A steam turbine thrust bearing bush failure damage optimization method and system based on lubrication and load bearing characteristic analysis

Abstract

This application discloses a method and system for optimizing the failure and damage of turbine thrust bearings based on lubrication and load-bearing characteristic analysis, relating to the field of mechanical lubrication. The method includes: establishing a thermo-fluid-structure interaction finite element model; deriving the actual axial load borne by the bearing by measuring the center displacement of the elastic gasket during operation; substituting the actual load into the model and performing lubrication and load-bearing characteristic analysis under typical operating conditions to determine the dominant factors of lubrication degradation; conducting parametric simulations of axial load and lubricating oil inlet temperature based on the dominant factors, summarizing the degradation law of increased load and temperature, and formulating optimization strategies to suppress the failure of turbine thrust bearings. This application improves the targeting and efficiency of turbine thrust bearing failure and damage optimization, and enhances the bearing's overload resistance.

Description

Technical Field

[0001] This application relates to the field of mechanical lubrication, and in particular to a method and system for optimizing the failure and damage of turbine thrust bearings based on lubrication and load-bearing characteristics analysis. Background Technology

[0002] As a core power source in industries such as power generation and chemical engineering, the operational stability of steam turbines is of paramount importance. Thrust bearings are key load-bearing components of the turbine rotor system, bearing the crucial responsibility of balancing the enormous steam thrust during operation and maintaining the rotor's axial positioning. Their performance directly affects the safe and long-term stable operation of the entire unit. Failure of the thrust bearing can lead to anything from minor issues like unit vibration and abnormal shaft displacement to more serious consequences such as severe wear of moving and stationary components or even catastrophic accidents, resulting in significant economic losses.

[0003] In some cases, during power plant overhauls, the working surface of turbine thrust bearing pads may become severely blackened, with surface cracks and Babbitt alloy material detachment. This may also be accompanied by turbine shaft displacement exceeding limits and causing tripping incidents during unit commissioning. Many factors influence the blackening / yellowing of thrust bearing pads, including load-bearing capacity, lubricating oil temperature / viscosity, parallelism, and unit start-up and shutdown processes.

[0004] Current technologies lack quantitative research on the lubrication and load-bearing characteristics of turbine thrust bearings, making it impossible to identify key factors contributing to premature bearing degradation and providing no clear direction for improvement. On one hand, conventional analyses often rely on design loads, leading to distorted load inputs and analytical conclusions that are disconnected from actual operating conditions, resulting in a lack of specificity. On the other hand, analytical methods typically fail to fully consider the coupling effects of multiple physical fields, making it difficult to accurately identify factors causing bearing surface degradation (such as blackening and cracking), thus obscuring subsequent optimization directions and compromising the effectiveness of measures. Summary of the Invention

[0005] The purpose of this application is to provide a method and system for optimizing the failure and damage of turbine thrust bearings based on lubrication and load-bearing characteristics analysis, which can improve the pertinence and efficiency of turbine thrust bearing failure and damage optimization, thereby improving the bearing's overload resistance.

[0006] To achieve the above objectives, this application provides the following solution: Firstly, this application provides an optimization method for the failure and damage of turbine thrust bearing bushes based on lubrication and load-bearing characteristic analysis. The method includes: constructing a thermo-fluid-structure interaction (TFI) finite element model based on the turbulent modified Reynolds equation, energy equation, and solid deformation equation; measuring the center displacement of the elastic washer ring of the failed turbine thrust bearing bush during operation to obtain single-bearing bush displacement data; the failure and damage include discoloration and / or cracking on the surface of the turbine thrust bearing bush; performing stiffness analysis on the elastic washer ring based on the single-bearing bush displacement data to inversely estimate the actual axial load borne by the turbine thrust bearing bush; and substituting the actual axial load into the TFI finite element model. Finite element analysis was performed on the lubrication and load-bearing characteristics of turbine thrust bearings under typical operating conditions. Based on these characteristics, the dominant factors for turbine thrust bearing failure were determined. When the dominant factors were axial load and inlet oil temperature, the influence of oil film characteristic indicators was analyzed through parametric simulation to obtain the degradation law of increased load and temperature. The oil film characteristic indicators included: minimum oil film thickness, maximum oil film pressure, and temperature field distribution. The degradation law of increased load and temperature was found to be: lubrication load and inlet oil temperature are inversely proportional to oil film thickness and directly proportional to bearing temperature. Based on the dominant factors and the degradation law of increased load and temperature, an optimization strategy was formulated with the goal of suppressing failure.

[0007] Secondly, this application also provides a computer system, including: a memory, a processor, and a computer program stored in the memory and executable on the processor, wherein the processor executes the computer program to implement the turbine thrust bearing failure and damage optimization method and system based on lubrication and load-bearing characteristic analysis described in the first aspect.

[0008] According to the specific embodiments provided in this application, the following technical effects are disclosed: This application first constructs a thermo-fluid-structure interaction finite element model based on the turbulent modified Reynolds equation, energy equation, and solid deformation equation. Then, it measures the displacement of the operating center of the elastic bearing ring in a failed bearing bush (e.g., surface discoloration or cracking) to obtain single-bearing displacement data, and uses this data to infer the actual axial load, achieving precise load quantification and improving the analytical focus. Substituting the actual load into the model, finite element analysis of lubrication and load-bearing characteristics is performed under typical operating conditions to efficiently identify the dominant failure factors. When the dominant factors are axial load and inlet oil temperature, the influence of oil film characteristics is analyzed through parametric simulation to determine the law of deterioration due to increased load and temperature: lubrication load and inlet oil temperature are inversely proportional to oil film thickness and directly proportional to bearing bush temperature. Based on the dominant factors and deterioration laws, strategies such as adjusting load distribution or controlling inlet oil temperature are formulated with the optimization goal of suppressing failure damage, thereby preventing faults, improving the bearing's overload resistance and operational reliability, and achieving extended life and performance optimization of the turbine thrust bearing bush. Attached Figure Description

[0009] 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.

[0010] Figure 1 This is a flowchart illustrating an optimization method for turbine thrust bearing failure based on lubrication and load-bearing characteristics analysis, provided as an embodiment of this application.

[0011] Figure 2 A schematic diagram of a failed or damaged turbine thrust bearing provided in an embodiment of this application.

[0012] Figure 3 A schematic diagram of the turbine thrust bearing structure provided in the embodiments of this application.

[0013] Figure 4 This is a schematic diagram of the elastic pad ring support structure provided in an embodiment of this application.

[0014] Figure 5 This is a schematic diagram of the load-bearing characteristic analysis results provided in an embodiment of this application.

[0015] Figure 6 Typical bearing oil film characteristic diagrams under typical operating conditions provided in embodiments of this application; wherein, Figure 6 In the diagram, 'a' represents the liquid film pressure analysis results of a typical working condition tile. Figure 6 In the diagram, b represents the liquid film thickness analysis results of a typical working condition tile. Figure 6 In the diagram, 'c' represents the temperature analysis results of a typical working tile. Figure 6 In the diagram, d represents the total elastic deformation analysis results of a typical working condition tile.

[0016] Figure 7 The minimum oil film thickness versus load curve provided in this application embodiment is shown in the figure.

[0017] Figure 8 The maximum oil film pressure versus load curve provided in this application embodiment.

[0018] Figure 9 The bearing temperature versus load curve provided in this application embodiment.

[0019] Figure 10 The friction coefficient versus load curve provided in this application embodiment is shown in the figure.

[0020] Figure 11 The power consumption versus load curve provided in this application embodiment.

[0021] Figure 12 The flow rate versus load curve provided in this application embodiment.

[0022] Figure 13 The elastic deformation versus load curve provided in the embodiments of this application.

[0023] Figure 14 The axial stiffness versus load curve provided in this application embodiment is shown.

[0024] Figure 15 The graph showing the relationship between film thickness and oil inlet temperature is provided for the embodiments of this application.

[0025] Figure 16 The maximum molding pressure versus oil inlet temperature curve is provided for the embodiments of this application.

[0026] Figure 17 The graph showing the relationship between bearing temperature and oil inlet temperature is provided for the embodiments of this application.

[0027] Figure 18 The graph showing the relationship between the friction coefficient and the oil inlet temperature is provided for the embodiments of this application.

[0028] Figure 19 The power consumption versus oil inlet temperature curve is provided for the embodiments of this application.

[0029] Figure 20 The graph showing the relationship between flow rate and inlet oil temperature is provided for the embodiments of this application.

[0030] Figure 21 The curve showing the relationship between elastic deformation and oil inlet temperature is provided for the embodiments of this application.

[0031] Figure 22 The graph showing the relationship between axial stiffness and oil inlet temperature is provided for the embodiments of this application.

[0032] Figure 23 This is an internal structure diagram of a computer system provided in an embodiment of this application. Detailed Implementation

[0033] 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.

[0034] 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.

[0035] Example 1, such as Figure 1 As shown in the figure, this embodiment provides an optimization method for turbine thrust bearing failure damage based on lubrication and load-bearing characteristic analysis. The method includes: S1. A thermo-fluid-structure interaction finite element model is constructed based on the turbulent modified Reynolds equation, energy equation, and solid deformation equation.

[0036] Furthermore, step S1 specifically includes: S11. The turbulence-corrected Reynolds equation is used as the flow field control equation, the energy equation is used as the temperature field control equation, and the solid elastic deformation equation is used as the structural field control equation.

[0037] Furthermore, the expression for the turbulence-corrected Reynolds equation is as follows: .

[0038] .

[0039] In the formula, Gx and Gy are both turbulence coefficients; x and y are the circumferential and radial directions of the bearing oil film, respectively; ρ is the density; h is the oil film thickness; p is the pressure; and μ is the fluid viscosity.

[0040] S12. Define the multiphysics coupling relationships between the flow field, temperature field, and structure field; the multiphysics coupling relationships include at least the coupling relationship between the flow field and the temperature field, the coupling relationship between the temperature field and the structure field, and the coupling relationship between the structure field and the flow field.

[0041] In practical applications, the thermo-elasto-fluidic coupling analysis of thrust bearings involves a complex multi-field coupling problem encompassing flow, thermal, and solid fields. This includes calculations of heat transfer within the lubricating oil, thermal structures within the bearing pads and mirror plates, and coupling calculations between the fluid and solid domains. This embodiment first introduces the main performance parameters of the thrust bearing under typical operating conditions, including oil film pressure field, oil film temperature field, bearing pad strain field, and temperature field analysis. Secondly, it explores the influence of changing parameters on the bearing lubrication characteristics.

[0042] 1) Typical working conditions Table 1 lists the main input parameters for non-radial line-supported thrust bearings, including geometric parameters (such as inner and outer diameters, wrap angle, number of bearing pads, etc.), operating parameters (such as speed, load, etc.), and lubricating oil parameters (such as grade, temperature, etc.).

[0043] Table 1 Main Input Parameters of Non-Radial Linear Support Thrust Bearings

[0044] Table 2 shows the main performance parameters of non-radial line-supported thrust bearings, including minimum film thickness, maximum molding pressure, highest bearing temperature, maximum deformation, stiffness, and damping.

[0045] Table 2 Main performance parameters of non-radial line support thrust bearings

[0046] Under typical operating conditions, the lubrication characteristics of the turbine thrust bearing are as follows: Figure 6 As shown in the figure, the calculation results of the liquid film pressure, liquid film thickness, tile surface temperature, and total elastic deformation of the tile are mainly displayed. Among them, Figure 6 In the diagram, 'a' represents the liquid film pressure analysis results of a typical working condition tile. Figure 6 In the diagram, b represents the liquid film thickness analysis results of a typical working condition tile. Figure 6 In the diagram, 'c' represents the temperature analysis results of a typical working tile. Figure 6 In the diagram, d represents the total elastic deformation analysis results of a typical working condition tile.

[0047] When the thrust bearing is under a load of 1590 kN and a speed of 1500 rpm, the maximum oil film pressure is 13.83 MPa, the minimum oil film thickness is 0.0336 mm, and the highest bearing temperature is 94.6℃ on the oil outlet side of the bearing. The highest oil film temperature is 101℃, and the outlet temperature is 76.2℃. This is because the lubricating oil temperature outside the outlet is lower, and convection reduces the lubricating oil temperature at the outlet. The maximum total elastic deformation of the bearing is 50.4 μm.

[0048] Under three-dimensional thermal-fluid-solid analysis, the surface temperature of the tile first increases and then decreases with the increase of the circumferential angle, reaching its maximum at approximately 19°, with a maximum temperature of 94.6℃. The liquid film thickness gradually decreases with the increase of the circumferential angle, but shows an increasing trend when the circumferential angle is 19°, with a minimum liquid film thickness of 33.6μm. The liquid film pressure first increases and then decreases with the increase of the circumferential angle, reaching its maximum at approximately 14°, with a maximum of 13.83MPa, and then gradually decreases with the increase of the circumferential angle, with zero oil film pressure at the edge of the tile.

[0049] Under three-dimensional thermo-fluid-solid-matter calculation conditions, the liquid film pressure first increases and then decreases with increasing radial position, reaching a maximum of 13.83 MPa at a radial position of approximately 420 mm. The tile surface temperature first increases and then decreases with increasing radial position, reaching a maximum of 94.6℃ at a radial position of approximately 430 mm, and then gradually decreases with increasing radial position. The liquid film thickness first decreases and then increases with increasing radial position.

[0050] In the two-dimensional liquid film pressure distribution of a single tile, the tile surface temperature distribution shows that the maximum oil film pressure occurs in the center of the tile, with a maximum value of 13.83 MPa. There is basically no pressure around the tile. The highest temperature on the tile surface occurs on the side of the oil outlet of the tile, with a maximum temperature of 94.6℃.

[0051] In the oil film and the three-dimensional and cross-sectional temperature field distributions of a single bearing, the highest temperature of the oil film is 100.6℃, which has a significant impact along its thickness. The highest temperature of the bearing is 94.6℃, occurring on the oil outlet side of the bearing.

[0052] The temperature distribution diagram and deformation diagram of the bearing pad along the temperature direction show that the highest temperature of the bearing pad is 94.6℃, occurring on the oil outlet side. The deformation diagram shows that the deformation of the bearing pad gradually increases along the temperature rise direction, with a maximum deformation of 50.3 μm. Since the bearing pad is a rigid body, it can resist large external loads. Heat has a relatively small effect on the deformation of the bearing pad; under adiabatic conditions, the heat from the lubricating oil is transferred to the bearing pad surface, and the effect of thermal expansion and contraction on the bearing pad is not significant.

[0053] In the bearing temperature field, the highest temperature is 94.6℃, which occurs near the oil outlet. The temperature gradually increases along the circumferential direction. Due to the hot upper surface and cold lower surface of the bearing pad, the middle of the pad bulges outwards, forming an arched bridge structure. This reduces the bearing area, increases the maximum oil film pressure, and decreases the minimum oil film thickness. The bulge in the middle of the pad can easily lead to oil film rupture and reduce the bearing's load-bearing capacity.

[0054] In the three-dimensional thermal fluid solidification of the bearing, the amount of deformation in the thickness direction and elastic deformation of the bearing pads will affect the lubrication characteristics of the bearing, change the oil film thickness, and cause the location of the maximum oil film pressure to change.

[0055] S13. Construct a thermo-fluid-structure coupled finite element model based on the geometric model of the turbine thrust bearing and the multiphysics coupling relationship.

[0056] In practical applications, the construction process of each equation is as follows: 1) The expression for the turbulent modified Reynolds equation is as follows: .

[0057] .

[0058] .

[0059] .

[0060] Here is the expression for oil film thickness; G x , G yAll are turbulence coefficients, obtained by fitting data from traditional lubricating oil; and The value is the moment of inertia. It should be noted that this paper uses the classical turbulence factor, which is obtained by fitting traditional lubricating oil data. θ The circumferential position angle, ω Let ω be the rotational angular velocity. U For circumferential linear velocity, Re Represents the Reynolds number.

[0061] 2) Integrating the pressure obtained from solving the Reynolds equation over the journal surface yields the fluid force. Integrating the oil film pressure over the entire oil film plane calculates the total bearing capacity of the oil film, as follows: .

[0062] The formula for calculating the bearing tangential force is as follows: .

[0063] Bearing power consumption can be determined by frictional resistance and rotational speed, as follows: .

[0064] .

[0065] By differentiating the pressure along the radial and circumferential directions and combining this with specific flow field parameters, the bearing lubricating oil flow rate can be calculated. .

[0066] .

[0067] 3) Solving the temperature field To assess the heat generation of a bearing and measure the safety margin of its temperature rise, it is necessary to solve for the bearing temperature field to investigate the bearing oil film temperature distribution, prevent bearing surface burn-out, and calculate the overall average temperature rise of the bearing. .

[0068] By combining the Reynolds equation, the energy equation is simplified to two dimensions, and the bearing oil film temperature distribution is obtained by solving the two-dimensional energy equation: .

[0069] .

[0070] By conserving overall energy and measuring the impact of initial oil supply temperature on the overall heat source of the bearing, and considering the leakage of lubricating oil at different locations, the oil return ratio and side-discharge heat can be determined. .

[0071] To consider the three-dimensional temperature field distribution, and further the thermal conductivity and thermal deformation of solids, it is necessary to solve the three-dimensional energy equation. By expanding the three-dimensional energy equation and substituting the bearing heat generation, combined with the Reynolds equation and the solution results of the two-dimensional oil film temperature field, the three-dimensional temperature field distribution, including the bearing bush, is further analyzed: .

[0072] .

[0073] 4) Calculation of thermoelastic deformation Because high-load thrust bearings undergo a certain degree of elastic deformation, when solving for the characteristic parameters of thrust bearings under actual working conditions, it is necessary to analyze the specific working conditions to measure the degree of influence of thermoelastic deformation and see if this influencing factor can be ignored. For the bearing structure and working parameters of this project, thermoelastic deformation has a significant impact on bearing characteristics and cannot be ignored.

[0074] The three-dimensional solid elastic deformation equation is as follows: .

[0075] .

[0076] .

[0077] .

[0078] Matrix D can be represented in the following form: .

[0079] To account for the thermal conduction and thermal deformation of solids, the following three-dimensional energy equation for solids needs to be solved: .

[0080] Some of the parameters are as follows: .

[0081] .

[0082] 5) Analysis of bearing parameters and calculation conditions A certain power plant nuclear power plant has a saturated steam, intermediate reheat, impulse steam turbine with a total of 10 support bearings and 1 thrust bearing. The thrust bearing is located in the intermediate pressure bearing housing at the electric end of the high-pressure cylinder.

[0083] The thrust bearing is used to balance the axial thrust during unit operation. It is installed in the intermediate pressure bearing housing. When the unit is operating at normal power, the thrust of the high-pressure steam on the rotor is directed towards the generator end of the unit.

[0084] like Figures 3-4 As shown, the thrust pads deflect around the support line, establishing a wedge-shaped oil film between the thrust disk and the pads to maintain fluid friction and withstand axial loads. When the thrust pads are under stress, both the radial and circumferential oil films change. The thrust pads are supported by elastic washers, whose flexibility is suitable for absorbing thrust during normal operation.

[0085] Table 3 shows the main parameters of the thrust bearing. After calculation, the specific pressure of the thrust bearing is 2.8 MPa, the sliding speed v is 12.6 m / s, and the pv value is 35.25 MPa·m / s. All of these values ​​meet the performance requirements of the bearing materials used in the "Mechanical Design Handbook".

[0086] Table 3 Main parameters of thrust bearings

[0087] The thrust bearing in the geometric model of the thrust bearing bush and elastic washer is a non-radial support type. The bush is mounted on the elastic washer, and the edge of the connection surface is parallel to the edge of the bush rather than along the radial line. The load of the bush is transferred to the elastic washer supporting the bush and causes deformation of the elastic washer. Since the actual load of this bearing cannot be measured, the actual load of the thrust bearing is inferred by measuring the displacement of the elastic washer and performing stiffness analysis. Figure 5 For the load-bearing characteristics analysis of the elastic washer ring, the displacement at its center position is the largest and gradually decreases towards the periphery. After measuring the center displacement, the stiffness analysis of the elastic washer ring can be performed to obtain the actual load on the bearing bush.

[0088] The bearing load analysis results are shown in Table 4. Under several working conditions, the actual working load of the designed thrust bearing is greater than the design calculation load. Subsequent calculations and analyses will use the actual working load.

[0089] Table 4 Bearing Load Analysis Results

[0090] S2. Measure the center displacement of the elastic washer ring of the failed or damaged turbine thrust bearing during operation to obtain single-bearing displacement data; such as... Figure 2 As shown, the failure damage includes: discoloration and / or cracking of the turbine thrust bearing surface.

[0091] Furthermore, step S2 specifically includes: S21. Determine the position of the thrust bearing and position the measuring point at the center of the elastic washer ring of the failed or damaged turbine thrust bearing bearing.

[0092] S22. Under different operating conditions of the unit, measure and record the displacement of the center position of the elastic gasket ring as the displacement data of a single gasket.

[0093] S3. Based on the displacement data of a single bearing, perform stiffness analysis on the elastic bearing to inversely estimate the axial load actually borne by the turbine thrust bearing.

[0094] Furthermore, step S3 specifically includes: S31. Perform finite element analysis on the geometric and material properties of the elastic washer of the failed turbine thrust bearing to establish the displacement-load stiffness relationship curve of the elastic washer.

[0095] S32. Substitute the displacement data of a single bearing into the stiffness relationship curve of the displacement-load of the elastic pad ring to deduce the actual axial load borne by a single bearing, and then summarize to obtain the actual axial load borne by the turbine thrust bearing.

[0096] S4. Substitute the actual axial load into the thermo-fluid-structure interaction finite element model, and perform finite element analysis on the lubrication and load-bearing characteristics of the turbine thrust bearing under typical working conditions.

[0097] Furthermore, step S4 specifically includes: S41. Substitute the actual axial load into the thermo-fluid-structure interaction finite element model.

[0098] S42. Set the high load condition, rated speed and specified lubricating oil inlet temperature as the typical operating conditions of the turbine thrust bearing.

[0099] S43. Under typical working conditions, the thermal-fluid-structure interaction finite element model is solved by finite element analysis to analyze the oil film pressure distribution, film thickness variation, temperature field and bearing deformation of the turbine thrust bearing.

[0100] S44. Based on the oil film pressure distribution, film thickness variation, temperature field and bearing deformation, analyze the lubrication characteristics and load-bearing characteristics of turbine thrust bearings.

[0101] S5. Determine the dominant factors for turbine thrust bearing failure based on lubrication and load characteristics.

[0102] Furthermore, step S5 specifically includes: S51. Based on lubrication and load characteristics, analyze the magnitude and direction of thermal and mechanical deformation of turbine thrust bearings and their influence on oil film shape. Compare and evaluate the contributions of thermal and mechanical deformation to lubrication performance, and determine the dominant deformation type and the mechanism of deterioration of lubrication.

[0103] S52. Correlation analysis is conducted between the dominant deformation type and the mechanism of deteriorated lubrication and the failure and damage state observed on site to determine the dominant factors of turbine thrust bearing failure and damage.

[0104] S6. When the dominant factors are axial load and oil inlet temperature, the influence of oil film characteristic indicators is analyzed by parametric simulation to obtain the law of deterioration due to increased load and temperature. The oil film characteristic indicators include: minimum oil film thickness, maximum oil film pressure and temperature field distribution. The law of deterioration due to increased load and temperature is: lubrication load and oil inlet temperature are inversely proportional to oil film thickness and directly proportional to bearing temperature.

[0105] Furthermore, step S6 specifically includes: S61. When the dominant factors are axial load and oil inlet temperature, select the axial load and oil inlet temperature, which directly affect the thermal load, as key variables according to the dominant factors, and set the range of variation.

[0106] S62. Based on typical working conditions, adjust key variables and perform multiple sets of thermal-fluid-structure interaction finite element simulation calculations to obtain simulation results.

[0107] S63. Analyze the simulation results, summarize the influence of oil film characteristic indicators, and obtain the law of deterioration due to increased load and temperature.

[0108] In practical applications, under typical operating conditions, the large temperature difference between the upper and lower surfaces of the thrust bearing pads causes the center of the pads to bulge upwards. Excessive thermal deformation reduces the bearing area of ​​the pads, worsening lubrication performance. Conversely, mechanical deformation caused by oil film pressure causes the center of the pads to concave downwards, resulting in more uniform oil film pressure and improved load-bearing capacity. Since thermal deformation is several times greater than mechanical deformation, overall, thermal deformation worsens the bearing's lubrication performance, increasing maximum oil film pressure, decreasing minimum oil film thickness, and raising maximum pad temperature. Parametric analysis shows that load and inlet oil temperature significantly affect key lubrication performance indicators, indicating poor overload resistance of the bearing.

[0109] In practical applications, the parametric analysis process is as follows: 1) Influence of Load: To investigate the influence of load on the turbine thrust bearing, theoretical calculations and analyses were performed for loads ranging from 900 to 2300 kN. The results are as follows. When the speed is 1500 rpm and the inlet oil temperature is 45℃, the calculation results for a load of 900 kN are as follows: the maximum oil film pressure is 7.12 MPa, located at the center of the bearing; the minimum oil film thickness is 0.0569 mm, located on the oil outlet side of the bearing; the highest bearing temperature is 72.9℃; the highest oil film temperature is 79.3℃; and the outlet temperature is 65.6℃. This is because the lubricating oil temperature outside the outlet is lower, and convection reduces the lubricating oil temperature at the outlet.

[0110] The relationship between minimum oil film thickness and load in non-radial line-supported thrust bearings is as follows: Figure 7As shown, the minimum oil film thickness gradually decreases with increasing load. This is because the increased external load raises the bearing's load-bearing requirements, and a smaller minimum oil film thickness results in a greater load-bearing capacity. The relationship between the maximum oil film pressure and load in a non-radial line-supported thrust bearing is shown in the figure. Figure 8 As shown, with increasing load, the pressure per unit area increases, the minimum oil film thickness decreases, and consequently the maximum oil film pressure increases.

[0111] In non-radial thrust bearings, the minimum oil film thickness gradually decreases, leading to increased bearing heat generation, elevated lubricating oil temperature, and consequently, a higher maximum oil film temperature. Figure 9 As shown, the maximum temperature of the bearing continues to increase, and the excessive load results in a thinner oil film, which will cause mixed lubrication and insufficient lubrication in areas with high roughness. Figure 10 The graph shows the relationship between the bearing friction coefficient and the load. It can be concluded from the graph that the friction coefficient gradually decreases as the load increases.

[0112] The relationship between power consumption of a non-radial line-supported thrust bearing and load is as follows: Figure 11 As shown, with increasing load, more heat is generated by friction between bearings, resulting in more work consumed. Increased load also leads to a thinner lubricating oil film, causing greater losses. Conversely, decreased load increases the oil film thickness, thereby increasing the maximum oil film pressure, which in turn reduces the bearing clearance and decreases the lubricating oil flow rate. The relationship between bearing flow rate and load is shown in the figure. Figure 12 As shown, with the increase of load, the leakage flow on the main bearing side decreases, and the inlet flow decreases.

[0113] Figures 13-14 The relationship between elastic deformation and axial stiffness of a non-radial line-supported thrust bearing and load variation. Figure 13 It can be seen that the maximum thermal deformation of the bearing gradually increases with increasing load, while the maximum force deformation and maximum total deformation of the bearing gradually decrease with increasing load. From... Figure 14 It can be seen that the axial stiffness of the bearing gradually increases with the increase of the load, and the increase is basically linear.

[0114] Load has a significant impact on the main lubrication performance indicators. As the load increases, the oil film thickness decreases, the oil film pressure increases, the power consumption increases, the flow rate decreases, the total deformation decreases, and the axial stiffness increases.

[0115] 2) The effect of oil inlet temperature In the calculation results of the non-radial line-supported thrust bearing under the conditions of 1590kN load, 1500rpm speed, and 35℃ oil inlet temperature, the maximum oil film pressure is 12.9MPa, the minimum oil film thickness is 0.0409mm, the highest pad temperature is 82.5℃, the highest oil film temperature is 88.4℃, and the outlet temperature is 66.4℃. This is because the lubricating oil temperature outside the outlet is lower, and the lubricating oil temperature at the outlet is reduced due to convection. The maximum total elastic deformation of the pad is 43.3μm.

[0116] The relationship between minimum oil film thickness and inlet oil temperature in non-radial line-supported thrust bearings is as follows: Figure 15 As shown, the minimum film thickness and the fulcrum film thickness gradually decrease with increasing inlet oil temperature. The relationship between the maximum film pressure of a non-radial line-supported thrust bearing and inlet oil temperature is as follows: Figure 16 As shown, with the increase of oil inlet temperature, the pressure per unit area increases, the minimum oil film thickness decreases, and the maximum film pressure increases.

[0117] In non-radial thrust bearings, the minimum oil film thickness gradually decreases, leading to increased bearing heat generation, elevated lubricating oil temperature, and consequently, a higher maximum oil film temperature. Figure 17 As shown, the maximum bearing temperature and outlet temperature increase continuously with the increase of the oil inlet temperature. Figure 18 The graph shows the relationship between the bearing friction coefficient and the oil inlet temperature. It can be concluded from the graph that the friction coefficient gradually decreases as the oil inlet temperature increases.

[0118] The relationship between power consumption of a non-radial line-supported thrust bearing and inlet oil temperature is as follows: Figure 19 As shown, as the inlet oil temperature increases, the lubricating oil viscosity decreases, reducing the frictional resistance of the lubricating oil on the bearing, and consequently decreasing the bearing power consumption. The relationship between bearing flow rate and inlet oil temperature is as follows: Figure 20 As shown, the relationship is basically linear; as the oil inlet temperature increases, the bearing inlet flow rate and side leakage flow rate decrease.

[0119] Figures 21-22 The relationship between elastic deformation and axial stiffness with oil inlet temperature is shown. Maximum force deformation and maximum thermal deformation gradually increase with increasing oil inlet temperature, while the maximum total deformation gradually decreases. The bearing axial stiffness gradually increases with increasing oil inlet temperature.

[0120] Oil inlet temperature has a significant impact on key lubrication performance indicators. Higher temperatures reduce oil film thickness, increase oil film pressure, decrease the coefficient of friction, reduce power consumption, decrease flow rate, and increase axial stiffness. To ensure safe operation, the oil inlet temperature should be controlled below 60℃.

[0121] S7. Based on the dominant factors and the degradation law of increased load and temperature, an optimization strategy is formulated with the goal of suppressing failure and damage.

[0122] Furthermore, the optimization strategies include: setting an upper limit for the oil inlet temperature during operation; improving the cooling design and thermal deformation resistance design of the turbine thrust bearing; and controlling the axial thrust to operate within a preset range.

[0123] Optionally, the upper limit of the oil inlet temperature is 60℃.

[0124] The technical effects of this application are as follows: This application first constructs a thermo-fluid-structure interaction finite element model based on the turbulent modified Reynolds equation, energy equation, and solid deformation equation. Then, it measures the displacement of the operating center of the elastic bearing ring in a failed bearing bush (e.g., surface discoloration or cracking) to obtain single-bearing displacement data, and uses this data to infer the actual axial load, achieving precise load quantification and improving the analytical focus. Substituting the actual load into the model, finite element analysis of lubrication and load-bearing characteristics is performed under typical operating conditions to efficiently identify the dominant failure factors. When the dominant factors are axial load and inlet oil temperature, the influence of oil film characteristics is analyzed through parametric simulation to determine the law of deterioration due to increased load and temperature: lubrication load and inlet oil temperature are inversely proportional to oil film thickness and directly proportional to bearing bush temperature. Based on the dominant factors and deterioration laws, strategies such as adjusting load distribution or controlling inlet oil temperature are formulated with the optimization goal of suppressing failure damage, thereby preventing faults, improving the bearing's overload resistance and operational reliability, and achieving extended life and performance optimization of the turbine thrust bearing bush.

[0125] Furthermore, this application uses the finite element method to analyze the lubrication characteristics of a turbine thrust bearing under typical operating conditions, identifies key parameters affecting bearing lubrication performance, assesses the safety of existing design parameters, and conducts parametric analysis of key parameters, load, and oil supply temperature affecting bearing performance. The results of the typical operating condition study indicate that bearing thermal deformation has a significant impact on bearing performance and may be the main cause of bearing performance degradation; therefore, the bearing should be optimized to reduce temperature. The parametric analysis results show that load has a significant impact on all major lubrication performance indicators: increased load reduces oil film thickness, increases oil film pressure, increases power consumption, decreases flow rate, reduces total deformation, and increases axial stiffness. Inlet oil temperature also has a significant impact on all major lubrication performance indicators: increased temperature reduces oil film thickness, increases oil film pressure, decreases the coefficient of friction, reduces power consumption, decreases flow rate, and increases axial stiffness. To ensure safe operation, inlet oil temperature should be strictly controlled.

[0126] Example 2: This example provides a computer system, which can be a server or a terminal, and its internal structure diagram can be as follows. Figure 23As shown, the computer system includes a processor, memory, input / output (I / O) interfaces, and a communication interface. The processor, memory, and I / O interfaces are connected via a system bus, and the communication interface is also connected to the system bus via the I / O interfaces. The processor provides computational and control capabilities. The memory includes non-volatile storage media and internal memory. The non-volatile storage media stores the operating system, computer programs, and a database. The internal memory provides the environment for the operating system and computer programs stored in the non-volatile storage media. The database stores forced oscillation samples and sub / supersynchronous oscillation samples. The I / O interfaces are used for information exchange between the processor and external devices. The communication interface is used for communication with external terminals via a network connection. When the computer program is executed by the processor, it implements the aforementioned method for rapid prediction and identification of the dominant frequency of sub / supersynchronous oscillations in new energy power systems based on transfer learning.

[0127] Those skilled in the art will understand that Figure 23 The structure shown is merely a block diagram of a portion of the structure related to the present application and does not constitute a limitation on the computer system to which the present application is applied. A specific computer system may include more or fewer components than those shown in the figure, or combine certain components, or have different component arrangements.

[0128] Those skilled in the art will understand that all or part of the processes in the methods of the above embodiments can be implemented by a computer program instructing related hardware. The computer program can be stored in a non-volatile computer-readable storage medium. When executed, the computer program can include the processes of the embodiments of the above methods. Any references to memory, databases, or other media used in the embodiments provided in this application can include at least one of non-volatile and volatile memory. Non-volatile memory can include read-only memory (ROM), magnetic tape, floppy disk, flash memory, optical memory, high-density embedded non-volatile memory, resistive random access memory (ReRAM), magnetic random access memory (MRAM), ferroelectric random access memory (FRAM), phase change memory (PCM), graphene memory, etc. Volatile memory can include random access memory (RAM) or external cache memory, etc. By way of illustration and not limitation, RAM can take many forms, such as Static Random Access Memory (SRAM) or Dynamic Random Access Memory (DRAM).

[0129] The databases involved in the embodiments provided in this application may include at least one type of relational database and non-relational database. Non-relational databases may include, but are not limited to, blockchain-based distributed databases. The processors involved in the embodiments provided in this application may be general-purpose processors, central processing units, graphics processing units, digital signal processors, programmable logic devices, quantum computing-based data processing logic devices, etc., and are not limited to these.

[0130] The various embodiments in this specification are described in a progressive manner, with each embodiment focusing on the differences from other embodiments. The same or similar parts between the various embodiments can be referred to each other.

[0131] 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 optimizing the failure and damage of turbine thrust bearings based on lubrication and load-bearing characteristics analysis, characterized in that, The method includes: A thermo-fluid-structure coupled finite element model is constructed based on the turbulence-modified Reynolds equation, energy equation, and solid deformation equation. The center displacement of the elastic washer ring of the failed or damaged turbine thrust bearing is measured during operation to obtain single bearing displacement data; the failure or damage includes: discoloration and / or cracking of the turbine thrust bearing surface; Stiffness analysis of the elastic bearing was performed based on single-bearing displacement data to infer the actual axial load borne by the turbine thrust bearing. Substituting the actual axial load into the thermo-fluid-structure interaction finite element model, the lubrication and load-bearing characteristics of the turbine thrust bearing are analyzed by finite element under typical working conditions. The dominant factors for turbine thrust bearing failure and damage were determined based on lubrication and load characteristics. When the dominant factors are axial load and oil inlet temperature, the influence of oil film characteristic indicators is analyzed through parametric simulation to obtain the law of temperature degradation under load. The oil film characteristic indicators include: minimum oil film thickness, maximum oil film pressure and temperature field distribution. The law of temperature degradation under load is: lubrication load and oil inlet temperature are inversely proportional to oil film thickness and directly proportional to bearing temperature. Based on the dominant factors and the degradation law of increased load and temperature, an optimization strategy is formulated with the goal of suppressing failure and damage.

2. The method for optimizing turbine thrust bearing failure and damage based on lubrication and load-bearing characteristics analysis according to claim 1, characterized in that, A thermo-fluid-structure interaction finite element model is constructed based on the turbulent modified Reynolds equation, energy equation, and solid deformation equation, specifically including: The turbulence-corrected Reynolds equation is used as the flow field control equation, the energy equation is used as the temperature field control equation, and the solid elastic deformation equation is used as the structural field control equation. Define the multiphysics coupling relationships between the flow field, temperature field, and structure field; the multiphysics coupling relationships include at least: the coupling relationship between the flow field and the temperature field, the coupling relationship between the temperature field and the structure field, and the coupling relationship between the structure field and the flow field; A thermo-fluid-structure coupled finite element model was constructed based on the geometric model of the turbine thrust bearing and the multiphysics coupling relationship.

3. The method for optimizing turbine thrust bearing failure and damage based on lubrication and load-bearing characteristics analysis according to claim 1, characterized in that, The center displacement of the elastic washer ring of the failed or damaged turbine thrust bearing was measured during operation to obtain single bearing displacement data, specifically including: Determine the position of the thrust bearing and locate the measurement point at the center of the elastic washer ring of the failed or damaged turbine thrust bearing bearing; Under different operating conditions of the unit, the displacement of the center position of the elastic gasket is measured and recorded as the displacement data of a single gasket.

4. The method for optimizing turbine thrust bearing failure and damage based on lubrication and load-bearing characteristics analysis according to claim 1, characterized in that, Stiffness analysis of the elastic bearing bush based on single-bearing displacement data is performed to infer the actual axial load borne by the turbine thrust bearing bush, specifically including: Finite element analysis was performed on the geometric and material properties of the elastic washer of the failed turbine thrust bearing to establish the displacement-load stiffness relationship curve of the elastic washer. By substituting the displacement data of a single bearing into the displacement-load stiffness relationship curve of the elastic bearing ring, the actual axial load borne by a single bearing can be deduced, and then the actual axial load borne by the turbine thrust bearing can be obtained by summing them up.

5. The method for optimizing turbine thrust bearing failure and damage based on lubrication and load-bearing characteristics analysis according to claim 1, characterized in that, Substituting the actual axial load into the thermo-fluid-structure interaction finite element model, finite element analysis was performed on the lubrication and load-bearing characteristics of the turbine thrust bearing under typical operating conditions, specifically including: Substitute the actual axial load into the thermo-fluid-structure interaction finite element model; The high load condition, rated speed and specified lubricating oil inlet temperature are set as the typical operating conditions of the turbine thrust bearing; Finite element analysis was performed on the thermo-fluid-structure interaction finite element model under typical working conditions to analyze the oil film pressure distribution, film thickness variation, temperature field and bearing deformation of the turbine thrust bearing. The lubrication and load-bearing characteristics of turbine thrust bearings are analyzed based on oil film pressure distribution, film thickness variation, temperature field, and bearing deformation.

6. The method for optimizing turbine thrust bearing failure and damage based on lubrication and load-bearing characteristics analysis according to claim 1, characterized in that, The dominant factors for turbine thrust bearing failure are determined based on lubrication and load characteristics, specifically including: Based on lubrication and load-bearing characteristics, the magnitude and direction of thermal and mechanical deformation of turbine thrust bearings and their influence on oil film shape are analyzed. The contributions of thermal and mechanical deformation to lubrication performance are compared and evaluated to determine the dominant deformation type and the mechanism of lubrication deterioration. By correlating the dominant deformation type and the mechanism of deteriorated lubrication with the failure and damage state observed on site, the dominant factors of turbine thrust bearing failure and damage were determined.

7. The method for optimizing turbine thrust bearing failure and damage based on lubrication and load-bearing characteristics analysis according to claim 1, characterized in that, When the dominant factors are axial load and inlet temperature, the influence of oil film characteristics is analyzed through parametric simulation to obtain the degradation law of increased load and temperature, which specifically includes: When the dominant factors are axial load and oil inlet temperature, the axial load and oil inlet temperature, which directly affect the thermal load, are selected as key variables based on the dominant factors, and the range of variation is set. Based on typical working conditions, multiple sets of thermal-fluid-structure interaction finite element simulations were performed by adjusting key variables to obtain simulation results. By analyzing the simulation results, we can summarize the influence of oil film characteristics and obtain the law of deterioration due to increased load and temperature.

8. The method for optimizing turbine thrust bearing failure and damage based on lubrication and load-bearing characteristics analysis according to claim 1, characterized in that, The optimization strategy includes: Set the upper limit of the oil inlet temperature during operation; Improve the cooling design and thermal deformation resistance design of the turbine thrust bearing; Control the axial thrust to operate within a preset range.

9. The method for optimizing turbine thrust bearing failure and damage based on lubrication and load-bearing characteristics analysis according to claim 1, characterized in that, The expression for the turbulent modified Reynolds equation is as follows: ； ； In the formula, Gx and Gy are both turbulence coefficients; x and y are the circumferential and radial directions of the bearing oil film, respectively; ρ is the density; h is the oil film thickness; p is the pressure; and μ is the fluid viscosity.

10. A computer system, comprising: A memory, a processor, and a computer program stored in the memory and executable on the processor, characterized in that the processor executes the computer program to implement the turbine thrust bearing failure and damage optimization method based on lubrication and load-bearing characteristic analysis as described in any one of claims 1-9.

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