JPH09329184A - Magnetic spring having damping characteristics and vibration mechanism having the magnetic spring - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a magnetic spring having zero or negative damping characteristics by utilizing permanent magnets, by changing a geometrical dimension between two permanent magnets separated from each other so as to converted into bounce, and setting output side bounce larger than input side bounce from the balanced position of two permanent magnets. SOLUTION: A magnetic spring having zero or negative damping characteristic is so formed that a geometrical dimension is changed by changing an opposed area of permanent magnets 2, 4. Then, when a certain input is applied to a base 6 so that the base 6 is moved toward a top plate 8, a magnet placing mount 18 is moved in a right direction along a linear way 16 by the inertia force of a balance weight 24. As a result, the opposed area of two permanent magnets 2, 4 is gradually increased, maximum bounce can be generated in a position where the permanent magnets 2, 4 are positioned near to each other most or in a position where this position is passed through, and also the base 6 is moved downward by the bounce. While the base 6 is reciprocated once against the top plate 8, the magnetic spring shows negative damping characteristic.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a magnetic spring having a plurality of permanent magnets, and more specifically, a magnetic spring having a positive, zero or negative damping characteristic utilizing the repulsive force of a plurality of permanent magnets, and The present invention relates to a nonlinear vibration mechanism or a coefficient excitation vibration mechanism that has the magnetic spring and does not have a damping function structure but is stable.

[0002]

2. Description of the Related Art At present, vibration models have been proposed and put into practical use in various forms. Generally, the vibration characteristics have a load mass dependency and an input dependency. It is considered that the load mass dependency has a correlation with the curvature of the load-displacement characteristic, and the input dependency has a correlation with the hysteresis of the load-displacement characteristic.

Taking an automobile suspension system as an example, road conditions, driving stability, and impedance (difference)
Conditions etc. are the points for adjusting the ride comfort such as the spring constant of the suspension system. If optimum is to be obtained under all conditions, active control is required. There is a big difference between low-frequency and high-amplitude driving in low-frequency and high-amplitude regions. When the damping property is low, the transmissibility of displacement is high, and the resonance point moves to the low frequency side. In order to improve the damping property, it is necessary to increase the damping ratio of the damper or decrease the spring constant. Therefore, the performance of the conventional passive vibration model is limited.

A suspension seat will be described as a specific example. The suspension seat is a seat that is mainly mounted on earth-moving machines, off-road vehicles such as recreational vehicles (RV), and long-distance traveling vehicles such as trucks and buses, and has a vibration isolation mechanism. As such, for example, a metal spring, an air suspension, an air damper, etc. are used. These sheets are about 1.5-1
Attention has been focused on improving vibration isolation in the 2 Hz range, especially in the 3-5 Hz range. Therefore, the resonance frequency of the suspension seat is set between 1 and 2.5 Hz.

FIG. 51 shows the vibration characteristics of a conventional suspension seat. In FIG. 51, (a) shows a rigid seat, (b) shows a suspension seat, (c) shows a spring-rich seat, and (d) shows a damper-less suspension seat.

A sheet having a small spring constant (soft spring) causes a large dynamic displacement when exposed to impact or low-frequency vibration. For example, the stroke of the suspension mechanism is usually limited to 100 mm or less so as not to interfere with the driver's driving operation such as pedaling. Therefore, when it is largely displaced, it reaches the end of the suspension, and a bottoming feeling occurs.

Further, in 1994, Stiles conducted an actual vehicle running evaluation test with a tractor in order to investigate the effect of bottoming on the performance of the suspension seat. As a result, 45% of the suspension seat increases the acceleration applied to the driver, and it has been clarified that the vibration insulation due to the bottoming is deteriorated. A shock absorber is used as a bottoming feeling and a coping method when a sudden and instantaneous shock load is applied to a vehicle.

Recently, an active suspension seat has been proposed in which an actuator is attached to the seat and vibration is actively controlled to improve a seating feeling.

[0009]

However, a vibration isolation mechanism using a metal spring, an air suspension, an air damper, etc., has a vibration of 4 to 20 Hz among the vibrations transmitted from the vehicle body floor.
It was not possible to further improve the sitting feeling or the feeling of use by lowering the frequency of the vibration.

In addition to the fact that the active suspension seat is heavy and expensive, it is necessary to keep the actuator constantly operating.
When set to F, the vibration is directly transmitted to the occupant via the actuator, and the seating feeling is impaired.

Further, in a suspension seat using a shock absorber, if the damping is too large, the damping range of the suspension is about 1.4 times the resonance frequency,
It may deteriorate the original vibration insulation of the seat.

The present invention has been made in view of the above problems of the prior art, and provides a magnetic spring having a positive, zero or negative damping characteristic by using a permanent magnet, and at the same time, provides the magnetic spring. The purpose of the present invention is to realize a dynamic characteristic control system or a high-efficiency engine with a cheap and simple structure by providing a nonlinear vibration mechanism or a coefficient excitation vibration mechanism that uses a spring but does not have a damping function structure but is stable. .

[0013]

In order to achieve the above object, the invention according to claim 1 of the present invention is characterized in that at least two permanent magnets are separated from each other, and a geometry between the at least two permanent magnets is provided. The mechanical dimensions are changed by input / output by a mechanism in the movement stroke or external force, and converted into a repulsive force within the movement system, so that the repulsive force on the output side from the repulsive force on the input side from the equilibrium position of the at least two permanent magnets is changed. It is a magnetic spring having a zero or negative damping characteristic characterized by a large repulsive force.

The invention according to claim 2 is characterized in that the maximum repulsive force is generated at the closest position of the at least two permanent magnets or at a position beyond the closest position.

Further, in the invention according to claim 3, the geometrical dimension is changed by changing any one of the separation distance, the facing area, the magnetic flux density and the magnetic field of the at least two permanent magnets. It is characterized by having done.

Further, in the invention according to claim 4, at least two permanent magnets are separated from each other, and the geometrical dimension between the at least two permanent magnets is changed by input / output by a mechanism in a motion stroke or an external force. A magnetic spring having a positive damping characteristic, which is characterized by a nonlinear damping / spring structure by extracting a damping term.

The invention described in claim 5 is the same as the invention in claim 4.
In the invention described in (3), the maximum repulsive force is generated at the closest position of the at least two permanent magnets.

The invention according to claim 6 is the same as the invention according to claim 4.
In the invention described in (3), the geometrical dimension is changed by changing any one of the separation distance, the facing area, the magnetic flux density, and the magnetic field of the at least two permanent magnets.

Further, the invention according to claim 7 is a magnetic spring exhibiting zero or negative damping characteristics by changing the geometrical dimension between at least two permanent magnets separated from each other by a mechanism in the movement stroke or an external force. And has a structure for converting energy into continuous or divergent vibration.

The invention according to claim 8 is the same as claim 7
In the invention described in (1), the geometrical dimension is changed by an external force so that the spring constant and the damping coefficient in the motion system of the mechanism are changed by the spring structure system.

The invention described in claim 9 is the same as claim 7
In the invention described in (1), the vibration characteristic is improved by changing the geometrical dimension by a mechanism in a movement stroke or an external force.

According to a tenth aspect of the invention, in the invention according to the seventh aspect, the excitation or the resonance frequency is made variable by changing the geometrical dimension by a mechanism in the movement stroke or an external force, and the resonance frequency is changed. It is characterized in that it operates so that the resonance frequency of the frequency characteristic is always lowered by tracking the frequency, or the fluctuation of the amplitude is reduced.

According to the invention described in claim 11, in the invention according to claim 7, the geometrical dimension is changed by a mechanism in a movement stroke or an external force, and when the displacement is small, the displacement is negative damping. It is characterized in that the vibration is made to be positive with the increase of, and the oscillation is made steady with an amplitude in which positive and negative damping are balanced.

Further, the invention according to claim 12 has a magnetic spring which exhibits a positive damping characteristic by changing the geometrical dimension between at least two permanent magnets separated from each other by a mechanism in the movement stroke or an external force. However, the non-linear vibration mechanism is characterized in that the damping characteristic is made larger than the spring characteristic of the mechanism in the movement stroke.

According to a thirteenth aspect of the invention, in the twelfth aspect of the invention, the vibration characteristics are improved by changing the geometrical dimensions by a mechanism in the movement stroke or an external force. .

According to a fourteenth aspect of the present invention, in the invention according to the twelfth aspect, the resonance frequency is made variable by changing the geometrical dimension by a mechanism in the movement stroke or an external force, and the resonance frequency is changed. Is operated so that the resonance frequency of the frequency characteristic is constantly lowered or the fluctuation of the amplitude is reduced.

Further, in the invention as set forth in claim 15, the resonance frequency is made variable by changing the geometrical dimension between at least two permanent magnets separated from each other by a mechanism in the movement stroke or an external force, and the resonance frequency is changed. Is a non-linear vibration mechanism characterized by generating a large acceleration or amplitude with a small input so that it can be operated at a location where the resonance frequency of the frequency characteristic is high or where there is a large fluctuation in amplitude. .

[0028]

BEST MODE FOR CARRYING OUT THE INVENTION Embodiments of the present invention will be described below with reference to the drawings. In the case of a magnetic spring structure having at least two permanent magnets spaced apart from each other and having the same magnetic poles opposed to each other, the separated permanent magnets are not in contact with each other, and therefore reversible when the friction loss of the structure itself is negligible, Its static characteristics are output nonlinearly (return) on the same line as when it is input (going), and by using the degree of freedom peculiar to the non-contact pair even and the instability of the levitation control system,
Negative attenuation is likely to occur by changing the static magnetic field (arrangement of magnets) with a small input.

The present invention has been made by paying attention to this fact, and the geometrical dimension between the two permanent magnets is changed on the input side (going) and the output side (returning) by the mechanism in the motion stroke or the external force. , The repulsive force on the output side is made larger than the repulsive force on the input side from the equilibrium position of the two permanent magnets by converting the repulsive force in the motion system.

The basic principle will be described below. FIG. 1 is a schematic diagram showing the equilibrium positions of the two permanent magnets 2 and 4 on the input side and the output side. FIG. 2 shows the load applied to one of the permanent magnets and the equilibrium of the two permanent magnets. The basic characteristics of the magnetic spring structure showing the relationship with the displacement amount from the position are shown.

As shown in FIG. 1, the input side equilibrium position and spring constant of the permanent magnet 4 with respect to the permanent magnet 2 are x ₀ and k ₁ , respectively, and the output side equilibrium position and spring constant are x ₁ and k ₁ , respectively. _{If the value is 2} , area conversion is performed between x ₀ and x ₁ , and the following relationship holds at each equilibrium position. _{-K 1 / x 0 + mg =} 0 -k 2 / x 1 + mg = 0 k 2> k 1

Therefore, the static characteristic shows a negative damping characteristic as shown in FIG. 2, and the potential difference between the position x ₁ and the position x ₀ can be considered as the potential energy of oscillation.

Further, when the model of FIG. 1 was manufactured and the relationship between the load and the displacement amount was actually measured while changing the load application time, a graph as shown in FIG. 3 was obtained. this is,
When the two permanent magnets 2 and 4 approach the closest position, a large repulsive force acts, and when the displacement amount from the equilibrium position changes slightly, friction loss occurs due to the damper effect of the magnetic spring. Is interpreted as the appearance of a damping term.

In FIG. 3, (a) is a graph when a constant load is applied, and the time when the load is applied becomes shorter in the order of (a), (b) and (c). That is, the static characteristics differ depending on how the load is applied, and the longer the load is applied, the greater the impulse.

In the rare earth magnet, the strength of magnetization does not depend on the magnetic field. In other words, since the internal magnetic moment is hardly affected by the magnetic field, the magnetization intensity hardly changes on the demagnetization curve, and the value of the saturation magnetization is almost maintained. Therefore, in the rare-earth magnet, input and output can be considered using a charge model that assumes that magnetic charges are uniformly distributed on the end face.

FIG. 4 shows the idea, in which a magnet is defined as a set of magnets having a minimum unit, and the relationship of forces between the unit magnets is classified into three and calculated. (A) withdrawing (r, since the same also m, defines the two types in one ^{1) f (1) = (} m 2 / r 2) dx 1 dy 1 dx 2 dy 2 f x (1) = f (1 _{^{) cosθ f z (1) =}} f (1) sinθ (b) repulsion _{^{f x (2) = f (}} 2) cosθ f z (2) = f (2) sinθ (c) rebound f _x ⁽³⁾ = f ⁽³⁾ cos θ f _z ⁽³⁾ = f ⁽³⁾ sin θ Therefore, −f _x = 2f _x ⁽¹⁾ −f _x ⁽²⁾ −f _x ⁽³⁾ −f _z = 2f _z ⁽¹⁾ −f _z ⁽²⁾ −f _z ⁽³⁾ Here, Coulomb's law is expressed as Can be determined force by integrating the -f _x, the -f _z in the range of dimensions of the magnet.

As shown in FIG. 5, this is completely deviated (x-axis movement amount = 5) from the state (x-axis movement amount = 0 mm) in which opposing magnets are completely wrapped for each magnetic gap.
0 mm) is calculated in the graph of FIG. However, it is defined as "the internal magnetic moment is constant", but when the magnetic gap is small, the disturbance is generated around the magnet, so the correction is made.

The above calculation results substantially match the measured values, and the force to move from point a to b in FIG. 2 is the load in the x direction, and the output is the load in the z direction. <The output relationship is clarified statically.

Further, in FIG. 7, the separation distance of the magnets shown in FIG. 5 is maintained at 3 mm, the magnet is moved from the completely displaced state to the completely wrapped state, and further moved from this state to the completely displaced state. It is a graph showing the relationship when doing. This graph shows that the absolute value of the load in the x direction is the same and the output direction is reversed, and the output direction is reversed. When approaching the fully wrapped state, resistance or attenuation occurs, and the state shifts from the completely wrapped state to a completely deviated state The case indicates that it will be accelerated. By utilizing this characteristic for a non-contact damper, it becomes possible to reduce vibration energy in low, medium, and high frequency regions (0 to 50 Hz) that can be perceived by a person who could not be achieved by the conventional damper, that is, to improve the vibration transmission rate. Was.

Further, as shown in FIG. 8, when the rotation angle of the opposing magnets was changed, a graph as shown in FIG. 9 was obtained. Naturally, as the facing area decreases, the maximum load decreases, indicating that the output can be changed via area conversion by applying a predetermined input.

FIG. 10 is a graph showing the relationship between the distance between magnets and the load when a neodymium magnet is used as the permanent magnet, and the repulsive force increases as the mass increases. Here, the repulsive force F is represented by F∝Br ² × (geometrical dimension) Br: strength of magnetization, and the geometrical dimension means a distance between opposing magnets, an opposing area, a magnetic flux density, and a magnetic field. Means the size determined by the strength of If the magnet material is the same, the strength of the magnetization (Br) is constant, so that the repulsive force of the magnet can be changed by changing the geometric dimensions.

FIG. 11 shows a first specific magnetic spring model in which the geometrical dimensions are changed by changing the facing areas of the permanent magnets 2, 4. FIG.
, The base 6 and the top plate 8 extending in parallel with each other,
They are connected to each other by a pair of left and right X links 10 composed of two links 10a and 10b. Link 10a, 1
One end of Ob is pivotally connected to the base 6 and the top plate 8, respectively, and the other end of the links 10a and 10b is slidable to the upper slider 12 slidably attached to the top plate 8 and the base 6. Are pivotally attached to the lower sliders 14 attached to the respective sliders.

The linear way 16 is fixed to the base 6, and the magnet mounting table 18 on which the permanent magnets 2 are mounted is slidably attached to the linear way 16, while the other permanent magnet 4 is mounted on the top plate. It is fixed at 8. A support 20 is further fixed to the base 6, and an L-shaped lever 22 including a first arm 22a and a second arm 22b is fixed to the support 20.
The center of is pivoted. One end of the first arm 22a is pivotally attached to the magnet mounting table 18, and the second arm 22a
The balance weight 24 is attached to b.

In the above structure, when an input is applied to the base 6 and the base 6 moves toward the top plate 8, the inertial force of the balance weight 24 causes the magnet mounting table 18 to move along the linear way 16 to the right in the figure. Move in the direction. as a result,
The opposing area of the two permanent magnets 2 and 4 gradually increases, and a maximum repulsive force is generated at a position where the permanent magnets 2 and 4 are closest to or past this position, and the base 6 is lowered by the repulsive force. Go to While the base 6 reciprocates once with respect to the top plate 8, the magnetic spring of FIG. 11 exhibits a negative damping characteristic as shown in FIG. Since the balance weight 24 has a slight phase delay with respect to the base 6, the position where the maximum repulsive force is generated is appropriately adjusted by moving the balance weight 24 with respect to the second arm 22b according to the input. be able to. Further, the timing and the facing area can be adjusted by linking the permanent magnet 4 with the upper slider 12.

FIG. 12 shows a base 6 having two permanent magnets 2 and 4.
And the top plate 8 respectively, and the other two permanent magnets 26,
The geometric dimensions are changed by making the gap (separation distance) between 28 variable.
In FIG. 12, the permanent magnet 28 is fixed to the top plate 8 with the S pole facing downward, opposite to the permanent magnet 4, while the permanent magnet 26 has the S pole facing upward, opposite to the permanent magnet 2. The swing lever 30 is fixed to one end of the swing lever 30 in the turned state. The swing lever 30 has a center portion swingably attached to the support base 20, and a balance weight 24 is attached to the swing lever 30 on the opposite side of the permanent magnet 26.

In the above structure, since the two permanent magnets 2 and 26 attached to the base 6 have opposite magnetic poles facing each other, an attractive force acts between the permanent magnets 2 and 26, and this attractive force acts as a balance spring. To work. When an input on the base 6 is applied and the base 6 moves toward the top plate 8, the inertia of the balance weight 24 causes the permanent magnet 26 to swing upward against the attraction with the permanent magnet 2. Move. As a result, the gap between the permanent magnets 26 and 28 gradually changes, and a maximum repulsive force is generated at the closest position or a position past this position, and the base 6 is moved downward by the repulsive force. While the base 6 makes one round trip to the top plate 8,
The magnetic spring shown in FIG. 12 exhibits a negative damping characteristic as shown in FIG. It should be noted that the position where the maximum repulsive force is generated can be appropriately adjusted by moving the balance weight 24 with respect to the swing lever 30 in accordance with the input, similarly to the model in FIG.

FIG. 13 shows a structure in which the rotary lever is used to change the geometrical dimensions of the two opposing permanent magnets 2 and 4. In FIG. 13, a permanent magnet 2 is fixed to a base 6, and a permanent magnet 4 facing the permanent magnet 2.
Is fixed to a mounting base 34 slidably mounted on a frame 32 erected on the base 6. Mounting stand 34
, One end of a link 36 is pivotally mounted, and the other end is pivotally mounted on a first support base 38 fixed to one side of the lower slider 14.

On the opposite side of the lower slider 14, a second
The support base 40 is fixed, and a pin 44 is attached to the other end of the lever 42 having one end pivotally connected to the second support base 40. The pin 44 is loosely inserted into a long hole 36 a formed in an intermediate portion of the link 36, and is attached to a lower end of an arm 46 pivotally attached to the top plate 8.

In the above structure, when an input is applied to the base 6 and the base 6 moves toward the top plate 8, the lever 4
2 rotates in the direction of the arrow in the figure, and the two permanent magnets 2 and 4 approach each other. Since the permanent magnets 2 and 4 face the same magnetic pole, the repulsive force gradually increases with the rotation of the lever 42, and when the permanent magnets 2 and 4 pass the closest position, the permanent magnets 2 and 4 Depart from each other. Base 6
13 makes one reciprocation with respect to the top plate 8, the magnetic spring of FIG. 13 exhibits a negative damping characteristic as shown in FIG. 3 as the lever ratio changes.

FIG. 14 shows a magnetic spring in which the geometrical dimension is changed by utilizing the pole conversion of a permanent magnet. In FIG. 14, a small-diameter pulley 48 is integrally fixed to the permanent magnet 2 rotatably attached to the base 6, and the pulley 48 is a large-diameter rotatably fixed to the base 6. It is connected to a pulley 50 by a belt 52. One end of a link 54 is fixed to the center of the pulley 50, and a lever 56 to which the balance weight 24 is attached is fixed to the other end of the link 54. In addition, the lower end position of the balance weight 24 is
And is regulated by a spring member 60 attached via the.

In the above structure, when an input is applied to the base 6 and the base 6 moves toward the top plate 8, the inertial force of the balance weight 24 causes the large-diameter pulley 50 to rotate in the direction of the arrow in the drawing, and the belt. The permanent magnet 2 rotates in the same direction via 52. As a result, the S pole of the permanent magnet 2 is attracted to the N pole of the permanent magnet 4 fixed to the top plate 8,
If the balance weight 24 follows the phase lag by a certain amount, the permanent magnet 2 will rotate in the opposite direction of the arrow.
The N pole of the permanent magnet 2 faces. When the same magnetic poles face each other, a repulsive force is generated, and the base 6 descends so as to separate from the top plate 8. While the base 6 makes one reciprocation, the magnetic spring of FIG. 14 is as shown in FIG. Shows a negative attenuation characteristic.

FIG. 15 shows a magnetic spring in which the geometrical dimension is changed by changing the magnetic flux density of the permanent magnet. In FIG. 15, a first support plate 62 fixed to the base 6 and the first support plate 62
Support plate 6 extending in parallel with a predetermined distance
4, both ends of the plurality of shielding plates 66 are pivotally connected. One end of the second support plate 64 is pivotally connected to an intermediate portion of an L-shaped lever 70 via an arm 68,
One end of the L-shaped lever 70 is connected to a support base 7 fixed to the base 6.
2 and a balance weight 24 is attached to the other end.

In the above structure, when an input is applied to the base 6 and the base 6 moves toward the top plate 8, the inertial force of the balance weight 24 moves the second support plate 64 in the direction of the arrow in the figure. The shield plate 66 is above the permanent magnet 2.
To some extent. As a result, the magnetic flux density of the permanent magnet 2 attached to the base 6 decreases, and the repulsive force with the permanent magnet 4 attached to the top plate 8 decreases.

When the balance weight 24 follows a slight phase delay and then follows, the second support plate 64 moves in the direction opposite to the arrow, so that the upper side of the permanent magnet 2 is opened and the repulsive force of the permanent magnets 2, 4 increases. Then, the base 6 descends so as to be separated from the top plate 8, but while the base 6 makes one reciprocation, the magnetic spring of FIG. 15 exhibits a negative damping characteristic as shown in FIG.

Next, the dynamic characteristics of the magnetic spring will be described with a simplified basic model shown in FIG.
The input F in FIG. 16 is a force caused by a geometric dimensional change such as the area conversion of the permanent magnet.

Further, FIG. 17 shows a facing area of 50 × 25 mm.
² , two permanent magnets with a thickness of 10 mm (Nd-Fe-B
2 shows the relationship between the distance between magnet surfaces (x) and the repulsive force (f) when they are opposed to each other so as to repel each other. The solid line is the result of regression analysis using the Levenberg-Marquardt algorithm, and shows that the relationship of f = 66 / x is well followed. That is, the repulsive force acting between the magnets is given by k / x.

In consideration of this, the function of the magnetic spring characteristic was converted into a function equation. Since the repulsive force acting between the magnets is given by k / z as described above, in FIG. 16, the mass including the mass of the upper permanent magnet 4 and the load applied to the magnet 4 is m, the spring constant is k, and the damping coefficient is Is r, and the harmonic vibration input to the mass m is F (t), the equation of state is

[Equation 1] It is expressed as

Here, when the equilibrium position is x ₀ and the displacement from the equilibrium position is y,

[Equation 2]

Here, if k / x ₀ ² = k 'is set,

(Equation 3)

[0060] The harmonic vibration ^F (t) = Fe iωt Distant, y
= Xe ^iωt ,

(Equation 4) Here, φ indicates a phase delay.

(Equation 5) Therefore, the natural frequency (resonance frequency) ω ₀ is

(Equation 6) Becomes The relationship between the natural frequency and the spring constant is opposite to that of the metal spring. That is, since it is non-linear, the resonance point can be made constant if the optimum curvature of the load-displacement characteristic can be obtained by adjusting the set position of the operating point and the magnetic circuit.

Here, the equation (2) can be further expressed as follows.

(Equation 7) When y is x and considering the third-order term,

(Equation 8)

In equation (3), the attenuation term of -bx ² appears in the quadratic term, but if equation (3) is replaced with a simpler image,

[Equation 9]

Here, if x = x ₀ cos ωt,

(Equation 10)

That is, in the minute vibration region, a constant repulsive force ((b / 2) x ₀ ² ) is constantly applied to the periodic external force, and the periodic external force is attenuated by the force. . That is, by adjusting the trajectory of the movement of the permanent magnet, it is possible to obtain a damping effect without adding a damper mechanism.

Then, when the dynamic characteristics of the magnet alone were examined using the apparatus of FIG. 18, the results shown in FIGS. 19 and 20 were obtained.

In the apparatus shown in FIG. 18, the two permanent magnets 2 and 4 are made to face each other, and the X link 10 is not converted.
Is a device that changes the separation distance via the.

19 and 20, the horizontal axis represents frequency (Hz) and the vertical axis represents vibration transmissibility (G / G). In FIG. 19, (a), (b),
(C), (d), (e) and (f) are each 50 × 5
0x10mm, 50x50x15mm, 50x50x2
0mm, 75 × 75 × 15mm, 75 × 75 × 20m
20 and the same load of 30 kg was applied using a magnet of 75 × 75 × 25 mm in FIG.
Different loads of 53 kg and 80 kg were applied using the same magnet of 0 × 50 × 20 mm.

19 and 20 show the non-linear characteristics of the magnetic spring. From both figures, when the load mass is the same,
When the facing area of the magnets increases, the distance between the magnets increases, the resonance point shifts to the low frequency range, and the vibration transmissibility also decreases.
That is, the behavior is opposite to that of the metal spring and the air spring. On the other hand, when the magnet size is the same, the resonance point does not change even if the load mass changes, and the vibration transmissibility decreases as the load mass increases. In other words, it can be seen that the vibration transmissibility at the resonance point varies depending on the weight of the load.

From the above, by designing a load-displacement curve having the same natural frequency, only in the low frequency region, the resonance point is kept constant even if the load mass changes, and the vibration transmissibility is maintained. Can be kept small. These are damping effects due to the spring constant k and the load mass m described in the equations (4), (5), and (6).

FIG. 21 is a graph showing the dynamic characteristics of a conventional passenger car seat as a comparative example, in which the vibration transmissibility is high as a whole, and both the resonance point and the vibration transmissibility fluctuate as the load changes. There is.

By the way, in the above formula (1), when the geometrical dimension between the opposing permanent magnets is changed by a mechanism within the movement stroke (a mechanism for moving the permanent magnet in the repulsion system) or an external force, the spring constant k 22 is a rectangular wave k (t) that changes with time as shown in FIG. 22, and takes the values of + k ′ and −k ′ alternately in every ½ cycle in the cycle T = 2π / ω. . Therefore, equation (1) is expressed as follows.

[Equation 11] (I) When 0 <t <π / ω,

(Equation 12) (Ii) When π / ω ≦ t <2π / ω,

(Equation 13)

When the equilibrium position when 0 <t <π / ω is x ₀ and the displacement from the equilibrium position is y ₁ ,

[Equation 14]

If (n−k ′) / x ₀ ² = k ₁ ′ is set,

(Equation 15)

[0074] The harmonic vibration ^F (t) = Fe iωt Distant, y
₁ = xe ^iωt ,

(Equation 16) Here, φ indicates a phase delay.

[Equation 17] Therefore, the resonance frequency ω ₀ is

(Equation 18)

Similarly, when π / ω ≦ t <2π / ω,

[Equation 19] Therefore, divergence occurs when y ₁ <y ₂ .

In general, the self-excited vibration system can be replaced with a spring-mass system having negative viscous damping, and vibration energy is introduced from the outside during vibration. However, the vibration actually generated is air resistance at the mass point. And various kinds of resistance are generated and energy is lost.

However, when vibration energy is introduced as an external force to the magnetic spring of the present invention having a negative damping characteristic, as described above, the energy diverges with y ₁ <y ₂ , and if the energy continues to diverge, the amplitude gradually increases. The system is destroyed, or
By adding a damping term that increases as the displacement increases to the above-mentioned equation of state, a positive damping acts and a steady oscillation is performed in a state of being balanced with the negative damping. That is, similarly to the spring constant k (t), the damping coefficient is also variable, and the equation (1) can be further rewritten as follows.

(Equation 20)

In the vibration system having the magnetic spring of the present invention, an energy change / conversion system for inducing continuous vibration and divergent vibration exists inside the vibration system, and a positive damping term is mechanically added to the above equation of state. As a result, the following equation of state can be obtained.

(Equation 21)

In this equation of state, when r ₂ ≠ 0, as x increases, the third term on the left side becomes larger, and the damping term of the spring term causes positive damping. Therefore, as the internal excitation characteristics of the permanent magnet, when the displacement is small, the damping is a negative damping, and the positive damping works as the displacement increases, so that the vibration becomes steady at an amplitude where the positive and the negative damping balance.

When the magnitude of at least one of the mass, damping coefficient and spring constant of the vibration system changes with time, the vibration caused by this is called coefficient excitation vibration. 7), (8), and (9) are coefficient excitation vibrations in which the excitation source itself vibrates, and non-oscillating energy in the system is converted into oscillatory excitation in the system to generate vibration.

Since the supplied energy is usually a part of the motive energy converted, if the kinetic energy has an upper limit, the supplied energy is also limited, and the amplitude is suppressed when this becomes equal to the consumed energy. The potential energy of the permanent magnet is independent of the power energy of the system and can widen the gap with the energy consumption. However, if the maximum energy product per mass of the permanent magnet increases, this gap can be further increased. Can be expanded to 1
By making the supplied energy due to negative damping greater than the energy consumed due to damping during the cycle, the vibration energy is increased.

As described above, in the formula (1), the damping coefficient r and the spring constant (coefficient) k can be freely controlled. For example, in the schematic view of FIG. 1, the permanent magnet 4 is at the lowermost end. At this time, the amplitude can be attenuated by maximizing the facing area with the permanent magnet 2, and it can be applied to a magnetic brake, a dynamic vibration absorber, or the like. In addition, since the repulsive force can be increased by maximizing the facing area after the permanent magnet 4 is separated from the lowermost end toward the uppermost end, the permanent magnet 4 can be applied to a generator, an amplifier, and the like.

As can be seen from the solution of the above equation of state, the coefficient-excited vibration system of the present invention reduces fluctuations in amplitude by moving the excitation frequency even if the natural frequency changes due to load fluctuations. can do. In other words, the excitation frequency is variable, and the resonance frequency can be manually or automatically tracked to operate at a position where the resonance frequency of the frequency characteristic always decreases, and used as an anti-vibration device for an automobile seat. Thereby, the vibration insulation can be improved, and the individual performance thereof can be improved. For example, the resonance point can be lowered to 4 Hz or less. Also, it is possible to improve the low frequency by using the negative attenuation and to absorb the weight difference by specializing the nonlinear characteristics of the permanent magnet.

Here, a vibration experiment was conducted using a pad combining urethane and fiber or a bed type vibration isolation unit adopting the magnetic spring structure of the present invention.
The result as shown in FIG. 23 was obtained.

As can be seen from the graph of FIG. 23, when the magnetic spring structure of the present invention is used together with the pad, the resonance frequency is reduced to 3 Hz, which is less than half of that when only the pad is used. It was found to be extremely effective. Furthermore, by performing the semi-active control, the vibration transmissibility at the resonance point could be reduced to about 1/3.

Further, the maglev (magnetic lev
When the dynamic characteristics of the unit were examined, the results as shown in FIG. 25 were obtained.

In the maglev unit shown in FIG. 24, a seat 78 is swingably supported on a base 74 via a plurality of swing levers 76, and two permanent magnets 80, 8 are provided on the upper surface of the base 74.
2 are fixed at a predetermined distance, and a permanent magnetic pole 84 having the same magnetic pole facing the permanent magnets 80 and 82 is fixed to the lower surface of the sheet 78. The permanent magnetic poles 80, 8
As 2,84, one having a size of 75 × 75 × 25 mm was used.

In this maglev unit, 53 kg, 75 k
Although different loads of g and 80 kg were applied, as shown in FIG. 25, the difference in the vibration transmissibility due to the fluctuation of the load could be suppressed small, and the resonance points could be substantially matched.

When the ride comfort evaluations of the passenger car seat, the suspension seat A, the suspension seat B, and the maglev unit according to the present invention were examined, the results shown in FIG. 26 were obtained. The load of the maglev unit was 53 kg, and a 75 × 75 × 25 mm permanent magnet was used. In the drawing, “fixed” indicates a state in which the seat is merely fixed to the suspension, and urethane, gel, and styrene indicate a cushion material mounted on the unit.

Here, as the riding comfort evaluation constant, "SAE
paper 820309 "and a ride comfort index R (Ride Number) expressed by the following equation: R = K / (A.B.fn) The variables A, B and fn are transfer functions of the seat (T.F.). .), And shows the following values: A: Maximum value of TF B: TF value at 10 Hz fn: Resonance frequency or frequency at which A appears K: A completely different sheet Expressed ride comfort coefficient (K value is set to "1" because various seats were used.) The ISO ride comfort evaluation indicates that the ride comfort is good with a small numerical value. The higher the value, the better the ride.

As can be seen from FIG. 26, among the seats for which the ride comfort was evaluated, the passenger car seats were 0.2 to 0.3 (all urethane type), 0.3 to 0.5 (spring type), and weight adjustment. The suspension seat subjected to the test showed a value of 0.5 to 0.7, and the riding comfort of the maglev unit of the present invention was better than other seats, and a riding comfort rating of 0.75 to 1.60 for a load of 53 kg was evaluated. The constant was obtained.

FIG. 27 shows the riding comfort evaluation constants of the maglev unit when the load is changed. As can be seen from this figure, a riding comfort evaluation constant of 0.7 or more is obtained for any load. In other words, the riding comfort of the maglev unit according to the present invention is shown.

FIG. 28 shows the dynamic characteristics of the passenger car seat, the suspension seat A, the suspension seat B, and the maglev unit according to the present invention.
In the drawing, (a) is a passenger car seat, (b) and (c) are suspension seats A with a load of 53 kg and 75 kg, respectively, and (d) and (e) are suspension seats B and 45 kg and 75 kg, respectively. (F) and (g) show the case where the cushioning material is changed in the maglev unit according to the present invention, and (h) shows the case where the maglev unit according to the present invention is semi-actively controlled. .

As can be seen from FIG. 28, the resonance point of the maglev unit is between 2 and 3 Hz, and the vibration transmissibility in the low and high frequency regions is also small. Furthermore, it has been confirmed that by performing the semi-active control, the resonance point can be further reduced and the vibration transmissibility can be reduced in a wide frequency range.

Further, the collision vibration can be utilized in the nonlinear vibration system or the coefficient excitation vibration system of the present invention. Collision is a typical nonlinear phenomenon of mechanical systems along with friction.
When a collision occurs, an object that suddenly hinders the movement, such as the deformation resistance of the object, acts, and thus decelerates rapidly to produce a very large acceleration. The magnetic spring also causes the same phenomenon (quasi) as the collision.

When an object collides with a certain kinetic energy, it deforms at the contact portion, that is, plastic deformation work, friction work on the contact surface, elastic waves inside the object, and is dissipated as acoustic energy to the outside, and the rest is elastic energy. Is converted into kinetic energy. As described above, in the case of the magnetic spring, there is no large loss due to non-contact, and the static characteristics return non-linearly on the same line and tend to cause negative damping.

For example, when the maglev unit does not hit the stopper, it exhibits vibration characteristics which are converted into acceleration and self-excited by a repulsive force of + α, or have low damping vibration due to non-contact but do not adversely affect the human body. Further, in combination with a metal spring, when the acceleration exceeds the damping, it is possible to induce a complete elastic collision by the hard spring to excite the self-excitation, thereby preventing the secondary resonance. The energy loss can be compensated by converting the potential energy of the magnetic field.

Further, as a general basic principle of anti-vibration, it is necessary to consider the directivity of mass effect, vibration insulation, vibration damping, vibration interference, and propagation. When elastically supported, vertical motion and roll are induced. Therefore, the vibration-proof foundation should be heavy and large, and the support span should be long. Further, when damping is given by using a viscous damper and a friction damper together, the energy given by the impact can be quickly dissipated by the damper or the like by the next impact, and the vibration can be attenuated.

Further, in order to suppress the frictional damping, the stopper is elastically supported to utilize the opposite impact to perform vibration isolation and energy conversion, thereby compensating for the insufficient repulsive force of the magnetic spring.

FIG. 29 shows a model in which the stopper is elastically supported. The spring constant k of the elastic support member is made variable and can absorb a predetermined acceleration or amplitude.
The resonance point can be adjusted by appropriately adjusting the spring constant k.

With this structure, when acceleration or amplitude of a predetermined value or less is applied to the stopper, the elastic support member elastically deforms to suppress friction damping, and the opposing impact is used to compensate for the insufficient repulsive force of the magnetic spring. In addition, the vibration isolation performance can be improved.

FIG. 30 shows experimental values of input and output of a slide-type principle model in which a facing area is 50 × 25 mm ² and a thickness is 10 mm, and friction loss is suppressed as small as possible. Was 3.135 kg.

Similarly, the area conversion rate is 80% (facing area: 2
Consider a case where the slide type of 50 → 1250 mm ² ) is changed to the rotary type in which the degree of area change is non-linear and the area conversion rate is changed to 50% (opposing area: 625 → 1250 mm ² ).

FIG. 31 shows experimental values of input and output of a rotary principle model in which a magnet having an opposing area of 50 × 25 mm ² and a thickness of 10 mm is opposed, and the area is converted with the center of gravity of one magnet as the center of rotation. FIG. 32 shows experimental values of input and output of the same rotation type principle model.

Energy extraction using a permanent magnet means that apparent energy is produced by increasing the difference in (apparent output / input). FIG. 33 shows the main points of the input / output principle model. Since it is a non-contact system, acceleration can also be used and larger energy can be apparently generated. The principle of virtual work is applied to the repulsive force acting between the magnets, and the amount of change in the stored magnetic energy due to the movement of the magnet is equal to the amount of work due to the movement of the magnet. How to extract magnetic energy is the key to boosting actuators.

That is, work is required to set a permanent magnet at infinity at a certain finite position. Once installed, for example, by changing the facing area of the permanent magnet of the repulsion system as a trigger, the work used for the installation, that is, the stored magnetic energy is released and used as an output, So that the power can be amplified.

Inputting electrical energy has the same effect as a transistor for amplification, but the characteristic of this amplifier is to efficiently convert the stored magnetic energy into mechanical energy and utilize it. That is, apparently
A small input (work) will produce a large output (work).

The work W is represented by W = W _g (h) + W _m (h) = mgh + W _m (h). The amount of change in energy ΔW is ΔW = mg · Δh + ΔW _m (h). Since mg · Δh >> ΔW, mg · Δh−ΔW = −ΔW _m (h)> 0 ΔW _m (h) indicates a decrease in the accumulated magnetic energy. The rotation type model has −ΔW _m (h) ≈ΔW and mg · Δh≈2ΔW.

Further, the accumulated magnetic energy is W _m (h) = 1 / 2BHV = B ² Sh / (2μ ₀ ) when the distance between the magnets is small and the magnetic flux density is uniform. At this time, B is the magnetic flux density of the void, H is the magnetic field of the void, V is the volume of the void, h is the distance of the void, and S is the magnet cross-sectional area. The change in the stored magnetic energy due to the movement of the magnet by Δh is ΔW _m (h) = B ² SΔh / (2μ ₀ ). When the repulsive force between the magnets is F, the work amount due to the movement is FΔh. ΔW _m (h) = FΔh, and the repulsive force F is represented by F = B ² S / (2μ ₀ ) (N). Br = 1.0T, facing area 100 × 100
Calculating mm ² and thickness 10 mm with the charge model,
34 is obtained.

FIG. 35 shows how the repulsive force changes, and the magnet used has a facing area of 50 × 25 m.
m ² and thickness 10 mm.

Similarly, consider a metal spring model. FIG. 36 shows the metal spring model shown in FIG. 37, where mg = 10.
N, k = 1 N / mm, L = 200 mm, no friction
Is the calculated value in the ideal state without mechanical deformation.

Statically, metal springs, air springs, and magnetic springs show the same tendency. However, the magnetic levitation pair is a lower order kinematic pair than the conventional mechanical kinematic pair, and when the aspects of non-linearity and utilization of acceleration are added, a great difference in practical use, including efficiency, occurs. 38 shows a rotary type principle model, and FIG. 39 shows a slide type principle model. In the rotary principle model of FIG. 38, the lower permanent magnet 2 has a base 90.
The upper permanent magnet 4 is attached to the slider 92 slidably in the vertical direction. Therefore, the two permanent magnets 2 and 4 facing each other are
By changing the separation distance or the facing area,
35 shows a load-displacement characteristic as shown in FIG.

On the other hand, in the slide type principle model of FIG. 39, the lower permanent magnet 2 is attached to the base 90 so as to be slidable in the horizontal direction, and the upper permanent magnet 4 is attached to the slider 92.
It is attached to the upside down so that it can slide freely. Therefore,
The two permanent magnets 2 and 4 facing each other show the input / output work characteristics as shown in FIG. 36 by changing the spacing distance or the facing area.

Therefore, when the characteristics of the magnetic spring model of FIG. 11 as an oscillating device or a driving device were examined, the results shown in FIG.
0 and the results shown in FIG. 41 were obtained. More specifically, in the magnetic spring model of FIG. 11, the thrust indicated by the dotted line in the figure is required to obtain the acceleration of FIG. 40 with the magnet mounting table 18, the L-shaped lever 22 and the balance weight 24 removed. there were. On the other hand, in the case of the magnetic spring model of FIG. 11 in which the magnet mounting table 18, the L-shaped lever 22 and the balance weight 24 are attached and the position of the balance weight 24 is adjusted, the acceleration can be obtained by the input shown by the solid line, and the frequency can be obtained. It was possible to generate a large (0.9 to 1 G) acceleration at the minimum input at about 5.5 Hz. Further, the amplitude could be considerably amplified as shown in FIG. 41.

That is, by changing the so-called geometrical dimension such as the facing area of the repulsive permanent magnet, the resonance frequency of the magnetic spring is used to make a small driving force (input).
A large acceleration and amplitude can be obtained at. In the magnetic spring model of FIG. 11, the gap amount and the facing area are variables. For example, by causing the facing area of the two permanent magnets 2, 4 to follow the balance weight 24 according to the change in the gap amount (input), it is possible to follow an arbitrary resonance point of the spring constant of the magnetic spring. .

Consider the application of magnetic force by electromagnetic induction in order to impart damping characteristics. First, the magnetic field in the metal conductor will be described with reference to FIG. In FIG. 42, (a)
Shows the coordinates of the cylindrical magnet and the metal conductor, (b) shows the cylindrical coordinates of the cylindrical magnet, and (c) shows the current density in the metal conductor.

As shown in FIG. 42 (a), an arbitrary point (x, y) on the lower surface of a cylindrical magnet having a radius a and a magnetization M is set to an arbitrary point (ξ, 0, z) in the conductor. make magnetic field dH ^L is,

(Equation 22) ds is a minute area including the point (x, y). The z component is

(Equation 23) Because

(Equation 24)

As shown in FIG. 42B, using cylindrical coordinates, x = r.cos (.psi.). Y = r.sin (.psi.) Ds = r
Since dψdr,

(Equation 25) The magnetic field H _z ^u created by the top surface is given by the thickness of the magnet h.

(Equation 26) become.

As described above, in the conductor (ξ, 0, z)
The vertical component H _z (ξ, 0, z) of the magnetic field at is Hz (ξ, 0, z) = H _z ^L (ξ, 0, z) + H _z ^u (ξ, 0, z)
z) can be obtained.

Next, the induced current in the conductor will be described.
When the magnet approaches, the downward (z direction) magnetic flux increases, so that an electromotive force e is generated in a direction that obstructs the magnetic flux.

[Equation 27] Here, φ (R, z) is the magnetic flux in the region surrounded by the radius R in the conductor. When the approaching speed is v, v = −dz / dt = −Δz / Δt ∴Δt = Δz / | v | From the equation (10), the voltage V along the circumference R is V = | e | = v · Δφ (R, z) / Δz = v · dφ (R, z) / dz (11)

The magnetic flux φ (R, z) is determined as follows. As shown in FIG. 42 (c), the magnetic field in the portion surrounded by the circumference of radius ξ and the circumference of radius ξ + dξ is H _z (ξ,
0, z) and its area is 2πξ · dξ, so Δφ (ξ, z) = μ ₀ H _z (ξ, 0, z) · 2πξ · d
ξ ∴φ (ξ, z) ＝ ∫ ₀ ^R μ ₀ H _z (ξ, 0, z) ・ 2πξ ・ d
ξ.

When the electric resistivity is ρ, the voltage is V, the current is I, the cross-sectional area of the circuit is S, and the length of the circuit is d = 2πR, the current density J is J (R, z) = 1 / s. = V / (ρd) = V / (2πR · ρ) (12) Substituting the formula (11) into the formula (12), J (R, z) = v / (2πR · ρ) · dφ (R , z) / dz (13)

Next, the interaction energy between the magnet and the conductor will be described. Due to the change of the magnetic flux, the increased current energy in the conductor, that is, the magnetic energy density u _m
Is

[Equation 28] It is.

The force acting on the current density J is f _z (R, z) = ∂υ _m (R, z) / ∂z. Therefore, F _z (R) acting on the total current I (R) of radius R
Is

(Equation 29) Where z ₁ is the distance from the magnet lower surface to the conductor upper surface, and z ₂ is the distance from the conductor lower surface.

From equations (13), (14) and (15)

[Equation 30] The force acting on the whole is

[Equation 31] Given in. Where φ (R, z) is the radius of the conductor R
In the region surrounded by, z ₁ and z ₂ are the coordinates of the upper and lower surfaces of the conductor, respectively, and F _z depends on the conductor thickness T = z ₂ −z ₁ .

FIG. 43 shows an example of application to a vibration isolation device for suspension seats. (A) shows the entire vibration isolation device.
(B) is a side view of the vibration isolator of (a), and (c) is a horizontal vibration isolation unit swingably attached to the upper part of the vibration isolator of (a). Is shown.
In the figure, 2, 4, 94 and 96 indicate permanent magnets, and 9
Reference numeral 8 indicates a copper plate as a conductor.

This anti-vibration device obtains the springiness in the vertical direction by the repulsion system by the permanent magnets 2 and 4, and uses it for the parallel link 10.
This is a vibration isolation device for suspension seats, which is supported by 0, 100 and can be attached and detached with a damping structure by electromagnetic induction in the front-back and up-down directions. Further, the damping force by electromagnetic induction can be changed by changing the plate thickness of the copper plate 98.

FIG. 44 is a graph comparing the vibration characteristics with and without the damping force effect in the front-rear direction, and the vibration transmissibility in the low frequency region is suppressed to some extent by electromagnetic induction.

In the magnetic spring structure according to the present invention, since the repulsive magnetic poles face each other, it means that the magnets are placed in a demagnetizing field with each other, and there is a fear of demagnetization during use. As a demagnetization measure, a pseudo degaussing structure, that is, the magnetic poles are arranged alternately so that the demagnetization field becomes small. At this time, a leakage magnetic field is generated between the alternating magnets corresponding to the domain walls, and a stronger repulsive force is obtained when the opposing magnets approach each other. Therefore, the repulsive force as a function of the distance between opposed magnets depends on the number of alternating magnets. FIG. 45 shows this state.

In FIG. 45, (a) single pole, (b) 2
The magnet arrangement of the poles, (c) 3 poles, (d) 4 poles of the Tagata is facing area (75 × 75 mm ² ), volume (75 × 75 × 25 m).
m ³ ) and Br value (11.7 KG) are the same, but the permeance coefficient is different as described below. In addition,
(E) is a view seen from the direction of the arrow in (d) Tagata 4-pole. Permeance coefficient (a) 0.10 (b) 0.37 (c) 0.54 (d) 0.49

FIG. 46 is a graph showing the relationship between the distance between magnets and the repulsive force in the respective magnet arrangements of (a) to (d). As can be seen from this graph, when the opposing magnets are close to each other, a leakage magnetic field is generated between the alternating magnets corresponding to the domain walls as described above, and thus the repulsive force is larger as the number of poles is larger.

Next, when the vibration characteristics of the Tagata 4-pole and 2-pole were compared by using the suspension seat vibration isolation device, the results shown in FIG. 47 were obtained. As can be seen from this graph, the 4-pole Tagata also has a damping effect due to the attractive force, and lowers the vibration transmissibility in the resonance frequency band of the internal organs and spinal column than the 2-pole. The vibration conditions are 0.3G with constant acceleration and LOG-SWE.
A sine wave of EP was used and the load mass was set to 53 kg.

Conventionally, k around the equilibrium point used in automobile seats is in the range of 10 to 30 N / mm, and tends to bottom when the load mass increases. Conversely, when the load mass decreases, the resonance frequency shifts in the resonance direction of the internal organs or the spinal column, and the vibration transmissibility increases. Therefore, a soft spring-rich structure is used with a metal spring by using urethane foam that provides a spring and a damping function for the pad layer. A shock absorber is used to enhance the damping function. By using these various functional parts, the automobile seat balances vibration insulation, damping, human body support pressure, and posture retention.

However, in the low frequency range of 2 to 3.5 Hz, the vibration transmissibility is 1.0 G / G or less, there is no secondary resonance in the high frequency range, and there is no fluffy feeling. It is considered difficult.

As a countermeasure against this, as shown in FIG. 48, a smooth natural frequency change and a load-deflection characteristic can be designed by using a magnetic spring characteristic exhibiting a behavior opposite to that of a metal spring and an air spring. . Further, in the suspension unit having the characteristics shown in FIG. 49, as shown in FIG. 50, the vibration transmissibility is 1.0 G / G at 2 to 3.5 Hz.
It was possible to achieve an ideal state of cutting 0.4 G / G between 3.5 and 50 Hz.

From the above, by incorporating the magnetic spring according to the present invention into the suspension seat, (1) the vibration transmissibility of 2.0 G / G in the low frequency range of 2 to 3.5 Hz.
The suspension seat under the following conditions does not require weight adjustment and damper hard / soft adjustment functions. (2) Vibration transmissibility 1.0G / G in the low frequency range 2 to 3.5Hz
Weight adjustment is required for the suspension seat under the following conditions. (3) By combining magnetic spring characteristics as a bottomed support function, the suspension unit operates in the high frequency range of 5 to 50H with a constant acceleration of 0.3G and a LOG-SWEEP sine wave.
High vibration isolation can be achieved with z.

On the other hand, by setting the input / output relationship of the negative attenuation characteristic to the boosting mechanism, an amplifier which produces a large output with a small input can be realized. When this is applied to active control, (1) Since the drive unit and the movable unit are not in contact with each other, power can be transmitted even when the partition is partitioned. (2) Since the drive section and the movable section are separate spaces, the degree of freedom in layout is widened. (3) Since it has a boosting function and the driving part and the movable part are almost incompatible, low noise and energy saving are achieved. (4) The structure has vibration insulation even if the actuator function is turned off. That is, the actuator has a flexible structure having a spring property and a damping property.

[0138]

Since the present invention is configured as described above, it has the following effects. The repulsive force on the input side from the equilibrium position of the permanent magnet is changed by changing the geometrical dimension between at least two facing permanent magnets on the input side and the output side by an external force and converting it into a repulsive force within its motion system. Since the repulsive force on the output side has been increased, passive control, semi-active control, and active control can all be handled with the same concept.

Further, since the maximum repulsive force is generated at the position where the permanent magnets face each other at the closest position or at a position beyond the closest position, it is possible to effectively use the magnetic field as a potential field, which is inexpensive. Magnetic brake, dynamic vibration absorber,
A generator and an amplifier can be realized.

Further, the non-linear vibration system or the coefficient excitation vibration system according to the present invention has a structure in which energy is converted into damping or continuous or divergent vibration by using a magnetic spring exhibiting positive, zero or negative damping characteristics. Therefore, by incorporating it into an anti-vibration device such as an automobile seat or an ambulance bed, it is possible to reduce the vibration transmissibility in the high frequency region, absorb the weight difference, and reduce the vibration energy in the low frequency region such as the lowering of the resonance point. effective.

[Brief description of drawings]

FIG. 1 is a schematic view showing an equilibrium position between an input side and an output side of two permanent magnets in a magnetic spring according to the present invention.

FIG. 2 is a graph of basic characteristics showing the relationship between the applied load and the displacement amount of the permanent magnet from the equilibrium position in the magnetic spring of FIG.

FIG. 3 is a graph showing a relationship between an actually measured load and a displacement amount.

FIG. 4 is a schematic diagram showing the concept of input and output in a charge model assuming that magnetic charges are uniformly distributed on the end face of a permanent magnet, where (a) shows attraction, (b) shows repulsion,
(C) shows the repulsion of a part different from (b).

FIG. 5 is a schematic diagram of a case where one of the permanent magnets whose magnetic poles are opposed to each other is moved relative to the other (the facing area is changed).

FIG. 6 is a graph showing a load in the X-axis and Z-axis directions with respect to an X-axis movement amount when calculated based on FIG. 5;

FIG. 7 is a diagram showing a state in which the distance between the permanent magnets shown in FIG.
6 is a graph showing the relationship between the displacement and the load when one is moved from a completely shifted state to the other from a completely wrapped state, and further moved from this state to a completely shifted state.

FIG. 8 is a schematic diagram of a case where one of the permanent magnets having the same magnetic poles is rotated with respect to the other (the facing area is changed).

FIG. 9 is a graph showing the maximum load with respect to the facing area when the permanent magnet is rotated based on FIG.

FIG. 10 is a graph showing a relationship between a distance between magnets and a load when a neodymium magnet is used as a permanent magnet.

FIG. 11 is a front view of the first magnetic spring model in which the geometrical dimension is changed by changing the facing area of the permanent magnet.

FIG. 12 is a front view of a second magnetic spring model in which the geometrical dimension is changed by changing the distance between the permanent magnets.

FIG. 13 is a front view of a third magnetic spring model in which geometrical dimensions are changed in combination with lever ratio conversion.

FIG. 14 is a front view of a fourth magnetic spring model in which a geometrical dimension is changed by polar conversion.

FIG. 15 is a front view of a fifth magnetic spring model in which geometrical dimensions are changed by magnetic circuit conversion.

FIG. 16 is a basic model for explaining the characteristics of the magnetic spring.

FIG. 17 is a graph showing the relationship between the repulsive force and the distance between two opposing permanent magnets.

FIG. 18 is a front view of an apparatus used to obtain the static / dynamic characteristics of a magnetic spring without area conversion.

FIG. 19 shows the dynamic characteristics of a magnetic spring obtained using the apparatus of FIG. 18, (a) 50 × 50 × 10 mm.
(B) is 50 × 50 × 15m when using the magnet of
(c) is 50 x 50 x 20 when using m magnet
(d) when a magnet of mm is used is 75 × 75 × 1
(E) when using a 5 mm magnet is 75 × 75 ×
(F) when a 20 mm magnet is used is 75 × 75
It is a graph at the time of using the magnet of x25mm.

20 is a graph showing dynamic characteristics of a magnetic spring obtained by using the device of FIG. 18, and is a graph when the same magnet is used and a load is changed.

FIG. 21 is a graph showing dynamic characteristics of a conventional passenger car seat as a comparative example.

FIG. 22 is a graph showing changes in spring constant and coefficient with time in the magnetic spring structure of the present invention.

FIG. 23 is a graph showing dynamic characteristics of the bed type vibration isolation unit when only the pad is used, when the pad and the magnetic spring are used, and when the semi-active control is further performed.

FIG. 24 is a front view of the maglev unit used to measure the dynamic characteristics of the magnetic spring.

FIG. 25 is a graph showing dynamic characteristics of the maglev unit measured using the maglev unit of FIG. 24;

FIG. 26 is a graph showing riding comfort evaluation constants measured using various seats including a maglev unit.

FIG. 27 is a graph showing riding comfort evaluation constants of a maglev unit measured by changing loads and cushion materials.

FIG. 28 is a graph showing dynamic characteristics measured using various sheets including a Maghreb unit.

FIG. 29 is a schematic view of a model in which a stopper and an elastic support member are incorporated in a magnetic spring.

FIG. 30 is a graph showing input / output work characteristics of the sliding principle model.

FIG. 31 is a graph showing experimental values of input and output of the rotary principle model.

FIG. 32 is a graph showing input / output work characteristics of the rotary principle model.

FIG. 33 is an explanatory diagram showing the main points of the input / output principle model.

FIG. 34 is a graph showing the relationship between the distance between magnets calculated by a certain charge model, the repulsive force, and the magnetic flux density.

FIG. 35 is a graph showing the relationship between the displacement and the repulsive force of the rotary principle model by area conversion.

FIG. 36 is a graph showing input / output work characteristics of a sliding metal spring model.

FIG. 37 is an explanatory diagram showing the main points of the metal spring model.

FIG. 38 is a perspective view of a rotary principle model.

FIG. 39 is a perspective view of a slide-type principle model.

FIG. 40 is a graph showing the relationship between the input or required thrust and the acceleration with and without the balance weight in the magnetic spring model of FIG. 11.

41 is a graph showing the relationship between the amplitude and the acceleration with respect to the frequency in the case where the balance weight is provided and the case where the balance weight is not provided in the magnetic spring model of FIG. 11.

FIG. 42 shows a magnetic field model in a metal conductor,
(A) is a figure which shows the coordinate of a cylindrical magnet and a metal conductor, (b) is a figure which shows the cylindrical coordinate of a cylindrical magnet, (c) is a figure which shows the current density in a metal conductor.

43A and 43B show a vibration isolation device for a suspension seat incorporating a magnetic spring according to the present invention, wherein FIG. 43A is a front view of the entire vibration isolation device, FIG. 43B is a side view thereof, and FIG.
FIG. 7A is a perspective view of a horizontal vibration isolation unit swingably attached to the upper portion of the vibration isolation device of FIG.

FIG. 44 is a graph showing vibration characteristics of a front-rear damping force effect by electromagnetic induction.

45A and 45B show magnet arrangements of magnetic springs having different numbers of poles, where FIG. 45A is a front view of a single-pole magnetic spring, FIG. 45B is a front view of a 2-pole magnetic spring, and FIG. The front view of the magnetic spring is shown in (d), and the front view of the Tagata 4-pole magnetic spring is shown in (e).
It is the top view seen from the arrow direction in.

FIG. 46 is a graph showing the relationship between the distance between magnets by pole number and the repulsive force.

FIG. 47 is a graph showing vibration characteristics of a vibration isolation device for a suspension seat that employs a 2-pole magnetic spring and a Tagata 4-pole magnetic spring.

FIG. 48 shows vibration characteristics of various springs, (a)
3 is a graph showing the vibration characteristics of a metal spring, (b) is a graph showing the vibration characteristics of an air spring, and (c) is a graph showing the vibration characteristics of a magnetic spring.

FIG. 49 is a graph showing static characteristics of a suspension unit that employs a magnetic spring.

FIG. 50 is a graph showing vibration characteristics of a conventional suspension unit and a suspension unit employing a magnetic spring.

FIG. 51 is a graph showing vibration characteristics of a conventional general suspension seat.

[Explanation of symbols]

 2,4 permanent magnet

 ─────────────────────────────────────────────────── ─── Continuation of front page (72) Inventor Hiroshi Nakahira 1-10-10 Yanoshinmachi, Aki-ku, Hiroshima-shi, Hiroshima Prefecture Delta Touring Co., Ltd. (72) Inventor Seiji Kawasaki, Yachishinmachi, Aki-ku, Hiroshima-shi, Hiroshima 2-chome, Delta Touring Co., Ltd. (72) Inventor Hiroki Honda, 1-chome, Yano-shinmachi, Aki-ku, Hiroshima-shi, Hiroshima Prefecture 2-chome, Delta Touring Co., Ltd. (72) Yoshimi Enoki, Yano, Aki-ku, Hiroshima-shi, Hiroshima 1-2-10 Shinmachi Inside Delta Touring Co., Ltd.

Claims (15)

[Claims] 1. At least two permanent magnets are separated from each other, and a geometrical dimension between the at least two permanent magnets is changed by input / output by a mechanism in a movement stroke or an external force, and converted into a repulsive force in the movement system. By so doing, the repulsive force on the output side is made larger than the repulsive force on the input side from the equilibrium position of the at least two permanent magnets, and the magnetic spring has a zero or negative damping characteristic. 2. The magnetic spring having a zero or negative damping characteristic according to claim 1, wherein the maximum repulsive force is generated at the closest position of the at least two permanent magnets or a position beyond the closest position. 3. The geometrical dimension is changed by changing any one of a separation distance, a facing area, a magnetic flux density, and a magnetic field of the at least two permanent magnets. Or a magnetic spring with negative damping properties. 4. Non-linear damping by deriving at least two permanent magnets from each other and changing the geometrical dimension between the at least two permanent magnets at the input and output by a mechanism in the motion stroke or external force to derive a damping term. A magnetic spring having a positive damping characteristic characterized by having a spring structure. 5. The magnetic spring having a positive damping characteristic according to claim 4, wherein the maximum repulsive force is generated at the closest position of the at least two permanent magnets. 6. The positive dimension according to claim 4, wherein the geometrical dimension is changed by changing any one of a separation distance, a facing area, a magnetic flux density, and a magnetic field of the at least two permanent magnets. Magnetic spring with damping characteristics. 7. A magnetic spring that exhibits zero or negative damping characteristics by changing the geometrical dimension between at least two permanent magnets that are separated from each other by a mechanism within a movement stroke or an external force, and is used for continuous or divergent vibrations. A coefficient-excited vibration mechanism having a structure for converting energy. 8. The coefficient-excited vibration mechanism according to claim 7, wherein the geometrical dimension is changed by an external force so that the spring constant and the damping coefficient in the motion system of the mechanism are changed by the spring structure system. 9. The coefficient-excited vibration mechanism according to claim 7, wherein the vibration characteristics are improved by changing the geometrical dimensions by a mechanism within a movement stroke or an external force. 10. Excitation or resonance frequency is made variable by changing the geometrical dimension by a mechanism in a movement stroke or an external force, and the resonance frequency is tracked to constantly decrease the resonance frequency of the frequency characteristic. Alternatively, the coefficient-excited vibration mechanism according to claim 7, wherein the coefficient-excited vibration mechanism operates at a place where fluctuations in amplitude are reduced. 11. The geometrical dimension is changed by a mechanism within a movement stroke or an external force, and when the displacement is small, a negative damping causes a positive damping with an increase in the displacement, and a vibration with an amplitude that balances the positive and negative damping. The coefficient-excited vibration mechanism according to claim 7, wherein 12. A magnetic spring having a positive damping characteristic by changing the geometrical dimension between at least two permanent magnets separated from each other by a mechanism in motion or external force, the spring of the mechanism in motion. A non-linear vibration mechanism characterized by having a larger damping characteristic than the characteristic. 13. The non-linear vibration mechanism according to claim 12, wherein the vibration characteristics are improved by changing the geometrical dimensions by a mechanism within a movement stroke or an external force. 14. A place where a resonance frequency is variable by changing the geometrical dimension by a mechanism in a movement stroke or an external force, and the resonance frequency is tracked so that the resonance frequency of the frequency characteristic always decreases, or The non-linear vibration mechanism according to claim 12, wherein the non-linear vibration mechanism is operated at a place where fluctuations in amplitude are reduced. 15. A resonance frequency is made variable by changing a geometrical dimension between at least two permanent magnets separated from each other by a mechanism in a motion stroke or an external force, and the resonance frequency is tracked so that a resonance of a frequency characteristic is constantly generated. Where the frequency is high,
Alternatively, a non-linear vibration mechanism characterized in that it produces a large acceleration or amplitude with a small input so that it operates in the presence of large fluctuations in amplitude.