characteristics such as high temperature, high pressure, and entrainment of solid particles or fibers. Valves are frequently opened and closed and subject to significant erosion and wear. Conventional soft-seal ball valves are prone to problems such as aging of the sealing material, cold flow, and ablation at high temperatures, leading to seal failure and internal leakage. While metal hard-seal ball valves can withstand temperature and pressure, the relative sliding friction between the metal sealing surfaces during opening and closing can easily cause scratches, seizing, increased torque, and accelerated wear, thus shortening their lifespan and increasing maintenance costs.

[0003] To reduce opening and closing friction and improve sealing reliability, eccentric structures are typically used in engineering to allow the sealing pair to disengage quickly during the initial opening phase, thereby reducing friction throughout the entire stroke. Simultaneously, pre-tightening with elastic elements and self-tightening under medium pressure are employed to establish and maintain the sealing specific pressure. However, under high-temperature conditions, thermal expansion and material softening alter the contact state of the sealing pair, leading to insufficient local specific pressure or the creation of opening gaps, increasing the risk of high-temperature internal leakage. Furthermore, critical dimensions of the sealing pair, matching of sealing components, and surface wear-resistant treatment significantly impact sealing performance. Without systematic thermo-solid coupling analysis and parameter optimization, insufficient design margins or overly conservative structures can easily result.

[0004] Therefore, there is an urgent need for a ball valve structure suitable for high temperature, high pressure and solid media conditions. This structure should achieve low friction through eccentric opening and closing, ensure sealing pressure through pre-tightening and media self-tightening, and enhance the wear resistance of the sealing pair by combining high temperature sealing components and wear-resistant surface treatment. At the same time, it should provide a method for optimizing the structural parameters of the sealing pair for thermo-solid coupling conditions to obtain stable and reliable sealing performance under different temperature and pressure conditions.

[0005] A search revealed a Chinese patent document disclosing a forced-sealing valve seat [Application No.: 202110934031.4, Publication No.: CN113586754B]. While the comparative patent achieves ball valve sealing and reduces leakage risk through the valve seat structure, featuring the arrangement of metal and non-metal sealing rings on the valve seat support and a forced-sealing structure at the valve body and valve seat assembly sealing point, this invention is based on the double-eccentric kinematics of a double-eccentric C-type ball valve and a floating valve seat sealing pair scheme driven by disc spring preload. It also optimizes the design of key dimensional parameters of the valve seat and ball by combining sealing pressure calculation, finite element simulation, and algorithms. This provides a comprehensive design path and structural combination for sealing pairs under high temperature, high pressure, wear-resistant, and particulate-containing conditions, a solution not found in the comparative patent. Summary of the Invention

[0006] To address the problems existing in the prior art, the purpose of this invention is to provide a double eccentric C-type wear-resistant ball valve and a method for optimizing the sealing pair structure.

[0007] A double-eccentric C-type wear-resistant ball valve includes a valve body, a C-shaped ball disposed within the valve body, a valve stem connected to the ball, and a valve seat assembly disposed within the valve body. The ball is a fixed ball structure, supported and positioned by an upper bearing and a lower bearing; the valve seat assembly is a floating valve seat assembly, capable of slight axial displacement relative to the valve body; a disc spring is disposed behind the valve seat to apply preload; a graphite ring and a graphite pressure ring are disposed between the valve seat and the disc spring; a pressure plate is disposed between the valve seat and the valve body; the rotation center of the ball is offset radially by a first eccentricity and axially by a second eccentricity relative to the centerline of the valve seat sealing surface; the sealing surface of the ball is coated with a Ni60 surface layer.

[0008] Preferably, the thickness of the Ni60 surface coating is 0.13mm-0.30mm.

[0009] Through the above technical solution, the present invention introduces a wear-resistant surface layer in the key contact area of ​​the sealing pair, making the sealing surface less prone to surface damage under conditions such as opening and closing friction, media scouring and particle entrainment, thereby helping to maintain the integrity of the sealing strip and the stability of the contact state.

[0010] Specifically, the Ni60 coating can form a relatively dense and hard surface structure on the spherical sealing surface, mitigating wear caused by micro-protrusion shearing and abrasive plowing, and reducing the probability of scratches or seizing on the metal substrate surface. By setting the coating thickness within the above-mentioned range, a relatively balanced effect can be achieved between wear resistance, bonding reliability, and processing controllability.

[0011] In practical applications, this coating thickness range is more suitable for high temperature, high pressure and media containing solid particles. Especially when valves are frequently opened and closed or are in a small opening adjustment condition for a long time, it can reduce the risk of the sealing surface condition fluctuating with the running time, and make it easier to achieve more predictable maintenance cycle and maintain sealing performance.

[0012] Preferably, the disc spring is made of Inconel X-750, and the disc spring has a thickness of 8mm, a compression of 5mm, an outer diameter of 340mm, and an inner diameter of 310mm.

[0013] Through the above technical solution, the present invention provides continuous pre-tightening behind the valve seat by using a high-temperature resistant elastic element, so that the valve can still establish initial sealing contact when the medium pressure is low or the pressure fluctuates, and maintain the availability of elastic output under heating conditions, thereby improving the adaptability of the sealing pair to changes in operating conditions.

[0014] Specifically, using Inconel X-750 helps maintain the stability of the mechanical properties of the elastic element at higher temperatures and reduces preload decay caused by thermal softening. Combined with the aforementioned thickness, compression, and geometry, the disc spring can provide a suitable load curve within the allowable assembly stroke, matching the slight axial movement of the valve seat with the preload output.

[0015] In practical applications, this flexible pre-tightening configuration is more suitable for high-temperature opening and closing or high-temperature pressure holding operation modes. It can slow down the change in the sealing contact state caused by thermal deformation during the process of the valve heating from cold to hot, and reduce the possibility of leakage in the low-pressure stage.

[0016] Preferably, the geometric parameters of the valve seat sealing pair include: a channel diameter of 303 mm, an outer diameter of the sealing surface of 330.8 mm, an inner diameter of the sealing surface of 324.2 mm, a ball radius of 256 mm, a sealing surface included angle of 39.12°, an outer diameter of the graphite ring of 342 mm, and a radial width of the sealing surface of 3.3 mm.

[0017] Through the above technical solution, the present invention clearly defines the key geometric relationship of the sealing pair, making it easier to control the position, width and normal loading path of the contact zone between the ball sealing surface and the valve seat sealing surface, which helps to maintain a consistent geometric benchmark in the process of structural design, processing and assembly and subsequent verification.

[0018] Specifically, the aforementioned parameters collectively determine the effective area of ​​the sealing strip, the contact angle, and the fit between the valve seat and the ball, thereby affecting the establishment of the sealing specific pressure and the stress distribution in the contact area. By combining these geometric parameters with valve seat floating, spring preload, and media self-tightening, the stress and contact state of the sealing pair can be made more interpretable, facilitating thermo-mechanical coupling analysis and structural optimization.

[0019] In practical applications, a well-defined geometric parameter system is beneficial for engineering implementation, especially when mass production and on-site maintenance and replacement of valve seat components are required. It can reduce the dispersion of sealing performance caused by the accumulation of geometric deviations and improve the repeatability of different batches of products under similar operating conditions.

[0020] Preferably, the valve seat has a five-step structure, and the radial thickness parameters H1-H5 and axial length parameters L1-L5 of the valve seat are both in millimeters, and the contact angle α between the valve seat and the sealing ring is in degrees; wherein: H1 is 4mm-18mm, H2 is 6mm-20mm, H3 is 3mm-15mm, H4 is 7mm-15mm, and H5 is 2mm-10mm; L1 is 4mm-20mm, L2 is 1mm-14mm, L3 is 10mm-25mm, L4 is 5mm-15mm, and L5 is 5mm-15mm. α is 10°-60°.

[0021] Through the above technical solution, the present invention abstracts the valve seat structure into a parameterizable step size and contact angle system, so that the influencing factors of the valve seat in terms of strength, stiffness, thermal deformation path and sealing contact zone position can be clearly expressed, thereby facilitating systematic size optimization and sensitivity analysis.

[0022] Specifically, H-type parameters mainly affect the radial thickness distribution and local support stiffness of the valve seat, L-type parameters mainly affect the axial structural length and force transmission path of the valve seat, and the angle parameter α affects the normal component of the sealing surface and the formation mode of the contact zone. By setting the units and value ranges of the above parameters, multivariate optimization can be carried out while ensuring the manufacturability and assemblability of the structure, and the infeasible design points caused by the lack of parameter constraints can be reduced.

[0023] In practical applications, this parametric valve seat structure facilitates rapid adjustment of the design scheme for different temperatures, pressures, and media abrasion conditions. It can also directly map the optimization results to the machining drawings and inspection elements, reducing the number of trial production iterations and improving the controllability of design changes.

[0024] Preferably, the C-shaped ball has a V-shaped opening, and the root of the V-shaped opening has a rounded transition; the ball has a thickened area in the weak part near the V-shaped opening; the front end of the valve seat has a continuous circumferential cutting edge, and the cutting edge is opposite to the V-shaped opening when the ball valve is closed.

[0025] Through the above technical solution, the present invention introduces a shear geometry structure in the mating area between the ball and the valve seat, so that the valve has a certain shearing and scraping effect on the fibers and particle agglomerates in the medium during the closing process, thereby helping to reduce the risk of local lifting of the sealing surface due to foreign matter clamping, and improving the self-cleaning ability of the sealing pair.

[0026] Specifically, the V-shaped port structure provides a discharge and cutting channel for the medium. The relative fit between the valve seat cutting edge and the V-shaped port in the closed position makes it easier to cut off or remove the entrained medium from the sealing zone. The rounded transition at the root of the V-shaped port and the thickened area in the weak part can reduce the stress concentration caused by geometric abrupt changes, making it less likely for the ball to form unfavorable high-stress areas in localized areas when subjected to torque and contact loads.

[0027] In practical applications, this structure is more suitable for working conditions containing solid particles and fibrous media, such as coal chemical slurry or papermaking black liquor. By reducing the retention of foreign objects at the sealing strip, the possibility of abnormal fluctuations in opening and closing torque and sudden changes in sealing state can be reduced, improving the maintainability of the valve in long-term operation.

[0028] A method for optimizing the sealing pair structure of a double-eccentric C-type wear-resistant ball valve, characterized by comprising the following steps: (1) Establish a three-dimensional model of the ball valve and sealing pair, and establish a thermo-solid coupling finite element analysis model of the sealing pair; (2) Under both normal temperature and high pressure and high temperature and high pressure conditions, analyze the influence of manufacturing process factors and dimensional parameters on sealing performance. The dimensional parameters include the radius of the sphere and the radial width of the sealing surface. (3) The valve seat size was optimized using the response surface optimization method. The radial thickness parameters H1-H5, axial length parameters L1-L5, and the contact angle α between the valve seat and the sealing ring were selected as optimization variables. (4) Thicken the weak parts of the C-shaped sphere and optimize the fillet at the root of the V-shaped opening; (5) Couple the valve seat size optimization result of step (3) with the ball structure optimization result of step (4) and perform the final analysis under the two working conditions.

[0029] Preferably, in step (3), a central composite design is used to design the test points, and a multi-objective genetic algorithm is used for multi-objective optimization; in step (1), the thermo-solid coupling finite element analysis adopts a sequential coupling method, first performing steady-state thermal analysis to obtain the temperature field, and then importing the temperature field into the static analysis; in step (2), the friction coefficient of the stainless steel contact surface between the valve seat and the ball is set to 0.15, and the friction coefficient of the graphite sealing ring contact surface is set to 0.30; in step (2), the thickness of the Ni60 surface coating of the ball is taken in the range of 0.13mm-0.30mm and the influence is evaluated.

[0030] Through the above technical solutions, this invention introduces standardized experimental point design and multi-objective optimization strategies into the optimization process, making the exploration of multi-parameter space more systematic, enabling more comprehensive coverage of design variable combinations with a limited number of calculations, and reducing blind spots caused by adjustments based solely on experience.

[0031] Specifically, the central composite design facilitates the establishment of a fitting relationship between variables and responses, while the multi-objective genetic algorithm can seek compromise solutions among multiple performance indicators. The sequential coupling approach, by obtaining the steady-state temperature field before structural analysis, makes the heat load transfer path clearer. Incorporating the friction coefficient and coating thickness as influencing factors in the evaluation helps to compare the effects of material properties and contact behavior on the sealing state within the same framework.

[0032] In practical applications, the above optimization strategies are suitable for rapid evaluation during product iteration or operating condition expansion. The calculated sensitive parameters and key combinations can be used to guide the setting of design changes and quality control points, thereby improving the consistency and interpretability of the sealing pair performance under different batches and operating conditions.

[0033] Compared with the prior art, the present invention has the following advantages: 1. This invention utilizes double-eccentric kinematics in conjunction with a fixed ball and a floating valve seat to facilitate easier disengagement of the sealing pair during the opening phase, reducing relative friction between the metal sealing surfaces throughout the entire stroke. A rear-mounted disc spring on the valve seat provides stable preload and, together with the medium pressure, creates a superimposed sealing effect, which helps maintain continuous sealing contact during pressure changes.

[0034] 2. This invention incorporates a graphite ring and a graphite pressure ring at the sealing joint, and employs a wear-resistant surface treatment to improve the adaptability of the sealing joint to particulate media and high-temperature environments. Through thermo-mechanical coupling analysis and parameter optimization of the key structural dimensions of the valve seat and ball, the contact state can be improved under different temperature and pressure combinations, making the structural design more targeted and interpretable. Attached Figure Description

[0035] Figure 1 This is the overall structural design drawing of the double eccentric C-type wear-resistant ball valve of the present invention; Figure 2 This is a schematic diagram of the valve body structure of the present invention; Figure 3 This is a schematic diagram of the valve seat structure of the present invention; Figure 4 This is a schematic diagram of the spherical structure of the present invention; Figure 5 This is a schematic diagram of the valve stem structure of the present invention; Figure 6 This is a partially enlarged schematic diagram of the sealing pair structure of the present invention; Figure 7 This is a simplified diagram of the forces acting on the floating valve seat of the present invention; Figure 8 This is a schematic diagram of the shell test boundary condition setting of the present invention; Figure 9 This is a schematic diagram of the shell test simulation cloud map of the present invention; Figure 10 This is a schematic diagram of the boundary condition setting for the high-pressure sealing test of the present invention; Figure 11 This is a schematic diagram of the simulated high-pressure sealing test intensity cloud map of the present invention; Figure 12 This is a schematic diagram of the contact cloud map of the sealing surface in the high-pressure sealing test simulation of the present invention; Figure 13 This is a schematic diagram of the deformation cloud of the valve body under normal working conditions according to the present invention; Figure 14 This is a schematic diagram of the deformation cloud diagram of the sealing pair under normal working conditions of the present invention; Figure 15 This is a schematic diagram of the sealing condition of the sealing pair under normal temperature conditions after the response surface optimization of this invention; Figure 16 This is a schematic diagram of the sealing condition of the sealing pair under normal temperature conditions after the structural optimization of this invention; Figure 17 This is a schematic diagram of the sealing condition of the sealing pair under normal operating conditions after the final coupling optimization of this invention.

[0036] In the diagram: 1. Upper valve stem; 2. Valve body; 3. Disc spring; 4. Graphite pressure ring; 5. Graphite ring; 6. Valve seat; 7. Pressure plate; 8. Upper valve cover; 9. Ball; 10. Lower valve stem. Detailed Implementation

[0037] The technical solutions of the embodiments of the present invention will be clearly and completely described below with reference to the accompanying drawings. Obviously, the described embodiments are only some embodiments of the present invention, and not all embodiments. Based on the embodiments of the present invention, all other embodiments obtained by those skilled in the art without creative effort are within the scope of protection of the present invention.

[0038] The present invention will be described in detail below through embodiments and comparative examples. These embodiments are intended to further illustrate the technical solutions and beneficial effects of the present invention, and are not intended to limit the scope of protection of the present invention.

[0039] Example 1 (1) Overview of Structure and Function like Figure 1 As shown, the double eccentric C-type wear-resistant ball valve of this embodiment mainly includes components such as a drive unit, valve stem, ball, valve body, and valve seat. The valve seat is used to establish a reliable sealing boundary and acts as a bottom cutter to cooperate with the ball crown with a V-shaped notch, achieving shearing and self-cleaning against media containing solid particles and fibers during the closing process. The partial structure of the sealing pair is as follows. Figure 6 As shown, a graphite ring and a graphite pressure ring are installed between the valve seat and the disc spring to achieve a seal; a pressure plate is installed between the valve seat and the valve body to achieve assembly constraint and sealing fit. The valve seat assembly can produce a slight axial displacement relative to the valve body, so that the elastic preload and the self-tightening of the medium pressure form a superimposed sealing effect.

[0040] (2) Design conditions and basic parameters (2.1) The design requirements are shown in Table 1; Table 1

[0041] (2.2) The main geometric parameters of the ball valve are shown in Table 2; Table 2

[0042] (3) Strength design and verification: (3.1) Calculation and determination of minimum wall thickness of valve body Equation (1): ; Equation (2): ; After rearranging the two formulas above, we can obtain: Equation (3): ;in: Equation (4): ; In the formula: Sb is the wall thickness including the additional margin; Sb′ is the wall thickness calculated based on strength; C is the additional margin; Dft is the maximum diameter of the valve body cavity; [σ] is the allowable stress of the material; P is the design pressure.

[0043] In this embodiment, the valve body material selected is ASTM A216 WCB, with a yield strength of 290.00 MPa and a tensile strength of 485.00 MPa. The safety factors are 2.30 and 4.25, respectively. Based on this, the allowable stress [σ] of the material is taken as the smaller of the two converted values.

[0044] In this embodiment, the test pressure of the shell is used for verification. P = 22.50 MPa (1.5 times the nominal pressure of 15.0 MPa). K0 is calculated by equation (4), and the wall thickness Sb is calculated by equations (1) to (3). Considering the processing and margin requirements in the structural design, the minimum wall thickness of the valve body is finally determined to be 45.00 mm. The valve body structure is as follows: Figure 2 As shown.

[0045] (3.2) Determining the size of the sphere The radius of the sphere is determined using an empirical formula: Equation (5): ; In the formula: R is the radius of the sphere; d is the diameter of the sphere's channel hole.

[0046] In this embodiment, the sphere radius was initially selected based on the aforementioned empirical relationships, and a geometric check was performed according to the inner and outer diameters of the sealing surface in this design to ensure that the sphere diameter is greater than the minimum check value; the final sphere diameter was determined to be 520mm, and the sphere structure is shown in [reference needed]. Figure 4 .

[0047] (3.3) Valve stem torsional shear strength check Valve stem verification is performed using: Equation (6): ; In the formula: Mm is the comprehensive maximum operating torque; W is the torsional resistance coefficient; [τN] is the allowable torsional stress of the valve stem material.

[0048] The parameters for this embodiment, calculated as Mm = 11,978,938.00 N·mm, and based on the valve stem structural dimension drawing, W = 61,949.6 mm³. This embodiment uses 17-4PH valve stems, with an allowable [τN] of 209.00 MPa. The calculated τN = 193.37 MPa meets the strength requirements. The valve stem structure is shown below. Figure 5 .

[0049] (4) Sealing force, sealing pressure and required sealing pressure (4.1) Calculation of sealing force like Figure 7 As shown, the total force exerted by the inlet valve seat on the ball consists of three parts: Equation (7): ; The thrust of the medium fluid on the valve seat assembly: Equation (8): ; Approximate calculation of disc spring clamping force: Equation (9): ; Where the correction factor is: Equation (10): ; Diameter ratio: Equation (11): ; Valve seat seal preload: Equation (12): ; In this embodiment, by substituting P=16.50 MPa and the dimensional parameters listed in Table 2, we obtain FZJ=88728.9N.

[0050] The disc spring material is Inconel X-750, with a thickness of 8 mm, a compression of 5 mm, a Poisson's ratio of 0.30, an outer diameter of 340 mm, and an inner diameter of 310 mm. The FTH is calculated to be 6823.5 N using equations (9) to (11).

[0051] Taking qMYmin=1.6, FMY=5432.25N is obtained by calculation using formula (12).

[0052] According to equation (7), FMZ = 100984.65N is obtained.

[0053] (4.2) Criteria for sealing specific pressure and required sealing specific pressure Sealing pressure criterion: Equation (13): ; The sealing specific pressure qb must be determined using an empirical formula: ; The coefficients of a and c in the formula are shown in Table 3. Table 3

[0054] The allowable sealing pressure q in the formula is shown in Table 4; Table 4

[0055] (4.3) Calculation of qb and q In this embodiment, the required sealing pressure qb is calculated according to formula (14), and the calculated q b =8.57MPa.

[0056] In this embodiment, the sealing pressure q is calculated based on the normal force of the sealing surface and the area of ​​the sealing strip, and the calculated value is q=8.97MPa.

[0057] The allowable specific pressure of the sealing material is selected as q = 80 MPa, which satisfies the criterion shown in equation (13).

[0058] (5) 3D modeling and finite element simulation settings (5.1) Three-dimensional model In this embodiment, a three-dimensional model of the ball valve is established and imported into finite element software for analysis.

[0059] (5.2) Material parameters The material parameters of the main components of the ball valve are shown in Table 5. Table 5

[0060] (5.3) Contact settings The ball valve contact pairs and contact type settings are shown in Table 6. Table 6

[0061] In this embodiment, the friction coefficient of the stainless steel contact surface between the valve seat and the ball is set to 0.15, and the friction coefficient of the graphite sealing ring contact surface is set to 0.30. The variation of the friction coefficient value will be studied in the subsequent influence analysis.

[0062] (5.4) Mesh Generation The valve body uses a tetrahedral mesh with an overall mesh size of 20.00 mm, 214,787 nodes, 142,997 elements, and a mesh quality of 0.83.

[0063] In the sealing pair structure, the ball adopts an 8mm volume mesh, the valve seat adopts a 4mm volume mesh, the contact area is densified and the contact surface is finely divided, the number of nodes is 637225, the number of elements is 436273, and the mesh quality is 0.84.

[0064] (6) Simulation verification of shell test and high pressure sealing test (6.1) Shell test simulation The shell test pressure is 22.50 MPa, and the boundary conditions are as follows: Figure 8 See the simulation cloud map. Figure 9Simulation results show that the maximum deformation of the valve body is 0.22109 mm and the maximum stress is 227.82 MPa, which is less than the yield strength of the valve body material of 290 MPa.

[0065] (6.2) Simulation of high-pressure sealing test The sealing test pressure is 16.50 MPa, and the boundary conditions are as follows: Figure 10 Intensity cloud map (see) Figure 11 See the contact cloud diagram of the sealing surface. Figure 12 Simulation results show that the maximum total deformation of the ball is 0.3656 mm, and the maximum total deformation of the valve seat is 0.3343 mm. The maximum equivalent stress occurs at the root of the contact between the ball and the valve stem, at 655.82 MPa. A continuous annular high-pressure band is formed on the sealing surface, with a maximum contact specific pressure of 102.32 MPa and a minimum gap of -0.0096492 mm.

[0066] (7) Performance analysis under temperature and pressure conditions (valve body and sealing pair) (7.1) Determination of heat transfer boundary and heat transfer coefficient The heat transfer coefficients are shown in Table 7. Table 7

[0067] In this embodiment, the convective heat transfer coefficient of the valve surface is taken as 10 W / m²·K.

[0068] (7.2) Valve body pressure and temperature ratings and deformation results Table 8 shows the pressure and temperature rating data for WCB carbon steel. Table 8

[0069] The CL900 has a value of 15.32 at -29°C to 38°C; 13.98 at 100°C; 13.14 at 200°C; and 11.95 at 300°C.

[0070] Deformation contour maps of the valve body under operating conditions of normal temperature, 100℃, 200℃, and 300℃ are shown below. Figure 13 The maximum deformations were 0.1558 mm, 0.9829 mm, 2.0914 mm and 3.1982 mm, respectively.

[0071] (7.3) Sealing pair pressure and temperature rating and deformation results Table 9 shows the pressure and temperature rating data for stainless steel. Table 9

[0072] The CL900 has a value of 14.89 at -29°C to 38°C; 12.66 at 100°C; 10.70 at 200°C; and 9.49 at 300°C.

[0073] Deformation contour maps of the sealing pair under operating conditions of normal temperature, 100℃, 200℃ and 300℃ are shown below. Figure 14 The maximum deformations were 0.3215 mm, 0.3328 mm, 0.6817 mm and 1.0904 mm, respectively.

[0074] (8) Influence of manufacturing process factors Analysis in this embodiment shows that reducing the roughness of the sealing surface reduces the opening of the sealing pair contact surface; applying a Ni60 surface coating to the spherical sealing surface can improve the wear resistance of the sealing pair. The impact of the Ni60 coating thickness was evaluated within the range of 0.13 mm to 0.30 mm.

[0075] (9) Analysis and optimization of the influence of dimensional parameters (9.1) Trend of influence of valve seat size Table 10 shows the trend data of the influence of valve seat size H-type parameters. Table 10

[0076] Note: - indicates decrease, + indicates increase.

[0077] Table 11 shows the trend data of the influence of L-type parameters on valve seat size. Table 11

[0078] Note: - indicates decrease, + indicates increase.

[0079] (9.2) Optimize the parameter value range The original values ​​and ranges of the optimized parameters are shown in Table 12. Table 12

[0080] (9.2) Optimize the result data The sealing condition of the sealing pair under normal temperature conditions after response surface optimization is shown in the figure. Figure 15 The maximum deformation is 0.29619 mm, the minimum gap is -0.000045425 mm, and the maximum permeability is 0.0014984 mm.

[0081] The sealing condition of the sealing pair under normal temperature conditions after structural optimization is shown in the figure. Figure 16 The maximum deformation is 0.31031 mm, and the minimum gap is -0.0018891 mm.

[0082] The final coupling optimization results for the sealing pair under ambient temperature conditions are shown in the figure. Figure 17 The maximum deformation is 0.27059 mm, the minimum gap is -0.0000040783 mm, and the maximum permeability is 0.00135 mm.

[0083] This embodiment completes the design and verification of the ball valve structure based on given operating conditions and parameters, and analyzes the deformation behavior of the valve body and sealing pair under different temperature conditions through finite element simulation, including shell test, high pressure sealing test, and dimensional parameter sensitivity analysis. At the same time, it verifies and improves the sealing reliability of the sealing pair under normal temperature and high temperature conditions by combining the influence of manufacturing process factors, dimensional parameter sensitivity analysis and optimization results.

[0084] The above description is merely a preferred embodiment of the present invention, but the scope of protection of the present invention is not limited thereto. Any equivalent substitutions or modifications made by those skilled in the art within the scope of the technology disclosed in the present invention, based on the technical solution and inventive concept of the present invention, should be covered within the scope of protection of the present invention.

Claims

1. A double eccentric C-type wear-resistant ball valve, comprising a valve body, a C-shaped ball disposed within the valve body, a valve stem connected to the ball, and a valve seat assembly disposed within the valve body, characterized in that: The ball is a fixed ball structure, supported and positioned by an upper bearing and a lower bearing; the valve seat assembly is a floating valve seat assembly, which can generate a slight displacement relative to the valve body along the axial direction; a disc spring is set behind the valve seat to apply a preload to the valve seat; a graphite ring and a graphite pressure ring are set between the valve seat and the disc spring; a pressure plate is set between the valve seat and the valve body; the rotation center of the ball is offset by a first eccentricity in the radial direction and a second eccentricity in the axial direction relative to the center line of the valve seat sealing surface; the sealing surface of the ball is coated with Ni60.

2. The double eccentric C-type wear-resistant ball valve according to claim 1, characterized in that: The thickness of the Ni60 surface coating is 0.13mm-0.30mm.

3. The double eccentric C-type wear-resistant ball valve according to claim 1, characterized in that: The disc spring is made of Inconel X-750, and has a thickness of 8mm, a compression of 5mm, an outer diameter of 340mm, and an inner diameter of 310mm.

4. The double eccentric C-type wear-resistant ball valve according to claim 1, characterized in that: The geometric parameters of the valve seat sealing pair include: a channel diameter of 303 mm, an outer diameter of the sealing surface of 330.8 mm, an inner diameter of the sealing surface of 324.2 mm, a ball radius of 256 mm, an included angle of the sealing surface of 39.12°, an outer diameter of the graphite ring of 342 mm, and a radial width of the sealing surface of 3.3 mm.

5. The double eccentric C-type wear-resistant ball valve according to claim 1, characterized in that: The valve seat has a five-step structure. The radial thickness parameters H1-H5 and axial length parameters L1-L5 of the valve seat are both in millimeters, and the contact angle α between the valve seat and the sealing ring is in degrees. Among them: H1 is 4mm-18mm, H2 is 6mm-20mm, H3 is 3mm-15mm, H4 is 7mm-15mm, and H5 is 2mm-10mm. L1 is 4mm-20mm, L2 is 1mm-14mm, L3 is 10mm-25mm, L4 is 5mm-15mm, and L5 is 5mm-15mm. α is 10°-60°.

6. The double eccentric C-type wear-resistant ball valve according to claim 1, characterized in that: The C-shaped ball has a V-shaped opening, and the root of the V-shaped opening has a rounded transition; the ball has a thickened area in the weak part near the V-shaped opening; the front end of the valve seat has a continuous circumferential cutting edge, and the cutting edge is opposite to the V-shaped opening when the ball valve is closed.

7. A method for optimizing the sealing pair structure of the double eccentric C-type wear-resistant ball valve according to any one of claims 1-6, characterized in that, Includes the following steps: (1) Establish a three-dimensional model of the ball valve and sealing pair, and establish a thermo-solid coupling finite element analysis model of the sealing pair; (2) Under both normal temperature and high pressure and high temperature and high pressure conditions, analyze the influence of manufacturing process factors and dimensional parameters on sealing performance. The dimensional parameters include the radius of the sphere and the radial width of the sealing surface. (3) The valve seat size was optimized using the response surface optimization method. The radial thickness parameters H1-H5, axial length parameters L1-L5, and the contact angle α between the valve seat and the sealing ring were selected as optimization variables. (4) Thicken the weak parts of the C-shaped sphere and optimize the fillet at the root of the V-shaped opening; (5) Couple the valve seat size optimization result of step (3) with the ball structure optimization result of step (4) and perform the final analysis under the two working conditions.

8. The method for optimizing the sealing pair structure according to claim 7, characterized in that: In step (3), a central composite design is used to design the test points, and a multi-objective genetic algorithm is used for multi-objective optimization. In step (1), the thermo-solid coupling finite element analysis adopts a sequential coupling method. First, a steady-state thermal analysis is performed to obtain the temperature field, and then the temperature field is imported into the static analysis. In step (2), the friction coefficient of the stainless steel contact surface between the valve seat and the ball is set to 0.15, and the friction coefficient of the graphite sealing ring contact surface is set to 0.

30. In step (2), the thickness of the Ni60 coating on the surface of the ball is taken in the range of 0.13mm-0.30mm and the influence is evaluated.

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