A method for analyzing fluid-structure coupling vibration characteristics of a rotating blade-casing rub system

Abstract

The application relates to a method for analyzing fluid-structure coupling vibration characteristics of a rotating blade-casing rub-impact system, and belongs to the technical field of aero-engine dynamics analysis. The method comprises the following steps: a rotating blade-rigid casing rub-impact dynamics model considering radial / angle misalignment is established, prestress modal analysis is performed on the blade to obtain modal characteristics of the blade; the rotating blade flow passage is divided into a structured grid, real aerodynamic force on the blade surface is obtained through fluid simulation and grid independence verification; in a one-way fluid-structure coupling mode, transient rub-impact analysis is performed on the aerodynamic force as external load combined with centrifugal force effect, and stable periodic results are extracted to obtain system vibration characteristics. The application solves the problems in the prior art, such as simplified aerodynamic force, poor adaptability of a rub-impact force model, and lack of misalignment fault analysis under fluid-structure coupling, improves analysis accuracy, provides precise theoretical support for reliability design and fault prevention of an aero-engine, and has remarkable industrial application value.

Description

Technical Field

[0001] This application relates to the field of aero-engine dynamics analysis technology, and in particular to a method for analyzing the fluid-structure interaction vibration characteristics of a rotating blade-casing rubbing system. Background Technology

[0002] The rapid development of the aerospace industry has driven the development of aero engines toward high performance, high efficiency, and high thrust-to-weight ratio. Reducing the rotor-stator clearance is a direct and effective way to improve the overall performance of aero engines. However, excessively small blade tip clearance can easily induce blade-casing rubbing failures. These failures can lead to abnormal rotor vibrations and even catastrophic accidents such as blade breakage, seriously affecting the engine's working performance and operational safety.

[0003] Blades are subjected to the combined effects of multiple loads, including fluid, solid, and thermal loads, during operation. Studying the fluid-structure interaction vibration characteristics of the rotating blade-casing system under rubbing conditions is crucial for the reliability design, structural optimization, and fault monitoring of aero-engines. Current research on blade-casing rubbing systems primarily focuses on the rubbing vibration characteristics of a single structural field, simplifying blade aerodynamic forces into uniformly distributed harmonic forces. This approach fails to adequately consider the influence of real flow field characteristics on rotating blades, leading to significant biases in rubbing response analysis. Furthermore, traditional rubbing force models are only suitable for single-tip nodes in cantilever beam blade models, failing to meet the computational requirements of multi-node chordal rubbing at the tips of solid element blades. Moreover, research on the impact of faults such as radial misalignment and angular misalignment on the vibration characteristics of the rubbing system under fluid-structure interaction conditions is scarce, making it difficult to accurately provide theoretical support for aero-engine design and fault prevention.

[0004] Therefore, there is an urgent need in related technologies to solve the problems of insufficient consideration of real flow field characteristics, poor adaptability of rubbing force models, and lack of analysis of the impact of misalignment faults on rubbing vibration characteristics in fluid-structure interaction environments in existing technologies, so as to achieve accurate analysis of the fluid-structure interaction vibration characteristics of the rotating blade-rigid casing rubbing system. Summary of the Invention

[0005] Therefore, it is necessary to provide a method for analyzing the fluid-structure interaction vibration characteristics of a rotating blade-casing rubbing system to address the aforementioned technical problems.

[0006] Firstly, this application provides a method for analyzing the fluid-structure interaction vibration characteristics of a rotating blade-casing rubbing system. The method includes: A dynamic model of the collision between the rotating blade and the rigid casing is established. The model takes into account the effects of radial misalignment and angular misalignment and modifies the collision force model. Fluid characteristics analysis was performed on the rotating blades to obtain the aerodynamic forces on the blade surface; The aerodynamic force is applied as an external load to the blade structure. Combined with the centrifugal force effect of the rotating blade, a fluid-structure interaction transient friction analysis is performed on the rotating blade-rigid casing friction system to obtain the vibration characteristics of the system.

[0007] Optionally, in one embodiment of this application, establishing the dynamic model of the rotating blade-rigid casing collision includes: The casing is set as a rigid casing with flexible support. The blades are fixed to the rigid blade disk. The influence of blade vortex and the thermal effect of the rubbing process are ignored. The blades are meshed to obtain the blade tip equivalent node. The casing is set with concentrated mass points along the rotation axis and each concentrated mass point is rigidly coupled. The initial clearance calculation method between the blade tip node and the concentrated mass point of the casing is defined.

[0008] Optionally, in one embodiment of this application, the modification of the friction force model includes: The sum of the intrusion amounts of all nodes at the blade tip is calculated and the total blade tip friction force is calculated. The friction force of a single blade tip node is calculated based on the proportion of the intrusion amount of each blade tip node. The normal friction force and tangential friction force of a single blade tip node are obtained by combining the friction coefficient between the blade and the casing.

[0009] Optionally, in one embodiment of this application, the step of performing fluid characteristic analysis on the rotating blade to obtain the aerodynamic forces on the blade surface includes: The flow channel of the rotating blade is divided into a structured mesh. The three-dimensional Reynolds-averaged Navier-Stokes equations are solved using a turbulence model. Boundary conditions for fluid simulation are set and mesh independence is verified. After the computational mesh is determined, the aerodynamic forces on the blade surface are obtained through unsteady fluid calculations.

[0010] Optionally, in one embodiment of this application, the method further includes: The motion equations of the rotating blade are established, and the motion equations incorporate the centrifugal stiffening matrix, the rotational softening matrix, and the Coriolis force matrix. Prestressed modal analysis is performed on the rotating blade to obtain the natural frequency, resonance point, and mode shape of the blade at different rotational speeds.

[0011] Optionally, in one embodiment of this application, the fluid-structure interaction transient rubbing analysis is a unidirectional fluid-structure interaction analysis, in which the aerodynamic force obtained from the fluid characteristic analysis is applied to the blade structure, and the blade rotation speed and the blade root fixed boundary conditions are applied simultaneously to perform transient analysis on the rotating blade-rigid casing rubbing system and extract the analysis results within the stable period.

[0012] Optionally, in one embodiment of this application, the influence of aerodynamic forces on the system's rubbing vibration characteristics is analyzed, and the tip displacement response, tip rubbing force, tip clearance, and casing center of mass displacement are compared under no aerodynamic force, peak efficiency operating condition aerodynamic force, and design point operating condition aerodynamic force.

[0013] Optionally, in one embodiment of this application, the influence of different radial misalignment amounts on the system's rubbing vibration characteristics is analyzed to obtain the number of blade tip rubbing nodes, the maximum blade tip rubbing force, and the variation law of blade tip displacement response under different radial misalignment amounts.

[0014] Optionally, in one embodiment of this application, the influence of different misalignment angles on the system's rubbing vibration characteristics is analyzed to obtain the number of blade tip rubbing nodes, the maximum blade tip rubbing force, and the blade tip displacement response variation law under different misalignment angles.

[0015] Compared with the prior art, the above-mentioned method for analyzing the fluid-structure interaction vibration characteristics of a rotating blade-casing rubbing system has the following advantages: First, considering the effect of real flow field characteristics on the blades, the aerodynamic forces on the blade surface are accurately obtained through fluid simulation, replacing the traditional method of simplifying aerodynamic forces into harmonic uniformly distributed surface forces. This significantly reduces the deviation of the collision response analysis and improves the accuracy of the system vibration characteristic analysis. Second, the traditional friction force model was modified to adapt to the calculation requirements of multiple nodes in the chord direction of the blade tip of the solid element, making the calculation results of friction force more consistent with the actual engineering scenario, and laying an accurate model foundation for the analysis of friction vibration characteristics. Third, the established dynamic model takes into account the effects of radial misalignment and angular misalignment, and combines the combined effects of aerodynamic force and centrifugal force through unidirectional fluid-structure interaction analysis, filling the gap in the existing technology for the study of the impact of misalignment faults on the rubbing vibration characteristics under fluid-structure interaction environment. Fourth, the system analyzes the influence of key parameters such as aerodynamic force, radial misalignment, and angular misalignment on the vibration characteristics of the rotating blade-rigid casing rubbing system. This provides specific and accurate theoretical support for the design of blade tip clearance, prevention and control of misalignment faults, and monitoring of rubbing faults in aero-engines, thus contributing to the reliability design and structural optimization of aero-engines. Attached Figure Description

[0016] Figure 1 This is a flowchart illustrating a method for analyzing the fluid-structure interaction vibration characteristics of a rotating blade-casing rubbing system in one embodiment. Figure 2 Here is a diagram of the Rotor37 blade analysis model in one embodiment; Figure 3 This is a schematic diagram of blade-casing rubbing in one embodiment; Figure 4 This is a schematic diagram of the blade-casing system tip-node clearance in one embodiment; Figure 5 Here is a Campbell diagram of the Rotor37 blade in one embodiment; Figure 6 Here is a blade mode shape diagram from one embodiment; Figure 7 This is a diagram showing the fluid characteristics of the rotor under different grids in one embodiment; Figure 8 This is a blade pressure contour plot in one embodiment; Figure 9 This is a displacement response diagram of blade tip 1 under different aerodynamic forces in one embodiment; Figure 10 This is a diagram showing the tip-to-blade contact response under different aerodynamic forces in one embodiment; Figure 11 This is a diagram of the friction force at the blade tip under different aerodynamic forces in one embodiment; Figure 12 The displacement response diagram of the lower blade tip 1 with different radial misalignment in one embodiment is shown. Figure 13 This is a diagram of the rubbing force of the lower blade tips with different radial misalignments in one embodiment; Figure 14 This is a displacement response diagram of the lower leaf tip 1 at different misaligned angles in one embodiment; Figure 15 This is a diagram showing the rubbing force of the lower blade tip at different angles in one embodiment. Detailed Implementation

[0017] To make the objectives, technical solutions, and advantages of this application clearer, the following detailed description is provided in conjunction with the accompanying drawings and embodiments. It should be understood that the specific embodiments described herein are merely illustrative and not intended to limit the scope of this application.

[0018] The following detailed implementation of the fluid-structure interaction vibration characteristic analysis method for a rotating blade-casing rubbing system, based on the Rotor37 transonic axial compressor rotor designed by NASA Lewis Research Center, provides a specific example. The Rotor37 rotor contains 36 blades arranged in a periodically symmetrical distribution. Some of its design parameters are: design speed 17188.7 rpm, blade tip clearance 0.356 mm, design flow rate 20.188 kg / s, total pressure ratio 2.106, total temperature ratio 1.27, and isentropic efficiency 0.877. The blade material is martensitic 200 with a density of 8000 kg / m³, an elastic modulus of 180 GPa, and a Poisson's ratio of 0.3.

[0019] In one embodiment, such as Figure 1 As shown, a method for analyzing the fluid-structure interaction vibration characteristics of a rotating blade-casing rubbing system is provided, including the following steps: S101: Establish a dynamic model of the collision between the rotating blade and the rigid casing. The model takes into account the effects of radial misalignment and angular misalignment and modifies the collision force model.

[0020] In one embodiment of this application, the model assumptions are set as follows: ① The casing is a rigid casing with flexible support, and the normal and tangential support stiffness and damping are evenly distributed according to the concentrated mass point; ② The blades are fixed on the rigid bladed disk, and their vortex effect is ignored; ③ The thermal effect generated during the collision process is ignored.

[0021] The blade is meshed with a high-quality hexahedral grid. The blade width is divided into 20 parts, and the length is divided into 30 parts. Initially, there are 42 nodes in the chord direction at the blade tip. The two nodes in the thickness direction are equivalent to one intermediate node, resulting in 21 equivalent nodes at the blade tip. Figure 2 As shown, the upper part is the rotor fluid model (selecting a single flow channel as the research object), and the lower part is the rotor blade structure model. The pressure and radial static deformation are greatest near the blade tip node 1, making it most prone to rubbing. 21 concentrated mass points are set along the rotation axis of the casing, retaining the radial and normal degrees of freedom of each mass point and rigidly coupling them to simulate the overall translational characteristics of the casing. Figure 3 As shown, Figure 3 (a) is a three-dimensional rubbing diagram. Figure 3 (b) is a two-dimensional rubbing diagram of each leaf tip node.

[0022] Considering radial misalignment e 0 and angle misalignment φ The influence of the blade tip node is calculated based on the positional relationship between the blade tip node and the concentrated mass point of the casing. I The initial gap, such as Figure 4 As shown, Figure 4 (a) is a top view of the blade. Figure 4 (b) is a diagram of the internode gaps at each leaf tip. Figure 4 (c) is the leaf tip node i The initial clearance diagram shows that the initial clearance is affected by parameters such as the offset between the blade disk center and the casing center, the blade mounting angle, the casing radius, and the blade tip trajectory radius. The larger the angular misalignment and the larger the radial misalignment, the smaller the initial clearance at the blade tip.

[0023] In one embodiment of this application, a multi-node adaptation correction is performed based on an analytical model of the friction force between a cantilever beam blade and a flexible casing. Solve for the intrusion amount of all leaf tip nodes. sum Intrusion amount The actual clearance between the blade tip node and the casing c i The actual clearance is determined to be a combination of the initial clearance, blade radial displacement, and casing radial displacement. Calculate the total blade tip friction force based on the proportion of intrusion at each blade tip node. Solving for a single leaf tip node iFriction force ; Combining blade-casing friction coefficient μ (In this embodiment, the value is 0.2), determined by the nodal friction force. Decomposition yields normal friction force and tangential friction force .

[0024] In one embodiment of this application, considering centrifugal stiffening, rotational softening, and Coriolis force effects, the motion equation of the rotating blade is established as follows:

[0025] in, For the structural mass matrix, For structural damping matrix, Coriolis force matrix, For the structural stiffness matrix, For the centrifugal stiffening matrix, For the rotation softening matrix, and Let C represent the displacement and external excitation force vectors, respectively.

[0026] in, , These are the mass, stiffness, and damping coefficients, respectively. , These are the first and second natural frequencies of the blade, respectively. , The damping ratios for the first and second natural frequency modes, respectively, can be determined experimentally. In this embodiment, we take... , .

[0027] Prestressed modal analysis was performed on the blades to obtain values ​​at 0 and 0.2. n d 0.4 n d 0.6 n d 0.8 n d 1.0 n d ( n d The first six natural frequencies of the blades at the rotor design speed were determined. The results show that the natural frequencies of each blade gradually increase with increasing speed. The Campbell diagram of the blades was plotted, and the resonance point was determined. Figure 5As shown, the relationship between the blade's natural frequency and rotational speed, as well as the distribution of resonance points, are illustrated. Within the design speed range, 0, 1, and 3 resonance points exist for engine orders EO of 1, 4, and 16, respectively. Simultaneously, the modal shape diagrams of the blade are plotted, as shown below. Figure 6 As shown, the first three modes of the blade are obtained as the first bending mode, the first torsional mode, and the second bending mode.

[0028] S102: Perform fluid characteristic analysis on the rotating blades to obtain the aerodynamic forces on the blade surface.

[0029] In one embodiment of this application, firstly, the Rotor37 rotor is structured and meshed using CFX TurboGrid software. The main runner is meshed using a HOH hybrid topology, the inlet and outlet are meshed using an H-type mesh, the blades are meshed using an O-type mesh, the tip clearance is meshed using a butterfly-shaped mesh, the casing and hub boundaries are finely meshed, and the height of the first layer of mesh near the wall is set to 3×10. - ³mm, ensuring y+ is less than 2, meeting the requirements of the turbulence model.

[0030] Subsequently, the three-dimensional Reynolds-averaged Navier-Stokes equations were solved using CFX software, and the SST k-ω two-equation turbulence model was selected for convergence calculation of the Reynolds stress terms. The spatial discretization scheme adopted the second-order accurate central difference method, and the fluid calculation residual value was set to 1×10⁻⁶. -6 Boundary conditions are set as follows: the working medium is an ideal gas, the inlet total temperature is 288.15K and the total pressure is 101.325kPa, the outlet is the average static pressure, the rotor speed is the design speed of 17188.7rpm, the circumferential sides of the single flow channel are periodic boundaries, and the solid walls of the casing, hub and other components are adiabatic and non-slip boundary conditions.

[0031] Subsequently, the mesh independence of five mesh numbers (3e6, 6e6, 9e6, 12e6, and 15e6) under three-dimensional unsteady conditions was verified. Steady calculation results were used as the initial values ​​for unsteady calculations. A single cycle was set with 512 time steps, and a total of 6 cycles were solved. The results are as follows: Figure 7 The results show that as the number of grids increases, the total pressure ratio, total temperature ratio, and static pressure at the blade tip monitoring point gradually stabilize. When the number of grids is greater than 6e6, the total pressure ratio error is ≤2% and the isentropic efficiency error is ≤4%. Considering both computational efficiency and accuracy, a 9e6 grid is selected as the optimal computational grid.

[0032] Finally, unsteady fluid calculations were performed based on the 9e6 grid to obtain the aerodynamic forces on the blade surface under peak efficiency and design point conditions. The maximum pressure region on the blade is located near the blade tip, and the minimum pressure region is located near the blade root. The blade surface pressure under the design point condition is greater than that under the peak efficiency condition.

[0033] S103: Apply the aerodynamic force as an external load to the blade structure, and perform fluid-structure interaction transient friction analysis on the rotating blade-rigid casing friction system in combination with the centrifugal force effect of the rotating blade to obtain the vibration characteristics of the system.

[0034] In one embodiment of this application, a one-way fluid-structure interaction analysis method is adopted, and the aerodynamic force obtained from the fluid characteristic analysis is applied as the initial load to the blade structure. At the same time, boundary conditions are applied: the blade root is fixed, the blade speed is the design speed of 17188.7 rpm, and the casing is a flexible support. Transient rubbing analysis is performed on the rotating blade-rigid casing rubbing system, and the results of 160 cycles are solved. The results of the last few stable cycles are used for subsequent analysis to obtain the vibration characteristic parameters of the system, such as the blade tip displacement response, blade tip clearance, rubbing force, and casing center of mass displacement.

[0035] In one embodiment of this application, three operating conditions are set: aerodynamic force without considering aerodynamic forces, peak efficiency operating condition aerodynamic force, and design point operating condition aerodynamic force. The pressure distribution of the blades under the latter two operating conditions is as follows: Figure 8 As shown, Figure 8 (a) shows the pressure distribution under peak efficiency conditions. Figure 8 (b) shows the pressure distribution under the design point condition. It can be seen that the area of ​​highest pressure is near the blade tip, and the area of ​​lowest pressure is near the blade root; the pressure under the design point condition is higher than the pressure under the peak efficiency condition; and the pressure distribution on the pressure surface is more complex than that on the suction surface. Comparative analysis of system vibration characteristics: such as... Figure 9 As shown, the system experiences periodic rubbing. Under aerodynamic conditions, the dominant frequency of the blade tip displacement response is the rotational speed (286Hz), while under non-aerodynamic conditions, the dominant frequency is three times the rotational speed. Figure 10 As shown, Figure 10 (a) shows the change in the leaf tip gap 1. Figure 10 (b) represents the change in the center of mass displacement of the casing. Figure 10 (c) shows the amplitude distribution of the blade tip rubbing force. The clearance at blade tip node 1 also changes periodically, gradually increasing from the initial minimum to the maximum value, and then falling back to the minimum value at the end of the period. When the blade tip clearance is negative, rubbing occurs in the blade-casing system, as shown in the shaded area in the figure. The radial displacement of the casing center of mass first increases rapidly from the initial minimum to the maximum value, and then falls back to the minimum value. This radial displacement reflects the overall degree of rubbing in the system, mainly affected by the resultant force of the blade tip normal rubbing force and the normal damping. The radial displacement of the casing center of mass indicates that the system rubbing is most intense under the aerodynamic conditions at the design point, and least intense under the conditions without considering aerodynamics. As the blade tip node number increases, the amplitude of the normal rubbing force first increases and then decreases. Under the design point conditions, the amplitude of the blade tip normal rubbing force is the largest; under the conditions without considering aerodynamics, the amplitude of the blade tip normal rubbing force is the smallest. Figure 11As shown, compared to not considering aerodynamic forces, the peak efficiency and the maximum tip friction force under steady aerodynamic forces at the design point increase by 22.00% and 23.75%, respectively.

[0036] In one embodiment of this application, three radial misalignment conditions of 0.26mm, 0.28mm, and 0.30mm are set. The analysis results show that: Figure 12 As shown, the greater the radial misalignment, the smaller the initial tip clearance, the greater the intrusion, the smaller the minimum tip displacement, and the greater the clearance fluctuation; as Figure 13 As shown, when the radial misalignment is 0.26mm, 0.28mm, and 0.30mm, the number of tip rubbing nodes is 11, 16, and 21, respectively, and the maximum tip rubbing force is 79.88N, 101.38N, and 123.62N, respectively. When the radial misalignment increases by 0.02mm from 0.26mm, the maximum tip rubbing force increases by 26.90%; when the radial misalignment increases by 0.02mm from 0.28mm, the maximum tip rubbing force increases by 21.94%. For every 0.02mm increase in radial misalignment, the increase in the maximum tip rubbing force decreases slightly.

[0037] In one embodiment of this application, three angular misalignment conditions of 0.10°, 0.12°, and 0.14° are set. The analysis results show that: Figure 14 As shown, the angular misalignment does not change the trend of the tip displacement response, but only reduces the minimum displacement. The larger the angular misalignment, the greater the fluctuation of the tip clearance. Figure 15 As shown, when the misalignment is 0.10°, 0.12°, and 0.14°, the number of blade tip rubbing points is 14, 16, and 17, respectively, and the maximum blade tip rubbing force is 89.07N, 101.38N, and 113.52N, respectively. The maximum blade tip rubbing force is the largest when the misalignment is 0.14°, and the minimum when the misalignment is 0.10°. Increasing the misalignment by 0.02° from 0.10° increases the maximum blade tip rubbing force by 13.82%; increasing it by 0.02° from 0.12° increases it by 11.98%; and the increase in maximum blade tip rubbing force gradually decreases with each 0.02° increase in misalignment.

[0038] It should be noted that the method for analyzing the fluid-structure interaction vibration characteristics of a rotating blade-casing rubbing system of the present invention can be directly applied to the dynamic analysis of the blade-casing rubbing system of rotating machinery such as aero-engines and gas turbines. It can accurately analyze the influence of key parameters such as aerodynamic force, radial misalignment, and angular misalignment on the fluid-structure interaction vibration characteristics of the system, and provide specific and accurate theoretical support for the design of blade tip clearance, prevention and control of misalignment faults, monitoring of rubbing faults and structural optimization of rotating machinery. It has significant industrial application value.

[0039] It should be understood that although the steps in the flowcharts of the embodiments described above are shown sequentially according to the arrows, these steps are not necessarily executed in the order indicated by the arrows. Unless explicitly stated herein, there is no strict order restriction on the execution of these steps, and they can be executed in other orders. Moreover, at least some steps in the flowcharts of the embodiments described above may include multiple steps or multiple stages. These steps or stages are not necessarily completed at the same time, but can be executed at different times. The execution order of these steps or stages is not necessarily sequential, but can be performed alternately or in turn with other steps or at least some of the steps or stages of other steps.

[0040] Those skilled in the art will understand that all or part of the processes in 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 described above. 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). 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.

[0041] The technical features of the above embodiments can be combined in any way. For the sake of brevity, not all possible combinations of the technical features in the above embodiments are described. However, as long as there is no contradiction in the combination of these technical features, they should be considered to be within the scope of this specification.

[0042] The embodiments described above are merely illustrative of several implementation methods of this application, and while the descriptions are specific and detailed, they should not be construed as limiting the scope of this patent application. It should be noted that those skilled in the art can make various modifications and improvements without departing from the concept of this application, and these all fall within the protection scope of this application. Therefore, the protection scope of this application should be determined by the appended claims.

Claims

1. A method for analyzing the fluid-structure interaction vibration characteristics of a rotating blade-casing rubbing system, characterized in that, The method includes: A dynamic model of the collision between the rotating blade and the rigid casing is established. The model takes into account the effects of radial misalignment and angular misalignment and modifies the collision force model. Fluid characteristics analysis was performed on the rotating blades to obtain the aerodynamic forces on the blade surface; The aerodynamic force is applied as an external load to the blade structure. Combined with the centrifugal force effect of the rotating blade, a fluid-structure interaction transient friction analysis is performed on the rotating blade-rigid casing friction system to obtain the vibration characteristics of the system.

2. The method for analyzing the fluid-structure interaction vibration characteristics of a rotating blade-casing rubbing system according to claim 1, characterized in that, The establishment of the dynamic model of the rubbing motion between the rotating blade and the rigid casing includes: The casing is set as a rigid casing with flexible support. The blades are fixed to the rigid blade disk. The influence of blade vortex and the thermal effect of the rubbing process are ignored. The blades are meshed to obtain the blade tip equivalent node. The casing is set with concentrated mass points along the rotation axis and each concentrated mass point is rigidly coupled. The initial clearance calculation method between the blade tip node and the concentrated mass point of the casing is defined.

3. The method for analyzing the fluid-structure interaction vibration characteristics of a rotating blade-casing rubbing system according to claim 1, characterized in that, The correction of the impact friction force model includes: The sum of the intrusion amounts of all nodes at the blade tip is calculated and the total blade tip friction force is calculated. The friction force of a single blade tip node is calculated based on the proportion of the intrusion amount of each blade tip node. The normal friction force and tangential friction force of a single blade tip node are obtained by combining the friction coefficient between the blade and the casing.

4. The method for analyzing the fluid-structure interaction vibration characteristics of a rotating blade-casing rubbing system according to claim 1, characterized in that, The process of analyzing the fluid characteristics of the rotating blades to obtain the aerodynamic forces on the blade surface includes: The flow channel of the rotating blade is divided into a structured mesh. The three-dimensional Reynolds-averaged Navier-Stokes equations are solved using a turbulence model. Boundary conditions for fluid simulation are set and mesh independence is verified. After the computational mesh is determined, the aerodynamic forces on the blade surface are obtained through unsteady fluid calculations.

5. The method for analyzing the fluid-structure interaction vibration characteristics of a rotating blade-casing rubbing system according to claim 1, characterized in that, The method further includes: The motion equations of the rotating blade are established, and the motion equations incorporate the centrifugal stiffening matrix, the rotational softening matrix, and the Coriolis force matrix. Prestressed modal analysis is performed on the rotating blade to obtain the natural frequency, resonance point, and mode shape of the blade at different rotational speeds.

6. The method for analyzing the fluid-structure interaction vibration characteristics of a rotating blade-casing rubbing system according to claim 1, characterized in that, The fluid-structure interaction transient friction analysis is a unidirectional fluid-structure interaction analysis. The aerodynamic force obtained from the fluid characteristic analysis is applied to the blade structure, and the boundary conditions of blade rotation speed and fixed blade root are applied at the same time. The transient analysis of the rotating blade-rigid casing friction system is performed and the analysis results within the stable period are extracted.

7. The method for analyzing the fluid-structure interaction vibration characteristics of a rotating blade-casing rubbing system according to claim 6, characterized in that, The influence of aerodynamic forces on the system's rubbing vibration characteristics is analyzed, and the tip displacement response, tip rubbing force, tip clearance, and casing center of mass displacement are compared under no aerodynamic force, peak efficiency operating condition aerodynamic force, and design point operating condition aerodynamic force.

8. The method for analyzing the fluid-structure interaction vibration characteristics of a rotating blade-casing rubbing system according to claim 6, characterized in that, The influence of different radial misalignment amounts on the system's rubbing vibration characteristics was analyzed, and the variation laws of the number of tip rubbing nodes, the maximum tip rubbing force, and the tip displacement response under different radial misalignment amounts were obtained.

9. The method for analyzing the fluid-structure interaction vibration characteristics of a rotating blade-casing rubbing system according to claim 6, characterized in that, The influence of different misalignment angles on the system's rubbing vibration characteristics was analyzed, and the variation law of the number of blade tip rubbing nodes, the maximum blade tip rubbing force, and the blade tip displacement response under different misalignment angles was obtained.

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