JP4293876B2 - Electromagnetic actuator for vibration control - Google Patents

Description

  The present invention relates to an electromagnetic actuator, and more particularly to a vibration-damping electromagnetic actuator that effectively attenuates rolling of a vehicle.

In railway vehicles, dampers are inserted between the bogie and the car body to absorb the rolling of the car body and improve riding comfort. Recently, since a high level of vibration control performance is required, an active actuator that can be electrically controlled is often used to measure and feed back the shaking of the vehicle body.
For example, Patent Literature 1 discloses an electric actuator that includes an electric motor that rotates and a rotation-linear motion conversion mechanism and operates so that both ends expand and contract in order to enable stable traveling of the vehicle and improve passenger comfort. Is disclosed.

  An electromagnetic actuator can be used as the active actuator. The electromagnetic actuator is a hollow cylindrical core wound with an electromagnetic coil. A mover consisting of an iron core is placed at the center of the stator, and the distance between the ends is adjusted by adjusting the amount of movement of the mover according to the direction and strength of the excitation current. Since it expands and contracts directly on the mechanism, the structure is simple.

  However, in the conventional electromagnetic actuator for vibration suppression, electromagnetic steel sheet thin plates are laminated in the circumferential direction to suppress eddy current loss induced in the core. For this reason, the number of parts is large and the manufacturing process is complicated. Therefore, many attempts have been made to improve the manufacturing method. For example, Patent Document 2 discloses a cylindrical type in which a wedge-shaped space is formed by skillfully utilizing a spacer and thin magnetic plates are stacked in a cylindrical shape. A method for manufacturing a laminated core is disclosed. However, since the disclosed method has not succeeded in reducing the number of parts, it has a limited effect.

In addition, when the power supply current is cut off due to a power failure or the like, the mover moves freely and loses the damper force, so a separate backup system is required.
In general, since a large thrust is required for the electromagnetic actuator for vibration control, the apparatus itself is large.
Japanese Unexamined Patent Publication No. 07-081561 JP-A-10-322945

  The problem to be solved by the present invention is to simplify the backup system by configuring the device so as to generate a certain damper force even if there is a failure such as a power failure, and to reduce the size by reducing the braking force. An electromagnetic actuator for vibration is provided.

In order to solve the above-described problems, an electromagnetic actuator for vibration suppression according to the present invention includes a stator core formed of a block-shaped conductive magnetic material, and while shifting a tile-shaped permanent magnet piece in the axial direction on the shaft of the mover. It is characterized by being wound .
When the mover moves relative to the stator, an eddy current is induced in the stator core, so that a damper force is generated due to the eddy current loss, and in addition to the damper force generated by the excitation current supplied to the electromagnetic coil, the damper force is Increase.

Unlike the conventional thin plate stack type, the stator core is in the form of a block, so that the generation of eddy currents is not hindered, and when the mover moves, the magnetic field formed by the magnet changes and the eddy currents are effectively generated. appear.
In the electromagnetic actuator for vibration suppression of the present invention, the damper force is generated due to the eddy current even if the damper force control becomes impossible due to a power failure or the like, so that the minimum braking effect is exhibited without providing a backup system. Safety can be ensured.
Therefore, the configuration of the vibration suppression electromagnetic actuator is simplified, and the size and cost can be reduced. In addition, there is no need to laminate electromagnetic steel sheets, which is advantageous in terms of productivity and cost.

The electromagnetic actuator for vibration suppression of the present invention may be provided with a slit in the stator core to adjust the damper force due to the eddy current. Limiting the current circulating by the slit can greatly reduce the eddy current. Moreover, if the block size is adjusted using an appropriate number of slits, the generated eddy current can be adjusted to properly control the damper force due to eddy current loss.
In addition, a tile-like permanent magnet piece can be wound around the shaft of the mover while shifting it in the axial direction to eliminate the detent phenomenon and generate a smooth braking force. In particular, when the slot pitch of the stator core at the start point and end point of one block of the tile-like magnet pieces is an integral multiple of the stator core, it is possible to configure an electromagnetic actuator that suppresses the detent force well and has excellent controllability.

  In the electromagnetic actuator of the present invention, since the appropriate damper force due to the eddy current loss is added, the maximum value of the damper force component due to the conventional excitation current can be suppressed, so the electromagnetic actuator for vibration suppression can be made smaller. Can be configured.

The electromagnetic actuator for vibration suppression according to the present invention also calculates a damper force generated by an eddy current, evaluates the insufficient damper force, and determines an electromagnetic current to be applied to the electromagnetic actuator to generate the insufficient damper force It is preferable to attach.
Furthermore, it is preferable to provide a position detection sensor for detecting the position of the mover so as to transmit a measurement signal to the excitation current control device to perform feedback control.

  Furthermore, the length of the stator core is preferably larger than the length of the mover and smaller than the sum of the mover length and the stroke length. Although a large thrust is required near the center of the stroke, a large thrust is often not required at the end, so that the stator core at the end can be omitted. If the stator core is shortened, it is advantageous in that the electromagnetic actuator can be further reduced in size and weight.

Hereinafter, the electromagnetic actuator for vibration suppression of the present invention will be described in detail with reference to examples.
FIG. 1 is an axial sectional view showing an electromagnetic actuator for vibration damping of an embodiment according to the present invention, FIG. 2 is a perspective view showing the shape of a unit body of a stator core used in this embodiment, and FIG. 4 is a partial cross-sectional view showing a state in which a tile-like permanent magnet is attached to the mover rod, FIG. 4 is a conceptual configuration diagram for explaining a state in which the vibration-damping electromagnetic actuator of this embodiment is attached to a railway vehicle, and FIG. FIG. 6 is a block diagram of a control device, and FIG. 6 is a block diagram of a control device used in the embodiment.

The electromagnetic actuator for vibration damping of the present embodiment is such that the mover 2 is reciprocated along the central axis of the case 3 in which the stator 1 is formed on the inner wall. The front end surface 21 is fixed to each of the relative-moving structure, in this embodiment, the vehicle body and the bogie of the railway vehicle, and suppresses vibration when the relative distance varies between them, that is, when vibration occurs. Is.
The stator 1 has a structure in which a plurality of electromagnetic coils 11 are sandwiched between stator cores 12 each made of a ferromagnetic material.

The stator core 12 is configured by stacking a plurality of stator core units formed of blocks made of a conductive magnetic material in the axial direction as shown in FIG.
For example, as shown in FIG. 2, the stator core unit body has a donut shape having a thin tooth 13 toward the center, and a space is formed between adjacent teeth 13 when the teeth are stacked. The electromagnetic coil 11 is disposed in the space. The teeth 13 may be provided with one or several slits 15 reaching an appropriate depth. The damper force can also be adjusted by making the material of the stator core 12 have an appropriate volume resistivity. Of course, the adjustment of the slit 15 and the adjustment of the stator core 12 can be performed together.
The electromagnetic coil 11 is composed of a ring-shaped winding and is connected to a plurality of phases, for example, U, V, and W phases.

The mover 2 includes a shaft 22 and a permanent magnet 23.
The shaft 22 is a cylinder made of a magnetic material, and a guide rod 32 provided in the case 3 is inserted into the core portion. The shaft 22 can be translated while maintaining the posture by the bearing 25 provided between the outer periphery of the guide rod 32 and the bearing 25 provided between the opening 33 of the case 3.
The permanent magnet 23 is formed by installing ring magnets on the outer periphery of the shaft 22 at an appropriate interval. Alternatively, a tile magnet may be fixed in a ring shape or a spiral shape.

The permanent magnet 23 is polarized on the inner and outer surfaces in the radial direction, and the permanent magnets adjacent in the axial direction are arranged so that opposite polarities appear on the surface. Therefore, a row of permanent magnets in which N poles and S poles are alternately arranged in the axial direction is formed.
Depending on the positional relationship between the teeth 13 and the permanent magnets 23, a wave may be generated in the braking force, so-called detent force may be generated and controllability may deteriorate. The permanent magnet 23 shown in FIG. 3 is obtained by shifting the tile-shaped magnet 26 in the axial direction so that the outer periphery of the shaft 22 is spirally overlapped by exactly one turn. The displacement amount at this time is the pitch P of the teeth 13. Alternatively, if an integer multiple of nP is selected, the amount of action between the shaft 22 and the stator is always constant even if an axial shift occurs between the tooth 13 and the permanent magnet 23, so that the detent force is suppressed. Thus, an electromagnetic actuator excellent in controllability can be obtained.

The guide rod 32 has a cylindrical shape, and, for example, a differential coil type position sensor 4 is set in a core portion. The differential coil type position sensor 4 can measure the axial position of the mover 2 through the position of a sensing core (not shown) fixed to the tip of a rod provided along the center line of the shaft 22. it can.
Since the position sensor 4 is incorporated in the mover shaft 22, the electromagnetic actuator can be easily miniaturized, and the influence of fluctuations in the magnetic field received by the position sensor 4 can be reduced.

FIG. 4 is a conceptual diagram when the electromagnetic actuator of the present embodiment is applied to a railway vehicle. A vehicle body 7 is placed on a carriage 6 having wheels. An air spring 8 and an electromagnetic actuator 5 are interposed between the carriage 6 and the vehicle body 7.
Since the vehicle body 7 is supported by the air spring 8, when the roll speed vb of the carriage 6 changes, the roll speed vc of the vehicle body 7 changes in the same cycle but has a phase delay. If the electromagnetic actuator 5 is not provided, the roll amplitude is usually larger in the vehicle body 7 than in the cart 6. If the electromagnetic actuator 5 is provided, the amplitude of the shaking of the vehicle body 7 is small and is suppressed to be equal to or less than the shaking of the carriage 6. However, even if the electromagnetic actuator 5 is present, the amplitude of the vehicle body 7 is larger than that of the carriage 6 in a transient response when an external force is received such as when passing through a rotary machine.

The total damper force F of the electromagnetic actuator 5 is the sum of a damper force Fd as a passive damper and a damper force Fa as an electromagnetic damper. That is,
F = Fa + Fd
Here, the damper force Fd of the passive damper depends on the eddy current loss and depends on the relative speed of the vehicle body 7 with respect to the carriage 6. Further, the damper force Fa as an electromagnetic damper depends on the current, and an arbitrary force can be generated as long as it is within the capacity of the power source.
A frictional force is generated in the electromagnetic actuator 5 corresponding to the relative speed between the carriage 6 and the vehicle body 7, but can be ignored here.

Here, the effect brought about by the passive damper is considered.
The following relationship is established among the roll speed vb of the carriage 6, the roll speed vc of the vehicle body 7, the angular frequency ω, the roll amplitude mb of the carriage 6, and the roll amplitude mc of the vehicle body 7.
mc = a × mb
vb = mbsinωt
vc = a × mbsin (ωt−φ)
a is a rolling amplitude ratio between the carriage and the vehicle body. When there is no damper or when it is shaken by a large force, a ≧ 1, and when the suppression of the electromagnetic actuator is effective, a ≦ 1. Further, the phase delay φ takes a value in the vicinity of 90 ° when the vehicle body 7 swings near the suspension resonance frequency.

It is said that people feel the strongest when the railway vehicle is shaking around 1 Hz. In many cases, the resonance frequency of the suspension is also in this vicinity.
In the theory of the skyhook damper, a desirable damper is one that attenuates the shaking of the vehicle body regardless of the carriage, and therefore generates a reaction force that suppresses the absolute roll velocity vc of the vehicle body 7. is there. Therefore, the desired damping force Fi is set as follows: Fi = −Di × vc = −Di × a × mbsin (ωt−φ)
far. Here, Di is a desirable damping constant according to the theory of the skyhook damper.

On the other hand, if the damping constant as an actual effect of the passive damper is D,
Fd = D (vb−vc) = D × mb (sin ωt−a × sin (ωt−φ))
Since the total damper force F = Fa + Fd of the electromagnetic actuator 5 is set to the desired damping force Fi according to the Skyhook damper theory, the damper force Fa to be generated by the current is:
Fa = Fi−Fd = −Di × a × mbsin (ωt−φ)
−D × mb (sin ωt−a × sin (ωt−φ))
= Di * mb (-a (1-k) sin ([omega] t- [phi])-ksin [omega] t)
= Di × mb (-(a (1-k) cosφ + k) sinωt
+ A (1-k) sinφcosωt)
= −Di × mb ((a (1-k) cos φ + k) ² + a ² (1-k) ² sin ² φ) ^1/2
× sin (ωt−δ)
= −Di × mb (a ² (1-k) ² + 2a (1-k) k cos φ + k ² ) ^1/2
× sin (ωt−δ)
It becomes. However,
k = D / Di
Is the ratio of the actual damping constant compared to the Skyhook damper. Also,
sinδ = a (1-k) sinφ
/ (A ² (1-k) ² + 2a (1-k) k cos φ + k ² ) ^1/2
cosδ = (a (1-k) cosφ + k)
/ (A ² (1-k) ² + 2a (1-k) k cos φ + k ² ) ^1/2
It is.

Here, when the electromagnetic actuator does not have the function of a passive damper as in the conventional product, it is necessary to supply all of the damping force Fi with current. At this time, the damper force Fa generated by the current is
Fai = Fi = −Di × a × mbsin (ωt−φ)
It becomes.
Therefore, by comparing Fa and Fai, the effect of adding a passive damper function to the electromagnetic actuator can be evaluated.
The comparison of both functions can be performed with the magnitude of the amplitude. The amplitude ratio r of both is
r = | Fa | / | Fai |
= ((1-k) ² +2 (1-k) (k / a) cosφ + (k / a) ² ) ^1/2 (1)
And given.

The smaller r <1, the more effective the passive damper works and the more the electric driving force can be saved. Organizing the conditions
k ((1-2 / acosφ + 1 / a 2) k-2 (1-1 / acosφ)) <0
In order for this to have a solution of k> 0,
It is necessary that cosφ / a <1. The range of k at this time is
0 <k <2a (a- cosφ) / (a 2 -2acosφ + 1)
It becomes.

For example, when φ = 90 °, a = 1, that is, when the rolling motion of the carriage and the vehicle body has a phase difference of 90 ° and the amplitude is the same, the amplitude ratio k of the damper force compared with the presence or absence of the passive damper is 0 <k <. If it is in the range of 1, the effect of the passive damper is recognized. Further, at this time, the amplitude ratio r of the damper force due to the presence or absence of the passive damper at k = ^1/2 becomes the minimum value 2 ^1/2 / 2. That is, by using a passive damper in combination, the maximum electric driving force can be suppressed to about 70%.

A method of selecting the damping constant ratio k that minimizes r representing the effect of the passive damper function will be described by taking the case of the phase difference φ = 90 ° as an example. In equation (1), if φ = 90 °,
r ² = (1−k) ² + (k / a) ²
So if you solve this for k,
k = a ² / (1 + a ² ) (1 ± ((1 / a ² +1) r ² −1 / a ² )) ^1/2
For this to have a meaningful solution,
r ² ≧ 1 / (1 + a ² )
Must. That is, when a certain amplitude ratio a between the carriage and the vehicle body is given, the active driving force is compared with the case where there is no function of the passive damper,
rmin = 1 / (1 + a ² ) ^1/2
Can be reduced. At this time, the effectiveness of the passive damper is
k = a ² / (1 + a ² )
Choose to be

Note that the above discussion is applied when a specific amplitude ratio is given, and r = rmin cannot be maintained when the amplitude ratio changes due to the attenuation of fluctuation.
For example, it is assumed that the amplitude changes to a1 after optimally setting for a certain amplitude ratio a0.
That is, when setting,
k0 = a0 ² / (1 + a0 ² )
rmin0 = 1 / (1 + a0 ² ) ^1/2
But when the amplitude changes to a1,
r = ((1+ (a0 / a1) ² a0 ² ) / (1 + a0 ² )) ^1/2 rmin0
Therefore, it can be seen that when (a0 / a1) is increased, r ≧ 1 in some cases, and the function of the passive damper is changed to hinder the active damping effect by the current.

However, when comparing the value of r multiplied by the amplitude,
a1r = ((a1 ² + a0 ⁴ ) / (1 + a0 ² )) ^1/2 rmin0
= ((((A1 ² / a0 ² ) + a0 ² ) / (1 + a0 ² )) ^1/2 a0rmin0
It becomes. here,
((((a1 ² / a0 ² ) + a0 ² ) / (1 + a0 ² )) ^1/2 <1
Therefore, the absolute value of the required damping force decreases as the oscillation amplitude decreases. Therefore, it can be understood that an optimum setting should be made when the amplitude is large to some extent.

The above examination was performed for φ = 90 °, but in a general case, for the amplitude ratio r = | Fa | / | Fai |
r = ((1-k) ² +2 (1-k) (k / a) cosφ + (k / a) ² ) ^1/2
If this is organized as an equation for k,
((1 / a) ² −2 / acos φ + 1) k ² − (1-1 / acos φ) k + 1−r ² = 0
It becomes. The coefficient of ^{k 2} will always be positive.

The necessary and sufficient condition for this equation to have a solution of k> 0 under the condition of r ≦ 1 is
(1-1 / acosφ)> 0
And the discriminant ≧ 0 is established.
From discriminant ≧ 0,
(1-2 acos φ + a ² ) r ² -sin ² φ ≧ 0
That is,
r ² ≧ sin ² φ / (1-2 acos φ + a ² )
It is.

Thus, the minimum value rmin ^{2 that r 2} can take is,
rmin ² = sin ² φ / (1-2 acos φ + a ² )
It becomes.
here,
1-rmin ² = 1−sin ² φ / (1-2 acos φ + a ² )
= (Cosφ-a) ² / (1-2acosφ + a ² ) ≧ 0
Therefore, rmin ² is always smaller than 1, and it can be seen that the passive damper has the effect of reducing the active thrust.

This is true because of the condition with a solution of k> 0 above.
a> cosφ
The damping constant k based on the skyhook damper is
k = a (a−cosφ) / (a ² −2acosφ + 1)
You can choose.
In addition, since the effect of suppressing the absolute value of the active thrust is obtained by applying the passive damper when the amplitude is large, the condition of a> cosφ is normally satisfied.

  In the above method, the passive damper has an effect of supplementing the active force due to the current when the amplitude is large. However, when the amplitude is small, the active damper rather hinders the active damping force. However, since the force used for vibration suppression is small when the amplitude is small, the required capacity of the active thrust is reduced as compared with the case where the passive damper function is not attached even if it is necessary to cancel the reverse action of the passive damper.

The electromagnetic actuator for vibration damping of the present embodiment must be operated based on the result of estimating the effect of the passive damper function using eddy current loss and canceling out the effect.
FIG. 5 is a block diagram showing a configuration of an estimator for estimating the action of the passive damper.
Damper force Fd generated by passive damper function,
Fd = D (vb−vc)
Need to be estimated.
Here, the damping constant D of the passive damper varies depending on the strength of the permanent magnet, the arrangement, the material of the stator core, the shape, the size of the block, the state of the slit, etc., but can be obtained in advance by design analysis or experiment. it can.

Since the position sensor 4 detects the expansion and contraction of the electromagnetic actuator 5, the relative displacement (xb-xc) between the carriage 6 and the vehicle body 7 is measured. By differentiating this, the speed difference between the carriage and the vehicle (vb -Vc) is obtained.
d / dt (xb-xc) = vb-vc
When the calculation function is used to actually calculate from the detection signal, there is a method of calculating a pseudo differential using a relative displacement (xb−xc) difference or the like.

Further, an absolute acceleration ac of the vehicle body 7 with respect to space can be obtained by an acceleration sensor attached to the vehicle body (not shown).
d / dt (vc) = ac
Note that these measured values are input as information necessary for control even in a conventional electromagnetic actuator having no passive damper function.

The estimator 41 shown in FIG. 5 obtains the relative speed (vc−vb) between the vehicle body 7 and the carriage 6 using the acceleration signal ac.
A measurement signal of acceleration ac is input from the acceleration sensor, integrated by an input integrator 42 provided outside the estimator 41, and input to the estimator 41 as the speed vc of the vehicle body 7.
The information signal of the relative position (xc−xb) supplied from the position sensor 5 is integrated by the first integrator 43 using the second coefficient K2 as an integration coefficient. The first integrator 43 includes a minor feedback circuit including a coefficient unit 44, and multiplies the feedback signal by a coefficient K1 and returns it to the input of the first integrator 43.

The feedback circuit formed by the first integrator 42 and the coefficient unit 44 is in the Laplace operator region.
Vbe = 1 / (K2s ² + K1s + 1) Vb
Given in. However, Vbe and Vb represent Laplace transforms of vbe and vb, respectively.
The speed difference (vc−vbe) obtained by subtracting the estimated speed vbe of the carriage 6 thus obtained from the speed measurement value vc of the vehicle body 7 obtained by the input integrator 42 becomes the output of the estimator 41.

  Further, the output signal (vc−vbe) of the estimator 41 is integrated through a major feedback circuit including the second integrator 45 to become a displacement signal (xc−xbe), and the relative position signal input of the first integrator 43 is obtained. Negative feedback to (xc-xb). Then, the position signal vc of the vehicle body 7 is canceled, and the input signal of the first integrator 42 is generated from the position signal component xb of the carriage 6 in the measurement signal of the position sensor and the estimated output signal (vc−vbe) of the estimator 41. The difference (xb−xbe) of the position signal component xbe of the cart 6 is obtained.

Since the output adjustment is performed so that the input signal becomes zero by the feedback operation, the difference (xb−xbe) of the position signal finally becomes zero, and the output signal vbe of the first integrator 43 at this time is the carriage. An accurate estimate of speed.
The coefficients K1 and K2 are parameters that determine the responsiveness of the estimator 41.
For example, in order to set the cutoff frequency to about 50 Hz so that the estimated value can follow the fluctuation of about 10 Hz, K1 = 4.5 × 10−3 and K2 = 1.0 × 10−5 may be selected. .

FIG. 6 is a block diagram of an electromagnetic actuator control apparatus in which such an estimator 41 is incorporated.
The input integrator 41 is supplied with the measurement signal ac of the acceleration sensor, and when integrated, the speed vc of the vehicle body 7 is obtained. By passing this through an ideal damper coefficient multiplier 46 that multiplies a damping constant Di based on the skyhook damper theory, a desirable damping force Fi based on the skyhook damper theory is obtained.

Further, the measurement output xc-xb of the position sensor 4 and the vehicle body speed vc are supplied to the estimator 41, output as a speed difference estimated value (vc-vbe), and a passive damper coefficient multiplier 48 that multiplies the damping constant D of the passive damper. Then, a damper force Fd as an effect of the passive damper function due to the eddy current is calculated.
The difference between the skyhook damper force Fi thus obtained and the damper force Fd of the passive damper determines the braking force Fa due to the current of the electromagnetic actuator.
Since the braking force Fa may include an error, the braking force Fa is adjusted by using a displacement adjustment coefficient unit 47 that outputs the displacement signal output (xc−xb) of the vehicle body 7 and the carriage 6 by multiplying by an appropriate coefficient Kx. be able to.

The magnet provided on the shaft of the vibration control mover of this embodiment may be an electromagnet. Since the shaft of the mover only reciprocates, an electric current can be easily supplied to the electromagnet attached to the shaft. When the electromagnet is used, the passive damper force can be controlled by adjusting the magnetic field intensity generated by the electromagnet by changing the supply current value.
Moreover, although the said Example demonstrated the railway vehicle as object, it cannot be overemphasized that it can utilize in the same way not only as a motor vehicle but as a damping device of a building.

In addition to adjusting the damper force by the coil current, the vibration-damping electromagnetic actuator of the present embodiment is passively generated by the eddy current induced in the stator core by the action of the permanent magnet provided on the shaft of the mover. Since the damper force is used, the electromagnetic actuator can be reduced in size. In addition, since the passive damper operates even in the event of a failure, the backup system may be small and may not be provided at all.
In addition, when making a stator core using block material to easily induce eddy currents, the number of parts is drastically reduced and the manufacturing process is simplified, so economic benefits such as improved productivity and reduced costs can be expected. .

It is an axial direction sectional view showing the electromagnetic actuator for vibration suppression of the example concerning the present invention. It is a perspective view which shows the shape of the unit body of the stator core used for a present Example. It is a partial cross section figure showing the state which attached the tile-shaped permanent magnet to the needle | mover rod of a present Example. It is a conceptual block diagram explaining the state which attached the electromagnetic actuator for damping | damping of a present Example to the railway vehicle. It is a block diagram of the estimator used for a present Example. It is a block diagram of the control apparatus used for a present Example.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Stator 2 Movable element 3 Case 4 Position sensor 5 Electromagnetic actuator 6 Carriage 7 Car body 8 Air spring 11 Electromagnetic coil 12 Stator core 13 Teeth 14 １５ 15 Slit 21 Front end surface of the mover 22 Shaft 23 Permanent magnet 24 Bearing 25 Bearing 26 Tile magnet 31 Case bottom 32 Guide rod 33 Opening 41 Estimator 42 Input integrator 43 First integrator 44 Coefficient multiplier 45 Second integrator 46 Ideal damper coefficient multiplier 47 Displacement adjustment coefficient multiplier 48 Passive damper coefficient multiplier

Claims (4)

An electromagnetic actuator for vibration damping composed of a stator core made of a ferromagnetic material, a stator made of an electromagnetic coil, and a mover made of a shaft made of a ferromagnetic material, wherein the stator core is electrically conductive A vibration-damping electromagnetic actuator characterized in that it is formed of a conductive block, and a tile-like permanent magnet piece is wound around the shaft while shifting in the axial direction .   2. The electromagnetic actuator for vibration damping according to claim 1, wherein one or more slits are formed in the stator core.   It is attached with an arithmetic unit that estimates the damper force due to the eddy current generated in the stator core, evaluates the insufficient damper force, and determines the electromagnetic current to be applied to the electromagnetic actuator to generate the insufficient damper force. The electromagnetic actuator for vibration suppression according to claim 1, characterized in that: Furthermore, a position detection sensor for detecting the position of the mover and an excitation current control device are provided, the position detection sensor transmits a position measurement signal to the excitation current control device, and the excitation current control device 4. The electromagnetic actuator for vibration suppression according to claim 3 , wherein feedback control of the electromagnetic actuator for vibration suppression is performed.

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