Plate-frame superstructure design method suitable for low-frequency vibration isolation of marine mechanical equipment
By introducing local resonators into the hull plate structure and optimizing their size and spacing, the problem of low-frequency vibration and noise isolation of marine mechanical equipment in the prior art has been solved, and effective isolation and vibration reduction effects of low-frequency vibration and noise have been achieved.
Patent Information
- Authority / Receiving Office
- CN · China
- Patent Type
- Patents(China)
- Current Assignee / Owner
- WUHAN UNIV OF TECH
- Filing Date
- 2022-08-15
- Publication Date
- 2026-04-28
AI Technical Summary
Existing technologies are insufficient to effectively isolate low-frequency vibration and noise from ship machinery and equipment without altering the basic hull frame structure. Furthermore, existing design methods are poorly suited for complex hull frame structures and cannot effectively isolate mid-to-low frequency elastic waves.
A plate-frame superstructure is designed. By introducing local resonators into the hull plate-frame structure, a model is built using COMSOL Multiphysics software, and parameter scanning and simulation calculations are performed to optimize the size and spacing of the local resonators, thereby achieving isolation of low-frequency vibration noise.
While ensuring the structural strength and stability of the hull, it effectively isolates the low-frequency vibration and noise of the ship's mechanical equipment, reducing vibration and underwater radiated noise throughout the ship.
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Figure CN115408833B_ABST
Abstract
Description
Technical Field
[0001] This invention belongs to the field of vibration isolation technology for marine machinery and equipment, and specifically relates to a plate frame superstructure design method suitable for low-frequency vibration isolation of marine machinery and equipment. Background Technology
[0002] With the booming development of the maritime transportation industry, the problem of ship vibration and radiated noise pollution is becoming increasingly serious. Ship vibration and noise not only affect the normal rest and work of crew members, but can also easily damage the hearing system, potentially causing noise-induced hearing loss, memory decline, and other symptoms. In 2014, the International Maritime Organization (IMO) adopted the "Noise Levels Code for Ships," which sets higher requirements for ship design and its vibration and noise reduction performance.
[0003] Ship vibration and noise excitation sources mainly include mechanical excitation, propeller excitation, and wave excitation, characterized by high power and a wide spectral distribution. Mechanical noise excitation sources primarily transmit vibrations through low-frequency bending waves in the hull structure and radiate noise into the air and water, severely impacting ship vibration and noise comfort and concealment. Currently, ship vibration reduction and noise reduction technologies mainly include passive control and active control technologies. Passive control technology can reduce vibration levels across the entire frequency band, but it remains extremely challenging in controlling the low-frequency vibration spectrum. Active control technology is an effective means of controlling the low-frequency vibration spectrum; however, it faces technical bottlenecks such as system stability, reliability, and environmental adaptability, and much research is still in the experimental stage, not yet widely applied in engineering.
[0004] To achieve low-frequency vibration isolation of ship hull mechanical equipment structures, phononic crystal structures capable of localized resonance can be introduced into the design of ship plate frame structures. A phononic crystal structure is a medium composed of unitary materials arranged periodically in space. This type of structure possesses a bandgap characteristic, meaning that elastic waves (vibration waves or sound waves) within a certain frequency range cannot propagate through the phononic crystal. The bandgap characteristics of different frequency bands can be adjusted by changing material properties, combination methods, and periodic distribution. When elastic waves within the bandgap frequency range enter the superstructure, due to the scattering effect of its internal periodic structure or the localized resonance effect, the elastic waves within the bandgap cannot propagate through the structure, forming an elastic wave bandgap. The concept of phononic crystal structures provides a new approach to solving low-frequency vibration and radiated noise control in structures.
[0005] Chinese patent CN2020108446478 discloses a load-bearing and vibration-isolation integrated plate-shell superstructure and its design method, comprising a plate-shell and several microstructure components. The microstructure components are block-shaped structures with uniform cross-sections, their cross-sectional outlines consisting of a base and two arcsine-like curves, the intersection of which forms a beak. The bottom surface of the microstructure components is connected to the plate-shell. Several microstructure components are periodically arranged to form an array that separates the vibration source from the protected object. When the elastic wave excited by the vibration source propagates along the plate-shell from any direction to the area around the array, the microstructure components undergo bending and torsional vibrations under the influence of the elastic wave. The microstructures exert forces and moments on the plate-shell, thereby suppressing the propagation of elastic waves within the plate-shell and achieving isolation of high-frequency elastic waves incident from any direction within the plate-shell. The structure is simple, easy to manufacture, and can simultaneously isolate the protected object and the vibration source, possessing both good load-bearing and vibration-isolation capabilities. However, the patent still has the following drawbacks: its microstructure components are relatively complex and require high processing precision; the above design method is mainly for bare plate structures and is not well applicable to more complex hull plate structures; and it is mainly for vibration isolation of mid-to-high frequency elastic waves, and cannot effectively isolate mid-to-low frequency elastic waves of ships. Summary of the Invention
[0006] The technical problem to be solved by the present invention is to provide a plate frame superstructure design method for low-frequency vibration isolation of ship machinery and equipment, which addresses the shortcomings of the existing technology. Without changing the basic plate frame structure of the hull to ensure its strength, stiffness, stability and reliability, the method can effectively isolate the low-frequency vibration and noise of ship machinery and equipment, thereby reducing vibration and underwater radiated noise throughout the ship.
[0007] The technical solution adopted by the present invention to solve the above-mentioned technical problems is as follows:
[0008] A plate frame superstructure design method for low-frequency vibration isolation of marine machinery and equipment, wherein the plate frame superstructure includes a hull plate frame structure and additional periodically arranged local resonators, and the design method includes the following steps:
[0009] S1. Analyze the vibration excitation sources of typical ship mechanical equipment and determine the required vibration isolation frequency range;
[0010] S2. For the hull plate structure, considering the continuity of the model, local resonators are arranged at intervals in the intersection area of longitudinal stiffeners and transverse stiffeners; a hull plate superstructure unit cell model and a hull plate superstructure finite periodic model are established respectively.
[0011] S3. Apply Floquet periodic boundary conditions in the x and y directions of the hull plate superstructure unit cell model, perform parameter scanning along the Γ-XM-Γ wave vector, solve for the eigenmodes and eigenfrequency of the structure under each wave vector, and obtain the band structure of the hull plate superstructure with the wave vector direction as the abscissa and the eigenfrequency as the ordinate; apply acceleration excitation to the left side of the finite periodic model of the hull plate superstructure to simulate the vibration of the vibration source, output the acceleration response on the right side, and output the vibration transmission characteristic curve; compare the band structure with the transmission characteristics to verify the correctness of the band gap position and width;
[0012] S4. By changing the width, height and longitudinal and transverse spacing of the local resonators, multiple simulation calculations were performed on the hull plate frame superstructure unit cell model. The structural dimensions of the hull plate frame superstructure were determined by comprehensively considering factors including structural lightweighting and the required band gap.
[0013] S5. Determine the relative position between the vibration source and the protected area. Add a local resonator with the structural dimensions and spacing determined in S4 to the hull plate structure in the middle of the vibration source and the protected area, and establish a superstructure model of the hull plate structure with actual dimensions.
[0014] S6. Apply acceleration excitation at the vibration source, calculate the intrinsic displacement field of the actual-sized hull plate superstructure model, verify the effectiveness of the local resonator arrangement, and complete the design of the hull plate structure.
[0015] In the above scheme, the frequency range of vibration control is 0-500Hz.
[0016] In the above scheme, the hull plate structure includes a deck and longitudinal and transverse reinforcing ribs installed on the bottom surface of the deck. The longitudinal reinforcing ribs are made of T-shaped profiles, and the transverse reinforcing ribs are made of L-shaped profiles.
[0017] In the above scheme, the lattice form of the plate-frame superstructure is rectangular, and the local resonator is a cuboid structure.
[0018] In the above scheme, the longitudinal lattice constant a and the transverse lattice constant b of the plate-frame superstructure are both in the range of 300-800 mm; the width l and height h of the local resonator are both in the range of 0.15a-0.35a; the longitudinal spacing m and the transverse spacing n of the local resonator satisfy m=a and n=b, respectively.
[0019] In the above scheme, both the local resonator and the hull plate frame structure are made of marine steel.
[0020] In the above scheme, step S2 uses COMSOL Multiphysics software to establish a unit cell model and a finite periodic model of the hull plate superstructure, respectively. The local resonator is modeled using solid elements, and the hull plate structure is modeled using shell elements. The shell elements are meshed using free triangular meshes, and the solid elements are meshed using a sweeping method. Material properties are added to the solid elements and shell elements, mainly including Young's modulus, Poisson's ratio, and density. Multiphysics coupling is added to realize the solid-shell connection.
[0021] The beneficial effects of this invention are as follows:
[0022] 1. This invention provides a plate-frame superstructure design method. By establishing two models of the hull plate-frame superstructure and calculating its band structure and transmission characteristics, the correctness of its bandgap position and width is verified. By introducing local resonators, effective isolation of vibration and noise in the mid-to-low frequency range below 500Hz for ship mechanical equipment is achieved. Vibration reduction in a specific frequency range can be achieved by designing the size and spacing of the local resonators, thereby improving the vibration reduction effect.
[0023] 2. The local resonator has a simple cuboid structure and is directly attached to the hull frame structure, making it easy to manufacture. The local resonator uses the same material as the hull frame structure, allowing for integral molding. The local resonator is attached to the same side of the intersection area between the T-shaped and L-shaped profiles of the hull frame structure, without altering the original structure. This ensures its strength, rigidity, stability, and reliability without affecting the installation of the original equipment. Attached Figure Description
[0024] The present invention will be further described below with reference to the accompanying drawings and embodiments. In the accompanying drawings:
[0025] Figure 1 This is a schematic diagram of the plate frame superstructure in an embodiment of the present invention (taking a 6×3 finite period array structure as an example);
[0026] Figure 2 These are three-view diagrams of the unit cell structure of the plate-frame superstructure in this embodiment of the invention;
[0027] Figure 3 (a) is the band structure diagram of the superstructure unit cell model of the hull plate frame in the embodiment of the present invention;
[0028] Figure 3 (b) is a transmission characteristic curve of the finite period model of the hull plate frame superstructure in the embodiment of the present invention;
[0029] Figure 4 These are superstructure model diagrams of two types of ship hull plates with actual ship dimensions in embodiments of the present invention;
[0030] Figure 5 yes Figure 4 The intrinsic displacement fields of the two types of full-size hull plate superstructures are shown.
[0031] In the diagram: 10. Hull plate frame structure; 11. Deck; 12. Longitudinal stiffener; 13. Transverse stiffener; 20. Local resonator; 30. Mechanical equipment. Detailed Implementation
[0032] To provide a clearer understanding of the technical features, objectives, and effects of the present invention, specific embodiments of the present invention will now be described in detail with reference to the accompanying drawings.
[0033] This invention addresses vibration isolation design for plate frame structures of marine machinery equipment, proposing a plate frame superstructure design method suitable for low-frequency vibration isolation of marine machinery equipment. The plate frame superstructure is as follows: Figure 1 As shown, the structure includes a hull frame structure 10 and additional periodically arranged local resonators 20. The hull frame structure 10 includes a deck 11 and longitudinal stiffeners 12 and transverse stiffeners 13 orthogonally distributed on the bottom surface of the deck. The longitudinal stiffeners 12 are T-shaped, and the transverse stiffeners 13 are L-shaped. The bottom of the local resonators 20 is connected to the ship's deck 11. The frame superstructure has a flexural bandgap in the low-frequency range below 500Hz, which can effectively isolate the vibration of the mechanical equipment 30 from further transmission to the surrounding structure.
[0034] This invention relates to a plate-frame superstructure design method for low-frequency vibration isolation of marine machinery and equipment, comprising the following steps:
[0035] S1. Analyze the vibration excitation sources of typical ship mechanical equipment and determine the required vibration isolation frequency range.
[0036] S2. For the hull plate structure, considering the continuity of the model, local resonators are arranged at intervals in the intersection area of longitudinal stiffeners and transverse stiffeners; a hull plate superstructure unit cell model and a hull plate superstructure finite periodic model are established respectively.
[0037] In this embodiment, COMSOL Multiphysics software was used to establish the superstructure unit cell model of the hull plate frame (e.g., Figure 2 (as shown) and the finite periodic model of the hull plate frame superstructure (as shown) Figure 1As shown, the finite-period model of the hull plate superstructure consists of 6×3 hull plate superstructure unit cell models. The local resonator is modeled using solid elements, while the hull plate structure is modeled using shell elements. Free triangular meshes are used for the shell elements, and a swept mesh is used for the solid elements. Material properties, primarily Young's modulus, Poisson's ratio, and density, are added to both solid and shell elements. Multiphysics coupling is added to achieve solid-shell connection.
[0038] The longitudinal lattice constant of the plate-frame superstructure is *a*, the transverse lattice constant is *b*, the plate thickness is *e*, the width, height, and longitudinal and transverse spacing of the local resonators are *l*, *h*, *m*, and *n*, respectively. The length, spacing, height, width, web thickness, and flange thickness of the T-sections are *a1*, *l1*, *H*, *B*, *t1*, and *t2*, respectively. The length, spacing, side width, and thickness of the L-sections are *b1*, *l2*, *c*, and *d*, respectively. Specific structural parameters: a = 600 mm; b = 500 mm; e = 6 mm; l = 100 mm; h = 100 mm; m = 600 mm; n = 500 mm; a1 = 600 mm; l1 = 250 mm; H = 50 mm; B = 50 mm; t1 = 5 mm; t2 = 7 mm; b1 = 500 mm; l2 = 300 mm; c = 30 mm; d = 3 mm. Material parameters: Low carbon steel (density 7800 kg / m³). 3 Elastic modulus 2×10 11 Pa; Poisson's ratio 0.3).
[0039] S3. Apply Floquet periodic boundary conditions in the x and y directions of the hull plate superstructure unit cell model, perform parameter scanning along the wave vector in the Γ-XM-Γ direction, solve for the eigenmodes and eigenfrequency of the structure under each wave vector, and obtain the band structure of the hull plate superstructure by plotting the wave vector direction as the abscissa and the eigenfrequency as the ordinate; apply acceleration excitation to the left side of the finite periodic model of the hull plate superstructure to simulate the vibration of the vibration source, output the acceleration response on the right side, and output the vibration transmission characteristic curve; compare the band structure with the transmission characteristics to verify the correctness of the band gap position and width.
[0040] In this embodiment, from Figure 3 (a) The band structure diagram of the unit cell reveals that the plate-frame superstructure unit cell has a complete band gap in the frequency range of 0-500 Hz, with the band gap located between 205-285 Hz; on the other hand, from Figure 3 (b) The transmission characteristic curves of the finite periodic structure reveal that the plate-frame superstructure exhibits significant attenuation characteristics for bending vibration in the 205-285Hz frequency range, with a bending vibration bandgap width of 80Hz, effectively achieving vibration reduction and noise reduction in both low and wide frequency bands. The bandgap position and width of the band structure and transmission characteristics match well.
[0041] S4. Change the width, height and longitudinal and transverse spacing of the local resonators, and perform multiple simulation calculations on the unit cell model according to steps S2-S3. Take into account factors including structural lightweighting and required band gaps to determine the structural dimensions of the hull plate superstructure.
[0042] S5. Determine the relative position of the vibration source and the protected area. Add local resonators with the structural dimensions and spacing determined in S4 to the hull plate structure in the middle of the vibration source and the protected area, and establish a superstructure model of the hull plate with actual dimensions.
[0043] S6. Apply acceleration excitation at the vibration source, calculate the intrinsic displacement field of the actual-sized hull plate superstructure model, verify the effectiveness of the local resonator arrangement, and complete the design of the hull plate structure.
[0044] For example Figure 4 Taking the two types of local resonator arrangements shown as examples, a real-size hull plate superstructure model is established. An acceleration excitation is applied at the vibration source (i.e., point A), and the acceleration response is output in the protected region (i.e., point B). The intrinsic displacement fields of the two types of hull plate superstructures are also output, such as... Figure 5 As shown. In Type I, a 6×4 local resonator is placed in the middle of a hull plate structure of full ship size. An excitation acceleration is applied to the left end of the plate structure to simulate the excitation generated by the ship's mechanical equipment, and the protected area is located at the right end of the plate structure. Figure 5 (a) and (b) represent the displacement fields at frequencies inside and outside the bandgap corresponding to type I, with frequencies of 229 Hz and 305 Hz, respectively. Figure 5 As can be seen from (a) and (b), when the excitation frequency of the excitation source is within the bandgap range (f = 229 Hz), the plate frame superstructure can effectively suppress vibrations near the excitation source and prevent the excitation source from propagating to the protected area. When the excitation frequency of the excitation source is outside the bandgap range (f = 305 Hz), the elastic wave can propagate smoothly from the excitation source to the protected area. In Type II, 3×4 local resonators are placed at both ends of the middle section, and excitation acceleration is applied to the middle of the plate frame structure to simulate the excitation generated by the ship's power equipment, while the protected area is located at both ends of the plate frame structure. From Figure 5 As can be seen from (c) and (d), when the excitation frequency of the excitation source is within the bandgap range (f = 263 Hz), the plate frame superstructure can effectively suppress the vibration near the excitation source; while when the excitation frequency of the excitation source is outside the bandgap range (f = 403 Hz), the elastic wave can smoothly propagate from the excitation source to the protected area.
[0045] The embodiments of the present invention have been described above with reference to the accompanying drawings. However, the present invention is not limited to the specific embodiments described above. The specific embodiments described above are merely illustrative and not restrictive. Those skilled in the art can make many other forms under the guidance of the present invention without departing from the spirit and scope of the claims. All of these forms are within the protection scope of the present invention.
Claims
1. A plate-frame superstructure design method for low-frequency vibration isolation of marine machinery and equipment, the plate-frame superstructure comprising a hull plate-frame structure and additional periodically arranged local resonators, the hull plate-frame structure comprising a deck and longitudinal and transverse stiffeners installed on the bottom surface of the deck; characterized in that, The lattice form of the plate-frame superstructure is rectangular, and the local resonator is a cuboid structure; both the local resonator and the hull plate-frame structure are made of marine steel; the design method includes the following steps: S1. Analyze the vibration excitation sources of typical ship mechanical equipment and determine the required vibration isolation frequency range; S2. For the hull plate structure, considering the continuity of the model, local resonators are arranged at intervals in the intersection area of longitudinal stiffeners and transverse stiffeners; a hull plate superstructure unit cell model and a hull plate superstructure finite periodic model are established respectively. S3. Apply Floquet periodic boundary conditions in the x and y directions of the hull plate superstructure unit cell model. Perform parameter scanning along the Γ-XM-Γ wave vector to solve for the eigenmodes and eigenfrequency of the structure under each wave vector. Plot the wave vector direction as the abscissa and the eigenfrequency as the ordinate to obtain the band structure of the hull plate superstructure. Apply acceleration excitation to the left side of the finite periodic model of the hull plate superstructure to simulate the vibration of the vibration source. Output the acceleration response and vibration transmission characteristic curve on the right side. Compare the band structure with the transmission characteristics to verify the correctness of the band gap position and width. S4. By changing the width, height and longitudinal and transverse spacing of the local resonators, multiple simulation calculations were performed on the hull plate frame superstructure unit cell model. The structural dimensions of the hull plate frame superstructure were determined by comprehensively considering factors including structural lightweighting and the required band gap. S5. Determine the relative position between the vibration source and the protected area. Add a local resonator with the structural dimensions and spacing determined in step S4 to the hull plate structure in the middle of the vibration source and the protected area, and establish a superstructure model of the hull plate structure with actual dimensions. S6. Apply acceleration excitation at the vibration source, calculate the intrinsic displacement field of the actual-sized hull plate superstructure model, verify the effectiveness of the local resonator arrangement, and complete the design of the hull plate structure.
2. The plate frame superstructure design method for low-frequency vibration isolation of marine machinery and equipment according to claim 1, characterized in that, The frequency range for vibration control is 0-500Hz.
3. The plate frame superstructure design method for low-frequency vibration isolation of marine machinery and equipment according to claim 1, characterized in that, The longitudinal stiffeners are made of T-shaped profiles, and the transverse stiffeners are made of L-shaped profiles.
4. The plate frame superstructure design method for low-frequency vibration isolation of marine machinery and equipment according to claim 3, characterized in that, The longitudinal lattice constant of the plate-frame superstructure a and transverse lattice constant b All are within the range of 300-800 mm; the width of the local resonator l and height h All are at 0.15 a -0.35 a Within the range; longitudinal spacing of local resonators m and horizontal spacing n Each satisfies m = a , n = b .
5. The plate frame superstructure design method for low-frequency vibration isolation of marine machinery and equipment according to claim 1, characterized in that, In step S2, COMSOL Multiphysics software is used to establish a unit cell model and a finite periodic model of the hull plate superstructure. The local resonator is modeled using solid elements, and the hull plate structure is modeled using shell elements. The shell elements are meshed using free triangular meshes, and the solid elements are meshed using a swept method. Material properties, including Young's modulus, Poisson's ratio, and density, are added to the solid elements and shell elements respectively. Multiphysics coupling is added to realize the solid-shell connection.