A rotor imbalance fault test bench and design optimization method

Abstract

The present invention discloses a rotor unbalanced fault test bench and a design optimization method. The fault test bench comprises a support base, a bearing support seat, a drive motor, a bearing, a turntable, a rotating shaft and a coupling. The support base is provided with a bearing support seat and a drive motor for providing support for the bearing support seat and the drive motor. The bearing support seat is provided with a bearing. The drive motor is connected to the rotating shaft through a coupling. The rotating shaft is connected to the bearing and the turntable. The turntable is a single-stage turntable or a plurality of stage turntables. An unbalanced mass block is provided on the turntable. The fault test bench design optimization method comprises: establishing a rotor dynamics model; setting the rotating shaft parameters; setting the unbalance parameters; and setting the blade parameters.

Description

Technical Field

[0001] The invention belongs to the field of rotating machinery performance testing and fault diagnosis, and in particular to a rotor unbalance fault simulation test bench and a design optimization method thereof. Background Art

[0002] The rotor system is the core component of rotating machinery. In the working process, typical faults represented by imbalance faults are difficult to completely eliminate, which seriously affects the operating safety and working efficiency of rotating machinery. Rotor imbalance is a fault caused by eccentricity of rotor parts or defects in rotor parts. It is the most common fault in rotating machinery. Therefore, fault diagnosis of rotating machinery rotor systems has always been a long-term research task in this field.

[0003] However, the rotor system of rotating machinery is an extremely complex nonlinear time-varying system. Due to the constraints of various factors such as manufacturing cost, test conditions, test cycle, and test technology, it is still difficult to carry out dynamic test research on it. At the same time, the fault state of the rotating machinery cannot be observed at any time during operation, and once a fault occurs, the rotating machinery will be greatly damaged. Therefore, designing a simulated rotor test bench and using the rotor test bench to obtain rotor operation data is an important means to realize fault diagnosis data research and algorithm verification.

[0004] At present, most of the research on rotor faults is limited to rotor dynamics research, fault models and fault diagnosis algorithms based on simulated rotor test benches, and there is little research on the optimization design method of rotor fault test benches. The optimal rotor test bench can achieve high-efficiency and low-cost rotating machinery data simulation, making fault diagnosis data analysis and algorithm verification more efficient. Summary of the invention

[0005] The purpose of the present invention is to overcome the deficiencies in the prior art, provide a rotor unbalance fault test bench and a design method, and realize the solution of the optimal parameters of the rotor unbalance fault test bench.

[0006] The objective of the present invention is achieved through the following technical solutions:

[0007] A rotor imbalance fault test bench comprises a support base, a bearing support seat, a drive motor, a bearing, a turntable, a rotating shaft and a coupling. The support base is provided with a bearing support seat and a drive motor for providing support for the bearing support seat and the drive motor. The bearing support seat is provided with a bearing. The drive motor is connected to the rotating shaft through a coupling. The rotating shaft is connected to the bearing and the turntable. The turntable is a single-stage turntable or a plurality of stage turntables. An unbalanced mass block is provided on the turntable.

[0008] The present invention also provides a rotor imbalance fault test bench design optimization method, comprising:

[0009] S1. Simplify the rotating mechanical rotor system consisting of bearings, a rotating shaft and several stages of rotating disks into a double-supported rotor system; establish a rotor dynamics model according to the lumped parameter method, the rotor dynamics model includes an elastic shaft, a pair of rolling bearings, and n rigid rotating disks;

[0010] S2. Set the axis parameters as follows:

[0011] (201) According to the size requirements of the experimental platform and the sensor installation requirements, determine the shaft material, total shaft length, distance between each shaft section, support point distance and number of turntable levels;

[0012] (202) According to the provisions of the mechanical design manual for the length-to-diameter ratio of ordinary shafts, based on the determined total length of the shaft, the shaft diameter traversal range is calculated according to the length-to-diameter ratio being greater than 5 and less than 30;

[0013] (203) Selecting 1 / 10 of the shaft diameter traversal range as the step size to traverse the shaft diameter, and calculating the shaft mid-span stiffness coefficient under each diameter;

[0014] (204) Determine the allowable value of shaft deflection for each diameter according to the Machinery Manual;

[0015] (205) Determine the maximum unbalanced force F at each diameter max ; The calculation formula is: F max =k mid y max -m k g, where k mid is the mid-span stiffness coefficient of the rotating shaft, y max is the allowable value of the shaft deflection; m k is the mass of the turntable when there is an unbalanced fault on the kth turntable, and g is the acceleration of gravity;

[0016] (206) solving the vibration response of the shaft system using the established rotor dynamics model, solving the vibration response of the shaft system in a normal state and a fault state according to the forced vibration differential equation and the parameters determined above, selecting a range with the largest vibration response value, and performing a diameter traversal within the range with a step length of 1 mm; the fault state refers to a state where the maximum unbalanced force is introduced;

[0017] (207) After traversing with a step length of 1 mm, the shaft diameter with the largest vibration response is obtained, and the shaft diameter, shaft stiffness, and maximum unbalanced force when the vibration response difference is the largest are selected as the optimal parameters;

[0018] S3. Setting unbalanced parameters, including maximum unbalanced force, unbalanced force caused by unbalanced block, unbalanced mass of unbalanced block, installation position of unbalanced block and rotational angular velocity of turntable;

[0019] S4. Set blade parameters.

[0020] Furthermore, the first-order resonance frequency of the blades of the rotating disk should be at least two orders of magnitude higher than the operating speed of the rotor, so that the tip clearance data is not coupled with the blade vibration information.

[0021] Furthermore, the mid-span stiffness coefficient of the rotating shaft must meet the following conditions:

[0022] ① Under normal conditions, the vibration response caused by gravity and initial unbalanced force should not be greater than the vibration response specified in the accuracy level determined according to ISO1940 before design.

[0023] ② Under the fault state, the vibration response caused by gravity and maximum unbalanced force should not be less than the minimum value of the vibration response under the fault determined before design.

[0024] Compared with the prior art, the technical solution of the present invention has the following beneficial effects:

[0025] 1. The present invention provides a method for optimizing the design of a rotor unbalance fault test bench. Through this method, the shaft parameters, unbalance parameters and blade parameters of a multi-stage rotor unbalance fault test bench can be obtained, and the test bench is designed according to the obtained parameters. The method in the present invention uses the concentrated mass method to establish a rotor dynamics model, and uses the rotor dynamics model to solve the vibration response of the rotor system under normal operation and fault conditions. Provide guidance for the design of an actual test bench. In addition, the present invention obtains the shaft parameters that can achieve the maximum fault response under the maximum unbalanced force by traversing the shaft diameter, that is, the optimal parameters of the rotor unbalance fault test bench shaft. The shaft parameters are used to design the unbalance parameters and blade parameters.

[0026] This method can achieve the maximum unbalanced fault response under the condition of safe operation of the test bench, provide theoretical guidance for the unbalanced fault test bench, improve the safety of the unbalanced rotor fault test bench, and ensure that the most obvious unbalanced fault response can be obtained, reducing the trial and error cost, optimizing the design process of the unbalanced fault test bench, and helping to achieve high-efficiency and low-cost rotating machinery data simulation.

[0027] 2. The present invention provides an unbalanced rotor fault test bench. The test bench can simulate the unbalanced fault of the rotor according to the parameters set in the design requirements, and provide test data for fault diagnosis data analysis and algorithm verification. Compared with the existing test benches, the rotating shaft, unbalanced force and blade parameters of the rotor unbalanced fault test bench are designed using a design optimization method. The maximum unbalanced fault response can be achieved while ensuring the safe operation of the test bench, and the generation of fault data can be better realized, providing effective test data for subsequent measurement systems.

[0028] 3. The present invention establishes a rotor dynamics model, utilizes shaft diameter traversal and unbalanced mass block traversal to obtain the optimal rotor parameters, realizes the optimal design of the rotor unbalance test bench, obtains the optimal shaft parameters, unbalanced fault parameters and blade parameters, thereby realizing high-efficiency and low-cost rotating machinery data simulation, making fault diagnosis data analysis and algorithm verification more efficient. BRIEF DESCRIPTION OF THE DRAWINGS

[0029] Figure 1 This is the structural diagram of the rotor imbalance fault test bench.

[0030] Figure 2 Schematic diagram of the blade disk structure.

[0031] Figure 3 It is a schematic diagram of a double-supported multi-rotor.

[0032] Figure 4 Schematic diagram of stiffness calculation for different shaft segments.

[0033] Figure 5 Schematic diagram of shaft imbalance.

[0034] Figure 6 Schematic diagram of turntable imbalance.

[0035] Figure 7 Flowchart for shaft parameter optimization. DETAILED DESCRIPTION

[0036] The present invention will be further described in detail below in conjunction with the accompanying drawings and specific embodiments. It should be understood that the specific embodiments described herein are only used to explain the present invention and are not used to limit the present invention.

[0037] like Figure 1 As shown, a rotor unbalance fault test bench includes a support base 1, a bearing support seat 2, a drive motor 3, a bearing 4, a turntable 5, a rotating shaft 6 and a coupling 7. The support base 1 is equipped with a bearing support seat 2 and a drive motor 3 for providing support for the bearing support seat 2 and the drive motor 3. The bearing support seat 2 is equipped with a bearing 4. The drive motor 3 is connected to the rotating shaft 6 through the coupling 7. The rotating shaft 6 is connected to the bearing 4 and the turntable 5. The turntable 5 is a single-stage turntable or a plurality of turntables. A single turntable unit 9 in the turntable 5 can be installed with an unbalanced mass block 8. During the rotation of the rotor, the unbalanced force provided by the unbalanced mass block 8 is used to add an unbalanced fault. The drive motor 3 drives the rotor system including the bearing 4, the rotating shaft 6 and the turntable 5, and the bearing support seat 2 supports the bearing 4.

[0038] In the above-mentioned rotor imbalance fault test bench, the rotating shaft 6 is designed by the rotating shaft parameter design optimization method of the present invention, the turntable blades in the single-stage or multi-stage turntable 5 are designed by the blade parameter design method of the present invention, and the installation position and mass of the unbalanced mass block 8 are designed by the unbalance parameter design method of the present invention.

[0039] The design optimization method of the rotor unbalance fault test bench includes the construction of the rotor dynamics model, the optimization design of the shaft parameters, the unbalance parameter design and the blade parameter design. The construction of the rotor dynamics model is the basis for the subsequent optimization design of the shaft parameters and unbalance parameters. The rotor dynamics model constructs the forced vibration differential equation of the rotor. When optimizing the shaft parameters, after the excitation force matrix has been obtained, the corresponding rotor shaft system vibration and axis trajectory response can be obtained by inputting the rotor dynamics model, and then the shaft system vibration and axis trajectory response can be constrained according to the obtained shaft system vibration and axis trajectory response, and the optimal parameters that can obtain the maximum fault response can be obtained. The unbalance parameter design can calculate the optimal unbalance parameter according to the maximum unbalance force obtained by the shaft parameter optimization design. In addition, in the blade parameter design, the resonant frequency of the blade can be constrained according to the rotation angular velocity obtained in the unbalance parameter design, so as to obtain the appropriate blade parameters.

[0040] The details are as follows:

[0041] 1. Establishing the rotor dynamics model

[0042] The rotor system of rotating machinery usually includes bearings, rotating shafts and multi-stage turntables, which can be simplified into a double-support multi-stage rotor system. Figure 1 According to the concentrated parameter method, a rotor dynamics model is established. The model includes an elastic shaft, a pair of rolling bearings, and n rigid turntables. The bearings and turntables are equivalent to (n+2) concentrated mass points, m1~m n+2 Among them, m1, m n+2 are the concentrated masses of the left and right bearings, m2~m n+1 are the concentrated masses of the n turntables. The elastic shaft mass is uniformly loaded between m2 and m n+1 On a turntable. 12 ~k n+1n+2 are the elastic axis stiffness of each section, c1~c n+2 Provide damping for each bearing or turntable.

[0043] Figure 2 The rotor system shown is a multi-degree-of-freedom system. When the rotor rotates, it will generate forced vibration responses in the x and y directions. Taking the x direction as an example, the forced vibration differential equation is:

[0044]

[0045] Among them, M is the mass matrix, C is the damping matrix, K is the stiffness matrix, X is the vibration displacement matrix in the x direction, and F x is the exciting force matrix in the x direction. Among them, the mass matrix M is:

[0046]

[0047] The damping matrix C is:

[0048]

[0049] Since there is stiffness coupling between adjacent turntables and bearings, or between turntables, the stiffness matrix K is:

[0050]

[0051] The x-direction vibration displacement matrix X is:

[0052] X=(x1 x2 … x n+1 x n+2 ) T (5)

[0053] When the rotor is in normal state, the force in the x direction is only the reaction force of the left and right bearings, which can be expressed as:

[0054] F x =(F lx 0 … 0 F rx ) T (6)

[0055] Among them, F lx 、F rx is the reaction force of the left and right bearings in the x direction, which is related to the bearing stiffness and the displacement of the concentrated mass point and can be expressed as:

[0056]

[0057] Among them, k1, k n+2 is the stiffness of the left and right bearings.

[0058] The bearing reaction force in the y direction is similar to that in the x direction. In addition, due to the gravity, the force on the rotor system in the normal state also includes the gravity of each bearing and the turntable. The exciting force matrix F in the y direction is Y It can be expressed as:

[0059] F Y =(F ly -m1g -m2g ... -m n+1 g by -m n+2g ) T (8)

[0060] Among them, F ly 、F ry is the support reaction force of the left and right bearings in the y direction, which is calculated in the same way as the x direction, and g is the gravitational acceleration at the location of the test bench. Substituting equations (2) to (8) into equation (1), the rotor system dynamics model based on the concentrated mass method is obtained. The model is a set of two ordinary differential equations in the x and y directions. Based on this model, numerical methods can be used to obtain the vibration and axis trajectory response of the rotor shaft system and analyze the dynamic characteristics of the rotor in different states.

[0061] The rotor unbalance fault can be equivalent to the mass eccentricity of the rotor. Taking the unbalance fault of the middle turntable of the mechanical rotor system as an example, the rotor rotation geometric center is offset from O to O', such as Figure 4 ; The rotor unbalance fault is equivalent to the distance e from the geometric center of the turntable m There is an unbalanced block at , with a mass of m, such as Figure 5 .

[0062] Figure 6 In the rotor unbalance fault shown, the geometric center is located at O', and the mass center is located at point C. The distance e between the geometric center and the mass center is the eccentricity, and the calculation formula is:

[0063]

[0064] Where M is the mass of the turntable. When the kth turntable has an unbalanced fault, M = m k .

[0065] The unbalanced force caused by adding an unbalanced mass block is:

[0066] F e =(M+m)eω 2 (10)

[0067] The unbalanced forces in the x and y directions can be expressed as:

[0068]

[0069] Where ω is the rotor angular velocity and t is the current moment. Taking the x direction as an example, the exciting force matrix under unbalanced fault can be expressed as:

[0070] F ex =(F lx 0 … 0 F ex 0 … 0 F rx ) T (12)

[0071] where F ex In the k+1th row of the matrix in formula (12).

[0072] Substituting equation (12) as the exciting force matrix into equation (1), the dynamic equation under rotor unbalance fault is obtained.

[0073] 2. Shaft parameter design

[0074] The shaft parameter design process is as follows: Figure 7 As shown in the figure, when optimizing the shaft parameters, the shaft stiffness and the maximum allowable deflection value are calculated by traversing the shaft diameter, and then the maximum unbalanced force F under each shaft diameter is obtained. max By solving the rotor system in normal state, fault state (i.e. the maximum unbalanced force F max The shaft vibration response value under the action of the shaft system is selected, and the shaft diameter, shaft stiffness, and maximum unbalanced force when the vibration response difference is the largest are selected as the optimal parameters. The specific process of shaft parameter design is as follows:

[0075] (201) According to the size requirements of the test bench and the sensor installation requirements, determine the shaft material, the total length of the shaft l, and the distance x from each shaft segment to the support point. nn+1 And the number of blade disks n.

[0076] (202) According to the provisions of the mechanical design manual for the length-to-diameter ratio of ordinary shafts, based on the determined total length of the shaft, the shaft diameter traversal range is calculated based on the length-to-diameter ratio being greater than 5 and less than 30.

[0077] (203) First, select 1 / 10 of the shaft diameter traversal range as the step size to traverse the shaft diameter, and calculate the shaft mid-span stiffness coefficient k under each diameter through the following process: mid .

[0078] The shaft stiffness is determined by the shaft material, length, diameter and other parameters, and is the most critical parameter of the rotor system. For a simply supported beam supported at both ends, the stiffness is expressed as:

[0079]

[0080] Among them, E is the elastic modulus of the shaft, l is the length of the shaft between the two supports, x is the distance from the left support point, and I is the section moment of inertia. The calculation formula is:

[0081]

[0082] Where d is the shaft diameter.

[0083] For the stiffness coefficient in the formula, the distances of different shaft segments from the left support point are as follows: Figure 3 As shown. According to the figure, the results can be obtained:

[0084]

[0085] Where w1 is the bearing width, w2 is the blade width, l1 is the distance from the bearing to the nearest blade, l2 is the distance between adjacent blades, and x nn+1 is the distance from the center of the nth shaft segment between the two bearings to the left support. Substituting the result into the stiffness calculation formula (13), the stiffness values ​​of different shaft segments can be obtained. The maximum deflection position is recorded as the mid-span stiffness coefficient k of the rotating shaft. mid .

[0086] Mid-span stiffness coefficient k mid Directly determines the maximum unbalanced force F max The shaft system vibration and axis trajectory response under the action of the rotor system. In order to distinguish the different operating states of the rotor system, the difference between the response of the rotor test bench in normal and fault conditions should be an order of magnitude. At the same time, the shaft must pass the safety check. In summary, the mid-span stiffness coefficient of the shaft should meet the following conditions:

[0087] ① Under normal conditions, the vibration response caused by gravity and initial unbalanced force should not be greater than the vibration response specified by the accuracy level determined according to ISO1940 before design. The stiffness obtained by the upper limit of the vibration response caused by gravity and initial unbalanced force is denoted as k in .

[0088] ② Under fault conditions, the vibration response caused by gravity and maximum unbalanced force should not be less than the minimum value of the vibration response under fault conditions determined before design. The lower limit of the vibration response caused by gravity and maximum unbalanced force determines the maximum value of stiffness, which is denoted as k. max .

[0089] ③ The stiffness calculated from the upper limit of the shaft deflection under the action of gravity and maximum unbalanced force is denoted as k de .

[0090] The k calculated by ①③ in and k de The maximum value of the two determines the minimum value of the shaft stiffness, so the minimum value of the stiffness is taken as k in and k de The maximum value of .

[0091] In summary, the mid-span stiffness coefficient k of the rotating shaft is mid The following relationship should be satisfied:

[0092] max(k in ,k de )≤k mid ≤k max (16)

[0093] In the process of shaft parameter optimization design, the shaft material and shaft length are generally given artificially before shaft design according to the size of the test bench, sensor installation and other conditions. From equations (15) and (13), it can be seen that the mid-span stiffness coefficient of the shaft is directly determined by the shaft diameter.

[0094] Therefore, by traversing the diameters, the shaft diameter range that meets formula (16) can be obtained.

[0095] (204) Determine the allowable value y of the shaft deflection at each diameter max The allowable value of shaft deflection is generally determined according to the machinery manual, that is, according to y max =0.0005·l calculation.

[0096] (205) Next, determine the maximum unbalanced force F at each diameter max The calculation formula is:

[0097] F max =k mid y max -m k g (17)

[0098] (206) The established rotor dynamics model is used to solve the vibration response of the shaft system. According to the parameters determined in the first four steps, the vibration response of the shaft system in the normal state and the fault state (when the maximum unbalanced force is introduced) is solved according to formula (1). The range with the largest vibration response value is selected, and a diameter traversal with a step length of 1 mm is performed within the range.

[0099] (207) After traversing with a step size of 1 mm, the shaft diameter with the largest vibration response is obtained, and the shaft diameter, shaft stiffness, and maximum unbalanced force when the vibration response difference is the largest are selected as the optimal parameters.

[0100] 3. Unbalanced parameter design

[0101] Under normal conditions, due to factors such as processing errors and uneven materials, the turntable also has a certain amount of imbalance, which is called initial imbalance. Initial imbalance mass m r The calculation formula is:

[0102]

[0103] Where, M is the mass of the turntable, in kg; G is the accuracy level selected, and the balance level is determined before design according to ISO1940; n is the rotor speed, in rpm; r is the correction radius, in mm, which is a parameter artificially determined according to the external dimensions of the rotating part during the dynamic balancing experiment, and is used to determine the position of adding or removing weight. m calculated according to formula (18) r According to the initial eccentricity determined by the selected balance level, the initial unbalanced force F can be obtained.r .

[0104] From formula (11), we can see that the unbalanced force caused by adding unbalanced blocks is:

[0105] F e =(M+m)eω 2 =me m ω 2 (19)

[0106] The maximum unbalanced force includes the unbalanced force caused by the unbalanced block and the unbalanced force caused by the initial unbalance, which can be expressed as:

[0107] F max =F e +F r (20)

[0108] According to the second step of the traversal process of the shaft parameter design, the maximum unbalanced force F is obtained. max According to formula (20), the unbalanced force F caused by the unbalanced block is obtained e , and then determine the unbalanced mass m and installation position e according to formula (19) m , parameter value of rotation angular velocity ω.

[0109] 4. Blade parameter design

[0110] For blade parameter design, considering that actual rotating machinery may use blade tip clearance data to diagnose rotor imbalance faults, it is required that the blade resonance frequency of the rotor test bench should be far away from the rotor operating speed so that the blade tip clearance data is not coupled with blade vibration information. This means ensuring that the blade does not vibrate and that the measured blade tip clearance data only reflects the vibration of the shaft, and does not include changes in the blade tip clearance introduced by blade vibration.

[0111] Generally, the blade of the unbalanced fault test bench is a rectangular blade, with a blade height of h, a width the same as the width of the turntable, also represented by w, and a thickness of b. The i-th order resonance static frequency of the blade is f si It can be expressed as:

[0112]

[0113] Among them, E b is the elastic modulus of the blade, I b is the moment of inertia of the blade section, and the calculation formula for the rectangular straight blade is I b =wb 3 / 12. S b is the cross-sectional area, ρ b is the leaf density, K i Calculate the coefficient for the i-th order natural frequency of the vibration system. Select the appropriate blade size by resonant frequency constraint.

[0114] The present invention is not limited to the embodiments described above. The above description of the specific embodiments is intended to describe and illustrate the technical solution of the present invention. The above specific embodiments are merely illustrative and not restrictive. Without departing from the scope of the present invention and the scope of protection of the claims, a person of ordinary skill in the art can also make many forms of specific changes under the guidance of the present invention, which all fall within the scope of protection of the present invention.

Claims

1. A rotor unbalance fault test bench design optimization method, based on the rotor unbalance fault test bench, includes a support base, a bearing support seat, a drive motor, a bearing, a turntable, a rotating shaft and a coupling, wherein the support base is equipped with a bearing support seat and a drive motor for providing support for the bearing support seat and the drive motor, a bearing is installed on the bearing support seat, the drive motor is connected to the rotating shaft through a coupling, the rotating shaft is connected to the bearing and the turntable, the turntable is a single-stage turntable or a plurality of stage turntables; an unbalanced mass block is installed on the turntable; it is characterized in that include: S1. Simplify the rotating mechanical rotor system consisting of bearings, a rotating shaft and several stages of rotating disks into a double-supported rotor system; establish a rotor dynamics model according to the lumped parameter method, the rotor dynamics model includes an elastic shaft, a pair of rolling bearings, and n rigid rotating disks; S2. Set the axis parameters as follows: (201) According to the size requirements of the experimental platform and the sensor installation requirements, determine the shaft material, total shaft length, distance between each shaft section, support point distance and number of turntable levels; (202) According to the provisions of the mechanical design manual for the length-to-diameter ratio of ordinary shafts, based on the determined total length of the shaft, the shaft diameter traversal range is calculated according to the length-to-diameter ratio being greater than 5 and less than 30; (203) Selecting 1 / 10 of the shaft diameter traversal range as the step size to traverse the shaft diameter, and calculating the shaft mid-span stiffness coefficient under each diameter; (204) Determine the allowable value of shaft deflection for each diameter according to the Machinery Manual; (205) Determine the maximum unbalanced force F at each diameter max ; The calculation formula is: F max =k mid y max -m k g, where k mid is the mid-span stiffness coefficient of the rotating shaft, y max is the allowable value of the shaft deflection; m k is the mass of the turntable when there is an unbalanced fault on the kth turntable, and g is the acceleration of gravity; (206) solving the vibration response of the shaft system using the established rotor dynamics model, solving the vibration response of the shaft system in a normal state and a fault state according to the forced vibration differential equation and the parameters determined above, selecting a range with the largest vibration response value, and performing a diameter traversal within the range with a step length of 1 mm; the fault state refers to a state where the maximum unbalanced force is introduced; (207) After traversing with a step length of 1 mm, the shaft diameter with the largest vibration response is obtained, and the shaft diameter, shaft stiffness, and maximum unbalanced force when the vibration response difference is the largest are selected as the optimal parameters; S3. Setting unbalanced parameters, including maximum unbalanced force, unbalanced force caused by unbalanced block, unbalanced mass of unbalanced block, installation position of unbalanced block and rotational angular velocity of turntable; S4. Set blade parameters.

2. According to claim 1, a rotor imbalance fault test bench design optimization method is characterized in that: In step S4, the first-order resonance frequency of the blades of the rotating disk should be at least two orders of magnitude higher than the operating speed of the rotor, so that the tip clearance data is not coupled with the blade vibration information.

3. According to claim 1, a rotor imbalance fault test bench design optimization method is characterized in that: The mid-span stiffness coefficient of the rotating shaft must meet the following conditions: ① Under normal conditions, the vibration response caused by gravity and initial unbalanced force is less than the vibration response specified by the accuracy grade; ② Under the fault state, the vibration response caused by gravity and the maximum unbalanced force is greater than the minimum value that the vibration response needs to reach under the fault.

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