A method for synergistically suppressing flow-induced vibration of an impeller and a guide vane

Abstract

The application discloses a kind of impeller-vane flow-induced vibration collaborative inhibition method, belong to the field of rotating fluid machinery vibration analysis.The method comprises: according to the structural characteristic parameters of target rotating machinery, the three-dimensional modeling software is used to establish the three-dimensional model of fluid domain and the three-dimensional model of solid domain of target rotating machinery, and the grid division is carried out by building automated process, and steady-state simulation analysis is carried out, and the grid independence and simulation accuracy are verified;Geometric parameter dimensionality reduction processing is carried out according to the profile characteristics;Optimization algorithm is selected, impeller-vane collaborative iterative optimization is carried out, and the optimization model of turbine expander is obtained;The results before and after optimization are compared, and the optimization effect is summarized.The method of the application can efficiently and low-cost optimize the flow-induced vibration characteristics of impeller and vane, provide scientific guidance for impeller design and optimization, and improve the overall stability and reliability of large-scale rotating equipment.

Description

TECHNICAL FIELD

[0001] The application belongs to the field of rotating fluid machinery vibration analysis, and particularly relates to a method for cooperatively inhibiting flow-induced vibration of an impeller-vane. BACKGROUND

[0002] Impeller machines are indispensable core equipment in industries such as energy production, chemical processes, aerospace, and ocean engineering. They convert energy into the kinetic energy or pressure change of fluids (gases or liquids) through rotating impellers. The working principle of such machines usually involves converting input energy (such as electrical or thermal energy) into rotational kinetic energy, which is then used to work on fluids, thereby achieving functions such as conveying, compressing, and pumping. However, due to characteristics such as high-speed operation, high temperature and pressure working conditions, and complex fluid-solid interactions, impeller machines are prone to vibration problems, which not only affect the reliability and lifespan of the equipment, but also can cause serious safety risks and economic losses.

[0003] Research has found that flow instability caused by fluid excitation during mechanical operation is also one of the main reasons for vibration of impeller machines. These vibrations can severely affect the efficiency of the entire impeller machine system and pose a significant risk to the structural integrity and stable operation. Current optimization design only optimizes the impeller or vane separately. However, the strong dynamic and static interference effect between the impeller and vane is the main cause of flow-induced vibration, and mutual influence can lead to changes in fluid state. Therefore, cooperative optimization design of the impeller and vane is necessary to ensure the optimal vibration performance of rotating machines. Therefore, it is of great significance to invent a method for cooperative optimization design of impeller and vane to reduce flow-induced vibration. SUMMARY

[0004] The purpose of the present application is to solve the problems existing in the prior art and provide a method for cooperatively inhibiting flow-induced vibration of an impeller-vane. The present application directly controls and optimizes model parameters through an optimization algorithm, achieving high-precision cooperative optimization of the impeller and vane, and providing a scientific reference for optimization design of the impeller and vane of an impeller machine.

[0005] In order to achieve the above-mentioned application purposes, the present application specifically adopts the following technical solutions:

[0006] A method for cooperatively inhibiting flow-induced vibration of an impeller-vane, comprising the following steps:

[0007] S1, taking the impeller and vane at the expansion end of a turboexpander as a target rotating machine, and establishing a three-dimensional model of the fluid domain and a three-dimensional model of the solid domain of the target rotating machine according to the structural characteristic parameters of the target rotating machine using a three-dimensional modeling software;

[0008] S2, finite element meshing is performed on the fluid domain three-dimensional model, then the turbine expander impeller finite element calculation grid and the turbine expander guide vane finite element calculation grid obtained by the finite element meshing are imported into the fluid domain simulation software, the fluid domain simulation conditions are set, the fluid domain calculation model is obtained, and the steady-state simulation of the fluid domain calculation model is performed to obtain the fluid domain steady-state simulation result;

[0009] S3, grid independence verification is performed on the fluid domain steady-state simulation result, experimental measurement is performed on the fluid domain three-dimensional model after the grid independence verification, the fluid domain experimental measurement result is obtained, and the fluid domain experimental measurement result is compared with the fluid domain steady-state simulation result to ensure the accuracy of the internal three-dimensional steady-state flow field simulation of the turbine expander;

[0010] S4, grid division is performed on the solid domain three-dimensional model, and the solid domain simulation conditions are set to obtain the solid domain calculation model;

[0011] S5, the fluid-solid coupling calculation model is established based on the solid domain calculation model and the fluid domain calculation model, the parameterization of the fluid-solid coupling calculation model is performed without changing the basic structure of the fluid domain three-dimensional model and the solid domain three-dimensional model, the impeller profile graph and the guide vane profile graph are obtained, and the angle and thickness in the impeller profile graph and the angle and thickness in the guide vane profile graph are parameterized to obtain the reduced parameterized model;

[0012] S6, the reduced parameterized model is used to minimize the energy to measure the flow-induced vibration strength parameter as the optimization objective, the performance of the target rotating machine is used as the constraint condition, the turbine expander original model is constructed, the optimization algorithm is used to iteratively optimize the constructed turbine expander original model, and the turbine expander optimization model is obtained;

[0013] S7, a bidirectional fluid-solid coupling data transmission module of CFX, Transient Structural and System Coupling is constructed in Ansys Workbench, and bidirectional fluid-solid coupling simulation is performed on the turbine expander optimization model and the turbine expander original model in the bidirectional fluid-solid coupling data transmission module;

[0014] S8, the bidirectional fluid-solid coupling simulation results of the turbine expander optimization model and the bidirectional fluid-solid coupling simulation results of the turbine expander original model are imported into the post-processing software for post-processing, the acceleration change of the impeller at the expansion end of the turbine expander is used as an evaluation index, the turbine expander optimization model whose flow-induced vibration characteristics meet the preset requirements is output, and the impeller-guide vane flow-induced vibration collaborative suppression is realized.

[0015] On the basis of the above scheme, each step can be implemented in the following preferred specific manner.

[0016] As preferred, in step S1, the three-dimensional modeling software adopts UG NX; the three-dimensional model of the fluid domain of the expansion end of the turbo expander includes four parts of the impeller, the guide vane, the inlet pipeline and the outlet pipeline, and each of the inlet pipeline and the outlet pipeline is extended to form an inlet extension section and an outlet extension section; the three-dimensional model of the solid domain of the expansion end of the turbo expander sets the blade surface and the hub surface as the fluid-structure coupling interface, and considers the influence of the earth gravity and ignores the mechanical vibration influence of the rest components, and an ideal displacement constraint is adopted at the shaft hole of the impeller.

[0017] As preferred, in step S2, the finite element grid division of the three-dimensional model of the fluid domain is realized by Turbo Grid and ICEM, the fluid domain simulation software adopts CFD simulation software, and the fluid domain simulation conditions include fluid material properties, a turbulent flow model, an inlet boundary condition, an outlet boundary condition, an impeller flow domain rotating speed, a dynamic-static interface, a wall surface setting and a calculation step number.

[0018] As preferred, in step S3, firstly, an initial grid model is established and simulation is performed, then the grid is gradually encrypted, and after each encryption of the grid, the simulation is re-run to complete the grid independence verification of the steady-state simulation result of the fluid domain; the measured outlet pressure, the measured inlet pressure, the measured outlet temperature and the measured inlet temperature are obtained from the experimental measurement result of the fluid domain, the isentropic efficiency is calculated based on the measured outlet pressure, the measured inlet pressure, the measured outlet temperature and the measured inlet temperature, and the isentropic efficiency calculation result is taken as the measured isentropic efficiency, the simulated outlet pressure, the simulated inlet pressure, the simulated outlet temperature and the simulated inlet temperature are obtained from the steady-state simulation result of the fluid domain, the isentropic efficiency is recalculated based on the simulated outlet pressure, the simulated inlet pressure, the simulated outlet temperature and the simulated inlet temperature, and the isentropic efficiency calculation result is taken as the simulated isentropic efficiency, the relative error of the isentropic efficiency between the measured isentropic efficiency and the simulated isentropic efficiency is calculated, if the relative error of the isentropic efficiency is less than a preset precision threshold, it is indicated that the three-dimensional steady-state flow field simulation of the turbo expander is correct; otherwise, the steady-state simulation result of the fluid domain is re-obtained until the three-dimensional steady-state flow field simulation of the turbo expander is correct.

[0019] As preferred, the isentropic efficiency η is calculated by the following formula:

[0020]

[0021] In the formula, κ is the gas constant of superheated steam; P in is the inlet pressure, P out is the outlet pressure, T in is the inlet temperature, T out is the outlet temperature.

[0022] As preferred, in step S4, the grid size of the fluid domain three-dimensional model is taken as the reference to divide the grid of the solid domain three-dimensional model by using the Ansys meshing tool in ANSYS Mechanical, the tetrahedral grid structure is used when the solid domain three-dimensional model is divided, and the solid domain simulation conditions include the solid material properties, constraint conditions, gravity settings, calculation time, calculation step, and fluid-structure coupling interface settings.

[0023] As preferred, in step S5, the Blade Editor in the Design Modeler software is used for parameterization, and a plurality of cross sections are equally divided between the hub surface and the cover surface in the impeller profile diagram, and each cross section corresponds to an angle curve in the following form:

[0024] Y Angle =AP M 2 +C

[0025] In the formula, Y Angle represents the angle curve matrix of each cross section, P M is the curve position matrix, A=[a1,a2,…,a m ] and C=[c1,c2,…,c m ] are coefficient matrices for dividing the number of cross sections; a1,a2,…,a m and c1,c2,…,c m are elements in the coefficient matrices A and C, respectively; and m represents the number of cross sections.

[0026] Then, the first point of the hub surface is taken as the first reference point, the last point of the hub surface is taken as the second reference point, the angle value of the first reference point and the angle value of the second reference point are taken as variables on the angle curve corresponding to the hub surface, when the angle value of the first reference point and the angle value of the second reference point change, the two angle values are substituted into the angle curve corresponding to the hub surface to obtain the optimized angle curve y1′ _Angle :

[0027] y1′ _Angle =a1′p M +c1′

[0028] In the formula, a1′ and c1′ are the quadratic coefficients of the optimized angle curve corresponding to the hub surface; p M represents the abscissa value of the angle curve corresponding to the hub surface.

[0029] In addition to the hub surface, the quadratic coefficients of the remaining cross sections are determined by the following formula:

[0030]

[0031] wherein a n and c n are the quadratic coefficients of the angle curve of the nth cross section, respectively. n and c n are the quadratic coefficients of the angle curve of the nth cross section, respectively.

[0032] As a preferred, in step S6, the optimization algorithm is a particle swarm optimization algorithm, and an equivalent entropy efficiency of the turboexpander is greater than 80% as a constraint condition, and a minimum of the flow velocity unevenness V s is an optimization target:

[0033]

[0034] wherein v k represents the velocity of the kth outlet node; v represents the average outlet velocity; and K represents the total number of outlet nodes.

[0035] As a preferred, in the two-way fluid-structure coupling data transmission module in step S7, the solid domain calculation model, the fluid domain calculation model, and the time step, the calculation step number, and the calculation time set in System Coupling are kept consistent, and the time for rotating the impeller of the expansion end of the turboexpander by 3° is taken as the time step.

[0036] As a preferred, in step S8, the acceleration variation time domain graph of the turboexpander optimization model and the acceleration variation time domain graph of the turboexpander original model are output by the post-processing software, the acceleration variation frequency spectrum graph of the turboexpander optimization model and the acceleration variation frequency spectrum graph of the turboexpander original model are obtained by the acceleration variation time domain graph of the turboexpander optimization model and the acceleration variation time domain graph of the turboexpander original model, the two frequency spectrum graphs are taken as a first frequency spectrum graph and a second frequency spectrum graph, respectively, when the acceleration amplitude of the impeller in the radial vibration direction at the peak frequency in the first frequency spectrum graph is less than the acceleration amplitude of the impeller in the radial vibration direction at the peak frequency in the second frequency spectrum graph and the acceleration amplitude difference between the two is less than a preset acceleration amplitude threshold, it is indicated that the flow-induced vibration characteristics of the turboexpander optimization model meet the requirements.

[0037] Compared with the prior art, the present application has the following beneficial effects:

[0038] The present application provides a method for synergistically suppressing impeller-vane flow-induced vibration. The dynamic and static interference effect between the impeller and the vane can cause strong vibration of the rotating machinery. However, the current optimization design only optimizes the impeller or the vane separately, and cannot effectively suppress the flow-induced vibration. The method of the present application conducts in-depth research on the dynamic and static interference effect between the impeller and the vane by introducing bidirectional fluid-structure coupling analysis, and considers the mutual influence between the two in the design stage, thereby realizing more accurate and effective optimization. Further, the calculation complexity is reduced and the simulation efficiency is improved by using geometric parameter dimension reduction processing. In addition, by selecting a suitable optimization algorithm and building a simulation optimization module, the synergistic iterative optimization of the impeller and the vane is realized. This method not only efficiently and low-costly improves the flow-induced vibration characteristics of the impeller and the vane, but also provides scientific guidance for the design of rotating machinery. At the same time, it helps to improve the overall stability and reliability of large rotating machinery, thereby prolonging the service life of the equipment, reducing maintenance costs, and reducing safety hazards. BRIEF DESCRIPTION OF DRAWINGS

[0039] Figure 1 A flowchart of the method of the present application;

[0040] Figure 2 A fluid domain three-dimensional model and a solid domain three-dimensional model schematic diagram of the expansion end of the turbo expander provided by the embodiment of the present application; wherein (a) is a fluid domain three-dimensional model schematic diagram, and (b) is a solid domain three-dimensional model schematic diagram;

[0041] Figure 3 A turbo expander impeller finite element calculation grid and a turbo expander vane finite element calculation grid schematic diagram provided by the embodiment of the present application;

[0042] Figure 4 A result graph when the turbo expander impeller is parameterized in the embodiment of the present application; wherein (a) is an impeller profile graph, (b) is an angle result graph of each section in the impeller profile graph, and (c) is a thickness result graph of each section in the impeller profile graph;

[0043] Figure 5 A bidirectional fluid-structure coupling data transmission path schematic diagram provided by the embodiment of the present application;

[0044] Figure 6 A time domain graph and a frequency spectrum graph of the impeller flow-induced vibration before and after optimization provided by the embodiment of the present application; wherein (a) is an acceleration change time domain graph of the original model of the turbo expander, (b) is an acceleration change time domain graph of the optimized model of the turbo expander, (c) is an acceleration change frequency spectrum graph of the original model of the turbo expander, and (d) is an acceleration change frequency spectrum graph of the optimized model of the turbo expander;

[0045] Figure 7The flow-induced vibration intensity comparison schematic diagram before and after optimization provided by the embodiment of the present application is shown. DETAILED DESCRIPTION

[0046] In order to make the above objectives, characteristics and advantages of the present application more apparent, specific embodiments of the present application are described in detail below with reference to the accompanying drawings. In the following description, numerous specific details are set forth in order to provide a thorough understanding of the present application. However, the present application can be practiced in a number of different manners other than those described herein, and it is understood that similar modifications of this description can be undertaken by one skilled in the art in the interest of the fair scope of the present application, and therefore the present application is not limited to the specific embodiments disclosed below. The technical features in each embodiment of the present application can be combined accordingly without conflict, provided that there is no conflict.

[0047] In the description of the present application, it should be understood that the terms "first", "second" are only used for distinguishing description purposes, and cannot be understood as indicating or implying relative importance or implicitly indicating the number of the technical features indicated. Therefore, the features defined with "first", "second" can explicitly or implicitly include at least one of the features.

[0048] In the present application, a kind of impeller-vane flow-induced vibration cooperative suppression method is provided, which is used for the optimization design of impeller and vane in rotating machinery, and can be specifically used for analyzing and optimizing the vibration intensity of impeller and vane caused by fluid, for example, the flow-induced vibration of turbine expander expansion end impeller and vane can be cooperatively optimized, and the vibration caused by the interaction between the remaining internal components of the impeller and vane can also be optimized.

[0049] As shown in the preferred implementation of the present application, Figure 1 In the present embodiment, the turbine expander expansion end impeller is selected for cooperative optimization to reduce the flow-induced vibration intensity, and the above-mentioned impeller-vane flow-induced vibration cooperative suppression method is used for analysis. The method includes the following S1-S8 steps, and the specific implementation process of each step is described below.

[0050] S1, the impeller and vane of the turbine expander expansion end are taken as the target rotating machinery, and the fluid domain three-dimensional model and the solid domain three-dimensional model of the target rotating machinery are established by using three-dimensional modeling software according to the structural characteristic parameters of the analyzed target rotating machinery.

[0051] It should be noted that in step S1 of the present application, the impeller and vane of the turbine expander expansion end are taken as the target rotating machinery, and the fluid domain three-dimensional model and the solid domain three-dimensional model of the turbine expander expansion end are established by using three-dimensional modeling software UG NX according to the structural characteristic parameters of the impeller and vane.

[0052] In step S1 of the present embodiment, as Figure 2(a) shown, the fluid domain three-dimensional model of the expansion end of the turbo expander includes four parts of the impeller, the guide vane, the inlet pipe and the outlet pipe, and each of the inlet pipe and the outlet pipe is extended, and an inlet extension section and an outlet extension section are formed correspondingly, so as to ensure that the boundary condition does not affect the numerical results in the fluid calculation domain. Figure 2 (b) shown, the solid domain three-dimensional model of the expansion end of the turbo expander sets the blade surface and the hub surface as the fluid-structure coupling interface, the solid domain three-dimensional model considers the influence of the earth gravity and ignores the mechanical vibration influence of the remaining components, and an ideal displacement constraint is adopted at the shaft hole of the impeller, that is, the XYZ direction displacement is limited.

[0053] S2, the fluid domain three-dimensional model is subjected to finite element grid division, then the turbo expander impeller finite element calculation grid and the turbo expander guide vane finite element calculation grid obtained by the finite element grid division are imported into the fluid domain simulation software, the fluid domain simulation conditions are set, the fluid domain calculation model is obtained, and the fluid domain calculation model is subjected to steady-state simulation, and the fluid domain steady-state simulation result is obtained.

[0054] It should be noted that in step S2 of the present application, the finite element grid division of the fluid domain three-dimensional model is realized by Turbo Grid and ICEM, the fluid domain simulation software is CFD simulation software, and the fluid domain simulation conditions include fluid material properties, turbulence model, inlet boundary condition, outlet boundary condition, impeller flow domain rotating speed, dynamic and static interface, wall surface setting and calculation step number.

[0055] In step S2 of the present embodiment, the models corresponding to the turbo expander impeller and the guide vane part are converted into TurboGrid for grid division, and the geometric models corresponding to the inlet pipe, the inlet extension section, the outlet pipe and the outlet extension section are subjected to grid division in ICEM-CFD. In order to meet the requirements of flow field analysis, the four parts of the fluid calculation domain are all converted into structured hexahedral elements, and the turbo expander impeller finite element calculation grid and the turbo expander guide vane finite element calculation grid are generated as shown in Figure 3 . Figure 3 The upper part of the figure is the turbo expander guide vane finite element calculation grid, and the lower part is the turbo expander impeller finite element calculation grid. At the same time, the grid quality is strictly controlled to ensure the reliability of numerical calculation, dense grids are arranged near the blade wall surface and the guide vane wall surface, a boundary layer is established, and complex flow structures are captured.

[0056] In step S2 of the embodiment, the CFD simulation software is used to perform steady-state simulation on the fluid domain calculation model under the rated design condition, the RNG k-ε model is selected as the turbulence model for steady flow field calculation, the turbine expander impeller area rotates at a speed of 24500 rpm, the remaining areas are static, the medium is compressible superheated water vapor, the rated mass flow of 100 t / h and the temperature of 698.15 K are used as the inlet boundary conditions, and the pressure of 2.2 MPa is used as the outlet boundary condition at the outlet.

[0057] S3, grid independence verification is performed on the steady-state simulation results of the fluid domain, after the grid independence verification, experimental measurement is performed on the three-dimensional model of the fluid domain, the experimental measurement results of the fluid domain are obtained, the experimental measurement results of the fluid domain are compared with the steady-state simulation results of the fluid domain, so as to ensure the accuracy of the three-dimensional steady-state flow field simulation inside the turbine expander.

[0058] It should be noted that in step S3, the initial grid model is first established and simulation is performed, then the grid is gradually encrypted, and the simulation is re-run after each encryption of the grid, and the grid independence verification of the steady-state simulation results of the fluid domain is completed.

[0059] It should be noted that in step S3, in order to study the independence of the steady-state simulation results of the fluid domain, five different size levels of prototype grids are generated and calculated under the rated condition. As can be seen from Table 1, with the increase of the number of grids, the outlet temperature T out and the value of the inlet pressure P in tend to be stable, the outlet temperature T out and the inlet pressure P in of scheme 3 and schemes 4 and 5 after scheme 3 have little difference and can be ignored. Therefore, considering the calculation accuracy and the calculation resources, the number of grids of the three-dimensional model of the fluid domain is set to 2.45×10 6 elements in the embodiment of the application.

[0060] Table 1. Grid independence verification

[0061]

[0062] The three-dimensional model of the fluid domain is experimented, and the accuracy of the numerical method is verified. Specifically, the measured outlet pressure, inlet pressure, outlet temperature and inlet temperature are obtained from the fluid domain experimental measurement results, the isentropic efficiency is calculated based on the measured outlet pressure, inlet pressure, outlet temperature and inlet temperature, and the isentropic efficiency calculation result is taken as the measured isentropic efficiency, the simulated outlet pressure, inlet pressure, outlet temperature and inlet temperature are obtained from the fluid domain steady-state simulation results, the isentropic efficiency is recalculated based on the simulated outlet pressure, inlet pressure, outlet temperature and inlet temperature, and the isentropic efficiency calculation result is taken as the simulated isentropic efficiency, the relative error of isentropic efficiency between the measured isentropic efficiency and the simulated isentropic efficiency is calculated, and if the relative error of isentropic efficiency is less than the preset accuracy threshold, it indicates that the three-dimensional steady-state flow field simulation inside the turboexpander is correct; otherwise, the steady-state simulation results of the fluid domain are obtained again until the three-dimensional steady-state flow field simulation inside the turboexpander is correct.

[0063] In this embodiment, the experimental test data and the numerical results of the three-dimensional model of the fluid domain are well fitted under the given rated operating condition. Under the design operating condition, the outlet temperature error is 0.63%, the inlet pressure error is 0.28%, and the relative error of isentropic efficiency is 4.69%. Since the numerical calculation does not consider factors such as gap leakage and mechanical friction loss, and the boundary conditions set in the rated operating condition are not completely consistent with the test results, there is a deviation between the numerical results and the experimental results. The overall calculation error basically meets the accuracy requirement, indicating that the numerical method adopted by the present application is acceptable.

[0064] Further, the isentropic efficiency η is calculated by the following formula:

[0065]

[0066] In the formula, κ is the gas constant of superheated steam, which is taken as 1.33 in this embodiment; P in is the inlet pressure, P out is the outlet pressure, T in is the inlet temperature, T out is the outlet temperature.

[0067] S4, the three-dimensional model of the solid domain is meshed, and the solid domain simulation condition is set to obtain the solid domain calculation model.

[0068] It should be noted that in step S4, the three-dimensional model of the solid domain is meshed according to the fluid-structure coupling theory under the premise of ensuring the grid mapping quality of the fluid-structure coupling interface. Specifically, the coupling strategy of CFX+ANSYS Mechanical+SystemCoupling is adopted, which is a node-based mapping relationship. In order to ensure the grid mapping quality of the fluid-structure coupling interface, the pressure at node i on the fluid domain grid is known as Pfi A circular region with radius r is determined with the jth node of the solid domain mesh as the center, and n nodes on the fluid domain mesh corresponding to the circular region are found. At this time, the pressure Pj at the jth node on the solid domain mesh sj may be expressed as:

[0069]

[0070] where d i is the spatial distance between the ith node of the fluid domain mesh and the jth node of the solid domain mesh.

[0071] The exciting load at the jth node on the solid domain mesh is calculated according to the following formula:

[0072] F sj =∫ S P sj dS

[0073] where S represents the finite element mesh area at the jth node of the solid domain mesh.

[0074] According to the above theory, the fluid-structure coupling interface mesh size should be as close as possible to better achieve the coupling effect and improve the simulation accuracy and speed. Therefore, the size of the mesh when the fluid domain three-dimensional model is divided into finite element meshes is taken as the basis, and the Ansys meshing tool in ANSYS Mechanical is used to divide the solid domain three-dimensional model into meshes. Due to the huge amount of calculation of two-way fluid-structure coupling, tetrahedral mesh structure is used to improve the calculation efficiency. The simulation conditions of the solid domain include the properties of the solid material, the constraint conditions, the gravity setting, the calculation time, the calculation step, and the fluid-structure coupling interface setting.

[0075] S5, based on the solid domain calculation model and the fluid domain calculation model, a fluid-structure coupling calculation model is established, without changing the basic structure of the fluid domain three-dimensional model and the solid domain three-dimensional model, the fluid-structure coupling calculation model is parameterized, the blade profile diagram and the guide vane profile diagram are obtained, and the angle and thickness in the blade profile diagram and the angle and thickness in the guide vane profile diagram are parameterized modeling, to obtain the reduced parameterized model.

[0076] It should be noted that in step S5, the Blade Editor in the Design Modeler software is used for parameterization, and the angle-thickness method in the Blade Editor is used for parameterized modeling of the blade and the guide vane. Among them, the angle refers to the angle between the tangent of the blade profile out-of-plane angle and the axial direction of the turbine at any point, and the thickness is defined as the maximum distance between the upper surface and the lower surface of the blade profile.

[0077] In the embodiment, the impeller profile diagram is as shown in Figure 4 (a). To ensure the accuracy of the parameterized modeling, the present application divides nine sections at equal intervals between the hub face (Span=0) and the shroud face (Span=1), the angle curve of each section is as shown in Figure 4 (b), and the thickness curve of each section is as shown in Figure 4 (c). The spline curve of the impeller airfoil is generated by using eight control points on the angle curve and the thickness curve of each section respectively. Therefore, in the optimization design process, the transverse and longitudinal coordinate values of each control point can be changed to change the impeller airfoil. To ensure the uniform distribution of the control points, the relative positions of the control points will not change in the optimization design process. This means that Figure 4 (b) and Figure 4 (c) transverse coordinate values remain unchanged.

[0078] It should be noted that in step S5, considering that the specific process of parameterizing the angle and thickness in the impeller profile diagram and the angle and thickness in the guide vane profile diagram is similar, the following will be described in detail by taking the parameterization of the angle in the impeller profile diagram as an example. First, the control point distribution of the impeller angle is curve-fitted. In the embodiment, the symmetric axis of the quadratic term fitting formula is y axis, and the regression coefficients R 2 of each section are all greater than 0.96, which indicates that the fitting effect is good. Here, other fitting formula forms can also be selected, as long as the regression coefficients R 2 of each section meet the preset accuracy requirement. Each section corresponds to an angle curve, which can be expressed as:

[0079] Y Angle = AP M 2 +C

[0080] In the formula, Y Angle represents the angle curve matrix of each section, P M is the curve position matrix, A=[a1, a2, …, a m ] and C=[c1, c2, …, c m ] are coefficient matrices for dividing the number of sections; a1, a2, …, a m and c1, c2, …, c m are elements in the coefficient matrices A and C respectively; and m represents the number of sections.

[0081] Then the first point (M-Prime=0) of the hub surface (Span=0) is taken as the first reference point, the last point (M-Prime=3.68) of the hub surface (Span=0) is taken as the second reference point, and the angle value θ1 of the first reference point and the angle value θ2 of the second reference point are taken as variables on the angle curve corresponding to the hub surface, when the angle value θ1 of the first reference point and the angle value θ2 of the second reference point change, the two angle values are substituted into the angle curve corresponding to the hub surface to obtain the optimized angle curve corresponding to the hub surface:

[0082] y1′ _Angle =a1′p M +c1′

[0083] In the formula, a1′ and c1′ are the quadratic coefficients of the optimized angle curve corresponding to the hub surface; p M represents the abscissa value of the angle curve corresponding to the hub surface;

[0084] In addition to the hub surface, the quadratic coefficients of the remaining sections are determined by the following formula:

[0085]

[0086] In the formula, a n ′ and c n ′ are the quadratic coefficients of the optimized angle curve of the nth section; a n and c n are the quadratic coefficients of the angle curve of the nth section.

[0087] Thus, the angle curve equation of all sections in the impeller profile diagram can be obtained, since the abscissa value of each section is unchanged during the optimization process, the corresponding angle curve equation is substituted to obtain the changed control point angle value. Thus, the impeller profile angle is parameterized modeled, and the parameters are geometrically reduced.

[0088] When the thickness of the impeller profile diagram is parameterized modeled, only the "impeller angle" in the foregoing process needs to be replaced by "impeller thickness". Specifically, first, the control point distribution of the impeller thickness is curve-fitted, and a thickness curve is obtained for each section. Then, the thickness value of the first reference point and the thickness value of the second reference point are taken as variables on the thickness curve corresponding to the hub surface, and the optimized thickness curve corresponding to the hub surface is obtained according to the foregoing process, and the coefficients of the remaining sections are determined to realize the parameterization modeling of the impeller profile thickness.

[0089] When the angle or thickness of the guide vane profile diagram is parameterized modeled, only the "impeller angle" in the foregoing process needs to be replaced by "guide vane angle", and the "impeller thickness" is replaced by "guide vane thickness", and the specific process is not described again.

[0090] S6, using the reduced dimension parameterized model, taking the parameter measuring the flow-induced vibration intensity as the optimization objective, and constructing a turboexpander original model with the performance of the target rotating machine as the constraint condition, and iteratively optimizing the constructed turboexpander original model by an optimization algorithm to obtain a turboexpander optimization model.

[0091] It should be noted that in step S6, the optimization algorithm is a particle swarm algorithm, which generally exhibits a fast convergence speed and effectively solves nonlinear problems. And taking the minimization of the non-uniformity V s as the optimization objective:

[0092]

[0093] In the formula, v k represents the speed of the kth outlet node; v is the average outlet speed; and K is the total number of outlet nodes.

[0094] In this embodiment, the joint simulation technology of Matlab and Ansys software is adopted to realize the optimization design of the internal flow characteristics of the fluid machine. In this process, a method of directly modifying the guide vane and impeller contour lines is adopted for structural optimization. First, the position of the particle swarm is initialized in the Matlab environment. Then, based on the parameterized modeling method in the foregoing steps, the variables in the optimization algorithm are converted into specific geometric parameters, and are imported into the Ansys software for geometric shape updating and mesh division. Subsequently, the finite element analysis (FEA) is performed according to the updated geometric model to evaluate the flow-induced vibration under different design schemes. After completing the finite element simulation, the key performance indicators are extracted from the simulation results, and are fed back to the particle swarm optimization algorithm in Matlab as the fitness function value. This iterative process will continue until the preset convergence criteria are met, that is, the variation amplitude of the optimal solution in the last ten iteration cycles is less than 10 -4 .

[0095] In this embodiment, the unsteady pressure fluctuation intensity is often used as a key indicator for measuring the water-induced vibration performance. The flow uniformity at the impeller outlet is directly related to the formation and evolution of the internal unsteady flow structure of the expander. Based on this, the expander outlet flow velocity non-uniformity is adopted to quantify the flow-induced vibration intensity.

[0096] It is worth noting that while pursuing the optimization of flow-induced vibration performance, the thermal performance of the turboexpander must be taken into account to avoid its deterioration due to adjustment of geometric characteristics. To this end, the present application sets a clear constraint: the isentropic efficiency of the turboexpander must be greater than 80%. In summary, the present application aims to explore and build an optimized model of the turboexpander that can minimize flow-induced vibration under the design condition of meeting the isentropic efficiency requirement (not less than 80%).

[0097] S7, a two-way fluid-structure coupling data transmission module of CFX, Transient Structural and System Coupling is built in Ansys Workbench, and the turboexpander optimization model and the turboexpander original model are simulated in two-way fluid-structure coupling in the two-way fluid-structure coupling data transmission module.

[0098] It should be noted that in step S7, the CFX+Transient Structural+System Coupling two-way fluid-structure coupling data transmission module is built in Ansys Workbench, and the data transmission path is as shown in Figure 5 , thereby performing two-way fluid-structure coupling simulation. In order to simulate the real flow field inside the turboexpander before and after optimization, in order to facilitate the convergence of transient calculation, the steady-state calculation results are used as the initial conditions of transient calculation after steady-state calculation. The unsteady analysis is still based on the commercial software CFX, and the mesh deformation is enabled in the rotating area to achieve two-way fluid-structure coupling. In the two-way fluid-structure coupling setting, the time step, the calculation step and the calculation time set in the System Coupling correspond to the solid domain calculation model and the fluid domain calculation model; the upper limit of each time step iteration is set to 20 steps, and the convergence tolerance is set to 1x10 -4 , the dynamic and static interface of the flow field is selected as the transient rotor model, and the remaining boundary conditions are consistent with the steady-state calculation; the time step is set to 3° of the impeller rotation at the expansion end of the turboexpander. Since the rotational speed of the turboexpander is 24500r / min, the time step is 2.04x10 -5 s, i.e. 120 time steps for one rotation period, the total sampling time is 15 rotation periods, and the total time steps are 1800, and the total simulation time is 0.03672s.

[0099] S8, the two-way fluid-structure coupling simulation results of the turboexpander optimization model and the two-way fluid-structure coupling simulation results of the turboexpander original model are imported into the post-processing software for post-processing, the acceleration change of the impeller at the expansion end of the turboexpander is used as an evaluation index, and the turboexpander optimization model that meets the preset requirements of flow-induced vibration characteristics is output, realizing the collaborative suppression of impeller-vane flow-induced vibration.

[0100] In the embodiment, in order to measure the vibration of the impeller, the acceleration change of the impeller is used as an evaluation index, and the acceleration change time domain of the optimized model of the turboexpander (i.e., the post-optimization model acceleration time domain in Figure 6 ) and the acceleration change time domain of the original model of the turboexpander (i.e., the baseline model acceleration time domain in Figure 6 ) are output by the post-processing software, as shown in Figure 6 . It can be seen from Figure 6 that the vibrations in the X direction and the Y direction are both radial vibrations, and the vibration amplitudes and laws are almost the same. In contrast, the vibration in the Z direction is axial vibration, and the vibration amplitude is much lower than the radial vibration due to the constraint condition. Therefore, in the flow-induced vibration of the impeller, the radial vibration occupies a dominant position. Then, the acceleration change frequency spectrum of the optimized model of the turboexpander (i.e., the post-optimization model acceleration frequency spectrum in Figure 6 ) and the acceleration change frequency spectrum of the original model of the turboexpander (i.e., the baseline model acceleration frequency spectrum in Figure 6 ) are obtained from the acceleration change time domain of the optimized model of the turboexpander and the acceleration change time domain of the original model of the turboexpander, and the two frequency domain graphs are correspondingly taken as a first frequency spectrum graph and a second frequency spectrum graph. According to the data analysis of Figure 6 (c) and Figure 6 (d), the radial vibration of the impeller is mainly characterized by 2f GV as the main characteristic frequency, and a peak value also appears at f GV . In contrast, the axial vibration of the impeller only has a peak value at f GV .

[0101] When the acceleration amplitude of the radial vibration of the impeller in the first frequency spectrum graph at the peak frequency is less than the acceleration amplitude of the radial vibration of the impeller in the second frequency spectrum graph at the peak frequency, and the difference between the two acceleration amplitudes is less than a preset acceleration amplitude threshold, it indicates that the flow-induced vibration characteristics of the optimized model of the turboexpander meet the requirements.

[0102] In the embodiment, in order to further quantify the vibration changes before and after optimization, the acceleration amplitudes of the above directions at the peak frequency are compared, and the results are shown in Figure 7 . It can be seen from Figure 7 that the amplitude of the radial vibration decreases significantly, and the acceleration amplitude of the optimized model of the turboexpander is reduced by 916.7 m / s 2, the descending ratio reached 74.3%. According to the previous analysis, the radial vibration plays a dominant role in the overall flow-induced vibration of the impeller. Therefore, by optimizing the radial vibration, the overall vibration level of the expander is significantly reduced. It should be noted that in some specific positions, the vibration amplitude appears a small rise, but still has little effect on the overall vibration. This phenomenon is due to the consideration of the overall effect in the optimization process, and the local optimum is properly compromised. This result again verifies the effectiveness and rationality of the method of the present application.

[0103] The above-described embodiments are only a preferred scheme of the present application, and are not intended to limit the present application. Those skilled in the art can make various changes and modifications without departing from the spirit and scope of the present application. Therefore, any technical solutions obtained by equivalent replacement or equivalent transformation shall fall within the protection scope of the present application.

Claims

1. A method for suppressing flow-induced vibration of an impeller-vane, characterized in that, The method comprises the following steps: S1, taking the impeller and guide vane of the expansion end of the turboexpander as a target rotating machine, a three-dimensional model of a fluid domain and a three-dimensional model of a solid domain of the target rotating machine are established by using a three-dimensional modeling software according to the structural characteristic parameters of the target rotating machine; S2, the three-dimensional model of the fluid domain is subjected to finite element meshing, then the turboexpander impeller finite element calculation grid and the turboexpander guide vane finite element calculation grid obtained by the finite element meshing are imported into a fluid domain simulation software, the fluid domain simulation conditions are set, a fluid domain calculation model is obtained, and the fluid domain calculation model is subjected to steady-state simulation to obtain a fluid domain steady-state simulation result; S3, the fluid domain steady-state simulation result is subjected to grid independence verification, the three-dimensional model of the fluid domain is subjected to experimental measurement after the grid independence verification, a fluid domain experimental measurement result is obtained, and the fluid domain experimental measurement result is compared with the fluid domain steady-state simulation result to ensure the accuracy of the three-dimensional steady-state flow field simulation inside the turboexpander; S4, the three-dimensional model of the solid domain is subjected to meshing, and the solid domain simulation conditions are set to obtain a solid domain calculation model; S5, a fluid-structure coupling calculation model is established based on the solid domain calculation model and the fluid domain calculation model, the fluid-structure coupling calculation model is parameterized without changing the basic structures of the three-dimensional model of the fluid domain and the three-dimensional model of the solid domain, the impeller profile and the guide vane profile are obtained, and the angle and thickness in the impeller profile and the angle and thickness in the guide vane profile are subjected to parameterized modeling to obtain a reduced parameterized model; S6, the reduced parameterized model is used to minimize the flow-induced vibration strength parameter as an optimization objective, the performance of the target rotating machine is taken as a constraint condition, a turboexpander original model is constructed, an optimization algorithm is used to iteratively optimize the constructed turboexpander original model, and a turboexpander optimization model is obtained; S7, a bidirectional fluid-structure coupling data transmission module of CFX, Transient Structural and System Coupling is constructed in Ansys Workbench, and the turboexpander optimization model and the turboexpander original model are subjected to bidirectional fluid-structure coupling simulation in the bidirectional fluid-structure coupling data transmission module; S8, the bidirectional fluid-structure coupling simulation result of the turboexpander optimization model and the bidirectional fluid-structure coupling simulation result of the turboexpander original model are imported into post-processing software for post-processing, the acceleration change of the impeller of the expansion end of the turboexpander is taken as an evaluation index, the turboexpander optimization model whose flow-induced vibration characteristics meet preset requirements is output, and the flow-induced vibration of the impeller-guide vane is cooperatively suppressed.

2. The method of claim 1, wherein the flow-induced vibration of the impeller and the guide vane is suppressed in cooperation. In step S1, the three-dimensional modeling software is UG NX; the three-dimensional model of the fluid domain of the expansion end of the turbo expander includes four parts of the impeller, the guide vane, the inlet pipeline and the outlet pipeline, and each of the inlet pipeline and the outlet pipeline is extended to form an inlet extension section and an outlet extension section; the three-dimensional model of the solid domain of the expansion end of the turbo expander sets the blade surface and the hub surface as the fluid-structure coupling interface, and considers the influence of the earth gravity and ignores the mechanical vibration influence of the rest of the components, and an ideal displacement constraint is adopted at the shaft hole of the impeller.

3. The method of claim 1, wherein the flow-induced vibration of the guide vane is suppressed by the guide vane and the impeller in cooperation. In step S2, the three-dimensional model of the fluid domain is divided into finite element grids by using Turbo Grid and ICEM, the fluid domain simulation software is CFD simulation software, and the fluid domain simulation conditions include fluid material properties, a turbulence model, inlet boundary conditions, outlet boundary conditions, impeller flow domain rotation speed, dynamic and static interface, wall surface setting and calculation step number.

4. The method of claim 3, wherein the flow-induced vibration of the guide vane is suppressed by the guide vane being arranged to have a shape that is different from the shape of the impeller. In step S3, an initial grid model is first established and simulation is performed, then the grid is gradually encrypted, and after each encryption of the grid, the simulation is re-run to verify the grid independence of the steady-state simulation results of the fluid domain; The measured outlet pressure, inlet pressure, outlet temperature and inlet temperature are obtained from the fluid domain experimental measurement results, the isentropic efficiency is calculated based on the measured outlet pressure, inlet pressure, outlet temperature and inlet temperature, and the isentropic efficiency calculation result is taken as the measured isentropic efficiency, the simulated outlet pressure, inlet pressure, outlet temperature and inlet temperature are obtained from the steady-state simulation results of the fluid domain, the isentropic efficiency is recalculated based on the simulated outlet pressure, inlet pressure, outlet temperature and inlet temperature, and the isentropic efficiency calculation result is taken as the simulated isentropic efficiency, the relative error of the isentropic efficiency between the measured isentropic efficiency and the simulated isentropic efficiency is calculated, and if the relative error of the isentropic efficiency is less than a preset precision threshold, it indicates that the three-dimensional steady-state flow field simulation inside the turbo expander is correct; otherwise, the steady-state simulation results of the fluid domain are re-acquired until the three-dimensional steady-state flow field simulation inside the turbo expander is correct.

5. The method of claim 4, wherein the flow-induced vibration of the guide vane is suppressed by the guide vane being arranged to have a shape that is different from the shape of the impeller. The isentropic efficiency η is calculated by the following formula: where k is the gas constant for superheated water vapor; P in is the inlet pressure, P out is the outlet pressure, T in is the inlet temperature, T out is the outlet temperature.

6. The method of claim 1, wherein the flow-induced vibration of the guide vane is suppressed by the guide vane and the impeller in cooperation. In step S4, the grid size during the finite element grid division of the three-dimensional model of the fluid domain is taken as the benchmark, the Ansys meshing tool in ANSYS Mechanical is used to divide the grid of the three-dimensional model of the solid domain, the tetrahedral grid structure is used when the three-dimensional model of the solid domain is divided into grids, and the simulation conditions of the solid domain include solid material properties, constraint conditions, gravity setting, calculation time, calculation step length and fluid-structure coupling interface setting.

7. The method of claim 1, wherein the flow-induced vibration of the guide vane is suppressed by the guide vane and the impeller in cooperation. In step S5, the Blade Editor in the Design Modeler software is used for parameterization, a plurality of cross sections are equally divided between the hub surface and the cover surface in the impeller profile diagram, and each cross section corresponds to an angle curve in the following form: Y Angle = AP M 2 + C where Y Angle represents each cross-section angle curve matrix, P M is a curve position matrix, A = [a1, a2, …, am] and C = [c1, c2, …, cm] are coefficient matrices for dividing the number of cross-sections; a1, a2, …, am and c1, c2, …, cm are elements in the coefficient matrices A and C, respectively; and m represents the number of cross-sections. m m m m ​​​​ Then the first point of the hub surface is taken as a first reference point, the last point of the hub surface is taken as a second reference point, and on the corresponding angle curve of the hub surface, the angle value of the first reference point and the angle value of the second reference point are taken as variables, when the angle value of the first reference point and the angle value of the second reference point change, the two angle values are substituted into the corresponding angle curve of the hub surface to obtain the optimized angle curve y1′ of the hub surface corresponding to the angle curve _Angle : y1′ _Angle = a1′p M + c1′ wherein a1' and c1' are the quadratic coefficients of the optimized angle curve corresponding to the hub face; p M represents the abscissa value of the angle curve corresponding to the hub face; The quadratic coefficient of the rest of the cross sections except the hub surface is determined by the following formula: In the formula, a n and c n are the quadratic coefficients of the angle curve of the nth cross section, respectively. n and c n are the quadratic coefficients of the angle curve of the nth cross section, respectively.

8. The method of claim 1, wherein the flow-induced vibration of the guide vane is suppressed by the guide vane and the impeller in cooperation. In step S6, the optimization algorithm is a particle swarm optimization algorithm, and the constraint condition is that the equivalent entropy efficiency of the turboexpander is greater than 80%, and the objective function is to minimize the outlet flow velocity non-uniformity V s The optimization objective is: where v k denotes the velocity of the kth exit node; is the average exit velocity; K is the total number of exit nodes.

9. The method of claim 1, wherein the flow-induced vibration of the guide vane is suppressed by the guide vane and the impeller in cooperation. In the two-way fluid-structure coupling data transmission module of step S7, the solid domain calculation model and the fluid domain calculation model correspond to the time step, the calculation step number and the calculation time set in System Coupling and keep consistent, and the time for rotating the impeller of the expansion end of the turboexpander by 3° is taken as the time step.

10. The method of claim 1, wherein the flow-induced vibration of the guide vane is suppressed in cooperation with the flow-induced vibration of the impeller. In step S8, the acceleration variation time domain graph of the turboexpander optimization model and the acceleration variation time domain graph of the turboexpander original model are output by the post-processing software, the acceleration variation frequency spectrum graph of the turboexpander optimization model and the acceleration variation frequency spectrum graph of the turboexpander original model are obtained by the acceleration variation time domain graph of the turboexpander optimization model and the acceleration variation time domain graph of the turboexpander original model, the two frequency spectrum graphs are taken as a first frequency spectrum graph and a second frequency spectrum graph, and when the acceleration amplitude of the impeller in the radial vibration direction at the peak frequency in the first frequency spectrum graph is less than the acceleration amplitude of the impeller in the radial vibration direction at the peak frequency in the second frequency spectrum graph and the acceleration amplitude difference between the two is less than a preset acceleration amplitude threshold, it is indicated that the flow-induced vibration characteristics of the turboexpander optimization model meet the requirements.

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