Motor vibration evaluation method and system without third support at the extended end on the drive side

Abstract

This invention discloses a method and system for evaluating the vibration of a motor without a third support at the excitation-side overhang, relating to the field of motor evaluation and manufacturing technology. The vibration evaluation method includes the following steps: S1, constructing a finite element simulation model based on the motor rotor parameters; S2, calculating the total rotor unbalance U according to the balance quality level; S3, distributing the position, magnitude, and phase of the total rotor unbalance U; S4, calculating the vibration value A at the excitation-side overhang; S5, obtaining the vibration acceptance value AH at the excitation-side overhang based on motor test simulation; S6, comparing the vibration value A and the vibration acceptance value AH at the excitation-side overhang; if A ≤ AH, no third support is needed; if A > AH, the motor is redesigned, and steps S1-S4 are repeated until A ≤ AH. This invention can accurately determine whether the vibration at the excitation-side overhang of a steam turbine generator is within a reasonable range, and can promptly detect potential vibration anomalies, thereby effectively ensuring the safe and stable operation of the motor.

Description

Technical Field

[0001] This invention relates to the field of motor evaluation and manufacturing technology, and in particular to a method and system for evaluating motor vibration based on vibration characteristics in motors without a third support at the excitation-side overhang. Background Technology

[0002] Currently, in motor design and manufacturing, for motors such as steam turbine generators, hydro turbine generators, and electric motors, when there is a certain length from the center line of the rotor body's excitation-side journal to the end of the excitation-side overhang, some designs choose to install a third support (stabilizing bearing) at the excitation-side overhang. During the acceptance phase, existing acceptance standards mainly focus on the overall performance indicators of the motor, such as power, efficiency, and speed stability. There is a lack of a clear and specific evaluation system for the vibration of the excitation-side overhang without a third support. Usually, only some general motor vibration standards are used for a rough assessment, which do not specifically consider the special structure and operating characteristics of the excitation-side overhang. Installing a third support has the following advantages:

[0003] 1. Ensure operational stability: Setting up a third support can enhance the structural rigidity of the motor's excitation end to a certain extent, reduce vibration, make the motor relatively stable during operation, reduce the risk of component damage caused by vibration, and extend the service life of the motor.

[0004] 2. Adapting to some complex operating conditions: For motors with complex operating conditions and large load fluctuations, the third support can provide additional support force to help the motor better adapt to these changes and maintain relatively stable performance.

[0005] However, setting up a third support also has the following drawbacks:

[0006] 1. Inaccurate acceptance standards: The lack of a specialized vibration assessment system for situations where there is no third support at the excitation side overhang makes it impossible to accurately determine whether the vibration of this part is within a reasonable range during motor acceptance. There may be excessive vibration that is not detected in time, posing a hidden danger to the safe operation of the motor.

[0007] 2. Increased structural complexity and cost: Adding a third support increases the structural complexity of the motor, requiring additional parts and installation processes, thus raising manufacturing costs. It also increases the potential for equipment failure; a malfunction in the third support could affect the normal operation of the entire motor.

[0008] 3. Inconvenient installation and maintenance, and wasted space: The installation of the third support requires a certain amount of space, and the installation process is complex, increasing the difficulty and time required for installation. During maintenance, disassembling and reinstalling the third support consumes a lot of manpower and time, reducing maintenance efficiency. Moreover, the space it occupies limits the installation and use of the motor in some space-constrained locations, resulting in wasted space and reduced economic benefits. Summary of the Invention

[0009] This invention aims to provide a vibration assessment method and system for motors without a third support at the excitation side overhang. It provides a specialized vibration characteristic assessment system for motors without a third support at the excitation side overhang, making motor assessment more accurate and targeted. It can accurately determine whether the vibration at the excitation side overhang of the turbine generator is within a reasonable range, and can promptly detect potential vibration anomalies, thereby effectively ensuring the safe and stable operation of the motor.

[0010] To achieve the above-mentioned objectives, the technical solution of the present invention is as follows:

[0011] A method for evaluating the vibration of a motor without a third support at the excitation-side overhang includes the following steps:

[0012] S1. Rotor three-dimensional dynamic modeling: Constructing a finite element simulation model based on motor rotor parameters;

[0013] S2. Calculate the total unbalance U of the rotor based on the balance quality level;

[0014] S3. Distribute the location, magnitude, and phase of the total unbalance U of the rotor.

[0015] S4. Unbalance Response Calculation: Based on the finite element simulation model in step S1, the distribution of the location, magnitude, and phase of the unbalance in step S3, and the rated speed n N Under the given boundary conditions, construct the dynamic equation for calculating the rotor-bearing unbalanced response and calculate the vibration value A at the excitation side overhang.

[0016] S5. The vibration acceptance value AH of the motor's excitation side overhang end is obtained based on the motor test simulation;

[0017] S6. Compare the vibration value A of the excitation side cantilever end in step S4 with the vibration acceptance value AH of the excitation side cantilever end in step S5; if A ≤ AH, then there is no need to set up a third support.

[0018] If A > AH, redesign the motor and repeat steps S1-S4 until A ≤ AH.

[0019] In step S1, the rotor three-dimensional dynamic modeling includes the following steps:

[0020] S11. Establish a finite element model based on the geometric parameters and material properties of each shaft segment of the rotor;

[0021] The geometric parameters include: length L, mass m, diameter D, and stiffness EI;

[0022] The turbine generator shaft section includes a coupling shaft section, a journal shaft section, a main fan shaft section, an oil baffle shaft section, a retaining ring and main body overlapping shaft section, a rotor main body bearing, a slip ring shaft section, a slip ring fan shaft section, a speed measuring gear disk shaft section, and a key phase disk shaft section.

[0023] The material properties include elastic modulus, shear modulus, density, Poisson's ratio, and damping characteristics.

[0024] S12. Component Equivalent Treatment: In the finite element model, the mass, moment of inertia, and stiffness of each rotor component are equivalent to concentrated mass points or distributed mass, and their axial positions are marked.

[0025] The rotor components include: coupling, fan, retaining ring, coil, insulating block, axial lead, fan seat ring, slot wedge, slip ring, slip ring insulation, slip ring fan, speed measuring gear disk, and key phase disk;

[0026] S13. Support boundary conditions: Combining the mass matrix [M], damping matrix [C], gyroscope matrix [G], linear stiffness matrix [K], nonlinear stiffness matrix [Ks], and excitation force {F(t)}, the core equation is constructed to solve for the critical speed, natural frequency, and vibration performance parameters;

[0027] The core equation is shown below:

[0028]

[0029] Where {F(t)} is the time-varying external excitation force vector, including the excitation force generated by the unbalanced mass, electromagnetic torque, bearing oil film force, or airflow disturbance force; {X ●●} represents the second derivative of displacement with respect to time, i.e., the instantaneous acceleration response vector of each degree of freedom of the rotor; {X ● {X} represents the first derivative of displacement with respect to time, i.e., the velocity response vector of each degree of freedom of the rotor; {X} represents the displacement response vector of each degree of freedom of the rotor.

[0030] S14. Model Verification: Modal tests are conducted on the finite element simulation model. The model parameters are corrected using the modal test data. The finite element simulation model is verified and iteratively optimized to ensure that the calculation error of the first critical speed (ncr1) and the second critical speed (ncr2) of the rotor is ≤5%.

[0031] The correction of model parameters using modal test data includes: stiffness parameter correction, mass distribution correction, and damping and gyroscopic effect correction;

[0032] The stiffness parameter correction is as follows: using the natural frequency obtained from the modal test as the target, adjust the elastic modulus of the central axis and the equivalent stiffness of the support in the model.

[0033] If the calculated critical speed is higher than the measured value, it indicates that the model stiffness is too high. Reduce the input value of the shaft's elastic modulus or decrease the stiffness coefficient of the support until the calculated frequency matches the experimental frequency.

[0034] The mass distribution correction includes: when the amplitude of the mode shape deviates from the test value by more than ±5%, adding an equivalent mass unit at the corresponding position for compensation or adjusting the input value of the material density to correct the mass distribution.

[0035] If concentrated masses such as fan blades and couplings are not considered in the model, they can be compensated for by adding equivalent mass elements at the corresponding locations. If the mass distribution deviation is uniform, the input value of the material density can be adjusted. Note that mass adjustment will affect the critical speeds at all orders simultaneously, requiring comprehensive verification at multiple frequencies.

[0036] The damping and gyro effect corrections are as follows: For high-speed rotors, the gyro effect will cause the critical speed to bifurcate into positive and negative precession frequencies. If this phenomenon is observed in the experiment, a gyro matrix needs to be introduced into the model. If the vibration amplitude deviation exceeds ±5%, the damping coefficient in the model is adjusted using the damping ratio data from the modal test to improve the calculation accuracy of the resonance zone.

[0037] The verification and iterative optimization process involves: after correcting the model parameters, recalculating the rotor critical speed and comparing it with the measured value. If the error is still greater than 5%, the model parameters need to be corrected again until the calculation error between the first-order and second-order critical speeds is ≤5%, and the correlation coefficient of the mode shape is not less than 0.9. This ensures that the model accurately reflects the actual dynamic characteristics of the rotor.

[0038] The model parameters are corrected through a closed-loop process of "modal test data correction - finite element model iteration - critical speed verification". The core is to adjust key parameters such as stiffness and mass distribution in the model in reverse based on the natural frequency and mode shape data of the modal test, so that the error between the calculated critical speed and the measured value is controlled within 5%.

[0039] In step S2, the total rotor imbalance U is calculated using the following formula:

[0040]

[0041] Where G represents the rotor balance quality grade, mm / s; M represents the rotor mass, kg; n N is the rated speed, rpm; e is the maximum residual eccentricity of the rotor, m.

[0042] In step S3, the distribution of the magnitude, position, and phase of the total rotor imbalance U is specified in standard API 684-2010.

[0043] In step S4, the dynamic equation for calculating the rotor-bearing imbalance response is shown in the following equation: 1111

[0044]

[0045] In the formula, ω It is the rotational angular frequency; M 1 represents the overall quality matrix; K 1 represents the overall stiffness matrix; G 1 represents the global rotation matrix; c ij The overall oil film equivalent damping matrix; k ij The overall oil film stiffness matrix; i, j=1, 2 ); U 1,2 It is the displacement vector; Q 1,c , Q 2,c The vector of the cosine component of the unbalanced force; Q 1,s , Q 2,s Ü is the sinusoidal component vector of the unbalanced force; 1,2 It is the second derivative of the displacement vector, i.e., the acceleration vector; 1,2 It is the first derivative of the displacement vector, i.e., the velocity vector;

[0046] In step S4, calculating the vibration value A at the excitation-side cantilever end includes the following steps:

[0047] S41. Obtain system parameters and unbalanced excitation parameters: System parameters include the overall mass matrix M1, stiffness matrix K1, and rotation matrix G1;

[0048] Unbalanced excitation parameters: cosine component vector Q of the unbalanced force 1c Q 2c The sinusoidal component vector of the unbalanced force Q 1s Q 2s ; Rotational angular frequency ω;

[0049] The unbalanced force vector is expressed by the following formula:

[0050]

[0051] Where t is time; Q is the vector of unbalanced force; Q 1,2Let Qc,s be the unbalanced force vector group; Qc,s are the cosine and sine components of the unbalanced force vector.

[0052] S42. Substituting the assumed form of the response: The unbalanced excitation varies sinusoidally with time. Assume the response of the displacement vector is as follows:

[0053]

[0054] Where U1 is the displacement vector of displacement + rotation in the x-direction; U2 is the displacement vector of displacement + rotation in the y-direction; {A1} and {A2} are the cosine component vectors of the response; {B1} and {B2} are the sine component vectors of the response.

[0055] S43. Substituting the unbalanced force vector from step S41 and the assumed form from step S42 into the rotor-bearing unbalanced response calculation equation, we obtain:

[0056]

[0057] in, ;

[0058] S44. Solve the system of algebraic equations to calculate the amplitude: Solve for {A1}, {B1}, {A2}, and {B2} by matrix inversion or numerical methods; substitute them into the assumed form to obtain the displacement vector, that is, the vibration value of each point in different directions, and then calculate the vibration value A of the excitation side cantilever end.

[0059] In step S5, the vibration acceptance value AH of the motor's excitation side overhang end is obtained based on the motor test simulation, including the following steps:

[0060] S51. Conduct test simulations on the motor to obtain the set of vibration values ​​H of the excitation side cantilever end under different operating conditions without a third support;

[0061] S52. Screen the vibration value set H to select the set A1 of the cantilever end vibration values ​​that meet the motor operation requirements: stable contact between the brush and the slip ring with no obvious sparks; appropriate brush wear; and stable operation for 6 months or more; and the set A2 of the cantilever end vibration values ​​where the shaft vibration and bearing vibration are within the allowable range.

[0062] S53. Perform statistical analysis on vibration value sets A1 and A2, and screen out vibration value sets A1 and A2 that indicate the need for brush replacement after 6 months of wear. 10 ; Select the set A of vibration values ​​at the excitation side overhang when the vibration of the motor journal and bearing is at the upper limit of the allowable value. 20 ;

[0063] S54, Using the vibration value set A from step S53 10 and vibration value set A 20Select vibration value set A 10 The minimum value A in 1min and vibration value set A 20 The minimum value A in 2min ;

[0064] S55, via A 1min and A 2min The vibration acceptance value AH of the excitation side cantilever end was calculated.

[0065] In step S51, the set of vibration values ​​H at the motor excitation side overhang end is obtained through the following steps:

[0066] S511. Determine the test object: Select the appropriate motor type as the test object for the type of motor to be evaluated;

[0067] S512. Conduct dynamic balancing and type tests on the test object under various simulated actual operating conditions.

[0068] The actual operating conditions include rated operating conditions, maximum load operating conditions, reduced load operating conditions, rated speed, over-critical speed, and 120% rated speed overspeed;

[0069] S513, the set of vibration values ​​H at the motor excitation side overhang during the measurement and statistical dynamic balancing test and type test.

[0070] In step S55, the vibration acceptance value AH of the excitation side cantilever end is calculated using the following formula:

[0071]

[0072] Where x is the safety margin, with a value ranging from 0.8 to 0.9.

[0073] A vibration assessment system for a motor without a third support at the excitation-side cantilever end, characterized in that: the system is used to implement the above-mentioned assessment method steps, including:

[0074] The data collection module is used to collect the geometric parameters and material properties of each shaft segment of the rotor;

[0075] The model building module, connected to the data collection module, is used to build a finite element simulation model;

[0076] The unbalance distribution module, connected to the model construction module, is used to distribute the position, magnitude, and phase of the total rotor unbalance U.

[0077] The unbalanced response calculation module, connected to the unbalanced quantity allocation module, is used to calculate the vibration value A at the excitation side cantilever end;

[0078] The motor test simulation module, connected to the model building module, is used to simulate motor tests to obtain the vibration acceptance value AH of the excitation side cantilever end;

[0079] The evaluation module is used to compare the vibration value A at the excitation side overhang with the vibration acceptance value AH at the excitation side overhang to evaluate the vibration of the motor.

[0080] The beneficial effects of this invention are:

[0081] 1) This invention proposes a specialized vibration characteristic assessment system for motors without a third support at the excitation-side overhang, making motor assessment more accurate and targeted. It can accurately determine whether the vibration at this location is within a reasonable range, promptly identify potential vibration anomalies, and thus effectively ensure the safe and stable operation of the motor, avoiding safety hazards caused by inaccurate assessment systems.

[0082] 2) This invention is based on a well-defined vibration characteristic evaluation system. When requirements are met, a third support is unnecessary, simplifying the motor structure. This reduces the number of parts and installation processes, lowers manufacturing costs, reduces potential failure points, and improves the motor's reliability and maintainability, saving companies significant costs in equipment procurement, installation, and maintenance.

[0083] 3) The motor evaluated in this invention does not require a third support, making the installation process simpler and faster, saving installation time and labor costs. During maintenance, the tedious steps of disassembling and reinstalling the third support are eliminated, improving maintenance efficiency, reducing motor downtime, ensuring production continuity, and improving the company's production efficiency and economic benefits.

[0084] 4) The motor evaluated in this invention does not require a third support, effectively saving installation space. In spaces with limited space, the motor can be installed more flexibly, improving space utilization. Simultaneously, it reduces equipment investment and operating costs, increases the motor's economic efficiency, and gives the company a greater competitive advantage in the market, thus contributing to its sustainable development. Attached Figure Description

[0085] Figure 1 This is a schematic diagram of a motor structure without a third support in an application example of the present invention.

[0086] Figure 2 This is a schematic diagram of a motor structure with a third support in an application example of the present invention.

[0087] Figure 3 This is a modeled data diagram of the relevant shaft segments of a steam turbine generator in an application example of the present invention.

[0088] Figure 4This is a diagram showing the bearing oil film stiffness and damping stiffness parameters of a steam turbine generator in an application example of the present invention.

[0089] Figure 5 This is a schematic diagram of the process for simulating and obtaining the vibration acceptance value AH of the motor's excitation side overhang end according to the present invention.

[0090] Among them, 1. Rotor; 2. Stator; 3. Bearing; 4. Exciter; 5. Shaft; 6. Excitation side overhang; 7. Third support. Detailed Implementation

[0091] The present invention will be further described in detail below with reference to embodiments, but the implementation of the present invention is not limited thereto.

[0092] Example:

[0093] This invention provides a method for evaluating the vibration of a motor without a third support at the excitation-side overhang, comprising the following steps:

[0094] S1. Rotor three-dimensional dynamic modeling: Constructing a finite element simulation model based on motor rotor parameters;

[0095] S2. Calculate the total unbalance U of the rotor based on the balance quality level;

[0096] S3. Distribute the location, magnitude, and phase of the total unbalance U of the rotor.

[0097] S4. Unbalance Response Calculation: Based on the finite element simulation model in step S1, the distribution of the location, magnitude, and phase of the unbalance in step S3, and the rated speed n N Under the given boundary conditions, construct the dynamic equation for calculating the rotor-bearing unbalanced response and calculate the vibration value A at the excitation side overhang.

[0098] S5. The vibration acceptance value AH of the excitation side cantilever end is obtained based on the motor test simulation.

[0099] S6. Compare the vibration value A of the excitation side cantilever end in step S4 with the vibration acceptance value AH of the excitation side cantilever end in step S5; if A ≤ AH, then there is no need to set up a third support.

[0100] If A > AH, redesign the motor and repeat steps S1-S4 until A ≤ AH.

[0101] This invention presents a specialized vibration characteristic assessment system for motors without a third support at the excitation-side overhang, making motor assessment more accurate and targeted. It can accurately determine whether the vibration at this location is within a reasonable range, promptly identify potential vibration anomalies, and effectively ensure the safe and stable operation of the motor, avoiding safety hazards caused by inaccurate assessment systems. It guarantees the safe and stable operation of the motor from the assessment level, eliminating the safety hazards caused by the inaccuracy of existing general standard assessments. Based on this assessment system, the third support can be eliminated when vibration requirements are met, significantly simplifying the motor structure, reducing parts and installation processes, lowering manufacturing costs, reducing failure points, and improving the motor's operational reliability and maintainability. Eliminating the third support also simplifies motor installation, eliminates disassembly and assembly steps during maintenance, saves manpower and time costs, reduces downtime, and ensures production continuity. Furthermore, the design without a third support effectively saves installation space, improves space utilization, allows for more flexible installation of the motor in space-constrained locations, reduces equipment investment and operating costs, enhances the company's market competitiveness, and contributes to the company's sustainable development.

[0102] In an optional embodiment of the present invention, step S1, the rotor three-dimensional dynamic modeling includes the following steps:

[0103] S11. Establish a finite element model based on the geometric parameters and material properties of each shaft segment of the rotor;

[0104] The geometric parameters include: length L, mass m, diameter D, and stiffness EI;

[0105] The turbine generator shaft section includes a coupling shaft section, a journal shaft section, a main fan shaft section, an oil baffle shaft section, a retaining ring and main body overlapping shaft section, a rotor main body bearing, a slip ring shaft section, a slip ring fan shaft section, a speed measuring gear disk shaft section, and a key phase disk shaft section.

[0106] The material properties include elastic modulus, shear modulus, density, Poisson's ratio, and damping characteristics.

[0107] S12. Component Equivalent Treatment: In the finite element model, the mass, moment of inertia, and stiffness of the coupling, fan, retaining ring, coil, insulating block, axial lead, fan seat ring, slot wedge, slip ring, slip ring insulation, slip ring fan, speed measuring gear disk, and key phase disk are equivalent to concentrated mass points or distributed mass, and their axial positions are marked.

[0108] S13. Support boundary conditions: Combining the mass matrix [M], damping matrix [C], gyroscope matrix [G], linear stiffness matrix [K], nonlinear stiffness matrix [Ks], and excitation force {F(t)}, the core equation is constructed to solve for the critical speed, natural frequency, and vibration performance parameters;

[0109] The core equation is shown below:

[0110]

[0111] Where {F(t)} is the time-varying external excitation force vector, including the excitation force generated by the unbalanced mass, electromagnetic torque, bearing oil film force, or airflow disturbance force; {X ●●} represents the second derivative of displacement with respect to time, i.e., the instantaneous acceleration response vector of each degree of freedom of the rotor; {X ● {X} represents the first derivative of displacement with respect to time, i.e., the velocity response vector of each degree of freedom of the rotor; {X} represents the displacement response vector of each degree of freedom of the rotor.

[0112] S14. Model Verification: Modal tests are conducted on the finite element simulation model. The model parameters are corrected using the modal test data. The finite element simulation model is verified and iteratively optimized to ensure that the calculation error of the first critical speed (ncr1) and the second critical speed (ncr2) of the rotor is ≤5%.

[0113] The correction of model parameters using modal test data includes: stiffness parameter correction, mass distribution correction, and damping and gyroscopic effect correction;

[0114] The stiffness parameter correction is as follows: using the natural frequency obtained from the modal test as the target, adjust the elastic modulus of the central axis and the equivalent stiffness of the support in the model.

[0115] If the calculated critical speed is higher than the measured value, it indicates that the model stiffness is too high. Reduce the input value of the shaft's elastic modulus or decrease the stiffness coefficient of the support until the calculated frequency matches the experimental frequency.

[0116] For complex rotors, local stiffness correction can be used, adjusting the area with large mode deformation separately.

[0117] The mass distribution correction includes: when the amplitude of the mode shape deviates from the test value by more than ±5%, adding an equivalent mass unit at the corresponding position for compensation or adjusting the input value of the material density to correct the mass distribution.

[0118] If concentrated masses such as fan blades and couplings are not considered in the model, they can be compensated for by adding equivalent mass elements at the corresponding locations. If the mass distribution deviation is uniform, the input value of the material density can be adjusted. Note that mass adjustment will affect the critical speeds at all orders simultaneously, requiring comprehensive verification at multiple frequencies.

[0119] The damping and gyro effect corrections are as follows: For high-speed rotors, the gyro effect will cause the critical speed to bifurcate into positive and negative precession frequencies. If this phenomenon is observed in the experiment, a gyro matrix needs to be introduced into the model. If the vibration amplitude deviation exceeds ±5%, the damping coefficient in the model is adjusted using the damping ratio data from the modal test to improve the calculation accuracy of the resonance zone.

[0120] The verification and iterative optimization process involves: after correcting the model parameters, recalculating the rotor critical speed and comparing it with the measured value. If the error is still greater than 5%, the model parameters need to be corrected again until the calculation error between the first-order and second-order critical speeds is ≤5%, and the correlation coefficient of the mode shape is not less than 0.9. This ensures that the model accurately reflects the actual dynamic characteristics of the rotor.

[0121] The model parameters are corrected through a closed-loop process of "modal test data correction - finite element model iteration - critical speed verification". The core is to adjust key parameters such as stiffness and mass distribution in the model in reverse based on the natural frequency and mode shape data of the modal test, so that the error between the calculated critical speed and the measured value is controlled within 5%.

[0122] In this invention, step S1 involves accurately collecting the geometric and material properties of each shaft segment of the rotor, performing equivalent mass processing on auxiliary components, and constructing the core dynamic equations using a multi-matrix approach to reconstruct the actual structure and operating characteristics of the rotor. Modal testing is then used to correct multi-dimensional parameters such as stiffness and mass distribution. Closed-loop iteration controls the calculation errors of the rotor's first and second critical speeds within 5%, and the mode shape correlation coefficient is no less than 0.9, ensuring a high degree of consistency between the finite element model and the actual dynamic characteristics of the rotor. The rigor of its modeling and verification effectively avoids vibration calculation errors caused by simulation deviations, providing reliable model support for subsequent unbalance distribution and accurate measurement of vibration values ​​at the excitation side overhang. It also provides precise analytical basis for subsequent parameter optimization, ensuring the feasibility of the design without a third support and the stability of motor operation from the outset.

[0123] In an optional embodiment of the present invention, the total rotor imbalance U is calculated using the following formula, referring to standard GB / T 9239.12-2021:

[0124]

[0125] Where G represents the rotor balance quality grade, mm / s; M represents the rotor mass, kg; n N is the rated speed, rpm; e is the maximum residual eccentricity of the rotor, m.

[0126] In step S3, the distribution of the magnitude, position and phase of the total rotor imbalance U is described in accordance with the standard API 684-2010.

[0127] In this invention, step S2 calculates the total rotor imbalance using a formula based on the balance quality level, rotor mass, and rated speed, quantifying the basic data of rotor imbalance during operation. Step S3 completes the distribution of the total imbalance according to industry standards, clarifying the rules for the magnitude, phase, and positional allocation of the imbalance. The combination of these two steps transforms the abstract imbalance into concrete parameters that can be substituted into the simulation model, ensuring that the input data for subsequent imbalance response calculations is accurate and realistic, laying a data foundation for the accurate measurement of vibration values ​​at the excitation-side overhang.

[0128] In step S4, the calculation equation for the rotor-bearing imbalance response is as follows:

[0129]

[0130] In the formula, ω It is the rotational angular frequency; M 1 represents the overall quality matrix; K 1 represents the overall stiffness matrix; G 1 represents the global rotation matrix; c ij The overall oil film provides equivalent damping; k ij The overall oil film stiffness matrix; i, j=1, 2 ); U 1,2 It is the displacement vector; Q 1,c , Q 2,c The vector of the cosine component of the unbalanced force; Q 1,s , Q 2,s Ü is the sinusoidal component vector of the unbalanced force; 1,2 It is the second derivative of the displacement vector, i.e., the acceleration vector; 1,2 It is the first derivative of the displacement vector, i.e., the velocity vector.

[0131] In step S4, calculating the vibration value A at the excitation-side cantilever end includes the following steps:

[0132] Vibration value refers to the amplitude of rotor displacement vibration under unbalanced mass excitation force (such as the maximum radial displacement vibration value at a certain point), and is a key indicator for judging whether the rotor is "balanced and qualified".

[0133] S41. Obtain system parameters and unbalanced excitation parameters: System parameters include the overall mass matrix M1, stiffness matrix K1, and rotation matrix G1;

[0134] Unbalanced excitation parameters: cosine component vector Q of the unbalanced force 1c Q 2cThe sinusoidal component vector of the unbalanced force Q 1s Q 2s ; Rotational angular frequency ω;

[0135] The unbalanced force vector is expressed by the following formula:

[0136]

[0137] Where t is time; Q is the vector of unbalanced force; Q 1,2 Let Qc,s be the unbalanced force vector group; Qc,s are the cosine and sine components of the unbalanced force vector.

[0138] The cosine and sine components of the unbalanced force are determined by the magnitude and location of the unbalanced mass. For example, if the unbalanced mass *m* and the eccentricity *e* are at a point on the rotor, then Q = meω. 2 Then decompose it into components in the x / y directions;

[0139] The rotational angular frequency ω is the angular frequency corresponding to the rotor speed, ω=2πn / 60, where n is the rotational speed r / min;

[0140] S42. Substituting the assumed form of the response: Because the unbalanced excitation varies sinusoidally with time (cosωt or sinωt), according to the "frequency preservation property" of linear systems, the steady-state response of the system is also a sinusoidal signal of the same frequency. Therefore, the following assumptions are attached:

[0141] The response forms of displacement vectors U1 (x-direction displacement + rotation angle) and U2 (y-direction displacement + rotation angle) are as follows:

[0142]

[0143] Where U1 is the displacement vector of displacement + rotation in the x-direction; U2 is the displacement vector of displacement + rotation in the y-direction; {A1} and {A2} are the cosine component vectors of the response; {B1} and {B2} are the sine component vectors of the response.

[0144]

[0145]

[0146] U1 corresponds to the vector group of displacement in the X direction plus positive rotation angle, representing the linear displacement (x1, x2, ..., xn) of n characteristic points of the rotor in the X direction. n The rotation angle (θ) of n feature points about the Y-axis (axis of rotation) y1 θ y2 , …, θ yn );

[0147] U2 corresponds to the vector set of displacement in the Y direction minus the negative rotation angle, representing the linear displacement (y1, y2, ..., y2) of n characteristic points of the rotor in the Y direction. n The rotation angle (-θ) of n feature points about the X-axis (axis of rotation) x1 , -θ x2 , …, -θ xn );

[0148] S43. Substituting the unbalanced force vector from step S41 and the assumed form from step S42 into the rotor-bearing unbalanced response calculation equation, we obtain:

[0149] Substituting the decomposition form of the unbalanced excitation and the assumed form of the response into the rotor-bearing dynamic equations, and utilizing the "orthogonality of trigonometric functions" (i.e., the coefficients of cosωt and sinωt must be equal), the time term t can be eliminated, resulting in a system of linear algebraic equations about {A1}, {B1}, {A2}, and {B2}:

[0150]

[0151] in:

[0152] • The T on the left 11 T 12 T 21 T 22 It is the coefficient matrix (composed of system parameters M1, K1, G1, C) ij K ij It is calculated from the rotational speed ω;

[0153] • The right side is the component matrix of unbalanced excitation (derived from Q in step S41). 1C Q 2C Q 1S Q 2S composition).

[0154] in, ;

[0155] S44. Solve the system of algebraic equations to calculate the amplitude: Solve for {A1}, {B1}, {A2}, and {B2} by matrix inversion or numerical methods; substitute them into the assumed form to obtain the displacement vector, that is, the vibration value of each point in different directions, and then calculate the vibration value A of the excitation side cantilever end.

[0156] The equations obtained in step S43 are a system of linear equations, with unknowns {A1}, {B1}, {A2}, and {B2}. These unknowns can be solved by matrix inversion or numerical methods (such as Gaussian elimination). After obtaining {A1}, {B1}, {A2}, and {B2}, they can be substituted into formula (3) to obtain the system displacement vector. The unbalanced response amplitude of each degree of freedom can be calculated by "combining sine / cosine components", such as the amplitude at point n in the X direction:

[0157] .

[0158] In an optional embodiment of the present invention, step S5, obtaining the vibration acceptance value AH of the excitation side cantilever end based on motor test simulation, includes the following steps:

[0159] S51. Conduct test simulations on the motor to obtain the set of vibration values ​​H of the excitation side cantilever end under different operating conditions without a third support;

[0160] S52. Screen the vibration value set H to select the set A1 of the cantilever end vibration values ​​that meet the motor operation requirements: stable contact between the brush and the slip ring with no obvious sparks; appropriate brush wear; and stable operation for 6 months or more; and the set A2 of the cantilever end vibration values ​​where the shaft vibration and bearing vibration are within the allowable range.

[0161] S53. Perform statistical analysis on vibration value sets A1 and A2, and screen out vibration value sets A1 and A2 that indicate the need for brush replacement after 6 months of wear. 10 ; Select the set A of vibration values ​​at the excitation side overhang when the vibration of the motor journal and bearing is at the upper limit of the allowable value. 20 ;

[0162] S54, Using the vibration value set A from step S53 10 and vibration value set A 20 Select vibration value set A 10 The minimum value A in 1min and vibration value set A 20 The minimum value A in 2min ;

[0163] S55, via A 1min and A 2min The vibration acceptance value AH of the excitation side cantilever end was calculated.

[0164] In step S51, the set of vibration values ​​H at the motor excitation side overhang end is obtained through the following steps:

[0165] S511. Determine the test object: Select the appropriate motor type as the test object for the type of motor to be evaluated;

[0166] S512. Conduct dynamic balancing and type tests on the test object under various simulated actual operating conditions.

[0167] The actual operating conditions include rated operating conditions, maximum load operating conditions, reduced load operating conditions, rated speed, over-critical speed, and 120% rated speed overspeed;

[0168] S513, the set of vibration values ​​H at the motor excitation side overhang during the measurement and statistical dynamic balancing test and type test.

[0169] In steps S52 and S53, the vibration of the motor journal and bearing is at the upper limit of the allowable value, as detailed in the standard "ISO 20816-2 Mechanical vibration — Measurement and evaluation of machine vibration — Part 2: Land-based gas turbines, steam turbines and generators in excess of 40 MW, with fluid-film bearings and rated speeds of 1500 r / min, 1800 r / min, 3000 r / min and 3600 r / min".

[0170] In step S55, the vibration acceptance value AH of the excitation side cantilever end is calculated using the following formula:

[0171]

[0172] Where x is the safety margin, with a value ranging from 0.8 to 0.9.

[0173] In this invention, vibration data is obtained through simulated full-condition tests of the motor. Statistical analysis is then used to extract the critical minimum values ​​that meet operational requirements. Finally, a safety margin is introduced to calculate the acceptance value, balancing motor operational safety with component lifespan. The established acceptance value AH closely matches the structural characteristics and actual operating conditions of a motor without a third support, avoiding component damage and operational failures caused by excessive vibration. By comparing vibration values, a scientific and reasonable quantitative benchmark is provided to determine the feasibility of the design without a third support, ensuring the safety and practicality of the motor design.

[0174] Application examples

[0175] This application example provides a vibration assessment system for a motor without a third support at the excitation-side overhang end, characterized in that: the system is used to implement the above-mentioned assessment method steps, including:

[0176] The data collection module is used to collect the geometric parameters and material properties of each shaft segment of the rotor;

[0177] The model building module, connected to the data collection module, is used to build a finite element simulation model;

[0178] The unbalance distribution module, connected to the model construction module, is used to distribute the position, magnitude, and phase of the total rotor unbalance U.

[0179] The unbalanced response calculation module, connected to the unbalanced quantity allocation module, is used to calculate the vibration value A at the excitation side cantilever end;

[0180] The motor test simulation module, connected to the model building module, is used to simulate motor tests to obtain the vibration acceptance value AH of the excitation side cantilever end;

[0181] The evaluation module is used to compare the vibration value A at the excitation side overhang with the vibration acceptance value AH at the excitation side overhang to evaluate the vibration of the motor.

[0182] In this application example, a steam turbine generator is selected as the evaluation object. The steam turbine generator includes a stator 2, a rotor 1, a pair of bearings 3, and an exciter 4. The rotor 1 is located inside the cavity of the stator 2; the shafts 5 at both ends of the rotor 1 extend out of the end walls of the stator 2, respectively; the bearings 3 are respectively supported on the shafts 5 at both ends of the rotor; the exciter 4 is mounted on the shaft 5 on one side of the rotor 1, and the shaft on the exciter side is the excitation-side overhang 6, as shown below. Figure 1 The diagram shown is a structural schematic of a steam turbine generator without the third support 7. Figure 2 The diagram shows a steam turbine generator with a third support 7.

[0183] In this application example, based on the actual structure of the rotor, it is divided into several shaft segments along the rotor axis, and each shaft segment is modeled. In this application example, the motor / generator rotor is simplified into 30 shaft segments according to the actual structure. Let be the length of the segment element of the i-th segment. Let be the diameter of the segment element for the i-th segment. Relevant segment modeling data is as follows: Figure 3 As shown.

[0184] The turbine generator rotor is a precision component made of multiple materials, its main structure ingeniously integrating epoxy insulation, aluminum, copper, and steel. Each material contributes its strengths to ensure stable rotor operation. Table 1 shows the materials and their properties used in this application example.

[0185] Table 1. Rotor materials and material properties in application examples.

[0186]

[0187] During the operation of a steam turbine generator, the characteristics of the bearing oil film change with the rotational speed, and its oil film stiffness and damping stiffness parameters are key indicators. Detailed measurements and analyses were conducted under different rotational speed conditions, and the specific details of the obtained bearing oil film stiffness and damping stiffness parameters are as follows: Figure 4 As shown;

[0188] Based on the aforementioned input data and related calculation steps, the rotor vibration of the generator in this application example at a rated speed of 3000 rpm was analyzed and calculated for the unbalanced condition of G2.5 balance level. The results show that the vibration amplitudes of the rotor at the steam-side journal, excitation-side journal, and slip ring tail end are 51 μm, 30 μm, and 46 μm, respectively.

[0189] As attached Figure 5 As shown, the process for determining the vibration acceptance value AH at the motor excitation side overhang is as follows: First, identify the type of motor (e.g., steam turbine generator, hydro turbine generator, etc.) and the specific operating conditions (including rated and maximum load conditions), and record the vibration value H at the motor overhang and its corresponding operating performance under different motor parameters and operating conditions; next, filter the recorded data based on the operating performance (e.g., stable contact between brushes and slip rings, no abnormal journal vibration, etc.); then, select the corresponding minimum A from the filtered data. 10min and A 20min Finally, considering a safety margin of 0.8–0.9, the vibration acceptance value AH of the excitation side cantilever end is calculated comprehensively.

[0190] To verify the reasonableness of the allowable vibration acceptance value AH at the excitation side overhang of a steam turbine generator using motors with different known vibration calculation values, the following steps can be taken:

[0191] 1) Define the basic data: Collect the existing allowable vibration acceptance value AH of the excitation side overhang of the steam turbine generator, and organize the known vibration calculation values ​​of different motors to ensure that all data units are consistent and the measurement conditions are consistent.

[0192] 2) Confirm operating conditions: For each turbine generator used for verification, ensure that it operates in a state where there is no abnormality in the vibration of the turbine and excitation shafts and bearings, and that the brushes and slip rings have stable contact, no obvious sparks, brush wear for more than 6 months, and brush wear can still be maintained for more than 6 months when the slip rings occasionally spark. Record the condition that each generator meets these conditions.

[0193] 3) Compare calculated values ​​with acceptance values: Compare the calculated vibration value of each motor with the acceptance allowable value AH one by one. If the calculated vibration value of most motors is less than or equal to AH, it initially indicates that the acceptance allowable value AH is reasonable within the existing motor group; if the calculated vibration value of most motors is greater than AH, it indicates that AH may be too small, and its reasonableness is questionable.

[0194] 4) Analyze special cases: For individual motors where the vibration calculation value differs significantly from the AH value, conduct an in-depth analysis of the specific reasons. Check whether there are any unique aspects in the motor's structural design, manufacturing process, or operating parameters that affect the accuracy of the vibration calculation value or render the original acceptance standards inapplicable.

[0195] 5) Comprehensive evaluation in conjunction with other parameters: In addition to comparing the vibration calculation values ​​with the acceptance values, it is also necessary to comprehensively consider the actual situation of each motor, such as shaft vibration, bearing vibration, and brush arcing. If the motor can still maintain a stable level of shaft vibration and bearing vibration when the vibration calculation value is close to or exceeds the AH, and the brush arcing is within an acceptable range, then the questioning of the rationality of AH can be appropriately relaxed; conversely, if the vibration calculation value does not exceed AH, but the motor exhibits abnormal shaft vibration, bearing vibration, or severe brush arcing, then the rationality of AH needs to be re-examined.

[0196] 6) Drawing Verification Conclusions: Based on the above comparisons, analyses, and evaluations, a final conclusion is drawn regarding the reasonableness of the allowable vibration acceptance value AH at the excitation side of the turbine generator. If, after comprehensive verification, AH effectively ensures the safe and stable operation of the generator under specific operating conditions, it is deemed reasonable; otherwise, AH needs to be adjusted or redefined.

[0197] Table 2 details the verification results of some motors. This table records the calculated vibration values, shaft vibration, bearing vibration, brush sparking, and other key parameters of different motors under specific operating conditions, providing intuitive and important data for a comprehensive evaluation of the rationality of the AH value.

[0198] Table 2 shows the verification results of some motors in the application examples.

[0199]

[0200] As can be seen from Table 2:

[0201] 1) Comparison of vibration calculation value and acceptance value: The vibration calculation value of most motors is less than or equal to the acceptance allowable value AH (100μm), but the vibration calculation value of some motors (such as motors 1, 2, 7, 8, and 11) exceeds AH.

[0202] 2) Motor operating status: Motors with vibration calculation values ​​exceeding AH are often accompanied by abnormal conditions such as excessive shaft vibration, bearing vibration, or brush sparking. For example, motors 7 and 8 have high maximum shaft vibration and excitation-side bearing vibration, and carbon brush sparking is also present.

[0203] 3) Remedial Measures: For motors whose vibration calculation values ​​exceeded AH and exhibited abnormal conditions, remedial measures such as adding a third support or redesigning the motor were taken. For example, motors 2, 7, and 8 were given a third support, while motors 1 and 11 were deemed to require redesign.

[0204] For motors that need to be redesigned (such as motors 1 and 11), adjustments can generally be made in the following aspects:

[0205] 1) Optimize the cantilever end structure:

[0206] Shape improvement: By changing the geometry of the overhang, such as changing the two-section shaft screw connection to a single shaft integrated type.

[0207] Size adjustment: Increase the diameter or thickness of the overhang to improve its moment of inertia, thereby enhancing the structural stiffness.

[0208] Material selection: Use materials with higher strength and higher modulus, such as alloy steel or composite materials, to improve the vibration resistance of the cantilever end.

[0209] 2) Add supporting structure:

[0210] Support point location: Add support points reasonably near the cantilever end to ensure that the support points can effectively disperse vibration energy.

[0211] Support structure type: Select the appropriate support structure type, such as elastic support or rigid support, according to the specific structure and operating conditions of the motor, in order to reduce the vibration level.

[0212] To reduce the vibration value A at the overhanging end of the excitation side, the following approaches can generally be taken:

[0213] 1) Enhance structural stiffness:

[0214] Increased cross-sectional area: By increasing the cross-sectional area of ​​the cantilever end, its bending stiffness and torsional stiffness are improved, thereby reducing vibration. Specific parameter adjustments include increasing the diameter or wall thickness of the cantilever end.

[0215] Reduce overhang length: While meeting the overall design and operation requirements of the motor, minimize the length of the overhang end to reduce its vibration sensitivity.

[0216] 2) Optimize dynamic balancing:

[0217] Improved balance accuracy: High-precision dynamic balancing adjustment of the motor rotor ensures that the unbalanced mass of the rotor is within the allowable range, reducing vibration caused by imbalance.

[0218] Balance position selection: Based on the rotor's structure and operating characteristics, select a suitable balance position, such as adding or removing weight, to achieve the best balance effect.

[0219] 3) Improve support conditions:

[0220] Support structure stability: Ensure the motor's support structure is stable and reliable, without any loosening or deformation. Conduct regular inspections and maintenance of the support structure to promptly identify and address any potential problems.

[0221] Support stiffness matching: Based on the operating conditions and vibration characteristics of the motor, the stiffness of the support structure is reasonably adjusted to ensure that it matches the overall stiffness of the motor, thereby reducing vibration caused by insufficient or excessive support stiffness.

[0222] It is understood that the present invention has been described through some embodiments, and those skilled in the art will recognize that various changes or equivalent substitutions can be made to these features and embodiments without departing from the spirit and scope of the invention. Furthermore, under the teachings of the present invention, these features and embodiments can be modified to adapt to specific situations and materials without departing from the spirit and scope of the invention. Therefore, the present invention is not limited to the specific embodiments disclosed herein, and all embodiments falling within the scope of the claims of this application are within the protection scope of the present invention.

Claims

1. A method for evaluating the vibration of a motor without a third support at the excitation-side cantilever end, characterized in that: Includes the following steps: S1. Rotor three-dimensional dynamic modeling: Constructing a finite element simulation model based on motor rotor parameters; S2. Calculate the total unbalance U of the rotor based on the balance quality level; S3. Distribute the location, magnitude, and phase of the total unbalance U of the rotor. S4. Unbalance Response Calculation: Based on the finite element simulation model in step S1, the distribution of the location, magnitude, and phase of the unbalance in step S3, and the rated speed n N Under the given boundary conditions, construct the dynamic equation for calculating the rotor-bearing unbalanced response and calculate the vibration value A at the excitation side overhang. S5. The vibration acceptance value AH of the excitation side cantilever end is obtained based on the motor test simulation. S6. Compare the vibration value A of the excitation side cantilever end in step S4 with the vibration acceptance value AH of the excitation side cantilever end in step S5; if A≤ AH, then there is no need to set up a third support. If A > AH, redesign the motor and repeat steps S1-S4 until A ≤ AH.

2. The method for evaluating motor vibration without a third support at the excitation-side cantilever end according to claim 1, characterized in that: In step S1, the rotor three-dimensional dynamic modeling includes the following steps: S11. Establish a finite element model based on the geometric parameters and material properties of each shaft segment of the rotor; S12. Component Equivalent Treatment: In the finite element model, the mass, moment of inertia, and stiffness of each rotor component are equivalent to concentrated mass points or distributed mass, and their axial positions are marked. S13. Support boundary conditions: Combining the mass matrix [M], damping matrix [C], gyroscope matrix [G], linear stiffness matrix [K], nonlinear stiffness matrix [Ks], and excitation force {F(t)}, the core equation is constructed to solve for the critical speed, natural frequency, and vibration performance parameters; The core equation is shown below: Where {F(t)} is the time-varying external excitation force vector; {X ●● } represents the second derivative of displacement with respect to time, i.e., the instantaneous acceleration response vector of each degree of freedom of the rotor; {X ● {X} represents the first derivative of displacement with respect to time, i.e., the velocity response vector of each degree of freedom of the rotor; {X} represents the displacement response vector of each degree of freedom of the rotor. S14. Model Verification: Modal tests are conducted on the finite element simulation model. The model parameters are corrected using the modal test data. The finite element simulation model is verified and iteratively optimized to ensure that the calculation error of the first critical speed (ncr1) and the second critical speed (ncr2) of the rotor is ≤5%.

3. The method for evaluating motor vibration without a third support at the excitation-side cantilever end according to claim 2, characterized in that: In step S11, the geometric parameters include: length L, mass m, diameter D, and stiffness EI.

4. The method for evaluating motor vibration without a third support at the excitation-side cantilever end according to claim 2, characterized in that: In step S11, the material properties include elastic modulus, shear modulus, density, Poisson's ratio, and damping characteristics.

5. The method for evaluating motor vibration without a third support at the excitation-side cantilever end according to claim 2, characterized in that: In step S14, the correction of model parameters using modal test data includes: stiffness parameter correction, mass distribution correction, and damping and gyro effect correction.

6. The method for evaluating motor vibration without a third support at the excitation-side overhang end according to claim 5, characterized in that: The stiffness parameter correction is as follows: taking the natural frequency obtained from the modal test as the target, adjust the elastic modulus of the shaft and the equivalent stiffness of the support in the model; when the calculated critical speed is higher than the measured value, it indicates that the model stiffness is too large, reduce the input value of the elastic modulus of the shaft, or reduce the stiffness coefficient of the support, until the calculated frequency matches the test frequency.

7. The method for evaluating motor vibration without a third support at the excitation-side overhang end according to claim 5, characterized in that: The mass distribution correction includes: when the amplitude of the mode shape deviates from the test value by more than ±5%, adding an equivalent mass unit at the corresponding position for compensation or adjusting the input value of the material density to correct the mass distribution.

8. The method for evaluating motor vibration without a third support at the excitation-side overhang end according to claim 5, characterized in that: The damping and gyro effect correction is as follows: For high-speed rotors, the gyro effect will cause the critical speed to bifurcate into positive and negative precession frequencies. If the gyro effect is observed in the experiment, a gyro matrix is ​​introduced into the model. If the vibration amplitude deviation exceeds ±5%, the damping coefficient in the model is adjusted by using the damping ratio data from the modal test to improve the calculation accuracy of the resonance zone.

9. The method for evaluating motor vibration without a third support at the excitation-side overhang end according to claim 5, characterized in that: The verification and iterative optimization are as follows: after correcting the model parameters, recalculate the rotor critical speed and compare it with the measured value. If the error is still greater than 5%, the model parameters need to be corrected again until the calculation error between the first-order critical speed and the second-order critical speed is ≤5%, and the correlation coefficient of the mode shape is not less than 0.

9.

10. The method for evaluating motor vibration without a third support at the excitation-side cantilever end according to claim 1, characterized in that: In step S2, the total rotor imbalance U is calculated using the following formula: Where G represents the rotor balance quality grade, mm / s; M represents the rotor mass, kg; n N is the rated speed, rpm; e is the maximum residual eccentricity of the rotor, m.

11. The method for evaluating motor vibration without a third support at the excitation-side overhang end according to claim 1, characterized in that: In step S4, the dynamic equation for calculating the rotor-bearing imbalance response is shown in the following equation: In the formula, ω It is the rotational angular frequency; M 1 represents the overall quality matrix; K 1 represents the overall stiffness matrix; G 1 represents the global rotation matrix; c ij The overall oil film equivalent damping matrix; k ij The overall oil film stiffness matrix; i, j=1, 2 ); U 1,2 It is the displacement vector; Q 1,c , Q 2,c The vector of the cosine component of the unbalanced force; Q 1,s , Q 2,s Ü is the sinusoidal component vector of the unbalanced force; 1,2 It is the second derivative of the displacement vector, i.e., the acceleration vector; 1,2 It is the first derivative of the displacement vector, i.e., the velocity vector.

12. The method for evaluating motor vibration without a third support at the excitation-side overhang end according to claim 11, characterized in that: In step S4, calculating the vibration value A at the excitation-side cantilever end includes the following steps: S41. Obtain system parameters and unbalanced excitation parameters: System parameters include the overall mass matrix M1, stiffness matrix K1, and rotation matrix G1; The unbalanced excitation parameters include the cosine component vector Q of the unbalanced force. 1c Q 2c and the sinusoidal component vector Q of the unbalanced force 1s Q 2s and rotational angular frequency ω; The unbalanced force vector is expressed by the following formula: Where t is time; Q is the vector of unbalanced force; Q 1,2 Let Qc,s be the unbalanced force vector group; Qc,s are the cosine and sine components of the unbalanced force vector. S42. Substituting the assumed form of the response: The unbalanced excitation varies sinusoidally with time. Assume the response of the displacement vector is as follows: Where U1 is the displacement vector of displacement + rotation in the x-direction; U2 is the displacement vector of displacement + rotation in the y-direction; {A1} and {A2} are the cosine component vectors of the response; {B1} and {B2} are the sine component vectors of the response. S43. Substituting the unbalanced force vector from step S41 and the assumed form from step S42 into the rotor-bearing unbalanced response calculation equation, we obtain: in, ; S44. Solve the system of algebraic equations to calculate the amplitude: Solve for {A1}, {B1}, {A2}, and {B2} by matrix inversion or numerical methods; substitute them into the assumed form to obtain the displacement vector, that is, the vibration value of each point in different directions, and then calculate the vibration value A of the excitation side cantilever end.

13. The method for evaluating motor vibration without a third support at the excitation-side overhang end according to claim 1, characterized in that: In step S5, the vibration acceptance value AH of the motor's excitation side overhang end is obtained based on the motor test simulation, including the following steps: S51. Conduct test simulations on the motor to obtain the set of vibration values ​​H of the excitation side cantilever end under different operating conditions without a third support; S52. Screen the vibration value set H to select the set A1 of the cantilever end vibration values ​​that meet the motor operation requirements: stable contact between the brush and the slip ring with no obvious sparks; appropriate brush wear; and stable operation for 6 months or more; and the set A2 of the cantilever end vibration values ​​where the shaft vibration and bearing vibration are within the allowable range. S53. Perform statistical analysis on vibration value sets A1 and A2, and screen out vibration value sets A1 and A2 that indicate the need for brush replacement after 6 months of wear. 10 ; Select the set A of vibration values ​​at the excitation side overhang when the vibration of the motor journal and bearing is at the upper limit of the allowable value. 20 ; S54, Using the vibration value set A from step S53 10 and vibration value set A 20 Select vibration value set A 10 The minimum value A in 1min and vibration value set A 20 The minimum value A in 2min ; S55, via A 1min and A 2min The vibration acceptance value AH of the excitation side cantilever end was calculated.

14. The method for evaluating motor vibration without a third support at the excitation-side overhang end according to claim 13, characterized in that: In step S51, the set of vibration values ​​H at the motor excitation side overhang end is obtained through the following steps: S511. Determine the test object: Select the appropriate motor type as the test object for the type of motor to be evaluated; S512. Conduct dynamic balancing and type tests on the test object under various simulated actual operating conditions. The actual operating conditions include rated operating conditions, maximum load operating conditions, reduced load operating conditions, rated speed, over-critical speed, and 120% rated speed overspeed; S513, Measurement and statistical analysis of the set of vibration values ​​H at the motor excitation side cantilever end during dynamic balancing and type testing; In step S55, the vibration acceptance value AH of the excitation side cantilever end is calculated using the following formula: Where x is the safety margin, with a value ranging from 0.8 to 0.

9.

15. A vibration assessment system for a motor without a third support at the excitation-side cantilever end, characterized in that: The system is used to implement the evaluation method steps according to any one of claims 1-14, including: The data collection module is used to collect the geometric parameters and material properties of each shaft segment of the rotor; The model building module, connected to the data collection module, is used to build a finite element simulation model; The unbalance distribution module, connected to the model construction module, is used to distribute the position, magnitude, and phase of the total rotor unbalance U. The unbalanced response calculation module, connected to the unbalanced quantity allocation module, is used to calculate the vibration value A at the excitation side cantilever end; The motor test simulation module, connected to the model building module, is used to simulate motor tests to obtain the vibration acceptance value AH of the excitation side cantilever end; The evaluation module is used to compare the vibration value A at the excitation side overhang with the vibration acceptance value AH at the excitation side overhang to evaluate the vibration of the motor.

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