A vertical vibration broadband control device and passive control method for a distributed drive vehicle

Abstract

The application provides a vertical vibration wide-frequency control device for a distributed drive vehicle and a passive control method thereof. An inertia coefficient switching component is arranged between a vehicle body and a suspension. The component comprises a driving shaft and a plurality of driven flywheels which are coaxially sleeved in a radial direction from inside to outside. A motion conversion mechanism is arranged between the driving shaft and a housing. Each driven flywheel is supported on the housing. A magnetic coupling pair is arranged between the outermost driven flywheel and the next outer driven flywheel. A friction coupling pair is arranged between the innermost driven flywheel and the driving shaft and between each adjacent driven flywheel between the next outer driven flywheel and the innermost driven flywheel. The transmission torque threshold of each driven flywheel gradually increases from outside to inside in the radial direction. The method makes each driven flywheel gradually disengage from outside to inside with the increase of vibration frequency, so that the device presents pure inertia, inertia and damping synergy and pure damping characteristics in sequence. The application can realize wide-frequency adaptive vibration isolation and energy recovery without external energy and control.

Description

Technical Field

[0001] This invention relates to the field of automotive suspension technology, and in particular to a distributed drive vehicle vertical vibration broadband control device and passive control method. Background Technology

[0002] Distributed drive vehicles integrate the drive motor directly into the wheel rim or hub, achieving a flattened and integrated chassis structure by eliminating traditional driveshafts and differentials. While this architecture significantly improves vehicle handling response and power transmission efficiency, it also introduces a substantial increase in unsprung mass. Because the drive motor, reduction gear, and braking components are all concentrated within the wheel rim, the unsprung mass of distributed drive vehicles is typically 40% to 80% higher than that of traditional centralized drive vehicles. This increased unsprung mass alters the mass distribution characteristics of the suspension system, making road excitation more easily transmitted to the vehicle body. The resulting negative vertical vibration effects adversely affect ride comfort and smoothness.

[0003] As a key force-transmitting component connecting the vehicle body and wheels, the suspension system's performance directly determines the vehicle's ride comfort and handling stability. Traditional suspensions use a combination of passive springs and dampers, buffering road impacts through elastic elements and dissipating vibration energy through damping elements. This type of "spring-damped" two-dimensional suspension structure lacks a mass element, resulting in limited impedance matching capability in the mechanical network, making it difficult to balance low-frequency vehicle posture control and high-frequency vibration isolation. To address this contradiction, engineers in related fields introduced inertial container elements. An inertial container is a two-terminal network element with mass characteristics; its output force is proportional to the relative acceleration at both ends, providing additional inertial capacity to the suspension system without significantly increasing unsprung mass. In 2002, Professor Smith of the University of Cambridge, UK, proposed the concept of inertial containers, enriching the theoretical system of mechanical networks and expanding the design of suspension systems from a "spring-damped" to a three-dimensional structure of "spring-damped-inertial capacity." Since then, various forms and structures of inertial containers have been proposed, and their effectiveness in improving suspension performance has been widely verified.

[0004] Existing vehicle suspension systems typically employ a parallel arrangement of an inertia reservoir and a spring damper between the sprung and unsprung masses, utilizing the mass effect of the inertia reservoir to improve vibration isolation performance. However, the inertia reservoirs used in these systems are passive inertia reservoirs with a fixed mass-inertia ratio. This ratio is fixed during the design phase and cannot be changed, thus providing optimal vibration isolation only around a specific frequency. Once the excitation frequency deviates from this design point, system performance significantly degrades. Another existing structure achieves stepped adjustment of the mass-inertia ratio by altering the flywheel's mass distribution or transmission ratio. However, this adjustment process relies on an external drive or manual operation and cannot automatically adapt parameters to changes in road excitation frequency. It falls into the category of semi-active or active adjustment, consuming external energy and increasing the complexity of the control system.

[0005] In terms of control strategies, ceiling damping control and acceleration-driven damping control in time-domain control have been widely studied due to their effectiveness and low energy consumption. These control methods achieve real-time adjustment of the suspension damping coefficient based on time-domain signal feedback. However, time-domain control inevitably suffers from time delays introduced by signal acquisition, processing, and actuator response, resulting in an inherent phase difference between the control target and the control action. This can lead to a decrease in control effectiveness or even negative effects in certain frequency bands. If the suspension parameters can be adjusted in the frequency domain according to the vibration frequency, the adverse effects caused by the phase difference can be avoided in principle. However, in existing inertial container technology, there are few schemes that passively achieve adaptive adjustment of the inertial mass coefficient based on the vibration frequency, and there is no known technical solution that can achieve a full-frequency passive adaptive transition from pure inertial capacity, through the combination of inertial capacity and damping, to pure damping.

[0006] Therefore, there is an urgent need to develop a suspension technology that can passively achieve wide-frequency adaptive vibration isolation and has energy recovery function in order to break through the performance bottleneck of vertical vibration control in distributed drive vehicles. Summary of the Invention

[0007] To address the shortcomings of existing technologies, this invention provides a distributed drive vehicle vertical vibration broadband control device and passive control method. Multiple driven flywheels are coaxially mounted radially from the inside out on the outside of the drive component. A magnetic coupling pair is provided between the outermost and second-outermost driven flywheels, and friction coupling pairs are provided between the innermost driven flywheel and the drive component, and between adjacent driven flywheels between the second-outermost and innermost layers. Each driven flywheel has a preset transmission torque threshold, and the transmission torque threshold of each driven flywheel increases progressively from the outside inwards radially. This solves the technical problem that the inertial container cannot passively adjust according to the vibration frequency, thereby improving suspension performance and broadening the effective vibration isolation frequency band.

[0008] The present invention achieves the above-mentioned technical objectives through the following technical means.

[0009] A distributed drive vehicle vertical vibration broadband control device includes an inertia-mass coefficient switching component between the vehicle body and the suspension.

[0010] The inertia-mass switching component includes a housing, a drive component, and a plurality of driven flywheels coaxially sleeved on the outside of the drive component from the inside to the outside in a radial direction; the drive component is rotatably supported in the housing, and a motion conversion mechanism is provided between the drive component and the housing to convert the vertical linear motion between the vehicle body and the suspension into the rotational motion of the drive component; each of the driven flywheels is supported on the housing.

[0011] A magnetic coupling pair is provided between the outermost driven flywheel and the adjacent second outermost driven flywheel located radially inside it; a friction coupling pair is provided between the innermost driven flywheel and the drive component; each driven flywheel located between the second outermost driven flywheel and the innermost driven flywheel is provided with a friction coupling pair with the adjacent driven flywheel located radially inside it.

[0012] Each of the driven flywheels has a preset transmission torque threshold, and the transmission torque threshold of each driven flywheel increases progressively from the outside to the inside in the radial direction.

[0013] Furthermore, a driving magnetic ring is provided on the outer circumferential surface of the driven flywheel located in the second outermost layer, and a driven magnetic ring is provided on the inner circumferential surface of the driven flywheel located in the outermost layer. A radial working air gap is maintained between the outer cylindrical surface of the driving magnetic ring and the inner cylindrical surface of the driven magnetic ring. The driven flywheel located in the outermost layer forms a non-contact magnetic coupling with the driven flywheel located in the second outermost layer through the radial working air gap.

[0014] Furthermore, a first-stage contact friction pair is provided between the inner circumferential surface of the innermost driven flywheel and the outer circumferential surface of the driving component, providing a first-stage static friction torque through a preset radial clamping force; when N≥3, an (i+1)th-stage contact friction pair is provided between the inner circumferential surface of the i-th driven flywheel between the second outermost driven flywheel and the innermost driven flywheel and the outer circumferential surface of its adjacent driven flywheel, providing an (i+1)th-stage static friction torque through a preset radial clamping force, where 1≤i≤N-2, and N is the number of driven flywheels; along the radial direction from the outside to the inside, the static friction torque of each stage increases progressively, and the first-stage static friction torque is greater than the static friction torque of the other stages.

[0015] Furthermore, the motion conversion mechanism is a ball screw pair, the nut of the ball screw pair is connected to the housing, and the screw of the ball screw pair is connected to the drive component or is an integral structure.

[0016] Furthermore, the friction interface of the contact friction pair is provided with a wear-resistant friction plate, and the wear-resistant friction plate is provided with an adjustable radial preload by a disc spring.

[0017] Furthermore, at least one of the aforementioned triboelectric coupling pairs is a triboelectric nano-power generation coupling pair, which is composed of triboelectric material layers respectively disposed on the surfaces of two relatively moving components and an electrode layer disposed on the back side of the triboelectric material layers, and outputs electrical energy when the two relatively moving components produce relative slippage.

[0018] A passive control method for a distributed drive vehicle vertical vibration broadband control device includes the following steps:

[0019] N driven flywheels are coaxially mounted radially from the inside out; the torque threshold values ​​of each driven flywheel radially from the outside in are T1, T2, ..., T. N And T1 <T2<…<T N Each transmission torque threshold corresponds to N switching frequencies f1, f2, ..., f N And f1 <f2<…<f N ;

[0020] The vertical excitation from the road surface is input into the drive unit, and each flywheel switches states according to the vibration frequency f of the excitation:

[0021] When f≤f1, all N driven flywheels rotate synchronously with the driving component, the inertia-mass coefficient switching component exhibits pure inertia-capacity characteristics, and the input vibration energy is temporarily stored in each flywheel in the form of rotational kinetic energy;

[0022] When f k <f≤f k+1 When k=1,2,…,N−1, the first k driven flywheels counted from the outside to the inside have already slipped relative to the driving component, while the remaining Nk driven flywheels still rotate synchronously. The inertia mass coefficient switching component exhibits both inertia-capacitance characteristics and damping characteristics, and the equivalent inertia mass coefficient decreases with increasing frequency, while the equivalent damping coefficient increases with increasing frequency.

[0023] When f>f N At that time, all N driven flywheels slip relative to the drive component, the inertia coefficient switching component exhibits pure damping characteristics, and the high-frequency vibration energy is dissipated by each slip interface.

[0024] Furthermore, the switching frequency f1 is determined by the transmission torque threshold T1 of the outermost magnetic coupling pair; the switching frequency f k+1 The maximum static friction torque threshold T of the friction coupling pair corresponding to the (k+1)th flywheel counted radially from the outside to the inside is respectively determined by... k+1 It has been determined.

[0025] Furthermore, the switching frequency f1 is determined by the following relationship:

[0026] ,

[0027] Where T1 is the maximum transmitted torque of the magnetic coupling pair; k t To convert linear motion into rotational motion, J1 is the moment of inertia of the outermost driven flywheel, and A is the vibration amplitude.

[0028] Furthermore, the switching frequency f k+1 They are determined through the following relationships respectively:

[0029] ,

[0030] Where: μ s,k+1 p is the static friction coefficient of the (k+1)th driven flywheel corresponding to the kth friction coupling pair, counted radially from the outside in; r,k+1 L represents the radial contact pressure of the k-th frictional coupling pair. f,k+1 R is the contact length of the k-th frictional coupling pair; f,k+1 Let k be the contact radius of the k-th frictional coupling pair; k t The transmission ratio for converting linear motion into rotary motion, J k+1 Let A be the moment of inertia of the (k+1)th driven flywheel counted radially from the outside in; let A be the vibration amplitude.

[0031] The beneficial effects of this invention are as follows:

[0032] 1. The distributed drive vehicle vertical vibration broadband control device of the present invention, by coaxially sleeved multiple driven flywheels radially from the inside to the outside of the drive component, and by setting a magnetic coupling pair between the outermost driven flywheel and the next outermost driven flywheel, and setting friction coupling pairs between the innermost driven flywheel and the drive component, and between each adjacent driven flywheel between the next outermost and the innermost layers, enables each driven flywheel to have a transmission torque threshold that increases radially from the outside to the inside, thereby achieving multi-level stepwise passive adjustment of the inertial mass coefficient. By sequentially combining the outer layer magnetic coupling threshold with the friction coupling threshold of each layer to form multiple preset switching points, the device can automatically switch the equivalent inertial mass level in four frequency ranges: low frequency, medium frequency, mid-high frequency, and high frequency. Compared with traditional fixed inertial containers that only have better performance near a single frequency point, it can significantly broaden the high-efficiency vibration isolation frequency band and maintain near-optimal vibration isolation performance in a broadband vibration environment.

[0033] 2. The distributed drive vehicle vertical vibration broadband control device of the present invention arranges magnetic coupling pairs between the outermost driven flywheel and the next outermost driven flywheel, and arranges each friction coupling pair radially from the outside to the inside. As the vibration frequency increases, the device automatically and sequentially completes the magnetic coupling disengagement of the outermost flywheel and the friction coupling disengagement of each layer of flywheels. The entire control process of "frequency identification - interval determination - state switching" is autonomously driven by vibration energy, without the need for sensors, controllers and external energy intervention. It fundamentally avoids the inherent problems of high energy consumption, complex structure, high cost and control delay of electronic control or semi-active systems. It is especially suitable for vehicle chassis applications with extreme requirements for reliability, response speed and durability.

[0034] 3. The distributed drive vehicle vertical vibration broadband control device of the present invention integrates multi-stage inertial mass switching function within a limited radial space by adopting a hybrid transmission method combining magnetic coupling pairs and friction coupling pairs, and a sleeve-type layout with flywheel nested coupling. The magnetic coupling pairs are non-contact transmissions, directly transmitting torque without magnetic field modulation, and the magnetic circuit is not affected by the contact state of the friction pairs; the friction coupling pairs rely on the normal pressure of the contact surface to transmit torque, and their working state is independent of the magnetic coupling pairs. The outermost layer uses non-contact magnetic coupling to reduce friction and wear, and the inner layer uses controllable friction pairs to achieve stable second and third switching thresholds. Combined with wear-resistant friction materials and a pre-tightening structure, the theoretical life of the device can be significantly extended.

[0035] 4. The distributed drive vehicle vertical vibration broadband control device of the present invention sets the transmission torque threshold of the magnetic coupling pair as the determining factor of the first switching frequency, and sets the static friction torque threshold of each friction coupling pair as the determining factor of the second and third switching frequencies. This allows the same basic structural platform to change the first switching frequency by adjusting the magnetic ring gap, magnetic field strength, and magnetic coupling area, and to change the second and third switching frequencies by adjusting the preload, friction coefficient, friction radius, and contact length of the friction pair. This flexibly adapts to the vibration isolation requirements of different vehicle models, from comfort cars to high-performance sports cars, and has good versatility and industrial application prospects.

[0036] 5. The distributed drive vehicle vertical vibration broadband control device of the present invention converts mechanical energy lost in the form of heat dissipation in traditional suspension into electrical energy by setting a triboelectric nano-power generation unit in at least one friction coupling pair, utilizing the relative motion inherent in the flywheel sliding interface. The triboelectric nano-power generation structure on the sleeve-type cylindrical sliding interface can achieve energy recovery under broadband vibration conditions. The recovered electrical energy, after being processed by the energy management circuit, can power low-power vehicle equipment such as tire pressure monitoring sensors and radio frequency identification nodes. At the same time, the frequency of the electrical signal output by the power generation unit is strictly consistent with the relative sliding frequency of the sliding interface, and the amplitude of the electrical signal is positively correlated with the sliding speed. This electrical signal can be directly used to provide feedback on the current vibration frequency range and the switching state of the inertial mass coefficient, without the need for additional sensors, thus achieving passive integrated integration of vibration reduction execution, energy recovery, and state feedback.

[0037] 6. The distributed drive vehicle vertical vibration broadband control device of the present invention, by configuring the transmission torque threshold of each driven flywheel to increase gradually from the outside to the inside in the radial direction, allows the device to successively exhibit pure inertial capacitance characteristics, inertial capacitance and damping synergistic characteristics, and pure damping characteristics in different frequency bands. In the low-frequency band, all flywheels maintain synchronous rotation, and the input vibration energy is temporarily stored in the flywheels in the form of rotational kinetic energy. The large inertial mass coefficient effectively suppresses the pitch and roll motion of the distributed drive vehicle during start-up, braking, and steering conditions. In the mid-frequency band, the outermost magnetically coupled flywheel slips first, the inertial capacitance characteristics begin to decrease, and the damping characteristics begin to appear. In the mid-to-high frequency band, the inner frictionally coupled flywheels slip sequentially, the inertial capacitance characteristics further decrease, and the damping characteristics gradually increase. This transition process fills the performance gap between pure inertial capacitance and pure damping, which precisely covers the frequency band where the human body is most sensitive to vertical vibration, thus helping to improve ride comfort. In the high-frequency band, all driven flywheels slip, and the device exhibits pure damping characteristics. The high-frequency vibration energy is dissipated by each slipping interface, effectively blocking the transmission of road impacts to the vehicle body.

[0038] 7. The distributed drive vehicle vertical vibration broadband control device of the present invention sets the adaptive vibration isolation units at each wheel end to work independently, allowing each unit to independently complete state switching based entirely on the real-time vibration frequency borne by its own wheel, without the need for a central controller to participate in decision-making. Under non-linear driving conditions such as vehicle steering and braking, the vibration frequencies of each wheel differ, and each unit automatically exhibits different mechanical characteristics: during steering, the vibration frequency of the outer suspension increases, and the device automatically switches to a high-damping state to suppress body roll; the vibration frequency of the inner suspension is lower, and the device maintains a high-inertia state to maintain the wheel's ground load, thus achieving adaptive adjustment of suspension characteristics to vehicle attitude at a physical level. Attached Figure Description

[0039] To more clearly illustrate the technical solutions in the embodiments of the present invention or the prior art, the drawings used in the description of the embodiments or the prior art will be briefly introduced below. The drawings described below are some embodiments of the present invention. For those skilled in the art, it is obvious that other drawings can be obtained from these drawings without creative effort.

[0040] Figure 1 This is a schematic diagram of the distributed drive vehicle vertical vibration broadband control method and its passive implementation device according to the present invention.

[0041] Figure 2 This is a schematic diagram of a sleeve-type inertia coefficient switching component.

[0042] Figure 3 This is a simplified axial diagram of a sleeve-type inertia coefficient switching component.

[0043] Figure 4 A simplified model of the inertial mass coefficient for a broadband control method for vertical vibration of distributed drive vehicles and its passive implementation device.

[0044] Figure 5 This is the original distributed drive vehicle chassis model and structural diagram.

[0045] Figure 6 This is a comparison diagram of the original vehicle vertical vibration model and the vehicle vertical vibration model with this device added.

[0046] Figure 7 This is a graph showing the change in inertial mass coefficient of a sleeve-type inertial-capacity device with frequency.

[0047] Figure 8 This is a schematic diagram illustrating the qualitative changes in the normalized inertial mass coefficient and damping coefficient of a sleeve-type inertial-capacitive device.

[0048] Figure 9 This is a sequence diagram of the working states of a sleeve-type inertial capacitive device.

[0049] Figure 10 A comparison graph showing the relationship between vehicle body acceleration gain and vibration frequency after incorporating this sleeve-type inertial capacitance device.

[0050] Figure 11 This is a schematic diagram of the switching process between inertial capacitance and damping states.

[0051] In the picture:

[0052] 1-Upper hanger; 2-Upper bearing; 3-Lower bearing; 4-Drive shaft; 5-Bolt; 6-First housing; 7-Driven component; 7a-First driven flywheel; 7b-Second driven flywheel; 7c-Third driven flywheel; 8-Lower hanger; 9-Inertia coefficient switching component; 10-Second housing; 11-Suspension system; 12-Tire; 13-Rotor; 14-Brake; 15-Mechanical device; 16-Motor; 17-Bearing; 18-Stator; 19-First triboelectric material layer; 20-Second triboelectric material layer; 21-Third triboelectric material layer; 22-Fourth triboelectric material layer; 23-Sleeve electrode layer; 24-Sleeve output terminal; 25-Energy management circuit. Detailed Implementation

[0053] Embodiments of the present invention are described in detail below, examples of which are illustrated in the accompanying drawings, wherein the same or similar reference numerals denote the same or similar elements or elements having the same or similar functions throughout. The embodiments described below with reference to the accompanying drawings are exemplary and intended to explain the present invention, and should not be construed as limiting the present invention.

[0054] In the description of this invention, it should be understood that the terms "center," "longitudinal," "lateral," "length," "width," "thickness," "upper," "lower," "axial," "radial," "vertical," "horizontal," "inner," and "outer," etc., indicating orientation or positional relationships based on the orientation or positional relationships shown in the accompanying drawings, are only for the convenience of describing the 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, and therefore should not be construed as a limitation of the invention. Furthermore, the terms "first" and "second" are used for descriptive purposes only and should not be construed as indicating or implying relative importance or implicitly specifying the number of indicated technical features. Thus, a feature defined with "first" or "second" may explicitly or implicitly include one or more of that feature. In the description of this invention, "a plurality of" means two or more, unless otherwise explicitly specified.

[0055] In this invention, unless otherwise explicitly specified and limited, the terms "installation," "connection," "linking," and "fixing," etc., should be interpreted broadly. For example, they can refer to a fixed connection, a detachable connection, or an integral connection; they can refer to a mechanical connection or an electrical connection; they can refer to a direct connection or an indirect connection through an intermediate medium; and they can refer to the internal connection of two components. Those skilled in the art can understand the specific meaning of the above terms in this invention according to the specific circumstances.

[0056] Frequency-adaptive inertial mass adjustment is a control algorithm that adjusts the inertial mass coefficient of an inertial container according to the vibration frequency band. Its control rule is shown in equation (1):

[0057] (1)

[0058] in, It is the vertical displacement of one end of the inertial container; It is the vertical displacement of the other end of the inertial container; It is the vertical acceleration at one end of the inertial container; The vertical acceleration at the other end of the inertial container; f is the vibration frequency, f0 is the frequency switching point, b is the inertial mass coefficient of the controllable inertial container, b max It is a large inertia mass coefficient, b min It is a small inertia coefficient.

[0059] Frequency-adaptive inertial mass adjustment identifies the vibration frequency band and adjusts the variable inertial mass coefficient, avoiding the phase influence problem of time-domain control and achieving complete consistency between the control result and the control objective. As can be seen from the control rule expressed in equation (1), the judgment condition of this control strategy essentially reflects the ratio between the relative acceleration and relative displacement of the suspension:

[0060] (2)

[0061] According to the physical definition of an inertial container, the force F between its two endpoints is equal to the product of the inertial mass coefficient and the relative acceleration between the two endpoints. Therefore, equation (2) can be rewritten as:

[0062] (3)

[0063] From equations (2) and (3), it can be seen that if a passive inertial container is used to achieve frequency-adaptive inertial mass adjustment, the following two conditions must be met:

[0064] Condition 1: The forces at both ends of the inertial container and f 2 Proportional; Condition 2: After F reaches a certain threshold, the inertial mass coefficient of the inertial container can switch.

[0065] Since the road surface input can be considered as a superposition of sinusoidal displacement inputs, we will now analyze it using sinusoidal displacement input as an example. For sinusoidal displacement input x = Asin(ωt) = Asin(2πft), where A is the input amplitude, ω is the angular frequency, t is the time variable, and the corresponding velocity... and acceleration They are respectively:

[0066] ,

[0067] This leads to the forces at both ends of the inertial container. It can be seen that F is proportional to f 2Therefore, the inertial container satisfies condition one for passively achieving frequency-adaptive inertial mass adjustment. To satisfy condition two, this embodiment proposes the following technical solution:

[0068] This embodiment relates to a broadband vertical vibration control structure for distributed drive vehicles and its corresponding passive control method. In practical applications, this structure is typically used as an additional inertial capacitance element in the suspension system, installed in parallel on both sides of the shock absorber and the coil spring. Alternatively, it can be integrated into the shock absorber assembly depending on the chassis space requirements.

[0069] like Figure 1 As shown, the vertical vibration broadband control structure for distributed drive vehicles of the present invention includes an upper hanger 1, an upper bearing 2, a lower bearing 3, a drive shaft 4, a bolt 5, a first housing 6, a driven component 7, a lower hanger 8, an inertia coefficient switching component 9, and a second housing 10. The upper hanger 1 is used to connect to a fixed point on the side of the vehicle body, and the lower hanger 8 is used to connect to the lower control arm of the suspension or a wheel-side component. The second housing 10 covers the outside of the inertia coefficient switching component 9. One end of the second housing 10 is detachably fixed to the upper hanger 1. The upper bearing 2 and the lower bearing 3 are respectively installed on both sides of the second housing 10. The two ends of the drive shaft 4 are supported on the upper bearing 2 and the lower bearing 3 respectively. The other end of the drive shaft 4 passes through the lower bearing 3 and is connected to the lower hanger 8 through the first housing 6.

[0070] The inertia-mass switching component 9 includes a driving component and a plurality of driven flywheels coaxially sleeved on the outside of the driving component in a radial direction from the inside to the outside. The driving component is the driving shaft 4, which is rotatably supported in the second housing 10. A motion conversion mechanism is provided between the driving component and the first housing to convert the vertical linear motion between the vehicle body and the suspension into the rotational motion of the driving component. Each of the driven flywheels is supported on the second housing 10. A magnetic coupling pair is provided between the outermost driven flywheel and the adjacent second outermost driven flywheel located radially inside it. A friction coupling pair is provided between the innermost driven flywheel and the driving component. Each driven flywheel between the second outermost driven flywheel and the innermost driven flywheel has a friction coupling pair with the adjacent driven flywheel located radially inside it. Each of the driven flywheels has a preset transmission torque threshold, and the transmission torque threshold of each driven flywheel increases progressively from the outside to the inside in a radial direction. A first-stage contact friction pair is provided between the inner circumferential surface of the innermost driven flywheel and the outer circumferential surface of the driving component, providing the first-stage static friction torque through a preset radial clamping force; when N≥3, an (i+1)th-stage contact friction pair is provided between the inner circumferential surface of the i-th driven flywheel between the second outermost driven flywheel and the innermost driven flywheel and the outer circumferential surface of its adjacent driven flywheel, providing the (i+1)th-stage static friction torque through a preset radial clamping force, where 1≤i≤N-2, and N is the number of driven flywheels; along the radial direction from the outside to the inside, the static friction torque of each stage increases progressively, and the first-stage static friction torque is greater than the static friction torque of the other stages.

[0071] In this embodiment, the drive shaft 4 is a ball screw. The nut portion of the ball screw is rigidly connected to the first housing 6, while the screw portion is an integral part of the drive shaft 4. When the suspension system generates a vertical relative displacement, the first housing 6 moves together with the lower hanger 8, driving the screw to rotate via the nut, thereby converting the vertical linear relative motion between the upper hanger 1 and the lower hanger 8 into the rotational motion of the drive shaft 4, and driving the driven component 7 to rotate synchronously. In some alternative embodiments, the drive shaft 4 can also be a straight rod structure with a helical guide groove, which, together with a guide pin fixed to the housing, achieves rotational output.

[0072] Taking three driven flywheels as an example, such as Figure 2 and Figure 3As shown, the driven member 7 includes a first driven flywheel 7a, a second driven flywheel 7b, and a third driven flywheel 7c; the drive shaft 4, the first driven flywheel 7a, the second driven flywheel 7b, and the third driven flywheel 7c are sequentially nested and arranged radially from the inside to the outside. To ensure stable rotation of the three sleeve-type driven flywheels and avoid axial play during operation, both ends of the first driven flywheel 7a, the second driven flywheel 7b and the third driven flywheel 7c are respectively supported on the second housing 10; the drive shaft 4 serves as the driving member, and the outer surface thereof directly acts as the active transmission surface. The first driven flywheel 7a is radially sleeved on the outer side of the drive shaft 4, and a contact friction pair is disposed therebetween, and the maximum static friction torque of the friction pair constitutes a third switching threshold. The second driven flywheel 7b is radially sleeved on the outer side of the first driven flywheel 7a, a contact friction pair is also disposed between the inner cylindrical surface of the second driven flywheel 7b and the outer cylindrical surface of the first driven flywheel 7a, and the maximum static friction torque of the friction pair constitutes a second switching threshold. The third driven flywheel 7c is radially sleeved on the outer side of the second driven flywheel 7b, a driving magnetic ring is fixed on the outer cylindrical surface of the second driven flywheel 7b, a driven magnetic ring is fixed on the inner cylindrical surface of the third driven flywheel 7c, a uniform radial working air gap is maintained between the outer cylindrical surface of the driving magnetic ring and the inner cylindrical surface of the driven magnetic ring, non-contact torque transmission is realized therebetween through magnetic field coupling, and the maximum transmission torque of the magnetic coupling pair constitutes a first switching threshold.

[0073] From the perspective of the sequence of torque transmission, the first switching threshold is determined by the outermost magnetic coupling pair, and the maximum transmission torque thereof is denoted as T1; the second switching threshold is determined by the friction coupling pair between the second driven flywheel 7b and the first driven flywheel 7a, and the maximum static friction torque thereof is denoted as T2; the third switching threshold is determined by the friction coupling pair between the first driven flywheel 7a and the drive shaft 4, and the maximum static friction torque thereof is denoted as T3. There is a clear magnitude relationship among the three thresholds, that is, T1<T2<T3. When the inertia torque transmitted by the drive shaft 4 gradually increases with the vibration frequency, the outermost magnetically coupled driven flywheel 8c first reaches its torque upper limit and slips; as the frequency continues to rise, after the inertia torque sequentially reaches the load limits of the second friction coupling pair and the first friction coupling pair, the corresponding driven flywheels slip sequentially, thereby forming a switching sequence of disengagement step by step from the outside to the inside.

[0074] Regarding the dynamic relationship between the drive shaft 4 and the driven flywheels, the following unified description can be made: let the maximum transmission torque between the drive shaft 4 and a certain driven flywheel be T max , and the inertia torque generated by the vibration acceleration during actual operation be T d . When T d <T max , the driven flywheel rotates synchronously with the drive shaft 4, the two have the same rotation angle and no relative motion; when T d ≥T maxAt this time, the driven flywheel no longer follows the drive shaft 4, and a speed difference and relative slippage occur between them. Due to the inertial torque T d Proportional to the square of the vibration frequency, the aforementioned boundary conditions for synchronization and slip naturally establish a quantitative relationship with the vibration frequency, thus providing a physical basis for passive frequency band identification.

[0075] like Figure 3 As shown, the solid dot at the very center represents the cross-section of the drive shaft 4. A first driven flywheel 7a is arranged around the drive shaft 4, with its inner cylindrical surface in contact with the outer cylindrical surface of the drive shaft 4, forming a first cylindrical surface friction pair. A second driven flywheel 7b is arranged around the first driven flywheel 7a, with its inner cylindrical surface in contact with the outer cylindrical surface of the first driven flywheel 7a, forming a second cylindrical surface friction pair. Both friction pairs are contact-type friction transmission interfaces and do not have magnetic coupling gaps. An outermost magnetically coupled driven flywheel 8c is arranged around the second driven flywheel 7b. A driving magnetic ring is fixedly mounted on the outer cylindrical surface of the second driven flywheel 7b, and a driven magnetic ring is fixedly mounted on the inner cylindrical surface of the outermost magnetically coupled driven flywheel 8c. An outermost magnetic coupling radial working gap A is formed between the outer cylindrical surface of the driving magnetic ring and the inner cylindrical surface of the driven magnetic ring. n .

[0076] In a preferred embodiment, the friction interface of the contact friction pair is provided with a wear-resistant friction plate, and the wear-resistant friction plate is provided with an adjustable radial preload by a disc spring.

[0077] To recover the mechanical energy generated during flywheel slippage, this invention integrates triboelectric nano-power generation units within a sleeve-type structure. The integration is located at the sliding interface between the drive shaft 4 and the driven component 7a, and between the driven component 7a and the driven component 7b. A first triboelectric material layer 19 is attached to the outer cylindrical surface of the driven component 7a, a second triboelectric material layer 20 is attached to the inner cylindrical surface of the driven component 7b, a third triboelectric material layer 21 is attached to the outer cylindrical surface of the drive shaft 4, and a fourth triboelectric material layer 22 is attached to the inner cylindrical surface of the driven component 7a. When the frictionally coupled driven flywheel slips, the first triboelectric material layer 19, the second triboelectric material layer 20, the third triboelectric material layer 21, and the fourth triboelectric material layer 22 slide relative to each other, generating electrical energy. An output terminal 24 is installed on the right side of the second housing 10, through which the electrical energy generated by each power generation unit is led out and connected to the energy management circuit 25.

[0078] exist Figure 4In the simplified mechanical model shown, the drive shaft 4 itself has a certain moment of inertia, and its contribution to the basic inertia coefficient is denoted as b0. The first driven flywheel 7a, the second driven flywheel 7b, and the third driven flywheel 7c contribute variable inertia coefficients b1, b2, and b3, respectively. When all three driven flywheels are synchronized with the drive shaft 4, the total inertia coefficient is the sum of the four, i.e., b = b0 + b1 + b2 + b3; when the outermost flywheel slips while the other flywheels remain synchronized, the total inertia coefficient is b = b0 + b1 + b2; when the outermost and second driven flywheels slip while the first driven flywheel remains synchronized, the total inertia coefficient is b = b0 + b1; when all driven flywheels slip, only the basic inertia coefficient b0 continues to play a role.

[0079] For ease of quantitative analysis, the vertical excitation of the road surface is simplified to a sinusoidal displacement input x = Asin(ωt) = Asin(2πft), where A is the excitation amplitude, f is the vibration frequency, and t is time. From this, the velocity can be obtained. and acceleration For an inertial container, the output force F at its two endpoints is equal to the product of the mass coefficient and the relative acceleration, and the magnitude of the relative acceleration is proportional to f. 2 Therefore, the inertial torque T borne by the driving end d with f 2 They are directly proportional.

[0080] The passive control method of the distributed drive vehicle vertical vibration broadband control device of the present invention includes the following steps:

[0081] N driven flywheels are coaxially mounted radially from the inside out; the torque threshold values ​​of each driven flywheel radially from the outside in are T1, T2, ..., T. N And T1 <T2<…<T N Each transmission torque threshold corresponds to N switching frequencies f1, f2, ..., f N And f1 <f2<…<f N ;

[0082] The vertical excitation from the road surface is input into the drive unit, and each flywheel switches states according to the vibration frequency f of the excitation:

[0083] When f≤f1, all N driven flywheels rotate synchronously with the driving component, and the inertia-mass coefficient switching component 9 exhibits pure inertia-capacity characteristics. The input vibration energy is temporarily stored in each flywheel in the form of rotational kinetic energy.

[0084] When f k <f≤f k+1When k=1,2,…,N−1, the first k driven flywheels counted from the outside to the inside have already slipped relative to the driving component, while the remaining Nk driven flywheels still rotate synchronously. The inertia mass coefficient switching component 9 simultaneously exhibits inertia-capacitance characteristics and damping characteristics, and the equivalent inertia mass coefficient decreases with increasing frequency, while the equivalent damping coefficient increases with increasing frequency.

[0085] When f>f N At that time, all N driven flywheels slip relative to the drive component, and the inertia-mass coefficient switching component 9 exhibits pure damping characteristics, with high-frequency vibration energy dissipated by each slip interface.

[0086] The following are quantitative expressions for the switching frequencies of the three examples of driven flywheels:

[0087] For the outer magnetically coupled driven flywheel 8c, its moment of inertia is denoted as J1, and the maximum transmitted torque of the magnetic coupling pair is denoted as T1 (i.e., the transmitted torque threshold of the first driven flywheel counted radially from the outside in). Under sinusoidal excitation, the inertial torque required for this flywheel to maintain synchronous rotation is:

[0088] ,

[0089] Where k t The transmission ratio of the ball screw pair is the proportionality coefficient that converts linear displacement into rotational angular displacement, and its unit is m / rad. When T... d,m equal to T m,max When the first switching condition is met:

[0090]

[0091] Therefore, the first switching frequency can be determined:

[0092] ,

[0093] For the second driven flywheel 7b (the second driven flywheel counted radially from the outside in), the maximum static friction torque threshold of the friction coupling pair between it and the first driven flywheel 7a is denoted as T2, and the moment of inertia of this flywheel is denoted as J2. The static friction coefficient of this friction coupling pair is μ. s,2 The radial contact pressure is p r,2 The contact length is L f,2 The contact radius is R f,2 The total normal force of this friction pair is:

[0094] ,

[0095] The maximum static friction torque that this friction pair can provide is:

[0096] ,

[0097] The inertial torque required for the flywheel to maintain synchronization is:

[0098] ,

[0099] Let T d,2 =T2, thus obtaining the second switching frequency:

[0100] ;

[0101] Similarly, the third switching frequency can be obtained: ;

[0102] In the above formulas, the units of J1, J2, and J3 are all kg⋅m. 2 The units for torques T1, T2, and T3 are all Nm, and the unit for A is m, k. t The unit is m / rad, and the units for f1, f2 and f3 are all Hz.

[0103] like Figure 5 As shown, Figure 5 This is a schematic diagram of the original distributed drive vehicle chassis model and its structure. The diagram fully presents all the components: the suspension system 11 includes shock absorbers, coil springs, and control arms, etc., connected to the vehicle body at the upper end and the wheel-side modules at the lower end, and is responsible for vertical force transmission and vibration damping. Tire 12 is mounted on the outer side of the rim, serving as the only contact interface between the vehicle and the road surface, transmitting vertical, longitudinal, and lateral forces; rotor 13 is rigidly connected to the wheel hub, serving as the rotating part of the motor, transmitting electromagnetic torque to the wheel; brake 14 is integrated inside the wheel hub, including a brake disc and brake caliper, used to provide braking torque; mechanical device 15 generally refers to a wheel-side reducer or transmission mechanism, used to match the motor characteristics and wheel torque requirements; motor 16, as the core power source of distributed drive, adopts an external rotor or internal rotor structure, directly or via a reduction mechanism to drive the wheel; bearing 17 supports the relative rotational motion between the rotor and stator, ensuring transmission efficiency and load-bearing capacity; stator 18 is fixed to the steering knuckle or support housing, cooperating with rotor 13 to form an electromagnetic circuit; wherein the distributed drive vehicle vertical vibration broadband control device described in this invention is installed inside the suspension system, specifically integrated into the shock absorber assembly or installed in parallel on both sides of the spring and damper as an independent inertial capacitance element, used to achieve adaptive vibration isolation of broadband road excitation without relying on external energy.

[0104] like Figure 6As shown, the lumped-parameter dynamic model for vertical vibration analysis of the original distributed-drive intelligent vehicle is compared with the model after incorporating this inertial capacitance device. The model of this invention abstracts the complex physical system into a combination of basic elements such as mass, spring, damping, and inertial capacitance, in order to theoretically and quantitatively analyze the "negative effect of vertical vibration" and verify the efficiency enhancement mechanism of the device of this invention. The key components of the model include: a mass block, elastic and energy-dissipating elements, and an inertial capacitance container. The mass block includes a sprung mass m. s Unsprung mass (m): Represents the vehicle body and its fixed mass; its motion directly affects ride comfort. u This represents the unsustainable mass, including the wheel hubs, brake discs, and, crucially, the distributed drive motor. The significant increase in this mass is a core issue for distributed drive vehicles; the model addresses this by setting a large m... u This characteristic is reflected in the value. Elastic and energy-dissipating components include the spring k. s : Represents suspension stiffness, primarily supporting static loads and storing vibrational energy. Damper c s : Represents the suspension damper, which dampens vibrations by dissipating energy. Tire stiffness k t Simplified as a spring, it represents the vertical stiffness of the tire. In the inertia diagram, it is represented by the symbol 'b'. In the model, it is connected in parallel to both the suspension spring and the damper, meaning it is dynamically coupled to both the sprung and unsprung masses simultaneously.

[0105] When the road surface excitation frequency f is not higher than the first switching frequency f1, the inertial torque generated by the drive shaft 4 is less than the maximum transmitted torque of the outer magnetic coupling pair. At this time, the outermost magnetically coupled driven flywheel 8c, the second driven flywheel 7b, and the first driven flywheel 7a all rotate synchronously with the drive shaft 4, and the device as a whole exhibits pure inertial-capacitive characteristics. Within this frequency band, the vibration energy input to the device is temporarily stored in each flywheel in the form of rotational kinetic energy, without substantial dissipation. The total inertial-mass coefficient during this stage remains b = b0 + b1 + b2 + b3.

[0106] When the vibration frequency f exceeds f1 but has not yet reached f2, the inertial torque generated by the drive shaft 4 exceeds the bearing capacity of the outer magnetic coupling pair. The outermost magnetically coupled driven flywheel 8c begins to slip relative to the drive shaft 4, but the second driven flywheel 7b and the first driven flywheel 7a, due to their high threshold values, still maintain synchronous rotation. At this time, the total mass coefficient drops to b = b0 + b1 + b2. Since the switching in this stage is achieved by non-contact magnetic coupling, the slip interface does not generate significant frictional energy loss. Therefore, the equivalent damping coefficient remains approximately constant in this frequency band, and the device mainly exhibits the inertial-capacitive characteristics after the mass coefficient decreases. This stage achieves a smooth transition from a low-frequency, high-mass state to the subsequent frictional energy-loss state.

[0107] When the vibration frequency f exceeds f2 but has not yet reached f3, the inertial torque generated by the drive shaft 4 exceeds the bearing capacity of the second friction coupling pair. The second driven flywheel 7b begins to slip relative to the first driven flywheel 7a, while the first driven flywheel 7a continues to rotate synchronously. At this time, the total mass coefficient drops to b = b0 + b1. The sliding interface of this friction pair generates frictional force related to the relative velocity, which is equivalent to introducing a damping component into the system. At the same time, the triboelectric nano-power generation unit arranged at this sliding interface converts part of the mechanical energy into electrical energy. The device exhibits both capacitive and damping characteristics in this frequency band, and as the frequency continues to increase, the mass coefficient gradually decreases while the equivalent damping coefficient gradually increases.

[0108] When the vibration frequency f continues to rise and exceeds the third switching frequency f3, the inertial torque borne by the first driven flywheel 7a also reaches its maximum static friction torque. This flywheel then slips from the drive shaft 4 and eventually comes to a stop. At this point, all three driven flywheels are disengaged from the drive shaft, and the device retains only the basic inertial mass coefficient b0 contributed by the rotational inertia of the drive shaft 4 itself. At this time, the inertial capacitance effect of the device is significantly weakened, and the frictional damping at the two friction-coupled sliding interfaces dominates the energy dissipation process. The frictional nano-power generation unit continuously outputs electrical energy, and the device's behavior approaches pure damping characteristics, effectively blocking the transmission of high-frequency road impacts to the vehicle body.

[0109] like Figure 7 As shown in the graph, this curve compares the performance of the sleeve-type adaptive inertial capacity device with that of a traditional fixed inertial capacity device and a traditional two-state switching inertial capacity device. By plotting the relationship between the equivalent mass coefficient b and the vibration frequency f, the intelligent adaptive characteristics of the device in different frequency bands are visually demonstrated. This device adopts a hybrid mechanism of friction coupling and magnetic coupling to achieve "step-like" mass adjustment, smooth transition, avoidance of impact, and widening of the effective vibration isolation frequency band. In contrast, traditional fixed inertial capacity devices have insufficient mass in the low-frequency band, resulting in poor vehicle body control performance; and excessive mass in the high-frequency band, resulting in poor impact isolation and inability to adapt to wide-frequency vibration environments. Traditional two-state switching inertial capacity devices generate impact at the moment of switching, causing a sharp drop in vibration isolation performance in the mid-frequency band (3.0-5.0Hz). In contrast, the adaptive inertial capacity device of this invention provides sufficient inertial force with a large mass coefficient of 500kg in the low-frequency vehicle body control range of 0-1.5Hz, corresponding to starting, braking, and low-speed steering, thus suppressing vehicle pitch and roll and improving driving stability. Within the mid-frequency comfort zone of 1.5-2.5Hz, since the human body's sensitive frequency range is 2-4Hz, the inertia is appropriately reduced to 400kg to balance vehicle control and ride comfort. In the mid-to-high frequency range (2.5-3.5Hz), the inertia is reduced to 250kg, smoothly transitioning to a high vibration isolation state. In the high-frequency impact isolation zone (≥3.5Hz), where there is high-frequency road impact, the inertia is reduced to a minimum of 50kg, effectively isolating high-frequency vibrations and preventing their transmission to the vehicle body.

[0110] See Figure 8 , which is a qualitative schematic diagram of the variation of the normalized inertia coefficient and damping coefficient of the sleeve-type magnetic-friction hybrid inertia coefficient switching component with frequency. In the figure, the horizontal axis represents the vibration frequency f, and the vertical axis represents the normalized coefficient value. The solid line represents the normalized equivalent inertia coefficient , and the dashed line represents the normalized equivalent damping coefficient . Since the equivalent inertia coefficient and the equivalent damping coefficient have different dimensions, to facilitate the representation of their relative variation trends with frequency in the same coordinate system, corresponding reference values are selected for normalization treatment respectively:

[0111] ,

[0112] where, b ref and c ref are the reference values of the inertia coefficient and the damping coefficient respectively, and the normalized and are both dimensionless quantities. When f<f1, all three driven flywheels remain synchronized with the drive shaft, the normalized inertia coefficient is at a relatively high level, and the normalized damping coefficient is at a relatively low level. When f1≤f<f2, the outermost magnetic coupling flywheel slips first, and the normalized inertia coefficient drops for the first time. Since this stage is non-contact magnetic coupling, the normalized damping coefficient remains approximately unchanged. When f2≤f<f3, the second driven flywheel starts to slip, the normalized inertia coefficient drops again, and the normalized damping coefficient rises at the same time. When f≥f3, the first driven flywheel also slips, the normalized inertia coefficient drops to the lowest level, and the normalized damping coefficient further increases.

[0113] As shown in Figure 9 , the passive adaptive adjustment of the inertia coefficient is realized through the synchronization-slip mechanism of the outer magnetic coupling driven flywheel and the intermediate friction coupling driven flywheels. As the frequency increases, the driving torque increases gradually, the outermost magnetic coupling driven flywheel starts to slip from the fully coupled state first, then the second driven flywheel and the first driven flywheel slip in sequence, and finally the inertia coefficient of the device is changed step by step.

[0114] As shown in Figure 10 , after adopting the structure of the present invention, the vehicle body acceleration gain is significantly improved. Figure 10shows the comparison results of the vertical acceleration gain of the vehicle body varying with frequency under the same vehicle parameters and road excitation conditions, when three strategies are adopted: the multi-stage adaptive inerter, the conventional two-state switching inerter, and the conventional fixed inerter. For the multi-stage adaptive inerter (solid line), the inertance coefficient is 500kg in the low frequency band (f<1.5Hz), presenting pure inerter characteristics, with all flywheels rotating synchronously and no electric energy output; in the middle frequency band (1.5Hz≤f<2.5Hz), the inertance coefficient is 400kg, the outermost magnetically coupled flywheel slips, and the equivalent damping coefficient remains approximately constant; in the middle-high frequency band (2.5Hz≤f<3.5Hz), the inertance coefficient is 250kg, the second driven flywheel slips, the friction nanogenerator unit starts to output electric energy, and the device presents synergistic effects of reduced inertance, enhanced damping and energy recovery; in the high frequency band (f≥3.5Hz), the inertance coefficient is 50kg, all driven flywheels are disengaged, and the friction power generation unit continuously outputs electric energy. The vehicle body acceleration gain of the multi-stage adaptive inerter is at the lowest level among the three curves in the wide frequency range of 0.1~15Hz, indicating that the device can provide superior vibration isolation performance in the wide frequency range.

[0115] As Figure 11 shown, the entire process of state switching between the inerter and damping of the present invention is autonomously driven by vibration frequency, without the intervention of sensors, controllers and external energy sources. After the road vertical vibration excitation is input, the state switching process is started: when f<f1, the drive shaft and each flywheel are in a synchronous state, presenting pure inerter characteristics; when f1<f<f2, the outermost magnetically coupled driven flywheel starts to slip. Since this stage is non-contact magnetic coupling, no significant friction energy consumption is generated, the equivalent damping coefficient remains approximately constant, and the device is still dominated by inerter characteristics but with reduced inertance; when f2<f<f3, the second driven flywheel starts to slip, presenting the synergy of inerter and damping, with decreasing inertance, increasing damping, partial energy recovery and partial energy dissipation; when f>f3, all driven flywheels slip, presenting pure damping characteristics, most of the energy is recovered, the remaining part is dissipated, and high-frequency vibration transmission is effectively blocked. The states of each frequency band are passively switched in sequence, the device matches the optimal mechanical characteristics, and efficient vibration isolation of wide-band vibration is completed.

[0116] It should be understood that although this specification is described in terms of various embodiments, not each embodiment only contains an independent technical solution. This narrative manner of the specification is merely for clarity, and those skilled in the art should regard the specification as a whole, and the technical solutions in each embodiment can also be appropriately combined to form other implementations that can be understood by those skilled in the art.

[0117] The series of detailed descriptions listed above are only specific descriptions of feasible embodiments of the present invention, and they are not intended to limit the protection scope of the present invention. All equivalent embodiments or modifications made without departing from the technical spirit of the present invention shall be included within the protection scope of the present invention.

Claims

1. A distributed drive vehicle vertical vibration broadband control device, characterized in that, An inertia coefficient switching component (9) is provided between the vehicle body and the suspension. The inertia-mass switching component (9) includes a housing, a drive component, and a plurality of driven flywheels coaxially sleeved on the outside of the drive component from the inside to the outside in a radial direction; the drive component is rotatably supported in the housing, and a motion conversion mechanism is provided between the drive component and the housing to convert the vertical linear motion between the vehicle body and the suspension into the rotational motion of the drive component; each of the driven flywheels is supported on the housing. A magnetic coupling pair is provided between the outermost driven flywheel and the adjacent second outermost driven flywheel located radially inside it; a friction coupling pair is provided between the innermost driven flywheel and the drive component; each driven flywheel located between the second outermost driven flywheel and the innermost driven flywheel is provided with a friction coupling pair with the adjacent driven flywheel located radially inside it. Each of the driven flywheels has a preset transmission torque threshold, and the transmission torque threshold of each driven flywheel increases progressively from the outside to the inside in the radial direction.

2. The distributed drive vehicle vertical vibration broadband control device according to claim 1, characterized in that, The outer circumferential surface of the driven flywheel located in the second outermost layer is provided with a driving magnetic ring, and the inner circumferential surface of the driven flywheel located in the outermost layer is provided with a driven magnetic ring. A radial working air gap is maintained between the outer cylindrical surface of the driving magnetic ring and the inner cylindrical surface of the driven magnetic ring. The driven flywheel located in the outermost layer forms a non-contact magnetic coupling with the driven flywheel located in the second outermost layer through the radial working air gap.

3. The distributed drive vehicle vertical vibration broadband control device according to claim 1, characterized in that, A first-stage contact friction pair is provided between the inner circumferential surface of the innermost driven flywheel and the outer circumferential surface of the driving component, providing the first-stage static friction torque through a preset radial clamping force; when N≥3, an (i+1)th-stage contact friction pair is provided between the inner circumferential surface of the i-th driven flywheel between the second outermost driven flywheel and the innermost driven flywheel and the outer circumferential surface of its adjacent driven flywheel, providing the (i+1)th-stage static friction torque through a preset radial clamping force, where 1≤i≤N-2, and N is the number of driven flywheels; along the radial direction from the outside to the inside, the static friction torque of each stage increases progressively, and the first-stage static friction torque is greater than the static friction torque of the other stages.

4. The distributed drive vehicle vertical vibration broadband control device according to claim 1, characterized in that, The motion conversion mechanism is a ball screw pair, the nut of the ball screw pair is connected to the housing, and the screw of the ball screw pair is connected to the drive component or is an integral structure.

5. The distributed drive vehicle vertical vibration broadband control device according to claim 3, characterized in that, The friction interface of the contact friction pair is provided with a wear-resistant friction plate, and the wear-resistant friction plate is provided with an adjustable radial preload by a disc spring.

6. The distributed drive vehicle vertical vibration broadband control device according to claim 1, characterized in that, At least one of the aforementioned triboelectric coupling pairs is a triboelectric nano-power generation coupling pair, which is composed of a triboelectric material layer disposed on the surface of two relatively moving components and an electrode layer disposed on the back side of the triboelectric material layer. When the two relatively moving components produce relative slippage, electrical energy is output.

7. A passive control method for a distributed drive vehicle vertical vibration broadband control device according to claim 1, characterized in that, Includes the following steps: N driven flywheels are coaxially mounted radially from the inside out; the torque threshold values ​​of each driven flywheel radially from the outside in are T1, T2, ..., T. N And T1 <T2<…<T N Each transmission torque threshold corresponds to N switching frequencies f1, f2, ..., f N And f1 <f2<…<f N ; The vertical excitation from the road surface is input into the drive unit, and each flywheel switches states according to the vibration frequency f of the excitation: When f≤f1, all N driven flywheels rotate synchronously with the drive component, and the inertia-mass coefficient switching component (9) exhibits pure inertia-capacity characteristics. The input vibration energy is temporarily stored in each flywheel in the form of rotational kinetic energy. When f k <f≤f k+1 When, where k=1,2,…,N−1, the first k driven flywheels counted from the outside to the inside have already slipped relative to the driving component, while the remaining Nk driven flywheels still rotate synchronously. The inertia mass coefficient switching component (9) simultaneously exhibits inertia-capacitance characteristics and damping characteristics, and the equivalent inertia mass coefficient decreases with increasing frequency and the equivalent damping coefficient increases with increasing frequency. When f>f N At that time, all N driven flywheels slip relative to the drive component, and the inertia mass coefficient switching component (9) exhibits pure damping characteristics, and the high-frequency vibration energy is dissipated by each slip interface.

8. The passive control method of the distributed drive vehicle vertical vibration broadband control device according to claim 7, characterized in that, The switching frequency f1 is determined by the transmission torque threshold T1 of the outermost magnetic coupling pair; the switching frequency f k+1 The maximum static friction torque threshold T of the friction coupling pair corresponding to the (k+1)th flywheel counted radially from the outside to the inside is respectively determined by... k+1 It has been determined.

9. The passive control method of the distributed drive vehicle vertical vibration broadband control device according to claim 7, characterized in that, The switching frequency f1 is determined by the following relationship: ， Where T1 is the maximum transmitted torque of the magnetic coupling pair; k t To convert linear motion into rotational motion, J1 is the moment of inertia of the outermost driven flywheel, and A is the vibration amplitude.

10. The passive control method of the distributed drive vehicle vertical vibration broadband control device according to claim 7, characterized in that, The switching frequency f k+1 They are determined through the following relationships respectively: ， Where: μ s,k+1 p is the static friction coefficient of the (k+1)th driven flywheel corresponding to the kth friction coupling pair, counted radially from the outside in; r,k+1 L represents the radial contact pressure of the k-th frictional coupling pair. f,k+1 R is the contact length of the k-th frictional coupling pair; f,k+1 Let k be the contact radius of the k-th frictional coupling pair; k t The transmission ratio for converting linear motion into rotary motion, J k+1 Let A be the moment of inertia of the (k+1)th driven flywheel counted radially from the outside in; let A be the vibration amplitude.

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