Damping device for heavy-lift helicopters

Abstract

The application relates to the field of damping technology, and discloses a damping device for heavy-load helicopters, which comprises a plurality of dampers used for connecting containers with helicopters, wherein each damper comprises a cylinder, a magneto-rheological fluid flow restrictor arranged in the cylinder, an actuating piston and a piston rod, and further comprises a tension-compression conversion assembly and a protection assembly; the tension-compression conversion assembly comprises a top plate, a bottom plate, a plurality of first connecting rods fixedly connected with the lower end of the top plate and a plurality of second connecting rods fixedly connected with the upper end of the bottom plate, the upper end of each second connecting rod is hingedly connected with a second hinge shaft of a corresponding second hinge piece on the helicopter, and the damper is connected with the helicopter through one second hinge piece; when the protection assembly comprises an upper crank and a lower crank arranged in an inclined mode, one end of the upper crank is hingedly connected with a first hinge shaft, the other end of the upper crank is hingedly connected with one end of the lower crank, and the other end of the lower crank is hingedly connected with the top plate. The application significantly improves the support, damping adaptability and reliability of the helicopter when flying under heavy-load and complex vibration conditions.

Description

Technical Field

[0001] This invention relates to the field of vibration reduction technology, and in particular to a vibration reduction device for heavy-duty helicopters. Background Technology

[0002] With the rapid development of aviation technology, heavy-duty helicopters are key transportation platforms. Their takeoff weight is usually heavy, and the equipment transported by internal or external sling loads has weight. The heavy-duty containers are sling-loaded onto the bottom of the fuselage and connected by connectors. Since the helicopter itself is subject to vibrations from many aspects such as the engine, tail system, and landing gear, the overall stiffness, mass distribution, and vibration characteristics of the helicopter will be affected as the counterweight increases after being loaded. This will exacerbate the complexity of vibration problems and significantly affect flight safety and the overall reliability of the aircraft. Therefore, a more reliable vibration damping device is needed between the heavy-duty container and the heavy-duty helicopter to reduce the threat posed by vibration.

[0003] Currently, vibration reduction devices commonly used in heavy-duty helicopters are generally divided into two main categories: passive vibration reduction and active vibration reduction. Passive vibration reduction includes technologies such as vibration isolation, vibration absorption, and vibration damping. However, due to the need to consider many factors such as strength stress, vibration reduction effect, and safety for heavy-duty helicopters, the application of active vibration reduction technology is dominant. It uses Active Controlled Structural Response (ACSR) technology to actively intervene in fuselage vibration by applying control forces at non-nodal locations of the helicopter's main modes. Active vibration reduction is usually achieved through magnetic rheostat dampers. These magnetic rheostat dampers include a cylindrical structure, a magnetorheological fluid current limiter, an actuating piston, and a piston rod. Connected within the cylindrical structure, the upper cavity of the cylinder is filled with high-pressure gas to act as an air spring buffer, while the bottom cavity is filled with magnetic metal liquid for passing through the magnetorheological fluid current limiter. At least four of these magnetic rheostats are used, typically connected sequentially to the front and rear sides of the container. The upper end of each magnetic rheostat is directly suspended from the fuselage, and the piston rod is connected to the container. When the helicopter is taking off, landing, stationary, or in flight, the damping force of the magnetic metal liquid is adjusted by regulating the magnetic field of the internal magnetorheological fluid current limiter. The internal air spring buffers the impact of compression and restores the elasticity of the internal actuating piston and piston rod, thus achieving the purpose of vibration reduction.

[0004] However, existing magnetic variable resistance vibration dampers are insufficient to cope with the wide range of load variations (1600kg to 10000kg) and complex operating conditions of helicopter containers. Specifically, although air springs have adjustable stiffness, their structure itself determines that they can only stably withstand compressive loads and cannot stably withstand tensile loads. This is because air springs rely on high-pressure gas in a sealed cavity to provide elastic restoring force, and the piston rod can only compress inward. Once under tension, the piston rod will dislodge from the cylinder, causing seal failure. Therefore, when the load increases, the air spring, as a variable stiffness compression buffer element, cannot continuously withstand tensile loads during helicopter flight, and may suffer serious damage or even functional failure. Existing magnetic variable resistance vibration dampers do not have effective ultimate overload protection measures and structures. When encountering extreme flight conditions under heavy loads, structural damage is likely to occur, increasing flight safety hazards. Summary of the Invention

[0005] This invention provides a vibration reduction device for heavy-load helicopters, which significantly improves the support, vibration reduction adaptability and reliability of helicopters when flying under heavy load and complex vibration conditions.

[0006] This invention provides a vibration damping device for heavy-duty helicopters, comprising multiple vibration dampers for connecting a container to the helicopter. Each vibration damper includes a cylinder, a magnetorheological fluid flow limiter housed within the cylinder, an actuating piston, and a piston rod. The vibration damper further includes a tension-compression conversion assembly and a protective assembly. The tension-compression conversion assembly includes a top plate, a bottom plate, multiple first connecting rods fixedly connected to the lower end of the top plate, and multiple second connecting rods fixedly connected to the upper end of the bottom plate. Each first connecting rod penetrates the bottom plate and is slidably connected to it. Each second connecting rod penetrates the top plate and is slidably connected to it. The top plate is fixedly connected to the upper end face of the cylinder. The lower end of the piston rod... The upper end of each first link is fixed to the bottom plate. The lower end of each first link is hinged to the first hinge shaft of the first hinge member of the container side wall. The upper end of each second link is hinged to the second hinge shaft of the corresponding second hinge member on the helicopter. It is connected to the helicopter through a second hinge member. When tension is applied, the high-pressure gas in the upper part of the cylinder is converted into pressure by the piston rod and is buffered and absorbed. The protective assembly includes an inclined upper crank and a lower crank. One end of the upper crank is hinged to the first hinge shaft, and the other end is hinged to one end of the lower crank. The other end of the lower crank is hinged to the top plate to provide tensile overload protection.

[0007] Preferably, the lower end of each first link and the upper end of each second link are fixedly connected with protective kits for pressure limiting protection.

[0008] Preferably, the number of the first link and the second link in the shock absorber is greater than or equal to two.

[0009] Preferably, the first hinge component includes a first hinge bracket, which is welded and fixed to the container, and the first hinge shaft is laterally connected to the first hinge bracket.

[0010] Preferably, the second hinge component includes a second hinge bracket, which is welded and fixed to the helicopter. The second hinge shaft is hinged to the second hinge bracket via a connecting arm, and the connecting arm is hinged to the second hinge bracket. The other end of the connecting arm is rotatably connected to the outside of the second hinge shaft.

[0011] Preferably, each first link is ball-jointed to a corresponding first hinge shaft, and the end of each second link is ball-jointed to a corresponding second hinge shaft.

[0012] Preferably, the protective kit has a wedge-shaped sleeve structure and is welded and fixed to the corresponding first or second link. When the top plate approaches the protective kit of the second link, the protective kit acts as a buffer and barrier to protect the top plate.

[0013] Preferably, the top plate is fitted with a protective sleeve to prevent the internal structure from being damaged by impact, and the lower end of the protective sleeve is fitted over the bottom plate and movably connected to the bottom plate.

[0014] Preferably, a floating piston is movably connected above the magnetorheological fluid current limiter inside the cylinder to block high-pressure air and magnetorheological fluid inside the cylinder, preventing gas from mixing into the magnetorheological fluid current limiter. The magnetorheological fluid current limiter dynamically changes the shear yield stress of the magnetorheological fluid by controlling the current of the internal excitation coil to cope with vibration force under overload.

[0015] Preferably, a pressure regulating valve is provided on the upper side of the cylinder to adjust the pressure of the high-pressure air inside the cylinder, which is used to achieve dual active control of damping stiffness with the magnetorheological fluid flow limiter.

[0016] Compared with existing technologies, the beneficial effects of this invention are as follows: This invention can achieve both tensile and compressive load bearing and vibration reduction. Specifically, through structures such as the top plate, bottom plate, first connecting rod, and second connecting rod, the high-pressure gas spring system, which could originally only withstand pressure, can convert external tension into internal pressure, thereby achieving buffering and absorption of tensile loads. This significantly improves the support, vibration reduction adaptability, and reliability of the helicopter under heavy load and complex vibration conditions. At the same time, this invention can achieve the effect of extreme overload protection. Specifically, through the upper and lower crank structures, when the tensile load exceeds a set threshold, it is activated to form a rigid connecting rod, limiting further displacement and effectively preventing the structure from being damaged due to the paralysis of the high-pressure gas spring system caused by extreme impacts, thus improving the overall system safety. Meanwhile, this invention also has the advantages of compact structure and reliable connection. Specifically, this invention achieves multi-point connection between the helicopter and the container through the combination of multiple connecting rods and hinges, distributing the load and enhancing the connection stiffness and stability.

[0017] During application, when the helicopter is subjected to an upward pull during flight, the pull is transmitted to the base plate through the first connecting rod, pushing the piston rod to move upward, compressing the high-pressure gas in the cylinder, and converting the pull into pressure, which is absorbed by the air spring and the magnetorheological fluid current limiter. Attached Figure Description

[0018] Figure 1 This is a partial structural schematic diagram of a vibration reduction device for a heavy-duty helicopter provided in an embodiment of the present invention; Figure 2 A schematic diagram of the application structure of a vibration reduction device for a heavy-duty helicopter provided in an embodiment of the present invention; Figure 3 A top view schematic diagram of the application of a vibration reduction device for a heavy-duty helicopter provided in an embodiment of the present invention; Figure 4 A schematic diagram of the internal structure of a cylinder for a vibration damping device used in a heavy-duty helicopter, provided as an embodiment of the present invention; Figure 5 This is a partial structural schematic diagram of a vibration reduction device for a heavy-duty helicopter provided in an embodiment of the present invention; Figure 6 This is a partial structural schematic diagram of a vibration reduction device for a heavy-duty helicopter provided in an embodiment of the present invention; Figure 7 This is a partial structural schematic diagram of a vibration reduction device for a heavy-duty helicopter provided in an embodiment of the present invention; Figure 8 This is a partial structural schematic diagram of a vibration reduction device for a heavy-duty helicopter provided in an embodiment of the present invention; Figure 9 This is a schematic diagram of the structure of the first hinge bracket in a vibration damping device for a heavy-load helicopter, provided in an embodiment of the present invention. Figure 10 A schematic diagram of the structure of the second hinge bracket in a vibration damping device for a heavy-load helicopter provided in an embodiment of the present invention; Figure 11 A schematic diagram of the structure of a protective kit in a vibration damping device for a heavy-duty helicopter, provided in an embodiment of the present invention; Figure 12 A schematic diagram of the connecting arm in a vibration damping device for a heavy-duty helicopter provided in an embodiment of the present invention; Figure 13 A schematic diagram of the relationship between shear stress and magnetic induction intensity in the shear characteristics of a magnetorheological fluid for a vibration damping device for a heavy-load helicopter, provided as an embodiment of the present invention; Figure 14 A schematic diagram of the constitutive relationship curve of a magnetorheological fluid for a vibration damping device for a heavy-duty helicopter, provided as an embodiment of the present invention; Figure 15 A schematic diagram of the maximum damping force curve of a magnetorheological fluid current limiter 12 for a vibration reduction device for a heavy-duty helicopter provided in an embodiment of the present invention; Figure 16 A schematic diagram of the stiffness curve of a single spring corresponding to different tensile requirements under full load conditions for a vibration damping device for a heavy-load helicopter provided in an embodiment of the present invention. Figure 17 A schematic diagram of the vibration isolation characteristic curve of a vibration reduction device for a heavy-load helicopter provided in an embodiment of the present invention; Figure 18 A vibration reduction device for heavy-load helicopters provided in this embodiment of the invention satisfies a frequency ratio requirement under no-load conditions. A schematic diagram of the design range curve for the stiffness of a single spring between 2 and 3; Figure 19 A vibration reduction device for heavy-load helicopters provided in this embodiment of the invention satisfies a frequency ratio under full-load conditions. A schematic diagram of the design range curve for the stiffness of a single spring between 2 and 3; Figure 20 A schematic diagram of the spring stiffness design range for a vibration reduction device for a heavy-load helicopter provided in an embodiment of the present invention, which satisfies a 60% vibration isolation effect under no-load conditions. Figure 21 A schematic diagram of the spring stiffness design range for a vibration damping device for a heavy-load helicopter provided in an embodiment of the present invention, which satisfies a 60% vibration isolation effect under full load conditions. Figure 22 A schematic diagram of the stiffness curves of a single spring corresponding to different stretching amounts under full load conditions for a vibration damping device for a heavy-load helicopter provided in an embodiment of the present invention. Figure 23 A schematic diagram showing the stiffness curves of a single spring corresponding to different stretching amounts under no-load conditions for a vibration damping device for a heavy-load helicopter provided in an embodiment of the present invention. Figure 24 The initial air pressure of a vibration damping device for a heavy-load helicopter under no-load conditions provided in this embodiment of the invention. With equivalent stiffness A schematic diagram of the relationship curve; Figure 25 The initial air pressure of a vibration damping device for a heavy-load helicopter under full load conditions, as provided in this embodiment of the invention. With equivalent stiffness A schematic diagram of the relationship curve; Figure 26 A schematic diagram of vibration isolation characteristic curves of a vibration reduction device for a heavy-load helicopter under different damping ratios, provided for an embodiment of the present invention; Figure 27A schematic diagram of dual-frequency excitation simulation of a vibration reduction device for a heavy-load helicopter provided in an embodiment of the present invention; Figure 28 An acceleration cloud map of the mission cabin floor when unloaded before vibration reduction, provided as an embodiment of the present invention for a vibration reduction device for a heavy-load helicopter; Figure 29 An acceleration cloud map of the mission cabin floor when unloaded after vibration reduction, provided as an embodiment of the present invention for a vibration reduction device for a heavy-load helicopter; Figure 30 An acceleration cloud diagram of the mission cabin floor under full load before vibration reduction, provided as an embodiment of the present invention for a vibration reduction device for a heavy-load helicopter; Figure 31 An acceleration cloud diagram of the mission cabin floor under full load after vibration reduction, provided as an embodiment of the present invention for a vibration reduction device for a heavy-load helicopter; Figure 32 A schematic diagram showing the acceleration comparison of the mission cabin floor in the time domain and when unloaded, for a vibration reduction device for a heavy-load helicopter provided in an embodiment of the present invention. Figure 33 A schematic diagram showing the acceleration of the mission cabin floor in the frequency domain and when unloaded, for a vibration reduction device for a heavy-load helicopter provided in an embodiment of the present invention. Figure 34 A schematic diagram showing the vibration reduction effect of a vibration reduction device for a heavy-load helicopter under full load in the time domain, as provided in an embodiment of the present invention; Figure 35 A comparative schematic diagram of the vibration reduction effect of a vibration reduction device for a heavy-load helicopter under full load, provided as an embodiment of the present invention; Figure 36 An acceleration cloud map of the mission cabin floor with a 5-ton load before and after vibration reduction, provided as an embodiment of the present invention, for a vibration reduction device for a heavy-load helicopter; Figure 37 An acceleration cloud map of the mission cabin floor with a 5-ton load behind the vibration reduction device for a heavy-load helicopter, provided as an embodiment of the present invention; Figure 38 A schematic diagram showing the acceleration comparison of a vibration reduction device for a heavy-load helicopter in the time domain with respect to the floor of a 5-ton load mission cabin, as provided in an embodiment of the present invention. Figure 39 A comparative diagram of acceleration of the floor of a 5-ton load mission cabin behind a vibration damping device for a heavy-load helicopter, provided as an embodiment of the present invention. Figure 40 The initial air pressure of a vibration damping device for a heavy-load helicopter under no-load conditions, as provided in an embodiment of the present invention. With itinerary A schematic diagram of the relationship curve; Figure 41 An embodiment of the present invention provides a vibration damping device for heavy-load helicopters, with initial air pressure under full load conditions. With itinerary A schematic diagram of the relationship curve.

[0019] Explanation of reference numerals in the attached figures: 1. Vibration damper; 11. Cylinder; 12. Magnetorheological fluid flow limiter; 13. Actuating piston; 14. Piston rod; 15. Tension-compression conversion assembly; 151. Top plate; 152. Bottom plate; 153. First connecting rod; 154. Second connecting rod; 155. First hinge bracket; 156. First hinge shaft; 157. Second hinge bracket; 158. Second hinge shaft; 159. Connecting arm; 16. Protective assembly; 161. Upper crank; 162. Lower crank; 2. Protective kit; 3. Floating piston; 4. Protective sleeve; 5. Pressure regulating valve. Detailed Implementation

[0020] The following detailed description of a specific embodiment of the present invention is provided in conjunction with the accompanying drawings. However, it should be understood that the scope of protection of the present invention is not limited to the specific embodiment.

[0021] In the description of this invention, it should be understood that the terms "center," "longitudinal," "lateral," "length," "width," "thickness," "upper," "lower," "front," "rear," "left," "right," "vertical," "horizontal," "top," "bottom," "inner," "outer," "axial," "radial," and "circumferential" indicate the orientation or positional relationship based on the orientation or positional relationship shown in the accompanying drawings. They are only for the convenience of describing the technical solution of this invention and simplifying the description, and do not indicate or imply that the device or element referred to must have a specific orientation, or be constructed and operated in a specific orientation. Therefore, they should not be construed as limitations on this invention.

[0022] refer to Figure 1 , Figure 2 and Figure 3 This invention provides a vibration damping device for heavy-duty helicopters, comprising multiple vibration dampers 1 for connecting a container to the helicopter. Each vibration damper 1 includes a cylinder 11, a magnetorheological fluid flow limiter 12 disposed within the cylinder 11, an actuating piston 13, and a piston rod 14. Figure 4 , Figure 5 and Figure 6As shown, the shock absorber 1 also includes: a tension-compression conversion assembly 15 and a protective assembly 16. The tension-compression conversion assembly 15 includes a top plate 151, a bottom plate 152, a plurality of first connecting rods 153 fixedly connected to the lower end of the top plate 151, and a plurality of second connecting rods 154 fixedly connected to the upper end of the bottom plate 152. Each first connecting rod 153 passes through the bottom plate 152 and is slidably connected to the bottom plate 152. Each second connecting rod 154 passes through the top plate 151 and is slidably connected to the top plate 151. The top plate 151 is fixedly connected to the upper end face of the cylinder 11. The lower end of the piston rod 14 is fixedly connected to the upper end face of the bottom plate 152. The lower ends of each first connecting rod 153 are connected to the first hinge of the container side wall. The first hinge pin 156 of the connector is hinged, and the upper end of each second connecting rod 154 is hinged to the second hinge pin 158 of the corresponding second hinge connector on the helicopter. It is connected to the helicopter through a second hinge connector. When the tension is applied, the high-pressure gas in the upper part of the cylinder 11 is converted into pressure by the piston rod 14 and is buffered and absorbed. The protective assembly 16 includes an inclined upper crank 161 and a lower crank 162. One end of the upper crank 161 is hinged to the first hinge pin 156, and the other end is hinged to one end of the lower crank 162. The other end of the lower crank 162 is hinged to the top plate 151 to provide tensile overload protection.

[0023] In the above embodiments, the present invention can achieve both tensile and compressive load bearing and vibration reduction. Specifically, through the top plate 151, bottom plate 152, first connecting rod 153, and second connecting rod 154, the high-pressure gas spring system, which could only withstand pressure, can convert external tension into internal pressure, thereby achieving buffering and absorption of tensile loads. This significantly improves the support, vibration reduction adaptability, and reliability of the helicopter under heavy load and complex vibration conditions. At the same time, the present invention can achieve the effect of extreme overload protection. Specifically, through the upper crank 161 and lower crank 162, the structure is activated when the tensile load exceeds a set threshold, forming a rigid connecting rod, which limits the further increase of displacement and effectively prevents the structure from being damaged due to the paralysis of the high-pressure gas spring system caused by extreme impacts, thus improving the overall system safety. Meanwhile, the present invention also has the advantages of compact structure and reliable connection. Specifically, the present invention achieves multi-point connection between the helicopter and the container through the combination of multiple connecting rods and hinges, distributing the load and enhancing the connection stiffness and stability.

[0024] During application, when the helicopter experiences an upward pull during flight, the pull is transmitted to the bottom plate 152 via the first connecting rod 153, pushing the piston rod 14 upward and compressing the high-pressure gas inside the cylinder, converting the pull into pressure. This pressure is absorbed by the air spring and the magnetorheological fluid flow limiter 12. The protective component 16 serves as the final safety barrier for the entire container system in the event of extreme impacts or accidental overloads. It is a purely mechanical linkage mechanism that remains in a non-load-bearing dormant state within its normal operating stroke. However, once the displacement of the top plate 151 and bottom plate 152 exceeds the safe range, the device is immediately activated and begins to bear load. Within the normal displacement range, the protective component 16 remains in a dormant state. Specifically, when the displacement... When the displacement exceeds 12mm, the crank mechanism gradually straightens, eventually forming a rigid limit at 16mm to prevent structural failure. If this safety range is exceeded, the device will be immediately activated and begin bearing load. When the displacement reaches the system's maximum design limit of 16mm, the upper crank 161 and lower crank 162 of the protective component 16 will be completely straightened to form a rigid connection, capable of withstanding tensile forces up to 300,000 Newtons. This allows the impact energy to be directly borne by the high-strength protective mechanism, preventing permanent damage to the core vibration damping unit and related structural components, ensuring flight safety. Specifically, for the protective component 16, this embodiment provides multiple actual state stages: initial state, dormant state, activated state, and locked state, wherein: In the initial state, the distance between the center of the upper crank 161 and the top plate 151 is 69.5 mm.

[0025] In the dormant state, the top plate 151 and the bottom plate 152 move 6mm relative to each other, and this distance is kept within the longitudinal height of 75.5. The overall vibration damper 1 is in the normal working range, and the protective component 16 does not intervene and is in the dormant state.

[0026] In the active state, when the displacement of the top plate 151 and the bottom plate 152 continues to increase, such that the top plate 151 and the bottom plate 152 move a distance of more than 6mm from each other, the protective component 16 is activated and begins to withstand external forces.

[0027] In the locked state, when the total distance the top plate 151 and the bottom plate 152 move relative to each other exceeds 16mm, the upper crank 161 and the lower crank 162 in the protective assembly 16 are completely straightened, and the distance between the center distance of the hole in the upper crank 161 and the top plate 151 is 79.5mm, at which point the protective locking is completed.

[0028] The high-pressure gas is used as an air spring module, and the pressure of the internal gas can be adjusted by opening and closing the pressure regulating valve 5 to achieve variable stiffness. Moreover, its magnetorheological fluid current limiter 12 module can achieve variable damping by adjusting the magnitude of the magnetic force. Combined with the structure of this invention, the optimal dynamic vibration isolation effect can be achieved. Specifically, to ensure absolute reliability, the pressure regulating valve 5 is selected as an aerospace-grade product with high pressure resistance and long service life. The core function of the air spring module is to provide basic bearing capacity to adapt to the wide weight variation of the container from 1600kg to 10000kg. By changing the system stiffness, the vibration isolation system can be matched to the optimal vibration isolation operating point under different loads. The magnetorheological fluid current limiter 12 module is located in the fluid cavity at the bottom of the cylinder 11. Based on the shear valve structure, it is filled with MRF-J25T magnetorheological fluid. By adjusting the current of the excitation coil in the current limiter through an external controller, the shear yield stress of the magnetorheological fluid can be changed in milliseconds, thereby realizing real-time and precise adjustment of the damping force to effectively suppress the specific frequency vibration of 16.2Hz.

[0029] Working principle of dynamic vibration buffering and isolation: First, in the static equilibrium stage, i.e., the initial load, when the container is under a stable load state during ground or flight, the air spring module plays a major supporting role. The external load can always be applied to the air spring in the form of compression force through the structure of this invention. The initial pressure in the air chamber generates a supporting force. When this supporting force reaches equilibrium with the external load, the entire device stabilizes at a certain equilibrium position. In order to achieve the function of variable stiffness, the pressure regulating valve 5 and the controllable gas channel allow the cavity pressure of the air spring inside the system to be flexibly adjusted according to the actual working conditions. By adjusting the internal pressure of the air chamber, the equivalent stiffness of the elastic element can be directly changed, thereby achieving dynamic matching of the system stiffness under different load conditions. In the static equilibrium stage, the magnetorheological fluid current limiter 12 is in a low-energy standby state. The damping force is mainly determined by the zero-field viscosity. Specifically, the basic damping when there is no external magnetic field is used. At this time, the excitation coil current I=0, and the magnetorheological fluid maintains a low-damping state to reduce static energy consumption.

[0030] Secondly, in the dynamic vibration suppression stage, i.e. forced vibration, during flight, when the container moves up and down due to external excitation, the piston rod 14 in the actuation assembly causes the actuating piston 13 to reciprocate within the cylinder 11. The magnetorheological fluid is forced to flow through the damping channel fixed in the middle of the cylinder 11. The magnetic induction intensity at the damping channel can be adjusted according to the change of the excitation coil current of the magnetorheological fluid current limiter 12, thereby changing the shear yield stress of the magnetorheological fluid flowing through the damping channel and realizing the control of the damping force. The vibration controller will analyze the amplitude and phase of the vibration in real time according to the sensor signal, and adjust the current of the excitation coil through the control algorithm to generate a precise damping force to resist the movement of the actuating piston 13, thereby realizing the active suppression of the 16.2Hz vibration.

[0031] Further, refer to Figure 6 , Figure 7 and Figure 11 Each of the lower ends of the first link 153 and the upper ends of the second link 154 is fixedly connected with a protective kit 2 for pressure limiting protection.

[0032] In the above embodiments, the protective kit 2 can provide cushioning when the top plate 151 or the bottom plate 152 is close to the limit position, avoid direct metal collision, achieve the purpose of pressure protection, and also reduce noise and wear. Through limit protection, it can prevent the structure from plastic deformation or fatigue damage due to overload impact.

[0033] Further, refer to Figure 1 The number of first link 153 and second link 154 in the shock absorber 1 is greater than or equal to two.

[0034] In the above embodiments, by providing each damper 1 with multiple first links 153 and second links 154, the load can be transmitted more evenly, avoiding excessive force at a single point, and improving the stability and reliability of the overall structure. Even if a single first link 153 or second link 154 malfunctions, the remaining first links 153 or second links 154 can still maintain the system function, improving the system robustness. In this embodiment, it is preferable to provide two first links 153 and two second links 154 in each damper 1.

[0035] Further, refer to Figure 1 and Figure 9 The first hinge component includes a first hinge bracket 155, which is welded and fixed to the container, and a first hinge shaft 156 is laterally connected to the first hinge bracket 155.

[0036] In the above embodiments, the first hinge bracket 155 is welded and fixed to the container, providing a solid connection foundation, adapting to heavy-load conditions, and improving the strength of the connection.

[0037] Further, refer to Figure 1 , Figure 10 and Figure 12 The second hinge component includes a second hinge bracket 157, which is welded and fixed to the helicopter. The second hinge shaft 158 ​​is hinged to the second hinge bracket 157 via a connecting arm 159. The connecting arm 159 is hinged to the second hinge bracket 157, and its other end is sleeved and rotatably connected to the second hinge shaft 158.

[0038] In the above embodiments, the second hinge bracket 157 is connected to the helicopter via the connecting arm 159, allowing for a certain range of adaptation to attitude changes during helicopter flight. Moreover, the hinge structure can reduce additional bending moments caused by misaligned installation and extend component life.

[0039] Further, refer to Figure 1 Each first link 153 is ball-jointed to the corresponding first hinge 156, and the end of each second link 154 is ball-jointed to the corresponding second hinge 158.

[0040] In the above embodiments, the ball joint connection allows the connecting rod to swing slightly in all directions, adapting to installation errors and dynamic deformation, and also has the effect of reducing stress concentration. Specifically, it avoids excessive local stress caused by rigid connection, improves fatigue strength, and the ball bearing of the ball joint connection can ensure that the device can withstand a pressure of 160,000 Newtons, thereby directly absorbing the impact energy by the high-strength protection mechanism, preventing permanent damage to the core vibration damping unit and related structural components, and ensuring flight safety.

[0041] Further, refer to Figure 6 The protective kit 2 has a wedge-shaped sleeve structure and is welded and fixed to the corresponding first link 153 or second link 154. When the top plate 151 approaches the protective kit 2 of the second link 154, the protective kit 2 plays a buffering and blocking role for the top plate 151.

[0042] In the above embodiments, the wedge structure has good guiding properties, which facilitates smooth contact when approaching the limit position and avoids jamming. It not only plays a buffering role, but also acts as a hard limiting mechanism to protect the structural safety.

[0043] Further, refer to Figure 8 The top plate 151 is covered with a protective sleeve 4 to prevent the internal structure from being damaged by impact. The lower end of the protective sleeve 4 is fitted over the bottom plate 152 and is movably connected to the bottom plate 152.

[0044] In the above embodiments, the protective cover 4 can prevent foreign objects from entering the moving parts, reduce wear and malfunctions, and at the same time, the outer cover structure is clean and prevents personnel from accidentally touching the moving parts.

[0045] Further, refer to Figure 4 and Figure 5Inside the cylinder 11, above the magnetorheological fluid current limiter 12, there is a floating piston 3 that is movably connected to block the high-pressure air and magnetorheological fluid inside the cylinder 11, preventing gas from mixing into the magnetorheological fluid current limiter 12. The magnetorheological fluid current limiter 12 dynamically changes the shear yield stress of the magnetorheological fluid by controlling the current of the internal excitation coil, in order to cope with the vibration force under overload load.

[0046] In the above embodiments, the floating piston 3 effectively isolates the high-pressure gas from the magnetorheological fluid, preventing excessive mixing of the high-pressure gas and the magnetorheological fluid from affecting the overall damping performance. By adjusting the excitation coil current, the shear stress of the magnetorheological fluid can be changed in real time, realizing active control of damping and adapting to different vibration frequencies and amplitudes.

[0047] Specifically, a cable channel for sealing and wrapping its wires is provided above the magnetorheological fluid choke. The cable channel extends out of the top plate 151 and is sealed with the top plate 151. The floating piston 3 is sleeved outside the cable channel and a dynamic sealing ring is provided between the two. The cable can be sealed without affecting the floating piston 3. The cable channel can prevent the wires from contacting the magnetorheological fluid and shortening the wire life.

[0048] Further, refer to Figure 4 A pressure regulating valve 5 is provided on the upper side of the cylinder 11 to regulate the pressure of the high-pressure air inside the cylinder 11, which is used to achieve dual active control of damping stiffness with the magnetorheological fluid flow limiter 12.

[0049] In the above embodiments, the pressure in the gas chamber is adjusted by the pressure regulating valve 5, thereby changing the overall stiffness of the high-pressure gas spring system. Combined with the magnetorheological fluid damping adjustment, the dynamic characteristics of the system are fully optimized.

[0050] Table 1 - Summary of Overall Parameters of the Device Table 2 - Parameters of Each Component Table 3 - Parameters of Main Components of Protective Component 2 Device load-bearing verification Tensile strength verification: According to GJB2150A-2015, as shown in Tables 1, 2, and 3, the typical yield stress of the 30CrMnSiA material used in the key components of protective assembly 16, after heat treatment, can reach [value missing]. .

[0051] According to design requirements, the protective component 16 of the vibration control system needs to withstand a total tensile force of 300,000 Newtons. The net cross-sectional area of ​​the upper crank 161 is: ; In the above formula: — Upper crank 161 outer diameter; —The upper crankshaft has a large bore diameter of 161mm; —Thickness of protective component 2.

[0052] The maximum tensile force that this component can theoretically withstand is: ; In the above formula: Yield strength of 30CrMnSiA.

[0053] The theoretical maximum load-bearing capacity is far greater than the tensile design requirement of 300,000 Newtons, and the strength meets the project requirements.

[0054] Stress verification: According to GJB2150A-2015, the yield stress of 30CrMnSiA material is... The net cross-sectional area of ​​the ball bearing limit is: ; In the above formula: —Diameter of first link 153 / second link 154, in mm.

[0055] If there are 2 rods, the maximum pressure that it can withstand is: ; The theoretical maximum bearing capacity is far greater than the 160,000 Newtons pressure design requirement, and the strength meets the project requirements.

[0056] Table 4 - Basic Parameters of Air Spring-Magnetorheological Fluid Current Limiter 12, etc. As shown in Table 4, the stiffness characteristic of air springs is non-linear; their stiffness increases with increasing compression. This non-linear characteristic results in air springs having low stiffness at small amplitudes, providing good comfort, and high stiffness at large amplitudes, providing sufficient support. Specifically: The magnetorheological fluid current limiter 12 utilizes the properties of magnetorheological fluid to adjust the rheological properties of the liquid according to the changes in the external magnetic field, thereby achieving precise control of vibration. Under the action of an external magnetic field, the magnetorheological fluid rapidly changes its viscosity or shear stress, enabling the overall structure to provide the required damping force under different working conditions. By selecting an appropriate magnetorheological fluid and designing a reasonable magnetorheological fluid current limiter 12 structure, the vibration response of the system can be effectively reduced, and the stability of the helicopter under different working conditions can be improved.

[0057] In this study, the commonly used MRF-J25T magnetorheological fluid was selected. The shear characteristics of the MRF-J25T magnetorheological fluid are as follows: Figure 13 As shown, this curve characterizes the existing relationship between shear stress and magnetic induction intensity. In the initial stage, the slope of the curve is small, corresponding to the process of particles beginning to form chains. Subsequently, it enters a relatively long, approximately linear growth region, at which point the engineering magnetic field (current) is almost proportional to the output force, such as... Figure 14 As shown in the diagram, T in the icon represents Tesla, the unit of magnetic flux density. The X-axis represents the velocity of fluid deformation, and the Y-axis (shear stress) represents the internal frictional resistance generated during fluid flow. As the magnetic flux density increases, the curve shifts upwards overall. This means that at the same shear rate, the stronger the magnetic field, the greater the shear stress generated by the magnetorheological fluid.

[0058] Table 5 - Performance parameters of magnetorheological fluids As shown in Table 5, the main components include ferromagnetic particles and hydrocarbon oil. Ferromagnetic particles change the fluidity of the liquid under the action of a magnetic field and provide adjustable shear stress, while hydrocarbon oil ensures the fluidity and stability of the liquid. The combination of these two gives the magnetorheological fluid excellent performance under different working conditions.

[0059] like Figure 15 As shown, when the system is in the critical resonance region, that is, when the frequency ratio is close to 1, the maximum damping force required is about 6398N to ensure that the vibration response of the container does not exceed 5 times the input excitation acceleration, thus meeting the project's control requirements for stability during the initial startup phase.

[0060] Table 6 - Design characteristics of magnetorheological fluid current limiter 12 The reference value for damping force is shown in Figure 7.

[0061] Table 7 - Reference values ​​for damping force (error ±7.5%) To verify whether the design can meet the dynamic damping force output requirements in the vibration isolation control of the system, as shown in Table 7, the damping force data obtained from the actual test, as shown in Table 6, were compared and analyzed with the dynamic design requirements of the system. It can be found that, as shown in Table 7, the maximum damping force is 6398N, and 6831N has been output at a speed of 1047mm / s, exceeding the design target. The overall curve is stable and monotonic, and the linear performance in the low-speed range is good. The results verify that the magnetorheological fluid current limiter 12 has the ability to meet the maximum damping requirements of the system. Its dynamic range, adjustment linearity and peak capacity all meet the control requirements of the vibration isolation system under extreme resonance conditions.

[0062] Regarding the static stiffness of the installation 0) Theoretical analysis and simulation verification 0) Key Indicators The stiffness of the entire vibration reduction system designed for the project can adapt to the variation of the container's design weight (empty weight) from 1600kg to the maximum design weight of 10000kg (a single device can adapt to a variation of 400kg-2500kg). Considering different weight variations, the stiffness is designed for vibration isolation, and the damping can be adjusted. Under the maximum weight of the container, the static stiffness is designed with a 12mm settlement after installation, and a protective component 16 (limiting structure, displacement limit of 4mm) is designed to prevent large deformation under loads such as accidental impacts. The vertical tensile strength of the protective component 16 structure (or other types of structure) is designed according to 300,000 Newtons, and the vertical compressive strength is designed according to 160,000 Newtons.

[0063] 2) The necessity of variable stiffness design Assuming a container's maximum mass of 10 tons (fully loaded), four air springs are needed to provide sufficient support force to achieve static equilibrium. At static equilibrium, the total force of the air springs must equal the weight of the container.

[0064] ; In the above formula: —Stiffness of a single spring; —Air spring tension; —Container quality; —Acceleration due to gravity, take .

[0065] like Figure 16 As shown, the required stiffness of a single air spring varies with different stretching amounts. Point A (12.00, 2046) indicates that when the air spring is fully loaded and the initial stretching amount is within 12mm, the stiffness of a single air spring should be designed to be above 2046N / mm. This is the stiffness requirement under static conditions and also serves as the basis for subsequent dynamic design.

[0066] The damping effect of multiple magnetorheological fluid current limiters 12 connected in parallel can be summed and expressed as the total damping coefficient. frequency ratio Damping ratio The transmissivity of the vibration isolation system is: ; like Figure 17As shown, in order to study the influence of damping ratio on the transmissibility curve, several representative values ​​were selected from 0.001 to 5, and multiple transmissibility curves were plotted. The vibration isolation curve shows that the frequency ratio corresponding to point A (0.99, 5.10) is about 0.99. At this time, the system resonates. If the damping ratio is lower, the vibration transmissibility of the system reaches about 5.10. That is, when the rotor starts and the excitation frequency instantaneously crosses the natural frequency of the airframe, the acceleration of the container will be instantly amplified to 5.1 times the acceleration of the airframe. If the damping ratio is further reduced, the vibration transmissibility will continue to increase. Therefore, when the excitation frequency is close to the natural frequency of the container during the start-up phase, it is very easy to cause severe resonance.

[0067] The natural frequency of the entire container connection system is determined by both the stiffness of the connecting springs and the mass of the container. When the mass is 1600kg (unloaded), if an air spring is used under the same full-load condition, then a single spring... The stiffness is 2046 N / mm, and the total equivalent stiffness is... The stiffness is 8184 N / mm, and the natural frequency of the system is: ; Combination such as Figure 17 As shown in the vibration isolation curve, when the ratio of the external excitation frequency to the system's natural frequency is... Greater than At point B, the system vibration isolation takes effect. For the single spring stiffness of 2046 N / mm that meets the requirement of 12 mm compression under full load in the preliminary design, the system frequency is higher than that under no-load conditions. The vibration is far below the vibration isolation threshold. Since the main rotor excitation is 16.2Hz, regardless of the damping adjustment, the vibration transmissibility remains greater than 1, exhibiting a significant vibration amplification phenomenon. Therefore, for the single spring stiffness in the above preliminary design... With a fixed spring stiffness of 2046 N / mm, vibration isolation fails. The core reason for this failure is that the excessive spring stiffness raises the system's natural frequency under no-load conditions, making the ratio of the excitation frequency to the natural frequency insufficient to enter the isolation zone. Although this stiffness design can meet the support requirements of the container under static load, its dynamic stiffness is significantly too high, resulting in a substantial decrease in vibration isolation capability. Therefore, to achieve good vibration isolation while ensuring static load safety, it is necessary to coordinately optimize the spring stiffness and damping parameters: ensuring that the system's natural frequency is lower than the main excitation frequency by a specific ratio, and further weakening vibration transmission through reasonable damping, thereby maintaining stable and efficient vibration isolation performance under different working conditions.

[0068] Optimal Stiffness Principle According to vibration isolation theory, when the ratio of the excitation frequency to the natural frequency... The vibration isolation effect is particularly significant when the frequency is between 2 and 3 Hz. Below this range, vibration isolation failure may occur, while excessively high frequencies will reduce the efficiency of vibration isolation. Therefore, for the main rotor excitation frequency of 16.2 Hz in this project, the initial consideration should be to control the system's natural frequency between 5.34 Hz ​​and 8.1 Hz to ensure that the vibration isolation zone remains effective and to suppress vibration transmission to the greatest extent possible. Figure 18 As shown, the spring stiffness range can be deduced from the natural frequency formula: when the load mass is 1600kg (unloaded), the design stiffness of a single spring is... The range is 451.81-1042.87 N / mm. When the load mass is 10000 kg (full load), the design stiffness of a single spring is... The range is 2806.33-6473.29 N / mm.

[0069] like Figure 19 As shown, under different load conditions, in order to meet the frequency ratio The requirements for individual spring design stiffness ranges do not overlap, indicating that using fixed-stiffness air springs makes it difficult to achieve vibration isolation performance under both operating conditions, posing a significant risk of failure. Therefore, to achieve effective vibration reduction control under all operating conditions, a programmed method must be adopted, combined with the transmissibility function, to comprehensively analyze the minimum 60% vibration reduction efficiency (i.e., transmissibility) that can be achieved under different damping ratios. The optimal spring stiffness range when ≤0.4).

[0070] Stiffness is maximized to meet design requirements (efficiency meets design requirements). Calculations show that the relationship between the stiffness coefficient and transmissivity of a single spring in the vibration isolation system, under both no-load and full-load conditions, satisfies a minimum 60% vibration isolation effect. Figure 20 The container shown is empty. Figure 21 The container shown is in a full state. Under two typical mass conditions and different damping ratios, the transmission rate is satisfied. The corresponding design range for the stiffness of a single air spring, such as Figure 20 As shown, when the container is empty, the damping ratio is... When equal to 0.05, 0.1 and 1, Only then can it be true. Point A (1135, 0.4) indicates that when the container is empty, when... When the value is 0.1, the maximum stiffness of a single spring cannot exceed 1135 N / mm to ensure the transmission rate. That is, the vibration reduction efficiency is above 60%, such as Figure 21 Point B (7097, 0.4) in the diagram indicates that when the container is fully loaded, when... When the value is 0.1, the maximum stiffness of a single spring cannot exceed 7097 N / mm in order to ensure that the vibration reduction efficiency is above 60%.

[0071] Minimum stiffness satisfies design (stroke satisfies design). To address the issues of inconsistent vibration isolation efficiency and difficulty in unifying the system's operating range under different load conditions, the vibration isolation system of this invention, while maintaining good support capabilities, can dynamically change its stiffness through internal air pressure adjustment, thereby expanding the vibration isolation frequency band and adapting to different load conditions.

[0072] Due to the adjustable stiffness of air springs, adjusting them to a suitable stiffness under different load conditions can meet vibration reduction requirements. However, to ensure that the container maintains static balance under different load conditions and that the system stroke does not exceed the limit, the four air springs need to provide sufficient support force within a given maximum extension limit (12mm) to achieve static balance. At this point, the total force of the springs is equal to the weight of the container. ; In the above formula: —Stiffness of a single air spring; —Air spring tension; —Container quality; —Acceleration due to gravity, take .

[0073] Considering the required stiffness of a single spring under different operating conditions (fully loaded and unloaded), and assuming a maximum container mass of 10 tons, the stiffness of a single spring corresponding to different amounts of tension was obtained, such as... Figure 22 As shown, point A (12.00, 2046) represents the point where the spring tension is... Stiffness of a single spring at 12mm The minimum design value should be 2046 N / mm. Therefore, under full load, the stiffness of a single air spring should be adjusted within the range of 2046 N / mm to 7097 N / mm to ensure that the system's tensile strength does not exceed 12 mm and the vibration reduction efficiency is higher than 60%.

[0074] Based on a minimum container mass of 1.6 tons, the required stiffness of a single air spring is as follows: Figure 23 As shown, point B (12.00, 326) represents the point where the spring tension is limited. When the thickness is 12mm, the stiffness of a single spring The minimum design value should be 326 N / mm. Therefore, under no-load conditions, the stiffness of a single air spring should be adjusted within the range of 326 N / mm to 1135 N / mm to ensure that the system's tensile strength does not exceed 12 mm and the vibration reduction efficiency is higher than 60%.

[0075] Implementation of static stiffness installation The core of an air spring lies in the ideal gas law and polynomial process model followed by its gas chamber. When the dynamometer piston produces a small displacement, the pressure and volume of the gas chamber undergo coupled changes. Near the static equilibrium position, the gas law can be approximated as: ; In the formula: —Initial pressure of the air chamber; —Initial volume of the air cavity; --constant; —The adiabatic coefficient of a gas reflects the temperature changes from an isothermal process ( From an approximately adiabatic process ( The different thermodynamic properties of ).

[0076] At the static design level, in order for the air springs to both bear the weight of the container under different load conditions and provide suitable initial stiffness for dynamic vibration isolation, it is usually necessary to solve for the "preload pressure" in reverse by combining the static equilibrium equations with the stiffness equations. "and cavity volume" "These two key parameters are: the total supporting force of the four parallel air springs must be equal to the weight of the cabin, and the small displacement stiffness of a single air spring must meet the design requirements for the natural frequency."

[0077] After entering the dynamic working condition, the stiffness of the air spring is not constant, but changes nonlinearly with the displacement of the actuating piston 13: as the gas volume increases or decreases, its stiffness will decrease or increase. It is this nonlinear characteristic that allows the natural frequency of the system to be smoothly tuned between the no-load and full-load ends by simply adjusting the air pressure.

[0078] Based on a maximum container weight of 10 tons carrying the mission items, four air springs are needed to provide sufficient support force to achieve static equilibrium. At static equilibrium, the total force of the air springs must equal the weight of the container. ; In the formula: —Effective cross-sectional area of ​​the air spring; —Initial absolute pressure; —Static off-center displacement of the air spring; —Initial volume of the internal air cavity; —Gas adiabatic coefficient; —Atmospheric pressure; —Container quality; —Acceleration due to gravity.

[0079] The state of the gas within the cavity approximately follows a polynomial process: ; In the formula: —Displacement, a positive value indicates upward compression by the piston; volume Decrease.

[0080] The dynamic stiffness of a single air spring is defined as the first derivative of force with respect to displacement: ; because ,have to: ; The expression for equivalent stiffness is: ; Assuming the initial air chamber height is 50mm and the cross-sectional diameter is... The thickness is 50 mm, and the gas insulation coefficient is... It is 1.4, atmospheric pressure The initial air pressure is 101325 Pa, which meets the requirements of both no-load and full-load operating conditions. With equivalent stiffness Relationship such as Figure 24 As shown, under no-load conditions, if the diameter If the value is 50mm, then the maximum displacement limit is met. The initial air pressure is 12 mm. It needs to be above 1.4 MPa, from which... Figure 24 As shown in the figure, by precisely adjusting the initial pressure This can achieve the equivalent stiffness of the device. It is continuously adjustable to achieve the vibration reduction stiffness requirement of more than 326 N / mm.

[0081] Under full load conditions, if the diameter If the value is 50mm, then the maximum displacement limit is met. The initial air pressure is 12 mm. It needs to be above 8.5 MPa, such as Figure 25 As shown in the figure, by precisely adjusting the initial pressure This can achieve the equivalent stiffness of the device. The continuous adjustability enables the vibration reduction stiffness requirement to reach over 2046 N / mm.

[0082] Vibration reduction efficiency estimation In the design of helicopter vibration isolation systems, effectively controlling the dynamic coupling and vibration transmission between the container and the fuselage is crucial to ensuring equipment performance and personnel comfort. This invention analyzes a typical vibration isolation system, revealing its damping force requirements during the rotor start-up and stable cruise phases, and estimating the corresponding vibration reduction efficiency.

[0083] To study the system's response to different damping forces under typical spring stiffness, the stiffness of a single spring in the no-load condition was set to [value missing]. The stiffness of a single spring under full load is set to Then the natural frequencies of the system under no-load and full-load conditions are: ; At a rotor frequency of 16.2Hz, the ratio of the natural frequencies under no-load and full-load conditions is... , ,like Figure 26 As shown, the transmissibility curve of the system under this stiffness is displayed, and the vibration transmission under different damping ratios is compared between no-load and full-load conditions.

[0084] For helicopter container systems, vibration control can be divided into two stages: 1. Startup and operation phase During the process of the rotor starting from rest and gradually accelerating to 16.2Hz, its excitation frequency will inevitably sweep across the system's natural frequency. , When the excitation frequency approaches the natural frequency, the system will enter a severe resonance state, such as... Figure 26 As shown, the frequency ratio is represented in the middle. Vibration at that time, with damping ratio In the micro-damped state, such as Figure 26 As shown, point A (0.99, 10.6) indicates that the vibration transmissibility at this point is... This means that a 1g vibration input to the machine body will be amplified more than 10 times, causing the container to experience a violent acceleration of 10.06g. This level of vibration is extremely harmful to equipment and personnel. To safely pass through the resonance zone, the damping must be increased, as shown at point B (0.97, 5.04) in the figure, by changing the damping ratio from... Increase to At that time, the resonant peak transmissibility dropped significantly to 5.04, and the vibration response decreased by nearly 50%, which proves that high damping is a necessary means to suppress resonance and protect the system during the startup phase.

[0085] 2. Stable cruise phase When the helicopter enters a stable flight state, the rotor frequency stabilizes at 16.2Hz, and the system operates above the vibration isolation zone of the resonance zone. At this point, the vibration isolation system begins to play its main role, such as... Figure 26 As shown, when the damping ratio is Under the given conditions, the vibration transmissibility of the system under no-load (point C) condition is 0.226, and the vibration reduction efficiency is 77.4%; under the full-load (point D) condition, the vibration transmissibility of the system is... The damping ratio is 0.158, resulting in a vibration reduction efficiency of 84.2%. However, at this stage, the situation is exactly the opposite of the resonance zone; a smaller damping ratio will achieve better vibration reduction efficiency, such as... Figure 26 As shown, as the damping ratio decreases, the transmission rate decreases, and the vibration reduction efficiency of the system will be further improved. This means that in order to achieve the ultimate vibration isolation performance during cruise, the project intends to control the damping to be as small as possible. Therefore, the ceiling control strategy of the magnetorheological fluid current limiter 12 is adopted to precisely adjust the required damping at different stages.

[0086] Simulation verification Verification and analysis of vibration reduction under equal load on containers To study the vibration reduction effect of the vibration damping device at a specific frequency, simulation analysis was conducted for two working conditions: no-load (1.6t) and full-load (10t), and the ideal stiffness range of a single spring under each working condition was determined: No-load condition (1.6t): To ensure effective vibration isolation and system stability, the stiffness of a single spring must be controlled within the range of 326 N / mm to 1135 N / mm.

[0087] Full load condition (10t): With the significant increase in load, in order to maintain sufficient load-bearing capacity and vibration reduction performance, the stiffness of a single air spring needs to be increased to the range of 2046N / mm to 7097N / mm.

[0088] The specific settings are as follows: No-load operating condition model: The damping coefficient of a single magnetorheological fluid current limiter 12 is set to ; The stiffness of a single air spring is set to .

[0089] Full load condition model: The damping coefficient of a single magnetorheological fluid current limiter 12 is set to ; The stiffness of a single air spring is set to .

[0090] In the simulation analysis platform, the project first defined a nonlinear damping unit that can accurately simulate the characteristics of the magnetorheological fluid current limiter 12. Subsequently, four sets of vibration reduction devices, which are integrated with air springs and magnetorheological fluid current limiters 12, were precisely arranged at the four predetermined connection points between the container and the fuselage in the simulation model.

[0091] Add a dual-frequency sine wave exciter to the fuselage Specifically, such as Figure 27 As shown, a loading duration of 1 second was selected for transient analysis. The velocity at the connection surface between the container and the fuselage was extracted for calculating the damping force, and the floor vibration response acceleration was used to calculate the vibration reduction efficiency. After the calculation, the floor acceleration contour maps before and after vibration reduction under no-load conditions at time 1 second were extracted, as shown below. Figure 28 As shown, the floor acceleration contour maps before and after vibration reduction under full load conditions at time 1 second are as follows. Figure 29 As shown, the time-domain analysis of the vibration reduction effect under no-load conditions is as follows. Figure 30 As shown, after the system vibration response tends to stabilize, at the rotor frequency... and tail fin frequency The vibration reduction effect at the location is as follows Figure 31 As shown.

[0092] like Figure 32 As shown, the vibration response of the container floor under no-load conditions exhibits a high peak acceleration (1.143g) without the dual-frequency canopy control vibration damping device. This is due to the superposition effect of the frequency excitations of the rotor and tail rotor. Meanwhile, the change in the steady-state time-domain response indicates that the dual-frequency canopy magnetorheological fluid current limiter 12 can quickly adapt to complex working conditions and respond promptly to the frequency excitation characteristics of the system. The optimized damping ratio ensures sufficient damping force output under different frequency excitations, thereby effectively suppressing the increase in vibration amplitude caused by the superposition of multimodal responses. Figure 33 As shown, with the introduction of the dual-frequency ceiling magnetorheological fluid current limiter 12, the vibration was significantly attenuated at both target frequencies, and the peak acceleration was reduced to 0.1046g (16.2Hz) and 0.0127g (71.7Hz), respectively, corresponding to 80.1% and 82.4% of the suppression amplitude.

[0093] Under full load conditions, the corresponding vibration reduction effect in the time domain is as follows: Figure 34 As shown, after the system vibration response tends to stabilize, at the rotor frequency... and tail fin frequency The vibration reduction effect at the location is as follows Figure 35 As shown.

[0094] like Figure 34 As shown, the vibration decay time under full load is slightly longer than under no-load conditions, but the vibration reduction device of this invention still has a good vibration reduction effect. Figure 35As shown, the vibration was significantly attenuated at both target frequencies, with peak accelerations reduced to 0.1536g (16.2Hz) and 0.0229g (71.7Hz), respectively, corresponding to suppression amplitudes of 71.7% and 79.4%. This significant vibration reduction effect demonstrates that the excitation of the helicopter rotor and tail rotor was effectively suppressed after parameter optimization, reflecting the efficient vibration control capability of the dual-frequency ceiling magnetorheological fluid current limiter 12 under multimodal coupling conditions.

[0095] Vibration reduction verification analysis of uneven load distribution in container front and rear To study effective vibration reduction under uneven load distribution, the working condition of container center of gravity shift was simulated. Under the premise that the container was unloaded (1.6t), an additional 5 tons of load was applied to the rear half of the container, while the front half remained unloaded. This asymmetrical loading caused the overall center of gravity of the container to shift backward, which put forward higher requirements for the adaptability and stability of the vibration reduction system.

[0096] Under this condition, the parameters of the two sets of vibration damping units were set asymmetrically: In the rear section (heavy load area), to effectively support the increased weight and suppress greater vibration energy, the damping coefficients of the two rear magnetorheological fluid current limiters 12 are both set to [value missing]. The stiffness of the two air springs at the rear of the fuselage was set to [stiffness value missing]. ; First half (light load zone): To match the lighter load and avoid excessively stiff dampers that could worsen vibration transmission, the total damping coefficient of the two front magnetorheological fluid current limiters 12 is set to [value missing]. The stiffness of the two front air springs was set to [value missing]. .

[0097] In the simulation platform, the COMBIN14 element is used to simulate the spring and the magnetorheological fluid current limiter 12. By assigning different stiffness and damping real constants to the two sets of elements respectively, an asymmetric vibration reduction model is constructed. Transient dynamic analysis is carried out by applying the same external excitation, focusing on the system's ability to suppress vertical vibration under the asymmetric load, so as to verify the adaptive effect of the vibration reduction system on uneven load.

[0098] After the calculation is completed, extract the floor acceleration contour maps before and after vibration reduction at time 1 second, as shown below. Figure 36 As shown, due to the lower mass at the front, the vibration amplitude is more severe compared to the rear. Adding a vibration damping device significantly improves the vibration. The corresponding time-domain analysis of the vibration damping effect is shown below. Figure 36 As shown, after the system vibration response tends to stabilize, at the rotor frequency... and tail fin frequency The vibration reduction effect at the location is as follows Figure 37 As shown.

[0099] like Figure 38 As shown, when the rear weight is heavier, the vibration response of the container floor exhibits a higher peak acceleration (1.136g) without the dual-frequency ceiling-controlled vibration damping device. Meanwhile, the change in the steady-state time-domain response indicates that the dual-frequency ceiling magnetorheological fluid current limiter 12 can quickly adapt to complex working conditions and respond promptly to the frequency excitation characteristics of the system. The optimized damping ratio ensures sufficient damping force output under different frequency excitations, thereby effectively suppressing the vibration amplitude enhancement caused by the superposition of multimodal responses. Figure 39 As shown, with the introduction of the dual-frequency ceiling magnetorheological fluid current limiter 12, the vibration was significantly attenuated at both target frequencies, with the peak acceleration reduced to 0.1136g (16.2Hz) and 0.0198g (71.7Hz), respectively, corresponding to 76.1% and 80.2% of the suppression amplitude. Under different operating conditions, by optimizing the stiffness parameter domain of the air spring and selecting an appropriate ceiling damping coefficient, the vibration reduction control effect of the helicopter structure under dual-frequency excitation can be significantly improved, verifying the effectiveness of dual-frequency ceiling damping control in the key frequency range.

[0100] Assuming the initial chamber height is 50mm, the cross-sectional diameter is 50mm, and the gas adiabatic coefficient is... It is 1.4, atmospheric pressure The initial air pressure is 101325 Pa, which meets the load requirements for both no-load and full-load conditions. With itinerary Relationship such as Figure 40 As shown, under no-load conditions, if the diameter If the value is 50mm, then the maximum displacement limit is met. The initial air pressure is 12 mm. It needs to be above 1.4 MPa, such as Figure 41 As shown, under full load conditions, if the diameter If the value is 50mm, then the maximum displacement limit is met. The initial air pressure is 12 mm. It needs to be above 8.5 MPa.

[0101] The above-disclosed embodiments are merely a few specific examples of the present invention. However, the embodiments of the present invention are not limited thereto, and any variations that can be conceived by those skilled in the art should fall within the protection scope of the present invention.

Claims

1. A vibration damping device for a heavy-duty helicopter, comprising a plurality of vibration dampers for connecting a container to the helicopter, each vibration damper comprising a cylinder, a magnetorheological fluid flow limiter disposed within the cylinder, an actuating piston, and a piston rod, characterized in that, The vibration damper also includes: The tension-compression conversion assembly includes a top plate, a bottom plate, a plurality of first connecting rods fixedly connected to the lower end of the top plate, and a plurality of second connecting rods fixedly connected to the upper end of the bottom plate. Each first connecting rod passes through the bottom plate and is slidably connected to the bottom plate. Each second connecting rod passes through the top plate and is slidably connected to the top plate. The top plate is fixedly connected to the upper end face of the cylinder. The lower end of the piston rod is fixedly connected to the upper end face of the bottom plate. The lower end of each first connecting rod is hinged to the first hinge shaft of the first hinge member of the container sidewall. The upper end of each second connecting rod is hinged to the second hinge shaft of the corresponding second hinge member on the helicopter. The assembly is connected to the helicopter through a second hinge member. When tension is applied, the high-pressure gas in the upper part of the cylinder is converted into pressure by the piston rod and is buffered and absorbed. The protective assembly includes an inclined upper crank and a lower crank, one end of the upper crank being hinged to a first hinge shaft and the other end being hinged to one end of the lower crank, and the other end of the lower crank being hinged to the top plate, for providing tensile overload protection.

2. A vibration damping device for heavy-duty helicopters as described in claim 1, characterized in that, The lower end of each of the first links and the upper end of each of the second links are fixedly connected with protective kits for pressure limiting protection.

3. A vibration damping device for heavy-duty helicopters as described in claim 1 or 2, characterized in that, The vibration damper has two or more first and second connecting rods.

4. A vibration damping device for heavy-duty helicopters as described in claim 3, characterized in that, The first hinge component includes a first hinge bracket, which is welded and fixed to the container, and the first hinge shaft is laterally connected to the first hinge bracket.

5. A vibration damping device for heavy-duty helicopters as described in claim 4, characterized in that, The second hinge component includes a second hinge bracket, which is welded and fixed to the helicopter. The second hinge shaft is hinged to the second hinge bracket via a connecting arm. The connecting arm is hinged to the second hinge bracket, and its other end is sleeved and rotatably connected to the outside of the second hinge shaft.

6. A vibration damping device for heavy-duty helicopters as described in claim 2, characterized in that, Each of the first links is connected to the corresponding first hinge axis ball joint, and the end of each of the second links is connected to the corresponding second hinge axis ball joint.

7. A vibration damping device for heavy-duty helicopters as described in claim 2, characterized in that, The protective kit has a wedge-shaped sleeve structure and is welded and fixed to the corresponding first or second link. When the top plate approaches the protective kit of the second link, the protective kit acts as a buffer and barrier to protect the top plate.

8. A vibration damping device for heavy-duty helicopters as described in claim 1, characterized in that, The top plate is covered with a protective sleeve to prevent the internal structure from being damaged by impact. The lower end of the protective sleeve is fitted over the bottom plate and is movably connected to the bottom plate.

9. A vibration damping device for heavy-duty helicopters as described in claim 1, characterized in that, A floating piston is movably connected above the magnetorheological fluid current limiter inside the cylinder to block high-pressure air and magnetorheological fluid inside the cylinder, preventing gas from mixing into the magnetorheological fluid current limiter. The magnetorheological fluid current limiter dynamically changes the shear yield stress of the magnetorheological fluid by controlling the current of the internal excitation coil to cope with vibration forces under overload loads.

10. A vibration damping device for heavy-duty helicopters as described in claim 9, characterized in that, The upper side of the cylinder is provided with a pressure regulating valve for adjusting the pressure of the high-pressure air inside the cylinder, which is used to achieve dual active control of damping stiffness with the magnetorheological fluid flow limiter.

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