Instrument damping control method and system based on double resonance

Abstract

The present application belongs to the technical field of vibration control, and relates to an instrument damping control method and system based on double resonance, which comprises the following steps: determining the resonance frequency of an instrument based on a finite element model of a pipeline system connected to the instrument; adding a resonance structure to the pipeline of the finite element model to establish a one-tuned model; performing harmonic response analysis on the one-tuned model to obtain the peak frequency of the one-tuned model, and adjusting the resonance structure parameters to optimize the one-tuned model when the peak frequency is inconsistent with the resonance frequency; installing a tuned mass damper on the resonance structure of the optimized one-tuned model to establish a two-tuned model; performing harmonic response analysis on the two-tuned model to obtain the valley frequency of the two-tuned model, and adjusting the tuned mass damper parameters to optimize the two-tuned model when the valley frequency is inconsistent with the resonance frequency. The present application efficiently suppresses the influence of vibration at a specific frequency on the instrument, and guarantees the stable performance and measurement accuracy of the instrument.

Description

Technical Field

[0001] The present invention belongs to the technical field of vibration control, and relates to structural vibration control technology. Specifically, it relates to a vibration damping control method and system for instruments based on double resonance. Background Technique

[0002] In fields such as chemical industry and energy, the pipeline system, as the core engineering structure for transporting flowing media such as liquids and gases, its operating stability and the metering accuracy of the media are directly related to the controllability of the production process, the accuracy of energy consumption monitoring, and the reliability of safe production. To achieve precise metering of the media flow rate in the pipeline, instruments for metering flow rate (such as Coriolis mass flow meters, vortex flow meters, ultrasonic flow meters, etc.) are widely deployed in the pipeline system.

[0003] The metering principle of the flow metering instrument highly depends on the capture and recognition of specific physical signals by the core sensing element or metering cavity, and such core components often show high sensitivity to external vibrations of specific frequencies. For example: The sensor of the vortex flow meter realizes flow rate conversion by detecting the vortex street frequency vibration generated by the fluid flow. A weak external vibration interference can cover up the true vortex street signal. The Coriolis mass flow meter is based on the Coriolis effect, and realizes mass flow measurement by driving the detection tube to generate vibrations of a specific frequency and monitoring its phase change. When the external vibration frequency is close to the driving frequency of the detection tube, it is extremely easy to cause resonance, which will in turn cause a significant increase in the measurement error, and in severe cases, even cause the metering data to fail.

[0004] In the on-site working conditions of the chemical and energy industries, the sources of vibration interference are extensive and the spectral characteristics are complex. For example, the vibration spectrum range generated during the operation of the pump body supporting the pipeline system is 0 - 2000Hz, and the maximum vibration acceleration is up to 4m / s²; the vibration frequency range of the motor equipment is 0 - 1000Hz, and the maximum vibration acceleration can also reach 1m / s². When the vibration frequencies of these mechanical equipment fall into the sensitive frequency range of the flow metering instrument (such as the driving frequency of the Coriolis flow meter, the vortex street characteristic frequency of the vortex flow meter), it will directly damage the normal working state of the instrument, cause a serious decline in the measurement accuracy, and cannot meet the high-precision requirements of industrial production for flow metering. To solve the problem of interference of external vibrations on the flow metering instrument, many vibration damping and protection schemes have been proposed in the prior art.

[0005] One type of solution focuses on the structural protection design of the instrument itself. For example, patent publication number CN207703281U discloses a vibration-resistant and interference-resistant Coriolis mass flow meter, which uses a multi-layer buffer structure with three protective sleeves to offset the transmission of external vibrations and reduce the impact of vibrations on the core components inside the instrument. Patent application publication number CN116678461A discloses a vibration-resistant mass flow meter, which strengthens the vibration isolation effect between the instrument and the pipeline by configuring a control base with vibration isolation function and combining it with a rigid pipeline connection isolation mechanism, thereby improving the operational stability of the flow meter in a vibration environment.

[0006] Another mainstream approach employs tuned mass dampers (TMDs), a typical passive vibration control device. By tuning the damper's natural frequency, it achieves efficient absorption of vibrations at specific frequencies. Related research shows that multi-tuned mass damper (MTMD) systems exhibit superior vibration reduction performance compared to single-tuned systems. For example, the paper "Research on Vibration Reduction of Pipeline Vibration Overtone Response Based on MTMD" uses multiple fixed-frequency TMDs to construct an MTMD system, achieving good vibration reduction effects in multiple resonant frequency bands of large pipelines in chemical enterprises. The paper "Analysis and Vibration Reduction Research of Abnormal Vibration in Pulleys of Ultra-High Voltage Converter Stations" reduces pipeline vibration acceleration by more than 90% by deploying TMDs at multiple locations in the lubricating oil supply pipeline. The paper "Using multiple tuned mass dampers to control offshore wind turbine vibrations under multiple hazards" further confirms that multiple TMDs can effectively reduce the displacement response of offshore wind turbines under multiple external excitations, significantly improving the system's robustness and engineering applicability.

[0007] However, existing vibration reduction and protection technologies all have significant limitations: whether it's multi-layer protective sleeves or vibration isolation base designs for the instrument body, or TMD / MTMD vibration reduction systems deployed on pipelines, they all adopt a direct installation method on the protected main body or adjacent pipelines. When vibration reduction measures act indirectly on flow metering instruments, the vibration energy is difficult to be efficiently absorbed by the vibration reduction device during the transmission process from the pipeline to the instrument. Especially for narrowband vibration interference near the instrument's operating frequency, the energy dissipation efficiency of existing indirect vibration reduction methods is greatly reduced, and the vibration reduction effect is significantly weakened. Summary of the Invention

[0008] This invention addresses the problems existing in the prior art by providing an instrument vibration reduction control method and system based on dual resonance. By introducing a resonant structure and a tuned mass damper, and optimizing the vibration reduction control based on dual resonance, the method can improve the external vibration reduction effect at the instrument's resonant frequency, ensuring the instrument's performance stability and measurement accuracy.

[0009] In a first aspect, the present invention provides an instrument vibration reduction control method based on dual resonance, the steps of which include:

[0010] S1. Establish a finite element model of the pipeline system, treat the finite element model as a non-vibration damping system model, and determine the resonant frequency of the instrument based on the finite element model;

[0011] S2. Add a resonant structure to the pipe in the finite element model, establish a first-tuned model, and set the initial stiffness and initial mass of the resonant structure.

[0012] S3. Perform harmonic response analysis on the first-tuned model to obtain the peak frequency of the first-tuned model;

[0013] S4. Determine whether the peak frequency of the first-tuned model is consistent with the resonant frequency of the instrument.

[0014] S5. When the peak frequency of the first-tuned model is inconsistent with the resonant frequency of the instrument, adjust the stiffness and mass of the resonant structure to optimize the first-tuned model so that the peak frequency of the first-tuned model is consistent with the resonant frequency of the instrument.

[0015] S6. Install a tuned mass damper on the resonant structure of the optimized single-tuned model to establish a double-tuned model. Set the mass ratio of the tuned mass damper to the resonant structure, as well as the mass, initial stiffness, and initial damping of the tuned mass damper.

[0016] S7. Perform harmonic response analysis on the double-tuned model to obtain the trough frequency of the double-tuned model;

[0017] S8. Determine whether the trough frequency of the double-tuned model is consistent with the resonant frequency of the instrument.

[0018] S9. When the trough frequency of the double-tuned model is inconsistent with the resonant frequency of the instrument, adjust the stiffness and damping of the tuning mass damper to optimize the double-tuned model so that the trough frequency of the double-tuned model is consistent with the resonant frequency of the instrument.

[0019] In some embodiments of the present invention, in step S1, the method for determining the resonant frequency of the instrument based on the finite element model is as follows: performing harmonic response analysis on the finite element model to obtain the amplitude-frequency characteristic curve of the finite element model, and taking the frequency corresponding to the peak value of the amplitude-frequency characteristic curve of the finite element model as the resonant frequency of the instrument.

[0020] In some embodiments of the present invention, the method for optimizing a single-tuned model in step S5 is as follows:

[0021] S51. By adjusting the length and wall thickness of the resonant structure, the stiffness and mass of the resonant structure are adjusted to obtain the optimized single-tuned model.

[0022] S52. Perform harmonic response analysis on the optimized single-tuned model to obtain the peak frequency of the optimized single-tuned model.

[0023] S53. Determine whether the peak frequency of the optimized single-tuned model is consistent with the resonant frequency of the instrument.

[0024] S54. If the peak frequency of the optimized first-tuned model is inconsistent with the resonant frequency of the instrument, repeat steps S51 to S53; if the peak frequency of the optimized first-tuned model is consistent with the resonant frequency of the instrument, end the optimization process.

[0025] In some embodiments of the present invention, the method for optimizing the double-tuning model in step S9 is as follows:

[0026] S91. Adjust the mass of the tuned mass damper by adjusting the mass ratio of the tuned mass damper to the resonant structure, and adjust the stiffness and damping of the tuned mass damper by adjusting the spring stiffness and damper damping of the tuned mass damper to obtain the optimized double-tuned model.

[0027] S92. Perform harmonic response analysis on the optimized double-tuned model to obtain the trough frequency of the optimized double-tuned model;

[0028] S93. Determine whether the trough frequency of the optimized double-tuned model is consistent with the resonant frequency of the instrument.

[0029] S94. If the trough frequency of the optimized double-tuned model is inconsistent with the resonant frequency of the instrument, repeat steps S91 to S93; if the trough frequency of the optimized double-tuned model is consistent with the resonant frequency of the instrument, end the optimization process.

[0030] In some embodiments of the present invention, the finite element model of the pipeline system includes:

[0031] Connecting pipes;

[0032] A first metal pipe is connected to the connecting pipe via a first standard flange;

[0033] The instrument is connected to the first metal pipe on one side via a second standard flange, and to the third standard flange on the other side.

[0034] The second metal pipe is connected at one end to the third standard flange and at the other end to the fourth standard flange.

[0035] In some embodiments of the present invention, the single-tuning model includes:

[0036] Connecting pipes;

[0037] A first metal pipe is connected to the connecting pipe via a first standard flange;

[0038] The first resonant structure has one side connected to the first metal pipe via the fifth standard flange and the other side connected to the sixth standard flange.

[0039] The third metal pipe has one end connected to the sixth standard flange and the other end connected to the second standard flange;

[0040] The instrument is connected to a third metal pipe on one side via a second standard flange, and to a third standard flange on the other side;

[0041] A third metal pipe, one end of which is connected to the third standard flange, and the other end of which is connected to the fourth standard flange;

[0042] The second resonant structure is connected to the fourth standard flange on one side and to the seventh standard flange on the other side.

[0043] The fourth metal pipe is connected at one end to the seventh standard flange and at the other end to the eighth standard flange.

[0044] In some embodiments of the present invention, the dual-tuning model includes:

[0045] Connecting pipes;

[0046] A first metal pipe is connected to the connecting pipe via a first standard flange;

[0047] The first resonant structure has one side connected to the first metal pipe via the fifth standard flange and the other side connected to the sixth standard flange.

[0048] A first tuned mass damper is mounted on the first resonant structure;

[0049] The third metal pipe has one end connected to the sixth standard flange and the other end connected to the second standard flange;

[0050] The instrument is connected to a third metal pipe on one side via a second standard flange, and to a third standard flange on the other side;

[0051] The second metal pipe has one end connected to the third standard flange and the other end connected to the fourth standard flange.

[0052] The second resonant structure is connected to the fourth standard flange on one side and to the seventh standard flange on the other side.

[0053] A second tuned mass damper is mounted on the second resonant structure;

[0054] The fourth metal pipe is connected at one end to the seventh standard flange and at the other end to the eighth standard flange.

[0055] In a second aspect, the present invention provides an instrument vibration reduction control system for implementing the instrument vibration reduction control method based on dual resonance described in the first aspect of the present invention, the system comprising:

[0056] The model building module is used to build finite element models, single-tuned models, and double-tuned models of pipeline systems.

[0057] The setting module is used to set the initial stiffness and initial mass of the resonant structure, the mass ratio of the tuned mass damper to the resonant structure, and the mass, initial stiffness, and initial damping of the tuned mass damper.

[0058] The analysis module determines the resonant frequency of the instrument based on the finite element model, performs harmonic response analysis on the single-tuned model to obtain the peak frequency of the single-tuned model, and performs harmonic response analysis on the double-tuned model to obtain the trough frequency of the double-tuned model.

[0059] The judgment module is configured to: determine whether the peak frequency of the first-tuned model is consistent with the resonant frequency of the instrument, and determine whether the trough frequency of the second-tuned model is consistent with the resonant frequency of the instrument.

[0060] The optimization module is configured to: when the peak frequency of the first-tuned model is inconsistent with the resonant frequency of the instrument, adjust the stiffness and mass of the resonant structure to optimize the first-tuned model so that the peak frequency of the first-tuned model is consistent with the resonant frequency of the instrument; when the trough frequency of the second-tuned model is inconsistent with the resonant frequency of the instrument, adjust the stiffness and damping of the tuning mass damper to optimize the second-tuned model so that the trough frequency of the second-tuned model is consistent with the resonant frequency of the instrument.

[0061] Compared with the prior art, the advantages and positive effects of the present invention are as follows:

[0062] (1) The instrument vibration reduction control method and system based on dual resonance provided by this invention firstly determines the resonant frequency of the instrument. Secondly, a resonant structure is introduced into the pipeline of the pipeline system to construct a first-order tuning model, and the peak frequency of the first-order tuning model is made consistent with the resonant frequency of the instrument through optimization of the resonant structure parameters, thus completing the first-order tuning. Thirdly, a tuned mass damper is installed on the tuning structure of the optimized first-order tuning model, and the trough frequency of the second-order tuning model is made consistent with the resonant frequency of the instrument through optimization of the tuned mass damper parameters, thus completing the second-order tuning. This invention performs vibration reduction control processing based on dual resonance optimization, so that the vibration energy at the resonant frequency of the instrument is effectively absorbed and dissipated by the tuned mass damper, thereby significantly suppressing the vibration transmitted to the instrument, improving the external vibration reduction effect at the resonant frequency of the instrument, and ensuring the performance stability and measurement accuracy of the instrument.

[0063] (2) The instrument vibration reduction control method and system based on dual resonance provided by the present invention has good versatility and can be applied to different types of instruments. Only the resonant frequency of the instrument needs to be determined to complete the parameter design of the resonant mechanism and the tuned mass damper, so as to achieve effective vibration protection control. Attached Figure Description

[0064] Figure 1 This is a schematic flowchart of the instrument vibration reduction control method based on dual resonance as described in an embodiment of the present invention;

[0065] Figure 2 This is a schematic diagram of the finite element model of the pipeline system described in an embodiment of the present invention;

[0066] Figure 3 This is a schematic diagram of the amplitude-frequency characteristic curve of the finite element model described in the embodiment of the present invention;

[0067] Figure 4 This is a schematic diagram of the structure of the single-tuning model described in an embodiment of the present invention;

[0068] Figure 5 This is a schematic diagram of the method for optimizing a single-tuned model according to an embodiment of the present invention;

[0069] Figure 6 This is a schematic diagram of the amplitude-frequency response curve of the single-tuned model before optimization in this embodiment of the invention;

[0070] Figure 7 This is a schematic diagram of the amplitude-frequency response curve corresponding to the optimized single-tuned model in this embodiment of the invention;

[0071] Figure 8 This is a schematic diagram of the structure of the dual-tuning model described in an embodiment of the present invention;

[0072] Figure 9This is a schematic diagram of the method for optimizing the double-tuning model according to an embodiment of the present invention;

[0073] Figure 10 This is a schematic diagram of the amplitude-frequency response curve of the double-tuned model before optimization in this embodiment of the invention;

[0074] Figure 11 This is a schematic diagram of the amplitude-frequency response curve corresponding to the optimized double-tuned model in this embodiment of the invention;

[0075] Figure 12 This is a structural block diagram of the instrument vibration reduction control system described in an embodiment of the present invention;

[0076] Figure 13 This is a schematic diagram of the finite element model of the pipeline system connected to the Coriolis mass flow meter in Embodiment 1 of the present invention;

[0077] Figure 14 This is a schematic diagram of the finite element model structure of the instrument vibration reduction control method and system based on dual resonance proposed in Embodiment 1 of the present invention.

[0078] Figure 15 This is a schematic diagram of the structure of a traditional vibration reduction system model in Embodiment 1 of the present invention;

[0079] Figure 16 This is a schematic diagram showing the results of harmonic response analysis of a non-damping system model under a displacement excitation of 0.1 mm in Embodiment 1 of the present invention.

[0080] Figure 17 This is a schematic diagram showing the results of harmonic response analysis of a single-tuned model in Embodiment 1 of the present invention;

[0081] Figure 18 This is a schematic diagram showing the results of harmonic response analysis of the optimized single-tuned model in Embodiment 1 of the present invention;

[0082] Figure 19 This is a schematic diagram showing the results of harmonic response analysis of the double-tuned model in Embodiment 1 of the present invention;

[0083] Figure 20 This is a schematic diagram showing the results of harmonic response analysis of the optimized double-tuned model in Embodiment 1 of the present invention;

[0084] Figure 21 This is a schematic diagram of the amplitude-frequency characteristic curve after double tuning at the flow meter monitoring point A in Embodiment 1 of the present invention;

[0085] Figure 22 This is a schematic diagram of the harmonic response results of the undamped system model at flowmeter monitoring point A in Embodiment 1 of the present invention;

[0086] Figure 23This is a schematic diagram of the harmonic response results of a traditional vibration reduction system model at flowmeter monitoring point A in Embodiment 1 of the present invention;

[0087] Figure 24 This is a schematic diagram showing the effect of different frequency ratios on the amplitude at monitoring point A of the Coriolis mass flow meter in Embodiment 1 of the present invention;

[0088] Figure 25 This is a schematic diagram of the time history curve of the harmonic excitation in Embodiment 2 of the present invention;

[0089] Figure 26 This is a schematic diagram of the steady-state vibration time history curve of monitoring point A under simple harmonic excitation in Embodiment 2 of the present invention without vibration damping system;

[0090] Figure 27 This is a schematic diagram of the steady-state vibration time history curve of monitoring point A under simple harmonic excitation using the traditional vibration reduction method in Embodiment 2 of the present invention;

[0091] Figure 28 This is a schematic diagram of the steady-state vibration time history curve of monitoring point A under simple harmonic excitation in Embodiment 2 of the present invention;

[0092] Figure 29 This is a schematic diagram of the time history curve of random excitation in Embodiment 2 of the present invention;

[0093] Figure 30 This is a schematic diagram of the steady-state vibration time history curve of monitoring point A under random excitation in Embodiment 2 of the present invention without vibration damping system;

[0094] Figure 31 This is a schematic diagram of the steady-state vibration time history curve of monitoring point A under random excitation in the conventional vibration reduction method of Embodiment 2 of the present invention;

[0095] Figure 32 This is a schematic diagram of the steady-state vibration time history curve of monitoring point A under random excitation in Embodiment 2 of the present invention.

[0096] In the diagram, 100 is the connecting pipe, 201 is the first metal pipe, 202 is the second metal pipe, 203 is the third metal pipe, 204 is the fourth metal pipe, 301 is the first standard flange, 302 is the second standard flange, 303 is the third standard flange, 304 is the fourth standard flange, 305 is the fifth standard flange, 306 is the sixth standard flange, 307 is the seventh standard flange, 308 is the eighth standard flange, 400 is the instrument, 401 is the Coriolis mass flow meter, 501 is the first resonant structure, 502 is the second resonant structure, 601 is the first tuned mass damper, 602 is the second tuned mass damper, 700 is the instrument vibration reduction control system, 701 is the model building module, 702 is the setting module, 703 is the analysis module, 704 is the judgment module, and 705 is the optimization module. Detailed Implementation

[0097] To make the technical problems to be solved, the technical solutions, and the beneficial effects of the present invention clearer, the present invention will be further described in detail below with reference to the accompanying drawings and embodiments. It should be understood that the specific embodiments described herein are merely illustrative of the present invention and are not intended to limit the present invention.

[0098] The prefixes such as "first" and "second" used in this application embodiment are merely for distinguishing different descriptive objects and do not limit the position, order, priority, quantity, or content of the described objects. The use of ordinal numbers and other prefixes used to distinguish descriptive objects in this application embodiment does not constitute a limitation on the described objects. The description of the described objects is given in the claims or the context of the embodiments, and should not constitute unnecessary restrictions due to the use of such prefixes. Furthermore, in the description of this embodiment, unless otherwise stated, "multiple" means two or more.

[0099] The technical solutions of the embodiments of this application will be described below with reference to the accompanying drawings. In the description of the embodiments of this application, unless otherwise stated, " / " means "or," for example, A / B can mean A or B; the term "and / or" in this document is merely a description of the relationship between related objects, indicating that three relationships can exist. For example, A and / or B can represent: A existing alone, A and B existing simultaneously, and B existing alone.

[0100] In the embodiments provided in this application, it should be understood that the disclosed systems and methods can be implemented in other ways. For example, the device embodiments described above are merely illustrative. For instance, the division of units is only a logical functional division, and in actual implementation, there may be other division methods. For example, multiple units or components may be combined or integrated into another system, or some features may be ignored or not executed. Furthermore, the coupling or direct coupling or communication connection shown or discussed may be through some interfaces; the indirect coupling or communication connection between devices or units may be electrical, mechanical, or other forms.

[0101] See Figure 1 The first aspect of this invention provides a method for instrument vibration reduction control based on dual resonance, the steps of which include:

[0102] S1. Establish a finite element model of the pipeline system, and use the finite element model as a model of a system without vibration reduction. Determine the resonant frequency of the instrument based on the finite element model.

[0103] In one embodiment of the present invention, see Figure 2 The finite element model of the pipeline system includes:

[0104] Connecting pipe 100;

[0105] The first metal pipe 201 is connected to the connecting pipe 100 via the first standard flange 301;

[0106] Instrument 400 is connected to the first metal pipe 201 on one side via the second standard flange 302, and to the third standard flange 303 on the other side;

[0107] The second metal pipe 202 is connected at one end to the third standard flange 303 and at the other end to the fourth standard flange 304.

[0108] In one embodiment of the present invention, the method for determining the resonant frequency of an instrument based on the finite element model is as follows: harmonic response analysis is performed on the finite element model to obtain the amplitude-frequency characteristic curve of the finite element model (see...). Figure 3 The frequency corresponding to the peak value of the amplitude-frequency characteristic curve of the finite element model is taken as the resonant frequency of the instrument.

[0109] By conducting harmonic response analysis on the finite element model and extracting the frequency corresponding to the peak value of the amplitude-frequency characteristic curve, the resonant frequency of the instrument under external vibration excitation can be directly located, overcoming the limitations of traditional physical test methods for resonant frequency testing, which have long cycles, high costs, and are greatly affected by on-site working conditions.

[0110] S2. Add a resonant structure to the pipeline in the finite element model, establish a first-tuned model, and set the initial stiffness and initial mass of the resonant structure.

[0111] In this embodiment of the invention, a resonant structure is added to the pipeline in the finite element model. Through the tuning effect of the resonant structure, the vibration response of the instrument at the resonant frequency is reduced.

[0112] In one embodiment of the present invention, see Figure 4 The single-tuning model includes:

[0113] Connecting pipe 100;

[0114] The first metal pipe 201 is connected to the connecting pipe 100 via the first standard flange 301;

[0115] The first resonant structure 501 is connected to the first metal pipe 201 on one side via the fifth standard flange 305, and to the sixth standard flange 306 on the other side.

[0116] The third metal pipe 203 is connected at one end to the sixth standard flange 306 and at the other end to the second standard flange 302;

[0117] Instrument 400 is connected to a third metal pipe 203 on one side via a second standard flange 302, and to a third standard flange 303 on the other side;

[0118] The second metal pipe 202 has one end connected to the third standard flange 303 and the other end connected to the fourth standard flange 304;

[0119] The second resonant structure 502 is connected to the fourth standard flange 304 on one side and to the seventh standard flange 307 on the other side.

[0120] The fourth metal pipe 204 is connected at one end to the seventh standard flange 307 and at the other end to the eighth standard flange 308.

[0121] S3. Perform harmonic response analysis on the first-tuned model to obtain the peak frequency of the first-tuned model.

[0122] S4. Determine whether the peak frequency of the first-tuned model is consistent with the resonant frequency of the instrument.

[0123] S5. When the peak frequency of the first-tuned model is inconsistent with the resonant frequency of the instrument, adjust the stiffness and mass of the resonant structure to optimize the first-tuned model so that the peak frequency of the first-tuned model is consistent with the resonant frequency of the instrument.

[0124] In one embodiment of the present invention, see Figure 5 The method for optimizing a single-tuned model is as follows:

[0125] S51. By adjusting the length and wall thickness of the resonant structure, the stiffness and mass of the resonant structure are adjusted to obtain the optimized single-tuned model.

[0126] S52. Perform harmonic response analysis on the optimized single-tuned model to obtain the peak frequency of the optimized single-tuned model.

[0127] S53. Determine whether the peak frequency of the optimized single-tuned model is consistent with the resonant frequency of the instrument.

[0128] S54. If the peak frequency of the optimized first-tuned model is inconsistent with the resonant frequency of the instrument, repeat steps S51 to S53; if the peak frequency of the optimized first-tuned model is consistent with the resonant frequency of the instrument, end the optimization process.

[0129] In this embodiment of the invention, on the one hand, by adjusting the length and wall thickness of the resonant structure to regulate its stiffness and mass, the natural frequency characteristics of the single-tuned model can be precisely altered. Combined with harmonic response analysis to extract the peak frequency and compare it with the instrument's resonant frequency for calibration, this ensures that the natural frequency of the single-tuned model is completely consistent with the instrument's resonant frequency. On the other hand, the parameter optimization method based on stiffness and mass regulation directly affects the core mechanical properties of the resonant structure. The optimized single-tuned model exhibits strong natural frequency stability and is less susceptible to changes in operating conditions such as temperature and pressure in chemical and energy environments. Multiple iterative calibrations ensure complete frequency consistency, avoiding the problem of vibration reduction efficiency attenuation caused by frequency deviation. This allows the tuned model to continuously and efficiently suppress external vibration transmission during long-term operation, guaranteeing the measurement accuracy and operational stability of instruments such as Coriolis mass flow meters and vortex flow meters.

[0130] Specifically, after introducing the resonant structure, harmonic response analysis was performed on the single-tuned model before and after optimization.

[0131] See the amplitude-frequency response curve of the unoptimized single-tuned model. Figure 6 At the instrument's resonant frequency, the amplitude-frequency response curve exhibits a double-peak characteristic, indicating a frequency matching deviation between the resonant structure's frequency and the instrument's resonant frequency. Further optimization of the resonant structure's stiffness and mass is needed. By adjusting the length and wall thickness of the resonant structure, its stiffness and mass can be adjusted, thereby optimizing the first-order tuning model.

[0132] The amplitude-frequency response curve corresponding to the optimized single-tuned model is shown in the figure. Figure 7 The two peaks before optimization have merged into a single resonance peak, located at the instrument's resonance frequency, indicating that the frequency of the resonant structure coincides with the instrument's resonance frequency, and their amplitude-frequency characteristics tend to be consistent.

[0133] S6. Install a tuned mass damper on the resonant structure of the optimized single-tuned model to establish a double-tuned model. Set the mass ratio of the tuned mass damper to the resonant structure, as well as the mass, initial stiffness, and initial damping of the tuned mass damper.

[0134] In one embodiment of the present invention, see Figure 8 The dual-tuning model includes:

[0135] Connecting pipe 100;

[0136] The first metal pipe 201 is connected to the connecting pipe 100 via the first standard flange 301;

[0137] The first resonant structure 501 is connected to the first metal pipe 201 on one side via the fifth standard flange 305, and to the sixth standard flange 306 on the other side.

[0138] The first tuned mass damper 601 is mounted on the first resonant structure 501;

[0139] The third metal pipe 203 is connected at one end to the sixth standard flange 306 and at the other end to the second standard flange 302;

[0140] Instrument 400 is connected to a third metal pipe 203 on one side via a second standard flange 302, and to a third standard flange 303 on the other side;

[0141] The second metal pipe 202 has one end connected to the third standard flange 303 and the other end connected to the fourth standard flange 304;

[0142] The second resonant structure 502 is connected to the fourth standard flange 304 on one side and to the seventh standard flange 307 on the other side.

[0143] The second tuned mass damper 602 is mounted on the second resonant structure 502;

[0144] The fourth metal pipe 204 is connected at one end to the seventh standard flange 307 and at the other end to the eighth standard flange 308.

[0145] S7. Perform harmonic response analysis on the double-tuned model to obtain the trough frequency of the double-tuned model.

[0146] S8. Determine whether the trough frequency of the double-tuned model is consistent with the resonant frequency of the instrument.

[0147] S9. When the trough frequency of the double-tuned model is inconsistent with the resonant frequency of the instrument, adjust the stiffness and damping of the tuning mass damper to optimize the double-tuned model so that the trough frequency of the double-tuned model is consistent with the resonant frequency of the instrument.

[0148] In one embodiment of the present invention, see Figure 9 The method for optimizing the double-tuned model is as follows:

[0149] S91. Adjust the mass of the tuned mass damper by adjusting the mass ratio of the tuned mass damper to the resonant structure, and adjust the stiffness and damping of the tuned mass damper by adjusting the spring stiffness and damper damping of the tuned mass damper to obtain the optimized double-tuned model.

[0150] S92. Perform harmonic response analysis on the optimized double-tuned model to obtain the trough frequency of the optimized double-tuned model;

[0151] S93. Determine whether the trough frequency of the optimized double-tuned model is consistent with the resonant frequency of the instrument.

[0152] S94. If the trough frequency of the optimized double-tuned model is inconsistent with the resonant frequency of the instrument, repeat steps S91 to S93; if the trough frequency of the optimized double-tuned model is consistent with the resonant frequency of the instrument, end the optimization process.

[0153] On the one hand, by adjusting the mass ratio of the TMD to the resonant structure, the TMD spring stiffness, and the damping parameters, the dynamic mechanical characteristics of the double-tuned model can be precisely shaped. By combining harmonic response analysis to extract the trough frequency and calibrating it with the instrument's resonant frequency, the vibration attenuation frequency band of the double-tuned model can completely coincide with the instrument's resonant frequency. Compared to the limitation of a single-tuned structure only achieving single-point frequency vibration reduction, this optimization method utilizes the coupling effect of dual resonances to form a deep vibration attenuation band at the instrument's resonant frequency, efficiently dissipating external vibration energy and fundamentally avoiding measurement errors caused by resonance. This ensures the measurement accuracy of sensitive instruments such as Coriolis flowmeters and vortex flowmeters in complex industrial vibration environments.

[0154] On the other hand, the iterative optimization process, centered on "parameter adjustment – ​​harmonic response analysis – frequency consistency judgment," directly correlates the design parameters of the double-tuned model with the target vibration reduction frequency. By coordinating the control of mass ratio, stiffness, and damping, it replaces the traditional trial-and-error method relying on experience, achieving quantitative and precise control over the inherent attenuation characteristics of the tuned model. The multi-iteration calibration mechanism effectively eliminates the influence of manufacturing errors and simulation deviations on frequency matching, ensuring that the optimized double-tuned model can stably perform its vibration reduction function in actual working conditions, significantly reducing the risk of vibration reduction failure due to parameter mismatch.

[0155] Specifically, after installing a tuned mass damper on the resonant structure, harmonic response analysis was performed on the double-tuned models before and after optimization.

[0156] See the amplitude-frequency response curve of the unoptimized double-tuned model. Figure 10 At the monitoring point of the resonant structure, the original single-peak curve with a large amplitude transforms into a double-peak curve with a smaller amplitude. The trough frequency f2 of the unoptimized dual-tuned model deviates from the instrument's resonant frequency f1, requiring further optimization of the stiffness and mass of the tuned mass damper. The mass of the tuned mass damper is adjusted by changing the mass ratio of the tuned mass damper to the resonant structure. The stiffness and damping of the tuned mass damper are then adjusted by changing the spring stiffness and damper damping, thus optimizing the stiffness and damping of the tuned mass damper and ultimately optimizing the dual-tuned model.

[0157] The amplitude-frequency response curve corresponding to the optimized double-tuned model can be found in [reference needed]. Figure 11 The optimized double-tuned model has a trough frequency that matches the resonant frequency of the instrument, while the double-peak frequency shifts.

[0158] The instrument vibration reduction control method of the present invention is based on dual resonance optimization for vibration reduction control processing, so that the vibration energy at the instrument's resonant frequency is effectively absorbed and dissipated by the tuned mass damper, thereby significantly suppressing the vibration transmitted to the instrument, improving the external vibration reduction effect at the instrument's resonant frequency, and ensuring the instrument's performance stability and measurement accuracy.

[0159] A second aspect of this invention provides an instrument vibration reduction control system for implementing the instrument vibration reduction control method based on dual resonance described in the first aspect of this invention. See also... Figure 12 The instrument vibration reduction control system 700 includes:

[0160] Model building module 701 is used to build finite element models, single-tuned models and double-tuned models of the pipeline system.

[0161] The setting module 702 is used to set the initial stiffness and initial mass of the resonant structure, the mass ratio of the tuned mass damper to the resonant structure, and the mass, initial stiffness, and initial damping of the tuned mass damper.

[0162] Analysis module 703 determines the resonant frequency of the instrument based on the finite element model, performs harmonic response analysis on the first-tuned model to obtain the peak frequency of the first-tuned model, and performs harmonic response analysis on the second-tuned model to obtain the trough frequency of the second-tuned model.

[0163] The judgment module 704 is configured to: determine whether the peak frequency of the first-tuned model is consistent with the resonant frequency of the instrument, and determine whether the trough frequency of the second-tuned model is consistent with the resonant frequency of the instrument.

[0164] The optimization module 705 is configured to: when the peak frequency of the first-tuned model is inconsistent with the resonant frequency of the instrument, adjust the stiffness and mass of the resonant structure to optimize the first-tuned model so that the peak frequency of the first-tuned model is consistent with the resonant frequency of the instrument; when the trough frequency of the second-tuned model is inconsistent with the resonant frequency of the instrument, adjust the stiffness and damping of the tuning mass damper to optimize the second-tuned model so that the trough frequency of the second-tuned model is consistent with the resonant frequency of the instrument.

[0165] The instrument vibration reduction control system described above is based on dual resonance optimization for vibration reduction control processing, which enables the vibration energy at the instrument's resonant frequency to be effectively absorbed and dissipated by the tuned mass damper, thereby significantly suppressing the vibration transmitted to the instrument. This improves the external vibration reduction effect at the instrument's resonant frequency and ensures the instrument's performance stability and measurement accuracy.

[0166] To verify the effectiveness of the instrument vibration reduction control method and system based on dual resonance described in the above embodiments of the present invention, the following specific embodiments are used for illustration.

[0167] Example 1: The instrument selected is the RHM15 Coriolis mass flow meter from RHIONIK, which has advantages such as high accuracy and wide measuring range, and has been widely used in various industrial fields. As a precision instrument based on the principle of vibration, this flow meter is susceptible to external vibration interference.

[0168] according to Figure 13 The finite element model of the pipeline system connected to the Coriolis mass flow meter is shown. This model is used as a model of a system without vibration damping. The finite element model of the instrument vibration damping control method and system based on dual resonance proposed in this invention (hereinafter referred to as: the method and system of this invention) is shown below. Figure 14 As shown. A resonant structure is added to the pipeline to construct a first-tuned model, and a TMD is further installed on the resonant mechanism to construct a second-tuned model.

[0169] To verify the effectiveness of the method and system of this invention, it is compared with traditional vibration reduction methods. Traditional vibration reduction methods are based on pipeline vibration reduction design concepts, see [link to relevant documentation]. Figure 15 The traditional vibration reduction system model shown typically involves installing a TMD directly on the connecting pipe to suppress vibration response by using damping energy dissipation.

[0170] The product manual states that the operating frequency of this Coriolis mass flow meter is 122.5Hz. In contrast, the pipe structure has higher rigidity, and its natural frequency is much higher than 122.5Hz. The significant frequency difference indicates that the Coriolis mass flow meter and the pipe are relatively independent in their dynamic characteristics and are less prone to coupled vibration. Therefore, this embodiment focuses on the frequency band where the Coriolis mass flow meter's own vibration characteristics are significant, in order to effectively suppress the influence of external vibrations. Simultaneously, as shown... Figures 13 to 15The flowmeter monitoring point A and the resonant structure monitoring point B are shown.

[0171] Harmonic response analysis was performed on the undamped system model under a displacement excitation of 0.1 mm, and the results are as follows: Figure 16 As shown, within the 50Hz to 150Hz frequency band, the amplitude-frequency response curve of the undamped system model exhibits a single resonance peak with a peak frequency of 122.5Hz, consistent with the operating frequency of the Coriolis mass flow meter specified in the product manual. The peak amplitude reaches 5.69mm, far exceeding the input displacement, indicating that the undamped system model experiences strong resonance at this frequency.

[0172] To suppress external vibrations at the resonant frequency of a Coriolis mass flow meter, the method and system of this invention are applied. The optimized design process includes the following steps: First, the resonant frequency of the Coriolis mass flow meter is determined. Second, a resonant structure is introduced into the connecting pipe to construct a first-order tuning model. Parameter optimization is then performed to ensure that the natural frequency of the first-order tuning model matches the resonant frequency of the Coriolis mass flow meter, thus completing the first-order tuning. Finally, a TMD (Transient Dynamic Modulation) is installed on the resonant structure to construct a second-order tuning model. The TMD parameters are further optimized to match the trough frequency of the second-order tuning model with the resonant frequency of the Coriolis mass flow meter, minimizing the vibration response and achieving second-order tuning.

[0173] Specifically, a resonant structure is added to the pipeline to construct a first-order tuned model. Parameter adjustments are made to ensure that the peak frequency of this first-order tuned model matches the resonant frequency of the Coriolis mass flow meter. The initial stiffness of the resonant structure... Set to 346.65 N / mm, initial mass Set it to 0.59kg.

[0174] After introducing the resonant structure, the finite element model becomes a single-tuned model. Harmonic response analysis of the single-tuned model is performed, and the results are as follows: Figure 17 As shown, the response curve exhibits a double resonance peak characteristic, compared to Figure 16 A single resonance peak was observed, with a new peak added. The first peak is located at 121.9 Hz, and the second peak is located at 122.5 Hz. The results indicate a frequency matching deviation between the resonant structure and the Coriolis mass flow meter at the resonant frequency, requiring further optimization of the resonant structure's stiffness and mass. The stiffness and mass of the resonant structure were adjusted by modifying its length and wall thickness. After numerical optimization, the stiffness of the resonant structure... The mass is 348.44 N / mm. The weight is 0.58 kg. Harmonic response analysis was performed on the optimized single-tuned model, and the results are as follows: Figure 18As shown, the two resonance peaks have merged into a single resonance peak at 122.5 Hz. The results indicate that the frequency of the resonant structure coincides with the resonant frequency of the Coriolis mass flow meter, and their amplitude-frequency characteristics are becoming consistent, indicating that the first tuning process is complete.

[0175] After installing a TMD as an accessory damping dissipation structure on the resonant structure, the finite element model becomes a double-tuned model. Setting the mass ratio of the TMD to the resonant structure to be 2%, the mass of the TMD in the double-tuned model can be obtained. Set the initial stiffness to 0.01 kg. The initial damping is 6.68 N / mm. for Harmonic response analysis was performed on the double-tuned model, and the results are as follows: Figure 19 As shown, at monitoring point B of the resonant structure, the original single-peak curve with a larger amplitude transforms into a double-peak curve with a smaller amplitude; the peak value of the low-order frequency is 108.3Hz, and the peak value of the high-order frequency is 136.6Hz. The trough frequency of the amplitude-frequency characteristic curve is 121.6Hz (orange dotted line), which deviates from the resonant frequency of the Coriolis mass flow meter, 122.5Hz (blue dashed line), indicating that frequency matching has not been fully achieved, and the vibration reduction effect can be further improved. The TMD parameters need to be optimized.

[0176] By analyzing the stiffness of the TMD and The damping was numerically optimized to obtain the optimal stiffness. The optimal damping is 6.81 N / mm. for The optimized double-tuned model harmonic response analysis results are as follows: Figure 20 As shown, the trough frequency of the amplitude-frequency response curve of the dual-tuned model is 122.5Hz (blue dashed line), which coincides with the resonant frequency of the Coriolis mass flow meter. Simultaneously, the dual-peak frequencies shift, with the lower-order peak at 108.8Hz and the higher-order peak at 137.4Hz. Through the above TMD parameter optimization, frequency matching of the dual-resonance system is achieved, making the trough frequency of the dual-tuned model coincide with the resonant frequency of the Coriolis mass flow meter, thus achieving optimal vibration reduction.

[0177] At flowmeter monitoring point A, the amplitude-frequency response curve after double tuning is as follows: Figure 21 As shown, its vibration amplitude at 122.5 Hz decreased to 0.058 mm. Therefore, the optimal combination of resonant structure and TMD structural parameters that minimizes the vibration response of the Coriolis mass flowmeter was determined through a double tuning process, marking the completion of the double tuning process.

[0178] The paper "Enhancement in the Seismic Performance of a Nuclear Piping System using Multiple Tuned Mass Dampers" investigated the application of TMDs in vibration reduction design of coolant piping systems in nuclear power plants. It proposed a numerical optimization framework based on frequency response analysis and employed a multi-objective genetic algorithm to optimize the stiffness and damping coefficients of the TMDs, minimizing the acceleration response and overall stress level of the piping system in three directions. This invention references this numerical optimization method to determine... Figure 15 The optimal TMD parameters for a traditional vibration reduction system model are determined. Setting the mass ratio of the TMD to the pipe to be 2%, the mass of the TMD in the traditional vibration reduction system model is 0.39 kg, the stiffness is 220.36 N / mm, and the damping is 0.05. .

[0179] The harmonic response results at flowmeter monitoring point A for the three schemes—the method and system of this invention, the model without vibration reduction, and the model of a traditional vibration reduction system—are as follows: Figures 21 to 23 As shown.

[0180] The harmonic response results of the undamped system model are as follows: Figure 22 As shown, the vibration amplitude is 5.69 mm at the Coriolis mass flow meter's resonant frequency of 122.5 Hz. This result serves as a baseline, reflecting the initial excitation level of the Coriolis mass flow meter by external vibration without any protective measures.

[0181] The harmonic response results of the traditional vibration reduction system model are as follows Figure 23 As shown, the vibration amplitude at 122.5 Hz is 2.60 mm, which is 3.09 mm lower than that of the model without vibration reduction, and the vibration reduction efficiency is 54.31%. Although it has a certain vibration reduction effect, it still does not reach the ideal level.

[0182] The harmonic response results of the method and system of the present invention are as follows: Figure 21 As shown, the vibration amplitude at 122.5 Hz is 0.058 mm, which is 5.632 mm lower than that of the model without vibration damping, and the vibration reduction efficiency reaches over 95%. This superior performance is attributed to the addition of a TMD as an additional damping unit to the resonant structure. Through dual tuning, precise frequency matching of the system is achieved, allowing the vibration energy at this frequency to be effectively absorbed and dissipated by the TMD, thereby significantly suppressing the vibration transmitted to the Coriolis mass flow meter.

[0183] Compared to the traditional vibration reduction method with a vibration reduction efficiency of 54.31%, the method and system of this invention, through more refined structural design, further improve the vibration reduction efficiency by 44.67%, demonstrating superior engineering application potential.

[0184] like Figure 17 As shown, the dynamic characteristics of the system change after a resonant structure is introduced into the pipeline, and its resonant response curve changes from a single resonance peak to a double resonance peak. To achieve optimal vibration reduction, the system needs to be restored to a single resonance peak state through a first tuning, that is, the natural frequency of the resonant structure needs to be numerically optimized to precisely match the resonant frequency of the Coriolis mass flow meter.

[0185] To verify the necessity and effectiveness of the first-order tuning process, the influence of the ratio of different resonant structure frequencies to the resonant frequency of the Coriolis mass flowmeter (i.e., different frequency ratios) on the amplitude at monitoring point A of the flowmeter was analyzed. The results are as follows: Figure 24 As shown.

[0186] When the frequency ratio deviates from 1, it indicates that the subsequent second tuning was performed without first-order tuning, and the resonant response curve still exhibits a bimodal response characteristic. For example, when the frequency ratio is 1.08, the amplitude at the monitoring point reaches 0.107 mm, indicating poor vibration reduction and suggesting that the resonant structure has not achieved effective tuning. When the frequency ratio equals 1, it indicates that first-order tuning has been completed, the bimodal effect is effectively suppressed, and the resonant response curve returns to a single resonance peak. At this time, the amplitude at the monitoring point drops to the global minimum of 0.058 mm, a decrease of 45.79% compared to the frequency ratio of 1.08. In this state, the resonant structure perfectly matches the resonant frequency of the Coriolis mass flow meter, which is equivalent to directly setting a TMD on the flow meter, thereby achieving the best vibration reduction effect (e.g., Figure 21 (As shown).

[0187] In summary, Figure 24 The study clearly reveals the variation law of amplitude at monitoring point A of the Coriolis mass flow meter with the frequency ratio, fully demonstrating the key role of primary tuning in the method and system of this invention. If the frequency of the resonant structure does not match the resonant frequency of the Coriolis mass flow meter, it will lead to a significant increase in amplitude and deterioration of vibration reduction performance. Achieving the merging of the two peaks into a single peak through primary tuning is a necessary prerequisite for ensuring the effectiveness of subsequent secondary tuning.

[0188] Example 2: Time-domain vibration analysis was performed on three schemes: a system without vibration damping, a traditional vibration damping method, and the method and system of this invention. A simple harmonic excitation with an amplitude of 0.1 mm and a frequency of 122.5 Hz (the resonant frequency of the flowmeter) was used. The time-history curves are shown below. Figure 25 As shown.

[0189] Figure 26 The steady-state vibration time history response of the undamped system at monitoring point A under the aforementioned harmonic excitation is presented. After the system enters steady state, the peak amplitude is 5.37 mm, which is much larger than the input amplitude, indicating that when the excitation frequency coincides with the resonant frequency of the Coriolis mass flowmeter, the system experiences strong resonance, and the displacement response is significantly amplified. This result serves as a benchmark, reflecting the sensitivity of the flowmeter to resonant excitation without any protective measures.

[0190] Figure 27 The steady-state amplitude response of monitoring point A under harmonic excitation using the traditional vibration reduction method is shown. The amplitude is 2.49 mm, which is 53.63% lower than that of the unreduced system (5.37 mm). The results indicate that the traditional vibration reduction method has a certain inhibitory effect on resonant frequency excitation, but the amplitude is still relatively large, and the vibration reduction effect is limited.

[0191] Figure 28 The steady-state amplitude response of the method and system of this invention at monitoring point A under the same harmonic excitation is demonstrated. The amplitude is 0.059 mm, which is more than 95% lower than that of the system without vibration reduction, showing excellent vibration isolation performance. Compared with traditional vibration reduction methods, the amplitude is reduced by 45%. The results show that the method and system of this invention have significant advantages over traditional vibration reduction methods.

[0192] Time-domain vibration analysis was conducted on three schemes under random excitation: a system without vibration damping, a traditional vibration damping method, and the method and system of this invention. The frequency range of the random excitation was 50–150 Hz, with a root mean square value of 0.1 mm. The time history curves are shown below. Figure 29 As shown.

[0193] Figure 30 The amplitude response of the undamped system at monitoring point A under random excitation is shown. It can be seen that the amplitude fluctuates continuously and irregularly, with the maximum amplitude exceeding 15 mm and the root mean square value being 5.74 mm. Due to the lack of an effective energy dissipation mechanism, the system cannot suppress the vibration response under continuous random excitation, resulting in a persistently high amplitude.

[0194] Figure 31 The amplitude response of monitoring point A under random excitation using a traditional vibration reduction method is shown. The amplitude is limited to within 10 mm, and the root mean square value is 2.37 mm, which is 58.71% lower than that of the unreduced system (5.74 mm). The results indicate that the traditional vibration reduction method has a certain inhibitory effect on random excitation, but the root mean square value is still relatively high, and the inhibitory effect is still limited.

[0195] Figure 32 The amplitude response of monitoring point A under the same excitation is shown in the method and system of this invention. The amplitude is significantly suppressed to within 0.2 mm, with a root mean square value of only 0.059 mm, a reduction of more than 95% compared to the system without vibration reduction. Compared to the vibration reduction efficiency of 58.71% of traditional vibration reduction methods, the method and system of this invention further improve the vibration reduction efficiency by nearly 40%.

[0196] The above results show that traditional vibration reduction methods have limited control effects on random excitation, while the method and system of this invention exhibit excellent random vibration control performance.

[0197] The above embodiments are used to explain the present invention, but not to limit the present invention. Any modifications and changes made to the present invention within the spirit and scope of the claims shall fall within the protection scope of the present invention.

Claims

1. A dual-resonance-based instrument damping control method, characterized by, The steps include: S1, establishing a finite element model of a pipeline system connecting instrument, taking the finite element model as a model without damping system, determining a resonance frequency of the instrument based on the finite element model; S2, adding a resonance structure on the pipeline of the finite element model, establishing a one-tuned model, setting initial stiffness and initial mass of the resonance structure; S3, performing harmonic response analysis on the one-tuned model to obtain a peak frequency of the one-tuned model; S4, judging whether the peak frequency of the one-tuned model is consistent with the resonance frequency of the instrument; S5, when the peak frequency of the one-tuned model is not consistent with the resonance frequency of the instrument, adjusting the stiffness and mass of the resonance structure, optimizing the one-tuned model, so that the peak frequency of the one-tuned model is consistent with the resonance frequency of the instrument; S6, installing a tuned mass damper on the resonance structure of the optimized one-tuned model, establishing a two-tuned model, setting a mass ratio of the tuned mass damper and the resonance structure, and mass, initial stiffness and initial damping of the tuned mass damper; S7, performing harmonic response analysis on the two-tuned model to obtain a valley frequency of the two-tuned model; S8, judging whether the valley frequency of the two-tuned model is consistent with the resonance frequency of the instrument; S9, when the valley frequency of the two-tuned model is not consistent with the resonance frequency of the instrument, adjusting the stiffness and damping of the tuned mass damper, optimizing the two-tuned model, so that the valley frequency of the two-tuned model is consistent with the resonance frequency of the instrument.

2. The dual-resonance-based instrument dither control method of claim 1, wherein, In step S1, the method for determining the resonance frequency of the instrument based on the finite element model is: performing harmonic response analysis on the finite element model to obtain an amplitude-frequency characteristic curve of the finite element model, taking a frequency corresponding to a curve peak value of the amplitude-frequency characteristic curve of the finite element model as the resonance frequency of the instrument.

3. The dual-resonance-based instrument dither control method of claim 1, wherein, In step S5, the method for optimizing the one-tuned model is: S51, adjusting the stiffness and mass of the resonance structure by adjusting the length and wall thickness of the resonance structure, to obtain the optimized one-tuned model; S52, performing harmonic response analysis on the optimized one-tuned model to obtain a peak frequency of the optimized one-tuned model; S53, judging whether the peak frequency of the optimized one-tuned model is consistent with the resonance frequency of the instrument; S54, if the peak frequency of the optimized one-tuned model is not consistent with the resonance frequency of the instrument, repeating steps S51 to S53; if the peak frequency of the optimized one-tuned model is consistent with the resonance frequency of the instrument, ending the optimization process.

4. The dual-resonance-based instrument dither control method of claim 1, wherein, In step S9, the method for optimizing the two-tuned model is: S91, adjusting the mass of the tuned mass damper by adjusting the mass ratio of the tuned mass damper and the resonance structure, adjusting the stiffness and damping of the tuned mass damper by adjusting the spring stiffness and damper damping of the tuned mass damper, to obtain the optimized two-tuned model; S92, performing harmonic response analysis on the optimized two-tuned model to obtain a valley frequency of the optimized two-tuned model; S93, judging whether the valley frequency of the optimized two-tuned model is consistent with the resonance frequency of the instrument; S94, if the trough frequency of the optimized double-tuned model is inconsistent with the resonance frequency of the instrument, repeating steps S91 to S93; if the trough frequency of the optimized double-tuned model is consistent with the resonance frequency of the instrument, ending the optimization process.

5. The dual-resonance-based instrument dither control method of claim 1, wherein, The finite element model of the pipeline system comprises: a connecting pipeline; a first metal pipeline connected to the connecting pipeline through a first standard flange; an instrument connected to the third standard flange through a second standard flange on one side and a third standard flange on the other side; a second metal pipeline connected to the third standard flange on one end and a fourth standard flange on the other end.

6. The dual-resonance-based instrument dither control method as claimed in claim 1, wherein, The single-tuned model comprises: a connecting pipeline; a first metal pipeline connected to the connecting pipeline through a first standard flange; a first resonant structure connected to the fourth standard flange on one side and a seventh standard flange on the other side; a fourth metal pipeline connected to the seventh standard flange on one end and an eighth standard flange on the other end. The double-tuned model comprises: a connecting pipeline; a first metal pipeline connected to the connecting pipeline through a first standard flange; a first resonant structure connected to the fourth standard flange on one side and a seventh standard flange on the other side; 7. The dual-resonance-based instrument dither control method as claimed in claim 1, wherein, a first tuned mass damper installed on the first resonant structure; a fourth metal pipeline connected to the seventh standard flange on one end and an eighth standard flange on the other end. The system comprises: a model construction module for constructing a finite element model of a pipeline system, a single-tuned model, and a double-tuned model; a setting module for setting the initial stiffness and initial mass of the resonant structure, the mass ratio of the tuned mass damper to the resonant structure, and the mass, initial stiffness, and initial damping of the tuned mass damper; an analysis module for determining the resonance frequency of the instrument based on the finite element model, performing harmonic response analysis on the single-tuned model to obtain the peak frequency of the single-tuned model, and performing harmonic response analysis on the double-tuned model to obtain the trough frequency of the double-tuned model; ​ ​ ​ ​ ​ 8. An instrument damping control system for implementing the dual resonance-based instrument damping control method according to any one of claims 1 to 7, characterized by ​ ​ ​ ​ The judging module is configured to judge whether the peak frequency of the one-tuned model is consistent with the resonance frequency of the instrument and whether the valley frequency of the two-tuned model is consistent with the resonance frequency of the instrument. The optimization module is configured to adjust the stiffness and mass of the resonance structure, optimize the one-tuned model, and make the peak frequency of the one-tuned model consistent with the resonance frequency of the instrument when the peak frequency of the one-tuned model is inconsistent with the resonance frequency of the instrument; and adjust the stiffness and damping of the tuned mass damper, optimize the two-tuned model, and make the valley frequency of the two-tuned model consistent with the resonance frequency of the instrument when the valley frequency of the two-tuned model is inconsistent with the resonance frequency of the instrument.

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