CN115031899A - Rotor unbalance fault experiment table and design optimization method - Google Patents
Rotor unbalance fault experiment table and design optimization method Download PDFInfo
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Abstract
The invention discloses a rotor unbalance fault experiment table and a design optimization method, wherein the fault experiment table comprises a supporting base, a bearing supporting seat, a driving motor, a bearing, a rotary table, a rotary shaft and a coupler, wherein the bearing supporting seat and the driving motor are arranged on the supporting base and used for providing support for the bearing supporting seat and the driving motor; and an unbalanced mass block is arranged on the rotary disc. The design optimization method of the fault experiment table comprises the following steps: establishing a rotor dynamics model; setting parameters of a rotating shaft; setting an unbalance parameter; and setting blade parameters.
Description
Technical Field
The invention belongs to the field of rotating machinery performance testing and fault diagnosis, and particularly relates to a rotor unbalance fault simulation experiment table and a design optimization method thereof.
Background
The rotor system is the most core component of the rotating machine, and in the working process, the typical faults represented by unbalanced faults are difficult to completely eliminate, and the operation safety and the working efficiency of the rotating machine are seriously influenced. Rotor imbalance is a failure due to rotor member mass eccentricity or rotor member defect, which is the most common failure of rotary machines. Therefore, development of fault diagnosis of rotating machine rotor systems has been a long-standing research task in this field.
However, the rotor system of the rotary machine is an extremely complex nonlinear time-varying system, and is limited by various factors such as manufacturing cost, test conditions, test period, test technology and the like, so that the dynamic test research on the rotor system is still difficult. Meanwhile, in the running process of the rotating machine, the fault state cannot be observed at any time, and once the rotating machine breaks down, the rotating machine is greatly damaged. Therefore, designing a simulation rotor experiment table, and acquiring the rotor operation data by using the rotor experiment table is an important means for realizing fault diagnosis data research and algorithm verification.
At present, most of researches on rotor faults are carried out on rotor dynamics research, fault models and fault diagnosis algorithms based on simulation rotor test tables, and researches on optimization design methods of the rotor fault test tables are less. The optimal rotor experiment table can realize high-efficiency and low-cost rotating machinery data simulation, and enables fault diagnosis data analysis and algorithm verification to be performed with half the effort.
Disclosure of Invention
The invention aims to overcome the defects in the prior art, provides a rotor imbalance fault experiment table and a design method thereof, and realizes the solution of the optimal parameters of the rotor imbalance fault experiment table.
The purpose of the invention is realized by the following technical scheme:
a rotor unbalance fault experiment table comprises a supporting base, a bearing supporting seat, a driving motor, a bearing, a turntable, a rotating shaft and a coupler, wherein the bearing supporting seat and the driving motor are installed on the supporting base and used for providing support for the bearing supporting seat and the driving motor; and an unbalanced mass block is arranged on the rotary disc.
The invention also provides a design optimization method of the rotor unbalance fault experiment table, which comprises the following steps:
s1, simplifying a rotary mechanical rotor system consisting of a bearing, a rotating shaft and a plurality of stages of turntables into a double-support rotor system; establishing a rotor dynamics model according to a lumped parameter method, wherein the rotor dynamics model comprises an elastic shaft, a pair of rolling bearings and n rigid turntables;
s2, setting the parameters of the rotating shaft, which are as follows:
(201) determining a rotating shaft material, the total length of the rotating shaft, the distance of each shaft section, the distance of a supporting point and the number of stages of the turnplate according to the size requirement of the experiment table and the installation requirement of the sensor;
(202) according to the specification of the length-diameter ratio of a common shaft in a mechanical design manual, according to the determined total length of the rotating shaft, the diameter traversal range of the rotating shaft is obtained by calculating according to the length-diameter ratio of more than 5 and less than 30;
(203) 1/10 in the traversing range of the diameter of the rotating shaft is selected as a step length to traverse the diameter of the rotating shaft, and the midspan stiffness coefficient of the rotating shaft under each diameter is calculated;
(204) determining a deflection allowable value of the rotating shaft under each diameter according to a mechanical manual;
(205) determining the maximum imbalance force F at each diameter max (ii) a The calculation formula is as follows: f max ＝k mid y max -m k g, in the formula, k mid Is the mid-span stiffness coefficient of the shaft, y max Is the allowable value of the deflection of the rotating shaft; m is k G is the gravitational acceleration, which is the mass of the turntable when the kth turntable has an imbalance fault;
(206) solving shafting vibration response by utilizing the established rotor dynamics model, solving shafting vibration response in a normal state and a fault state according to a forced vibration differential equation and the determined parameters, selecting a range with the maximum vibration response value, and traversing the diameter with the step length of 1mm in the range; the fault state refers to a state in which a maximum unbalanced force is introduced;
(207) obtaining the diameter of a rotating shaft with the maximum vibration response after traversing by 1mm step length, and selecting the diameter of the rotating shaft, the rigidity of the rotating shaft and the maximum unbalanced force when the vibration response difference is maximum as optimal parameters;
s3, setting unbalance parameters including maximum unbalance force, unbalance force caused by an unbalance block, unbalance mass of the unbalance block, mounting position of the unbalance block and rotation angular velocity of the turntable;
and S4, setting blade parameters.
Furthermore, the first-order resonant frequency of the blade of the rotating disc is at least two orders of magnitude higher than the working rotating speed of the rotor, so that the blade tip clearance data is not coupled with the vibration information of the blade.
Further, the mid-span stiffness coefficient of the rotating shaft needs to meet the following conditions:
in the normal state, the vibration response due to gravity, the initial unbalance force, should not be greater than that specified by the accuracy class determined according to ISO1940 before design.
Secondly, in a fault state, the vibration response caused by gravity and the maximum unbalanced force is not less than the minimum value which needs to be reached by the vibration response under the fault determined before design.
Compared with the prior art, the technical scheme of the invention has the following beneficial effects:
1. the invention provides a design optimization method for a rotor unbalance fault experiment table, which can be used for obtaining rotating shaft parameters, unbalance parameters and blade parameters of a multistage rotor unbalance fault experiment table and designing the experiment table according to the obtained parameters. The method of the invention utilizes a concentrated mass method to establish a rotor dynamic model, and utilizes the rotor dynamic model to solve the vibration response of a rotor system under the normal operation state and the fault state. And guidance is provided for the design of an actual experiment table. In addition, the invention obtains the rotating shaft parameters which can realize the maximum fault response under the maximum unbalanced force through traversing the diameter of the rotating shaft, namely the optimal parameters of the rotating shaft of the rotor unbalance fault experiment table. And carrying out unbalance parameter design and blade parameter design by using the parameters of the rotating shaft.
The method can achieve the maximum unbalance fault response under the condition that the experiment table runs safely, provides theoretical guidance for the unbalance fault experiment table, improves the safety of the unbalance rotor fault experiment table, ensures that the most obvious unbalance fault response can be obtained, reduces the trial and error cost, optimizes the design flow of the unbalance fault experiment table, and is beneficial to achieving high-efficiency and low-cost rotating machinery data simulation.
2. The invention provides an unbalanced rotor fault experiment table, which can realize the simulation of unbalanced faults of a rotor according to parameters set in design requirements, provides test data for fault diagnosis data analysis and algorithm verification, and compared with the existing experiment table, changes a rotating shaft of the unbalanced rotor fault experiment table, designs unbalanced force and blade parameters by using a design optimization method, can achieve the maximum unbalanced fault response under the condition of ensuring the safe operation of the experiment table, can better realize the generation of fault data, and provides effective test data for a subsequent measurement system.
3. According to the invention, a rotor dynamics model is established, the optimal parameters of the rotor are obtained by traversing the diameter of the rotating shaft and the unbalanced mass block, the optimal design of the unbalanced experimental table of the rotor is realized, and the optimal parameters of the rotating shaft, the unbalanced fault parameters and the blade parameters are obtained, so that the high-efficiency and low-cost rotating machinery data simulation can be realized, and the fault diagnosis data analysis and the algorithm verification are performed with half the effort.
Drawings
FIG. 1 is a block diagram of a rotor imbalance fault laboratory bench.
Fig. 2 is a schematic view of a leaf disc structure.
FIG. 3 is a schematic view of a dual support multiple rotor.
FIG. 4 is a schematic diagram of the calculation of stiffness of different shaft sections.
FIG. 5 is a schematic view of an imbalance of a rotating shaft.
FIG. 6 is a schematic view of a turntable imbalance.
FIG. 7 is a flow chart of spindle parameter optimization.
Detailed Description
The invention is described in further detail below with reference to the figures and specific examples. It should be understood that the specific embodiments described herein are merely illustrative of the invention and are not intended to limit the invention.
As shown in fig. 1, a rotor unbalance fault experiment table comprises a supporting base 1, a bearing supporting seat 2, a driving motor 3, a bearing 4, a turntable 5, a rotating shaft 6 and a coupler 7, wherein the bearing supporting seat 2 and the driving motor 3 are installed on the supporting base 1 and used for providing support for the bearing supporting seat 2 and the driving motor 3, the bearing 4 is installed on the bearing supporting seat 2, the driving motor 3 is connected with the rotating shaft 6 through the coupler 7, the rotating shaft 6 is connected with the bearing 4 and the turntable 5, and the turntable 5 is a single-stage turntable or a plurality of stages of turntables; the mounting of the unbalance masses 8 can be carried out by a single rotor disc unit 9 in the rotor disc 5, the addition of unbalance faults being achieved by the unbalance forces provided by the unbalance masses 8 during the rotation of the rotor. The driving motor 3 drives a rotor system comprising a bearing 4, a rotating shaft 6 and a rotating disc 5, and the bearing 4 is supported by the bearing support seat 2.
In the rotor unbalance fault experiment table, the rotating shaft 6 is designed by the rotating shaft parameter design optimization method, the rotating disc blades in the single-stage or multi-stage rotating disc 5 are designed by the blade parameter design method, and the installation position and the mass of the unbalance mass block 8 are designed by the unbalance parameter design method.
The rotor unbalance fault experiment table design optimization method comprises the steps of rotor dynamics model building, rotating shaft parameter optimization design, unbalance parameter design and blade parameter design. And establishing a rotor dynamics model as a basis for subsequent optimization design of the parameters of the rotating shaft and the unbalance parameters. The rotor dynamics model constructs a forced vibration differential equation of the rotor, when the parameters of the rotating shaft are optimized, after an obtained exciting force matrix is obtained, corresponding rotor shafting vibration and axis locus response can be obtained through inputting the rotor dynamics model, and then the parameters of the rotating shaft can be constrained according to the obtained shafting vibration and axis locus response, so that the optimal parameters of the maximum fault response can be obtained. And the maximum unbalance force obtained by the unbalance parameter design according to the rotating shaft parameter optimization design can be calculated to obtain the optimal unbalance parameter. In addition, in 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 that an appropriate blade parameter can be obtained.
The method comprises the following specific steps:
firstly, establishing a rotor dynamics model
A rotary machine rotor system typically includes bearings, a shaft, and multiple stages of turntables, which can be simplified to a double-supported, multiple stage rotor system. Referring to fig. 1, a rotor dynamics model is established according to a lumped parameter method, and the model comprises an elastic shaft, a pair of rolling bearings and n rigid rotating disks. The bearing and the turntable are equivalent to (n +2) concentrated mass points which are m 1 ～m n+2 . Wherein m is 1 、m n+2 The concentrated mass m of the left and right bearings respectively 2 ～m n+1 Respectively the concentrated mass of n turntables. The elastic shaft is loaded with mass m uniformly 2 ～m n+1 On a rotating disk. k is a radical of formula 12 ～k n+1n+2 Respectively, the stiffness of each section of elastic shaft, c 1 ～c n+2 Damping each bearing or turntable.
The rotor system shown in fig. 2 is a multi-degree-of-freedom system, and when the rotor rotates, forced vibration responses are generated in the x direction and the y direction. Taking the x direction as an example, the forced vibration differential equation is:
wherein M is a mass matrix, C is a damping matrix, K is a stiffness matrix, X is a vibration displacement matrix in the X direction, F x Is an x-direction excitation force matrix. Wherein the quality matrix M is:
the damping matrix C is:
because adjacent turntables and bearings exist rigidity coupling, or turntables and turntables, the rigidity matrix K is:
the X-direction vibration displacement matrix X is:
X＝(x 1 x 2 … x n+1 x n+2 ) T (5)
in the normal state of the rotor, the x-direction stress is only the support reaction force of the left and right bearings, and can be expressed as:
F x ＝(F lx 0 … 0 F rx ) T (6)
wherein, F lx 、F rx The support reaction force of the left bearing and the right bearing in the x direction is related to the bearing rigidity and the displacement of the concentrated mass point, and can be expressed as follows:
wherein k is 1 ，k n+2 The stiffness of the left and right bearings.
The y-direction bearing reaction force is similar to the x-direction. In addition, due to the action of gravity, the stress of the rotor system in a normal state also comprises the gravity of each bearing and each turntable, and a y-direction excitation force matrix F Y Can be expressed as:
F Y ＝(F ly -m1g -m2g … -m n+1 g F by -m n+2g ) T (8)
wherein, F ly 、F ry The support reaction force of the left bearing and the right bearing in the y direction is calculated in the same way as the x direction, and g is the gravity acceleration of the position of the experiment table. And (3) substituting the expressions (2) to (8) into the expression (1) to obtain a rotor system dynamic model based on the concentrated mass method. The device isThe model is two ordinary differential equations in the x direction and the y direction. Based on the model, the vibration of a rotor shafting and the axle center track response can be obtained by using a numerical method, and the dynamic characteristics of the rotor in different states are analyzed.
The rotor unbalance fault can be equivalent to the mass eccentricity of the rotor, and taking the unbalance fault existing in the middle rotating disc of the mechanical rotor system as an example, the geometric center of the rotor rotation is shifted from O to O', as shown in FIG. 4; rotor imbalance fault is equivalent to a distance e from the geometric center of the turntable m There is an unbalanced mass, m, as shown in fig. 5.
In the rotor unbalance fault shown in fig. 6, the geometric center is located at O', 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:
wherein M is the turntable mass. When the k-th rotary table has unbalance fault, M is M k 。
The unbalanced force caused by the additional installation of the unbalanced mass block is as follows:
F e ＝(M+m)eω 2 (10)
the x and y direction imbalance forces can be expressed as:
where ω is the rotor angular velocity and t is the current time. Taking the x direction as an example, the excitation force matrix under the unbalanced fault can be expressed as:
F ex ＝(F lx 0 … 0 F ex 0 … 0 F rx ) T (12)
wherein F ex In the (k + 1) th row of the matrix of equation (12).
And (3) substituting the formula (12) as an excitation force matrix into the formula (1) to obtain a dynamic equation under the condition of the unbalanced rotor fault.
Second, design of parameters of the rotating shaft
The design flow of the parameters of the rotating shaft is shown in fig. 7, when the parameters of the rotating shaft are optimized, the rigidity and the maximum allowable deflection value of the rotating shaft are calculated by traversing the diameter of the rotating shaft, and then the maximum unbalanced force F under each diameter of the rotating shaft is obtained max . By solving for the rotor system in the normal, fault state (i.e. maximum out of balance force F) max Under the action), and selecting the diameter of the rotating shaft, the rigidity of the rotating shaft and the maximum unbalanced force as optimal parameters when the vibration response difference is maximum. The specific process of the rotating shaft parameter design is as follows:
(201) according to the size requirement of the experiment table and the installation requirement of the sensor, the material of the rotating shaft, the total length l of the rotating shaft and the distance x between each shaft section and the supporting point are determined nn+1 And the number n of the leaf disc stages.
(202) According to the specification of the mechanical design manual for the length-diameter ratio of a common shaft, according to the determined total length of the rotating shaft, the traversing range of the diameter of the rotating shaft is obtained by calculating according to the length-diameter ratio of more than 5 and less than 30.
(203) Firstly, 1/10 in the traversing range of the diameter of the rotating shaft is selected as a step length to traverse the diameter of the rotating shaft, and the rigidity coefficient k of the span of the rotating shaft under each diameter is calculated through the following process mid 。
The rigidity of the rotating shaft is determined by the material, length, diameter and other parameters of the rotating shaft, and is the most key parameter of the rotor system. For a simply supported beam supported at both ends, the stiffness is expressed as:
wherein E is the elastic modulus of the rotating shaft, l is the length of the rotating shaft between the double bearings, x is the distance from the left bearing point, I is the section moment of inertia, and the calculation formula is as follows:
wherein d is the diameter of the rotating shaft.
For the stiffness coefficients in the formula, the distance of the different shaft segments from the left support point is shown in fig. 3. From the results available in the figure:
in the formula, w 1 Is the bearing width, w 2 Is the width of the leaf disk, /) 1 Distance of bearing to nearest blade disc,/ 2 For adjacent leaf disc spacing, x nn+1 And substituting the result into a rigidity calculation formula (13) for the distance from the shaft center of the nth shaft section between the two bearings to the left support, so as to obtain the rigidity values of different shaft sections. The maximum deflection position is recorded as the mid-span rigidity coefficient k of the rotating shaft mid 。
Mid-span stiffness coefficient k of rotating shaft mid Directly determines the maximum unbalance force F max Shafting vibration and axle center track response under action. In order to distinguish different operating states of the rotor system, the difference between the responses of the rotor experiment table in normal and fault states should be an order of magnitude difference. Meanwhile, the rotating shaft must pass safety check. In summary, the mid-span stiffness coefficient of the rotating shaft should satisfy the following conditions:
in the normal state, the vibration response due to gravity, the initial unbalance force, should not be greater than that specified by the accuracy class determined according to ISO1940 before design. The stiffness, k, calculated from the upper limit of the vibrational response due to gravity and initial imbalance in 。
Secondly, in a fault state, the vibration response caused by gravity and the maximum unbalanced force is not less than the minimum value required to be reached by the vibration response under the fault determined before design. The maximum value of the rigidity is determined by the lower limit of the vibration response caused by gravity and maximum unbalanced force and is recorded as k max 。
Thirdly, the rigidity obtained by calculating the upper limit of the deflection of the rotating shaft under the action of gravity and maximum unbalanced force is recorded as k de 。
And thirdly k obtained by calculation in And k de The maximum value of the two determines the minimum value of the rigidity of the rotating shaft, so the minimum value of the rigidity is taken as k in And k de Of (c) is calculated.
In conclusion, the mid-span stiffness coefficient k of the rotating shaft mid The following relationship should be satisfied:
max(k in ,k de )≤k mid ≤k max (16)
in the process of optimization design of the rotating shaft parameters, the rotating shaft material and the rotating shaft length are generally manually set according to the conditions of the size of an experiment table, sensor installation and the like before the rotating shaft is designed. From the equations (15) and (13), the shaft span stiffness coefficient is directly determined by the shaft diameter.
Thus, a range of spindle diameters conforming to equation (16) can be obtained by traversing the diameters.
(204) Determining the allowable value y of the deflection of the rotating shaft at each diameter max . The permissible deflection of the rotating shaft is generally determined according to the machine manual, i.e. according to y max Calculated at 0.0005 · l.
(205) The maximum imbalance force F at each diameter is then determined max . The calculation formula is as follows:
F max ＝k mid y max -m k g (17)
(206) and solving the shafting vibration response by utilizing the established rotor dynamics model, solving the shafting vibration response in a normal state and a fault state (when the maximum unbalanced force is introduced) according to the parameters determined in the previous four steps and a formula (1), selecting a range with the maximum vibration response value, and traversing the diameter with the step length of 1mm in the range.
(207) And obtaining the diameter of the rotating shaft with the maximum vibration response after traversing by 1mm step length, and selecting the diameter of the rotating shaft, the rigidity of the rotating shaft and the maximum unbalance force when the vibration response difference is maximum as optimal parameters.
Design of three, unbalanced parameters
In a normal state, due to factors such as machining errors and material unevenness, the turntable also has a certain amount of unbalance, which is called initial unbalance. Initial unbalanced mass m r The calculation formula is as follows:
wherein M is a turntableMass in kg; g is selected as the precision grade, and the balance grade is determined before design according to ISO 1940; n is the rotor speed in rpm; r is a correction radius, and the unit is mm, and is a parameter artificially determined according to the external dimension of the rotating member when a dynamic balance experiment is carried out, and is used for determining the weighting or de-weighting position. M calculated according to equation (18) r The initial imbalance force F can be obtained from the initial eccentricity determined from the selected balance level r 。
As can be seen from equation (11), the unbalanced force caused by adding the unbalanced mass is:
F e ＝(M+m)eω 2 ＝me m ω 2 (19)
the maximum imbalance force includes the imbalance force caused by the imbalance mass and the imbalance force caused by the initial imbalance, expressed as:
F max ＝F e +F r (20)
calculating the maximum unbalanced force F according to the ergodic process of the second step of the rotating shaft parameter design max The unbalance force F caused by the unbalance mass is obtained from the equation (20) e Further, the unbalanced mass m and the mounting position e are determined by the equation (19) m And a parameter value of the rotational angular velocity ω.
Fourth, blade parameter design
For the design of blade parameters, the consideration that the actual rotating machinery may use the blade tip clearance data to diagnose the imbalance fault of the rotor is required, so that the resonance frequency of the blades of the rotor experiment table is far away from the working rotating speed of the rotor, and the blade tip clearance data is not coupled with the vibration information of the blades. The method ensures that the blade does not vibrate, ensures that the measured blade tip clearance data only reflects the vibration condition of the rotating shaft, and does not include the change of the blade tip clearance caused by the vibration of the blade.
The blade of the general unbalance fault experiment table is a rectangular blade, the height of the blade is h, the width of the blade is the same as the width of a rotating disc, the width is also expressed by w, and the thickness is b. Blade ith order resonance static frequency f si Can be expressed as:
wherein E is b Is the modulus of elasticity of the blade, I b The calculation formula of the rectangular straight plate blade is I b ＝wb 3 /12。S b Is the cross-sectional area, ρ b Is the blade density, K i Coefficients are calculated for the ith order natural frequency of the vibration system. The appropriate blade sizing is performed by resonance frequency constraints.
The present invention is not limited to the embodiments described above. The foregoing description of the specific embodiments is intended to describe and illustrate the technical solutions of the present invention, and the above specific embodiments are merely illustrative and not restrictive. Those skilled in the art can make many changes and modifications to the invention without departing from the spirit and scope of the invention as defined in the appended claims.
Claims (4)
1. A rotor unbalance fault experiment table is characterized by comprising a supporting base, a bearing supporting seat, a driving motor, a bearing, a turntable, a rotating shaft and a coupler, wherein the bearing supporting seat and the driving motor are installed on the supporting base and used for providing support for the bearing supporting seat and the driving motor; and an unbalanced mass block is arranged on the rotary disc.
2. A design optimization method for a rotor unbalance fault experiment table is characterized by comprising the following steps:
s1, simplifying a rotary mechanical rotor system consisting of a bearing, a rotating shaft and a plurality of stages of turntables into a double-support rotor system; establishing a rotor dynamics model according to a lumped parameter method, wherein the rotor dynamics model comprises an elastic shaft, a pair of rolling bearings and n rigid turntables;
s2, setting the parameters of the rotating shaft, which are as follows:
(201) determining a rotating shaft material, the total length of the rotating shaft, the distance of each shaft section, the distance of a supporting point and the number of stages of the turnplate according to the size requirement of the experiment table and the installation requirement of the sensor;
(202) according to the length-diameter ratio of a common shaft specified in a mechanical design manual, according to the determined total length of the rotating shaft, the diameter traversal range of the rotating shaft is obtained by calculating according to the length-diameter ratio of more than 5 and less than 30;
(203) 1/10 in the traversing range of the diameter of the rotating shaft is selected as a step length to traverse the diameter of the rotating shaft, and the midspan stiffness coefficient of the rotating shaft under each diameter is calculated;
(204) determining the allowable value of the deflection of the rotating shaft under each diameter according to a mechanical manual;
(205) determining the maximum imbalance force F at each diameter max (ii) a The calculation formula is as follows: f max ＝k mid y max -m k g, in the formula, k mid Is the mid-span stiffness coefficient of the shaft, y max Is the allowable value of the deflection of the rotating shaft; m is k G is the gravitational acceleration, which is the mass of the turntable when the kth turntable has an imbalance fault;
(206) solving the shafting vibration response by utilizing the established rotor dynamic model, solving the shafting vibration response in a normal state and a fault state according to a forced vibration differential equation and the determined parameters, selecting a range with the maximum vibration response value, and traversing the diameter with the step length of 1mm in the range; the fault state refers to a state in which a maximum unbalanced force is introduced;
(207) obtaining the diameter of the rotating shaft with the maximum vibration response after traversing by 1mm step length, and selecting the diameter of the rotating shaft, the rigidity of the rotating shaft and the maximum unbalanced force when the vibration response difference is maximum as optimal parameters;
s3, setting unbalance parameters including maximum unbalance force, unbalance force caused by an unbalance block, unbalance mass of the unbalance block, mounting position of the unbalance block and rotation angular velocity of the rotary table;
and S4, setting blade parameters.
3. The method for optimizing design of a rotor imbalance fault experimental table according to claim 2, wherein in step S4, the first-order resonant frequency of the blades of the rotary table should be at least two orders of magnitude higher than the operating speed of the rotor, so that the blade tip clearance data is not coupled with the vibration information of the blades.
4. The design optimization method of the rotor unbalance fault experiment table according to claim 2, wherein the mid-span rigidity coefficient of the rotating shaft needs to meet the following conditions:
under a normal state, the vibration response caused by gravity and initial unbalanced force is smaller than the vibration response specified by the precision grade;
and secondly, in a fault state, the vibration response caused by gravity and the maximum unbalanced force is greater than the minimum value required by the vibration response under the fault.
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