A design and verification method of a vibration isolation scheme of a robotic arm rotor system

Abstract

This invention discloses a design and verification method for vibration isolation schemes in an airframe rotor system, comprising two stages: The first stage is support design, which involves constructing a finite element model, applying static loads, performing topology optimization with the goal of minimizing flexibility, and conducting static, modal, and critical origin dynamic stiffness analyses on the reconstructed model to ensure that the support meets the basic dynamic conditions for the vibration isolator's operation; the second stage is virtual verification of system-level vibration isolation performance, which involves constructing two parallel virtual vibration test platforms containing the vibration isolator and rigid connections based on the qualified support model, performing dynamic simulations under the same vibration load spectrum, and calculating the overall system isolation rate by extracting and comparing the vibration responses of the two platforms at the same location, thereby quantitatively evaluating the effectiveness of the vibration isolation scheme before the physical prototype is manufactured. This invention achieves fully digital design and performance prediction of the vibration isolation scheme, fundamentally avoiding the risk of vibration isolation failure caused by the soft support effect.

Description

Technical Field

[0001] This invention belongs to the field of aircraft structural design and vibration control technology, and in particular relates to a design and verification method for a vibration isolation scheme for an airframe rotor system. Background Technology

[0002] As a new type of aircraft that combines vertical takeoff and landing with high-speed cruise capabilities, tiltrotor aircraft experience harsh dynamic environments under complex flight conditions. The vibrations generated by the rotor operation have significant characteristics of multiple frequency bands (such as the blade passing frequency and its harmonics) and high magnitude (large root mean square value of total acceleration). If such vibrations are not effectively isolated, they will lead to fatigue damage to the arm structure, deterioration of the performance of airborne equipment, or even failure, seriously affecting flight safety and mission reliability.

[0003] Traditionally, the development of vibration isolation schemes for airframe rotor systems (including isolator selection and support design) relies heavily on a physical iterative model of "design-prototype-testing." This involves first designing the support and initially selecting isolators based on experience, then manufacturing expensive metal prototypes, and finally verifying the vibration isolation effect through vibration table testing. This method has the following inherent drawbacks: (1) Design blind spots and iteration risks: The design of the support and the selection of the vibration isolator are often separated. Engineers usually focus on the static strength, modal frequency and lightweight of the support, but generally ignore a key indicator, namely the dynamic stiffness of the origin of the connection interface between the support and the vibration isolator. If the dynamic stiffness of the support itself is insufficient, even the best vibration isolator will cause the vibration isolation performance of the arm rotor system to deteriorate sharply due to the soft support effect, and even amplify the vibration. This problem is only exposed in the physical test stage, leading to the design overturning and huge cost waste. (2) Performance evaluation is delayed and costly: The vibration isolation performance of the system must be tested on an expensive physical vibration table and a complete prototype to obtain quantitative results (vibration isolation rate). If the results are not up to standard, it means that redesign, manufacturing and testing are required, which takes a long time (in months) and is extremely expensive. (3) Lack of accurate virtual prediction methods: Existing computer-aided engineering (CAE) methods are mostly used for component-level strength and modal analysis, or for response calculation of single system models. There is a lack of a standard method that can directly and quantitatively compare and calculate the system-level vibration isolation rate in a virtual environment, and it is impossible to make a high-confidence prediction and optimization of the effectiveness of the design scheme before the physical prototype.

[0004] Therefore, there is an urgent need for an innovative digital approach that can ensure the support meets the basic dynamic conditions for vibration isolator operation during the design phase, and can accurately predict the vibration isolation performance of the system through virtual testing, thereby reducing physical iterations, shortening the development cycle, and lowering R&D costs.

[0005] Therefore, it is necessary to provide a design and verification method for vibration isolation schemes of airframe rotor systems to solve the above-mentioned technical problems. Summary of the Invention

[0006] The main objective of this invention is to provide a design and verification method for vibration isolation schemes of airframe rotor systems. This method addresses the core pain points of traditional vibration isolation schemes, such as lagging performance evaluation, numerous design blind spots, and reliance on physical iteration. It achieves the goal of ensuring the dynamic conditions of the support structure and accurately predicting the vibration isolation performance of the system during the design phase.

[0007] This invention achieves the above objective through the following technical solution: a design and verification method for a vibration isolation scheme of an airframe rotor system, comprising the following steps: Phase 1: Design of brackets that meet the installation requirements of vibration isolators, including the following steps: Step S1: Construct a finite element model of the support structure for the arm-rotor system and apply a six-force static load to the arm-rotor system. Step S2: With the goal of minimizing structural flexibility, perform topology optimization on the support under volume fraction constraints, and reconstruct the geometric model of the support based on the optimization results; Step S3: Perform static and modal analysis on the geometric model of the reconstructed support to verify its static stress σ_ s It is less than the material fatigue limit, and its first natural frequency avoids the preset excitation frequency band of the vibration source; Step S4: At the installation interface connecting the bracket and the vibration isolator, perform origin dynamic stiffness analysis, apply excitation force to perform frequency response simulation, calculate the dynamic stiffness curve of the installation interface in the preset working frequency band, and determine whether the minimum dynamic stiffness reaches the preset dynamic stiffness threshold. If it does not reach the threshold, return to step S1 to modify the design, and repeat steps S1 to S4 until the minimum dynamic stiffness of the bracket is greater than the dynamic stiffness threshold. Phase Two: Virtual verification of system-level vibration isolation performance based on qualified supports, including the following steps: Step S5: Based on the support model that has been verified and approved in step S4, construct two parallel virtual vibration test platforms, namely the first platform and the second platform; Step S6: Under the same boundary conditions and vibration load spectrum excitation simulating the real environment, perform dynamic simulations on the first platform and the second platform respectively; Step S7: Extract the frequency domain data of the vibration response of the mass point at the same position from the simulation results of the first platform and the second platform respectively; Step S8: Based on the extracted response frequency domain data, calculate the root mean square value (RMS) of the acceleration of the first platform. i The root mean square value of acceleration of the second platform (RMS(A)) j ); Step S9: According to the formula: Vibration isolation ratio β = [1 - RMS(A)] i ) / RMS(A j The overall vibration isolation rate of the computer arm rotor system is calculated as 100% to evaluate the effectiveness of the vibration isolation scheme for the arm rotor system.

[0008] Furthermore, in step S3, the first-order natural frequency and the excitation frequency band of the vibration source must maintain an interval of at least 5%. Furthermore, in step S4, the preset operating frequency band is 15~2000Hz; the dynamic stiffness threshold is 1000 N / mm.

[0009] Furthermore, in step S5, the first platform is a model with a vibration isolator, and an elastic connection unit simulating the vibration isolator characteristics is established between the support and the mass point of the simulated arm rotor system; the second platform is a rigid connection model, and the elastic connection unit is replaced with a connection unit simulating a rigid connection, while the rest remains consistent with the settings of the first platform.

[0010] Furthermore, the elastic connection unit is a Busing unit, whose parameters are set according to the triaxial stiffness and damping of the actual vibration isolator.

[0011] Furthermore, the connection elements for simulating rigid connections are spring elements or constraint elements that directly couple degrees of freedom.

[0012] Furthermore, in step S6, the vibration load spectrum is an acceleration power spectral density curve that is a superposition of broadband random and narrowband random.

[0013] Furthermore, following the simulation in step S6, step S10 is also included: extracting the dynamic stress σ_ of the support from the simulation results of the first platform and the second platform, respectively. d The total stress σ of the support in the first platform is verified to be less than the fatigue limit of the material in order to evaluate the total strength of the support. The total stress σ of the support is the static stress σ_ s With dynamic stress σ_ d The sum of these, i.e., the total stress σ = σ_ s +σ_ d .

[0014] Furthermore, in step S5, when constructing the first platform, the theoretical natural frequency of the vertical single degree of freedom of the arm rotor system is pre-calculated. The pre-calculated result is compared with the vertical natural frequency of the system extracted by simulation in step S6 for verification. The deviation between the two must be less than 10%. If the deviation exceeds 10%, the parameters of the elastic connection unit or the connection relationship of the overall model need to be calibrated to ensure the dynamic equivalence of the virtual model.

[0015] Furthermore, after calculating the overall vibration isolation rate in step S9, step S11 is also included: based on the vibration response frequency domain data extracted in step S7, acceleration power spectral density curves of the first platform and the second platform are generated respectively; comparative analysis is performed on the attenuation of the peak value of the acceleration power spectral density of the first platform relative to that of the second platform within the preset narrow band of vibration source excitation, so as to evaluate the isolation effect of the vibration isolator on specific discrete frequency vibration components.

[0016] Compared with existing technologies, the beneficial effects of the design and verification method for vibration isolation schemes of airframe rotor systems proposed in this invention are as follows: Through an innovative digital process, it successfully solves the core pain points in the development of traditional vibration isolation schemes, such as lagging performance evaluation, numerous design blind spots, and reliance on physical iterations. It achieves the goal of ensuring the dynamic conditions of the support structure and accurately predicting the vibration isolation performance of the system during the design phase, resulting in significant comprehensive effects of improving design quality, shortening the cycle, reducing costs, and ensuring safety. Specifically: (1) Completely eliminate design blind spots and ensure the foundation conditions for vibration isolation from the root: Innovatively, the dynamic stiffness of the support origin is incorporated as the core design index into the first stage of optimization and verification process. By establishing a clear dynamic stiffness threshold (e.g., ≥1000 N / mm) and virtual frequency response simulation verification, potential vibration isolation failure caused by soft support effect can be discovered and corrected in the early stage of design. This avoids the risk of design fragmentation in traditional methods and major defects being exposed only in the physical test stage, thus fundamentally ensuring the dynamic basis for the normal operation of the vibration isolator. (2) Achieve early, accurate, and quantitative virtual prediction of system-level vibration isolation performance: A parallel virtual vibration test platform with "rigid connection" and "with vibration isolator" was constructed. By applying a vibration load spectrum that conforms to the real working conditions and conducting comparative simulation, the overall vibration isolation rate of the system can be directly calculated in the digital environment. This method provides a standardized and quantifiable virtual prediction method for the first time. The effectiveness of the vibration isolation scheme can be evaluated with high confidence in the design stage, realizing a fundamental shift in performance evaluation from "after physical test" to "during design". (3) Significantly reduce physical iterations and reduce R&D costs and cycle: Through a series of fully digital processes such as topology optimization, static and dynamic mechanical analysis, dynamic stiffness verification and virtual vibration isolation test, multiple rounds of design optimization and performance verification can be completed before manufacturing expensive metal prototypes. The examples show that a design scheme that meets the requirements of static strength, dynamic stiffness, frequency avoidance, lightweight and has excellent vibration isolation performance can be completely closed-loop verified in virtual space. This greatly reduces the number of physical iterations of "design-prototype-test", and reduces the original monthly and costly development cycle and expenses to the minimum. (4) Provides multi-dimensional and comprehensive safety and performance evaluation capabilities: The method not only focuses on the vibration isolation rate, but also simultaneously completes the comprehensive verification of the static strength, dynamic strength, total stress, fatigue limit, frequency avoidance and dynamic equivalence of the vibration isolator model. In addition, by comparing and analyzing the acceleration power spectral density attenuation of key excitation frequency bands (such as the blade passing frequency and its harmonics), the isolation effect of the vibration isolator on specific discrete frequency components can be quantitatively evaluated, providing refined data support for scheme optimization and comprehensively ensuring the structural safety and equipment reliability of the aircraft. (5) Improve design quality and engineering efficiency, and have universal application value: This method organically integrates structural design, dynamic analysis and vibration isolation performance prediction to form a complete and rigorous digital design-verification specification. It is not only applicable to tiltrotor aircraft arm systems, but its core ideas and processes can also be extended to other aviation, aerospace and high-end equipment fields with severe vibration environments and high requirements for vibration isolation. It provides an innovative technical path and reliable engineering practice method for achieving efficient and low-cost advanced equipment research and development. Attached Figure Description

[0017] Figure 1 This is a logic block diagram of the design and verification method for vibration isolation scheme of boom rotor system according to an embodiment of the present invention; Figure 2 This is a simulation diagram of the first platform in an embodiment of the present invention, which includes a vibration isolator model; Figure 3 This is a simulation diagram of the second platform in an embodiment of the present invention, which is a rigid connection model; Figure 4 This is a logic block diagram of a bracket design that meets the installation conditions of a vibration isolator according to an embodiment of the present invention; Figure 5 This is a logic block diagram for virtual verification of the system-level vibration isolation performance based on a qualified bracket, according to an embodiment of the present invention. Detailed Implementation

[0018] Please refer to Figures 1-5 This embodiment is a design and verification method for a vibration isolation scheme of an airframe rotor system, which includes the following steps: Phase 1: Design of brackets that meet the installation requirements of vibration isolators, including the following steps: Step S1: Construct a finite element model of the support structure for the arm-rotor system and apply a six-force static load to the arm-rotor system. Step S2: With the goal of minimizing structural flexibility, perform topology optimization on the support under volume fraction constraints, and reconstruct the geometric model of the support based on the optimization results; Step S3: Perform static and modal analysis on the geometric model of the reconstructed support to verify its static stress σ_ sIt is less than the material fatigue limit, and its first natural frequency avoids the preset excitation frequency band of the vibration source; Step S4: At the installation interface connecting the bracket and the vibration isolator, perform origin dynamic stiffness analysis, apply excitation force to perform frequency response simulation, calculate the dynamic stiffness curve of the installation interface in the preset working frequency band, and determine whether the minimum dynamic stiffness reaches the preset dynamic stiffness threshold. If it does not reach the threshold, return to step S1 to modify the design, and repeat steps S1 to S4 until the minimum dynamic stiffness of the bracket is greater than the dynamic stiffness threshold. Phase Two: Virtual verification of system-level vibration isolation performance based on qualified supports, including the following steps: Step S5: Based on the support model verified in step S4, construct two parallel virtual vibration test platforms. The first platform is a model with vibration isolators, and the second platform is a rigid connection model. Step S6: Under the same boundary conditions and vibration load spectrum excitation simulating the real environment, perform dynamic simulations on the first platform and the second platform respectively; Step S7: Extract the frequency domain data of the vibration response of the mass point at the same position from the simulation results of the first platform and the second platform respectively; Step S8: Based on the extracted response frequency domain data, calculate the root mean square value (RMS) of the acceleration of the first platform. i The root mean square value of acceleration of the second platform (RMS(A)) j ); Step S9: According to the formula: Vibration isolation ratio = [1 - RMS(A)] i ) / RMS(A j [ ] × 100%, the overall vibration isolation rate of the computer arm rotor system, to quantitatively evaluate the effectiveness of the vibration isolation scheme of the arm rotor system.

[0019] Each step is explained in detail below.

[0020] Specifically, for step S1, a finite element model of the arm-rotor system support is constructed, and six static loads on the arm-rotor system are applied. A 3D CAD software (such as CATIA) is used to establish an initial conceptual model of the support, including material constitutive modeling, property assignment, boundary condition simulation, and rotor system mass characteristic simulation. For example, the support material is a titanium alloy (e.g., TC4), with a density of 440 kg / m³. 3With a Poisson's ratio of 0.30 and an elastic modulus of 91 GPa, this model is a solid block that meets the basic installation space requirements. After importing it into finite element software (such as ANSYS Workbench) and meshing it, a MASS21 mass point is created at the center of the upper surface of the support; for example, the mass attribute is set to 4.3. kg, and assign it to represent the rotational inertia of the rotor system. The MASS21 mass point is rigidly coupled to the mounting area on the upper surface of the support to transfer the load. All six static loads are applied to the MASS21 mass point, and all degrees of freedom of the four bolt holes at the bottom of the support connected to the arm are constrained. The six static loads include tension, drag, lateral force, pitch moment, roll moment and anti-torsional moment. In this embodiment, the drag is 100N, the lateral force is -1500N, the tension is 10000N, the roll moment is 2000N·m, the pitch moment is 1500N·m and the anti-torsional moment is -1200N·m. In other embodiments, the magnitude of the six forces can be set according to the actual situation, and the magnitude of the six forces is not limited here.

[0021] Specifically, for step S2, with the goal of minimizing structural flexibility, topology optimization is performed on the support under volume fraction constraints, and the geometric model of the support is reconstructed based on the optimization results. In the structural optimization module of the finite element software, the support entity (excluding non-design areas such as bolt holes) is set as the design area, the optimization objective is set as minimizing structural flexibility, i.e. maximizing stiffness, and the constraint condition is that the material volume fraction is ≤ 30%. The calculation is submitted for topology optimization, and the optimization results are displayed in the form of a material density cloud map, clearly indicating the optimal material distribution path. Material is retained in high stress transmission paths, and material is removed in low stress areas. Based on this cloud map, geometric modeling is re-performed in CAD software to generate a smooth and manufacturable new support model, which has initially possessed the optimal force transmission configuration.

[0022] Specifically, for step S3, static and modal analyses are performed on the geometric model of the reconstructed support to verify its static stress σ_ s The stress must be less than the material fatigue limit, and its first natural frequency must avoid the preset excitation frequency band of the vibration source. The first natural frequency and the excitation frequency band of the vibration source must maintain a distance of at least 5%. For the reconstructed support model, perform finite element modeling (mesh refinement) again, and apply the same static load for linear static analysis. For example, calculate the maximum equivalent static stress σ_ of the support. sThe static strength is 200 MPa, which is less than the fatigue limit of TC4 (315 MPa), thus meeting the requirements. Next, modal analysis was performed to extract the first few natural frequencies and mode shapes. For example, the first bending mode frequency was obtained as f1 = 226 Hz. The relationship between this frequency f1 and the input key vibration source frequency bands (95-105 Hz, 190-210 Hz, 285-315 Hz) was checked: the interval between 226 Hz and the nearest 210 Hz is approximately 7.6%, which is greater than the 5% requirement, successfully achieving frequency avoidance.

[0023] Specifically, for step S4, at the installation interface connecting the bracket and the vibration isolator, a dynamic stiffness analysis is performed at the origin. For example, a unit excitation force is applied to perform frequency response simulation, the dynamic stiffness curve of the installation interface within the preset operating frequency band is calculated, and it is determined whether the minimum dynamic stiffness reaches the preset dynamic stiffness threshold. If it does not reach the threshold, the design is modified by returning to step S1, and steps S1 to S4 are repeated until the minimum dynamic stiffness of the bracket is greater than the dynamic stiffness threshold. Specifically, in step S4, the preset operating frequency band is 15~2000Hz; the dynamic stiffness threshold is 1000 N / mm. Local coupling nodes are created at the center nodes of the four bolt holes at the bottom of the bracket, and the degrees of freedom of these four coupling nodes are coupled to a master node, which represents the entire installation interface. At this master node, a simple harmonic force with an amplitude of 1 N is applied along the vertical (Z-direction), where the formula for the simple harmonic force is: F = 1 × sin(ωt), where ω is the angular velocity of the force change and t is the time of force application. A frequency sweep range of 15~2000Hz is set for harmonic response analysis. After solving, the Z-direction displacement frequency response function H_ of the master node is extracted. Z (ω) (unit: m / N), the origin dynamic stiffness IPI is the reciprocal of the frequency response function of this displacement, origin dynamic stiffness IPI = 1 / |H_z(ω)|. Plot the dynamic stiffness curve in the 15-2000Hz frequency band. The curve shows, for example, that the minimum dynamic stiffness in the entire frequency band is IPI_ dmin = 2000 N / mm, occurring at approximately 180 Hz, with a minimum dynamic stiffness of IPI_ dmin = 2000 N / mm is much greater than the preset threshold of 1000 N / mm. Therefore, the support design passes the dynamic stiffness verification and is judged to be a qualified support. Moreover, the weight calculation of the support is 3.5 kg, which meets the requirement of ≤4.0 kg, and it can proceed to the second stage. If IPI_ dminIf the dynamic stiffness IPI is less than 1000 N / mm, it indicates that the support is too flexible in the vicinity of that frequency point. It is necessary to return to step S1 and strengthen the stiffness by increasing the cross-sectional size and adjusting the stiffener layout. Then repeat steps S2-S4 until the dynamic stiffness IPI of the support meets the standard. In order to design a qualified support that is resistant to multiple frequency bands, has a high g value, and is lightweight, the qualified support has good structural dynamic characteristics and meets the application requirements.

[0024] Specifically, for step S5, based on the support model verified in step S4, two parallel virtual vibration test platforms are constructed. The first platform is a model containing vibration isolators, and the second platform is a rigid connection model. Specifically, in step S5, the first platform, the vibration isolator model, establishes elastic connection units simulating the characteristics of vibration isolators between the support and the mass points of the simulated boom rotor system. The second platform, the rigid connection model, replaces the elastic connection units with connection units simulating rigid connections, while maintaining the same settings as the first platform. Specifically, the elastic connection units are bushing elements, with parameters set according to the triaxial stiffness and damping of the actual vibration isolator; the connection units simulating rigid connections are spring elements with extremely high stiffness or constraint elements with directly coupled degrees of freedom. Specifically, the first platform is a model including a vibration isolator: The support model and mass point are retained. A rushing element is inserted between the mass point and the mounting area on the upper surface of the support. The properties of the rushing element include stiffness values ​​in three translational directions and the damping coefficient calculated based on the damping ratio. The stiffness values ​​in the three translational directions are as follows: X-direction: 308 N / mm, Y-direction: 308 N / mm, Z-direction: 294 N / mm, with a damping ratio of 0.3. The second platform is a rigid connection model: All settings of the first platform are copied, except that the rushing element is replaced with Rb2. The stiffness of Rb2 is approximately infinite to simulate the state of a direct rigid connection between the motor and the support. The boundary conditions of the first and second platforms are completely identical, meaning the constraints of the main nodes at the bottom of the support are completely identical.

[0025] Specifically, for step S6, under the same boundary conditions and vibration load spectrum excitation simulating a real environment, dynamic simulations are performed on the first and second platforms respectively. In step S6, the vibration load spectrum is an acceleration power spectral density curve superimposed with broadband and narrowband random signals. This broadband can reflect the vibration characteristics of a specific aircraft. Identical basic excitations are applied to the main nodes at the bottom of the supports of both platforms. This excitation is defined in the form of an acceleration power spectral density (PSD) curve, whose magnitude and shape precisely correspond to the input nine times the gravitational acceleration (9g RMS) and the "broadband + narrowband" random vibration environment spectrum. Random vibration analysis is performed on the first and second platforms respectively. After the simulation in step S6, step S10 is also included: extracting the dynamic stress σ_ of the support from the simulation results of the first and second platforms respectively. dFirst, confirm that its dynamic strength is less than the material's evaluation limit, and verify whether the total stress σ of the support in the first platform is less than the material's fatigue limit, in order to evaluate the total strength of the support. The total stress σ of the support is the static stress σ_ s With dynamic stress σ_ d The sum of these, i.e., the total stress σ = σ_ s +σ_ d For example, from the random vibration analysis results of the first platform, the maximum dynamic stress σ_ of the support can be read. d = 40MPa, then the total stress σ = 200 + 40 = 240MPa. This value is much less than the material fatigue limit of 315 MPa. Therefore, the total strength of the support under vibration isolation condition fully meets the requirements and the safety margin is sufficient.

[0026] Specifically, in step S5, when constructing the first platform, it also includes pre-calculating the vertical single-degree-of-freedom theoretical natural frequency of the arm rotor system, and comparing the pre-calculation result with the vertical natural frequency of the system extracted by simulation in step S6. The deviation between the two should be less than 10%. If the deviation exceeds 10%, the parameters of the elastic connection unit or the connection relationship of the overall model should be calibrated to ensure the dynamic equivalence of the virtual model.

[0027] Specifically, for step S7, extract the vibration response frequency domain data of the mass point at the same position from the simulation results of the first platform and the second platform respectively; after the simulation is completed, extract the Z-axis acceleration frequency domain data of the MASS21 mass point position from the analysis results of the two platforms respectively. Each frequency domain data should contain enough data points to ensure statistical significance.

[0028] Specifically, for step S8, based on the extracted response frequency domain data, the root mean square value (RMS) of the acceleration of the first platform is calculated. i The root mean square value of acceleration of the second platform (RMS(A)) j For example, the RMS response acceleration of the first platform (including vibration isolators): RMS(A) i = 2.15 g; Second platform (rigid connection) response acceleration RMS: RMS(A) j ) = 16.82 g.

[0029] Specifically, regarding step S9, according to the formula: Vibration isolation ratio β = [1 - RMS(A)] i ) / RMS(A jThe overall vibration isolation rate of the computer arm rotor system is calculated as β = [1 - 2.15 / 16.82] × 100% to quantitatively evaluate the effectiveness of the vibration isolation scheme. Substituting the above results into the formula: Vibration isolation rate β = [1 - 2.15 / 16.82] × 100% ≈ 87.2%, this result clearly shows that, based on the currently designed qualified support and selected vibration isolators, the entire arm rotor system can achieve a vibration isolation effect of approximately 87.2%, exceeding the design target of 80%, indicating that the scheme is effective.

[0030] Specifically, after calculating the overall vibration isolation rate in step S9, step S11 is also included: based on the vibration response frequency domain data extracted in step S7, the acceleration power spectral density curves of the first platform and the second platform are generated respectively; the peak value of the acceleration power spectral density of the first platform is compared with the attenuation of the second platform in the preset narrow band of vibration source excitation, so as to evaluate the isolation effect of the vibration isolator on specific discrete frequency vibration components. The acceleration time history data extracted in step S7 were subjected to Fourier transform to generate acceleration power spectral density (PSD) curves for the first and second platforms. Comparing the two curves, the focus was on the three narrow frequency bands of 95-105 Hz, 190-210 Hz, and 285-315 Hz. Analysis revealed that on the PSD curve of the rigid connection model of the second platform, there was a sharp peak of 0.15 g² / Hz near 100 Hz. However, on the PSD curve of the model with the vibration isolator on the first platform, the peak value of this frequency band had decayed to below 0.01 g² / Hz, with a decay of more than 20 dB. This quantitatively proves that the vibration isolator has an excellent isolation effect on the vibration component of the blade passing frequency (1P). Correspondingly, the analysis of other harmonics (2P, 3P) also showed significant peak decay.

[0031] As can be seen from the detailed steps demonstrated in the above embodiments, the method of the present invention successfully completed a complete design and verification closed loop in the digital space. It not only designed a lightweight, high dynamic stiffness qualified bracket, but also accurately predicted the system vibration isolation rate of up to 87.2% in advance through innovative virtual comparative tests, and verified the structural safety. The entire process does not require the manufacture of any physical prototype, achieving efficient and low-cost research and development, and has significant engineering practical value.

[0032] The above descriptions are merely some embodiments of the present invention. For those skilled in the art, various modifications and improvements can be made without departing from the inventive concept of the present invention, and these all fall within the protection scope of the present invention.

Claims

1. A design and verification method for a vibration isolation scheme of an airframe rotor system, characterized in that: It includes the following steps: Phase 1: Design of brackets that meet the installation requirements of vibration isolators, including the following steps: Step S1: Construct a finite element model of the support structure for the arm-rotor system and apply a six-force static load to the arm-rotor system. Step S2: With the goal of minimizing structural flexibility, perform topology optimization on the support under volume fraction constraints, and reconstruct the geometric model of the support based on the optimization results; Step S3: Perform static and modal analysis on the geometric model of the reconstructed support to verify its static stress σ_ s It is less than the material fatigue limit, and its first natural frequency avoids the preset excitation frequency band of the vibration source; Step S4: At the installation interface connecting the bracket and the vibration isolator, perform origin dynamic stiffness analysis, apply excitation force to perform frequency response simulation, calculate the dynamic stiffness curve of the installation interface in the preset working frequency band, and determine whether the minimum dynamic stiffness reaches the preset dynamic stiffness threshold. If it does not reach the threshold, return to step S1 to modify the design, and repeat steps S1 to S4 until the minimum dynamic stiffness of the bracket is greater than the dynamic stiffness threshold. Phase Two: Virtual verification of system-level vibration isolation performance based on qualified supports, including the following steps: Step S5: Based on the support model that has been verified and approved in step S4, construct two parallel virtual vibration test platforms, namely the first platform and the second platform; Step S6: Under the same boundary conditions and vibration load spectrum excitation simulating the real environment, perform dynamic simulations on the first platform and the second platform respectively; Step S7: Extract the frequency domain data of the vibration response of the mass point at the same position from the simulation results of the first platform and the second platform respectively; Step S8: Based on the extracted response frequency domain data, calculate the root mean square value (RMS) of the acceleration of the first platform. i The root mean square value of acceleration of the second platform (RMS(A)) j ); Step S9: According to the formula: Vibration isolation ratio β = [1 - RMS(A)] i ) / RMS(A j The overall vibration isolation rate of the computer arm rotor system is calculated as 100% to evaluate the effectiveness of the vibration isolation scheme for the arm rotor system.

2. The design and verification method for a vibration isolation scheme of an airframe rotor system as described in claim 1, characterized in that: In step S3, the first-order natural frequency and the excitation frequency band of the vibration source must maintain an interval of at least 5%.

3. The design and verification method for a vibration isolation scheme of an airframe rotor system as described in claim 1, characterized in that: In step S4, the preset operating frequency band is 15~2000Hz; the dynamic stiffness threshold is 1000 N / mm.

4. The design and verification method for a vibration isolation scheme of an airframe rotor system as described in claim 1, characterized in that: In step S5, the first platform is a model with a vibration isolator, and an elastic connection unit simulating the vibration isolator characteristics is established between the support and the mass point of the simulated arm rotor system; the second platform is a rigid connection model, and the elastic connection unit is replaced with a connection unit simulating a rigid connection, while the rest is consistent with the settings of the first platform.

5. The design and verification method for a vibration isolation scheme of an airframe rotor system as described in claim 4, characterized in that: The elastic connection unit is a Busing unit, and its parameters are set according to the triaxial stiffness and damping of the actual vibration isolator.

6. The design and verification method for a vibration isolation scheme of an airframe rotor system as described in claim 4, characterized in that: The connection elements for simulating rigid connections are spring elements or constraint elements with directly coupled degrees of freedom.

7. The design and verification method for a vibration isolation scheme of an airframe rotor system as described in claim 1, characterized in that: In step S6, the vibration load spectrum is an acceleration power spectral density curve that is a superposition of broadband random and narrowband random.

8. The design and verification method for a vibration isolation scheme of an airframe rotor system as described in claim 1, characterized in that: Following the simulation in step S6, step S10 is also included: extracting the dynamic stress σ_ of the support from the simulation results of the first platform and the second platform, respectively. d The total stress σ of the support in the first platform is verified to be less than the fatigue limit of the material in order to evaluate the total strength of the support. The total stress σ of the support is the static stress σ_ s With dynamic stress σ_ d The sum of these, i.e., the total stress σ = σ_ s +σ_ d .

9. The design and verification method for a vibration isolation scheme of an airframe rotor system as described in claim 1, characterized in that: In step S5, when constructing the first platform, the theoretical natural frequency of the vertical single degree of freedom of the arm rotor system is pre-calculated. The pre-calculated result is compared with the vertical natural frequency of the system extracted by simulation in step S6 for verification. The deviation between the two should be less than 10%. If the deviation exceeds 10%, the parameters of the elastic connection unit or the connection relationship of the overall model should be calibrated to ensure the dynamic equivalence of the virtual model.

10. The design and verification method for a vibration isolation scheme of an airframe rotor system as described in claim 1, characterized in that: After calculating the overall vibration isolation rate in step S9, step S11 is also included: based on the vibration response frequency domain data extracted in step S7, the acceleration power spectral density curves of the first platform and the second platform are generated respectively; the peak value of the acceleration power spectral density of the first platform is compared with the attenuation of the second platform in the preset narrow band of vibration source excitation, so as to evaluate the isolation effect of the vibration isolator on specific discrete frequency vibration components.

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