JP6146463B2 - Head assembly and magnetic disk drive - Google Patents

Description

  The present invention relates to a head assembly and a magnetic disk device having a precision positioning mechanism for a head in a hard disk drive.

  In recent years, the recording density of a magnetic disk provided in a magnetic disk device has been increasing with each passing day. Patent Document 1 proposes a head support mechanism in which a head support spring mechanism is finely driven by a piezoelectric element to position a magnetic head on a recording track with high accuracy.

  Further, Patent Document 2 proposes a configuration in which the slider rotates around a fulcrum protrusion provided on the load beam, and the inertial axis of the rotating portion including the slider substantially coincides with the fulcrum protrusion.

  However, in the configuration of Patent Document 1, the reaction force when the slider is slightly displaced by the pair of displacement members resonates the head support spring mechanism. For this reason, when the head element is positioned at a high speed, there is a problem that the control band cannot be expanded by this resonance.

  Further, the configuration of Patent Document 2 has a rotation mode in the yaw direction of the slider when the slider is driven to rotate around the fulcrum protrusion of the load beam. In order to further expand the head positioning control band in the future, it becomes necessary to set the resonance of the yaw mode of the slider higher.

JP-A-2-227886 Japanese Patent No. 5360129

  In the conventional configuration, when the head element is positioned on the track on the disk, the operating frequency of the actuator excites the resonance frequency of the head support mechanism, causing unnecessary vibration in the head support mechanism. Therefore, the conventional configuration has a problem that only a low control band that is not affected by resonance can be obtained.

  The present invention has been made in view of the above circumstances, and an object of the present invention is to suppress the resonance generated in the head assembly and enhance the control characteristics of the head positioning.

  The head assembly according to the present invention is provided at a slider having a head element, a slider support plate for holding the slider, a load beam for holding the slider support plate, and a tip end portion of the load beam, and the slider support plate can be rotated. A fulcrum protrusion supported on the fulcrum, a driving means for rotating the slider support plate around the fulcrum protrusion, and a dynamic vibration absorber provided on the slider support plate. It is located on the rear end side, and has a degree of freedom of vibration with respect to the rotation direction of the slider support plate.

  According to the present invention, the dynamic vibration absorber provided on the slider support plate is positioned on the rear end side of the load beam with respect to the fulcrum projection and has a degree of freedom of vibration in the rotation direction of the slider support plate. It is possible to suppress the resonance peak value of the rolling mode (Sway mode) of the beam and to suppress the rotational vibration of the slider in the yaw direction at the antiresonance point of the dynamic vibration absorber. As a result, the head positioning control characteristics can be improved, and the weight itself can be reduced to reduce the weight.

  Preferably, the dynamic vibration absorber may include a weight portion, a spring portion that connects the weight portion to the slider support plate, and a damping portion that suppresses an amplitude at which the weight portion vibrates with respect to the slider support plate. Thereby, damping property can be added to the operation of the dynamic vibration absorber, and a stable characteristic can be obtained.

  Preferably, the weight portion has a first resonance point that resonates due to vibration in the same direction as the direction in which the head element of the slider rotates with respect to the fulcrum protrusion, and the first resonance point is a roll of the load beam. The frequency is preferably higher than the resonance frequency of the mode. Thereby, the rolling mode of the load beam can be suppressed.

  Preferably, the weight portion, the spring portion, and the damping portion are provided by etching the flexure laminate. Thereby, a dynamic vibration absorber can be configured on the slider support plate easily and at low cost.

  Preferably, the weight part has a mass adjusting part. As a result, the anti-resonance frequency of the dynamic vibration absorber can be accurately adjusted to a frequency that needs to be suppressed.

  A magnetic disk device according to the present invention is equipped with the above-described head assembly. According to the present invention, it is possible to obtain a magnetic disk device in which resonance occurring in the head assembly is suppressed and the head positioning control characteristics are improved.

  The present invention can suppress the resonance generated in the head assembly and improve the control characteristics of the head positioning. In addition, it is possible to improve the positioning accuracy of the head element with respect to the recording track at a low cost without requiring a new additional process.

1 is a schematic plan view of a magnetic disk device on which a head assembly according to a preferred embodiment of the present invention is mounted. 1 is a perspective view of a head assembly according to a preferred embodiment of the present invention. 1 is an exploded perspective view of a head assembly according to a preferred embodiment of the present invention. It is a disassembled perspective view of the flexure with which the head assembly which concerns on suitable embodiment of this invention is provided. It is a top view of the 1st drive means with which the head assembly which concerns on suitable embodiment of this invention is provided. It is AA sectional drawing in FIG. 5a. It is BB sectional drawing in FIG. 5a. It is the top view which looked at the front-end | tip principal part of the head assembly which concerns on suitable embodiment of this invention from the upper surface side. It is the top view which looked at the front-end | tip principal part of the head assembly which concerns on suitable embodiment of this invention from the lower surface side. It is CC sectional drawing in FIG. It is DD sectional drawing in FIG. It is FF sectional drawing in FIG. It is GG sectional drawing in FIG. It is HH sectional drawing in FIG. It is EE sectional drawing in FIG. It is a figure which shows the operation state of a dynamic vibration absorber. It is JJ sectional drawing in FIG. It is II sectional drawing in FIG. It is a model figure of one Example of this invention which simplified the structure of FIG. In one Example of this invention, it is the model figure which simplified and showed a mode that the slider rotationally moved centering on the fulcrum protrusion by the 1st and 2nd drive means. In one Example of this invention, it is the model figure which simplified and showed a mode that the slider rotationally moved centering on the fulcrum protrusion by the 1st and 2nd drive means. In one Example of this invention, it is a model figure which further simplified the slider, the slider support plate, and the dynamic vibration absorber. In one Example of this invention, it is the figure which showed the operation state of the model which further simplified the slider, the slider support plate, and the dynamic vibration absorber. It is a control block diagram for performing head positioning control. Board characteristic diagram when there is no gain margin in head positioning characteristics. In the head positioning characteristic, the board characteristic diagram when the gain margin is 10 dB. In the head positioning characteristic, the board characteristic diagram when the gain margin is 20 dB. It is a simple model figure for demonstrating the behavior of the dynamic vibration damper in one Example of this invention. It is a figure which shows the frequency response characteristic for demonstrating the effect | action of a dynamic vibration absorber. It is a figure which shows the frequency response characteristic for demonstrating the effect | action of a dynamic vibration absorber. It is a figure which shows the frequency response characteristic for demonstrating the effect | action of a dynamic vibration absorber. It is a figure which shows the positioning frequency response characteristic of the head element of the head assembly in one Example of this invention. It is a figure of the frequency response characteristic which shows the damping characteristic effect in the head assembly in one Example of this invention. It is the top view which looked at the front-end | tip principal part of the head assembly of the 1st prior art example from the upper surface side. It is a structural model figure of the head assembly of a 1st prior art example. It is an operation | movement figure in the structural model of the head assembly of a 1st prior art example. It is a simple diagram explaining resonance of a load beam in the first conventional example. It is a figure which shows the positioning frequency response characteristic of the head element in the head assembly of a 1st prior art example. It is the top view which looked at the front-end | tip principal part of the head assembly of the 2nd prior art example from the upper surface side. It is a structural model figure of the head assembly of the 2nd prior art example. It is an operation | movement figure in the structural model of the head assembly of a 2nd prior art example. In the 2nd prior art example, it is a model figure which further simplified the rotation part and counter balance of a slider. FIG. 10 is a simplified diagram illustrating a configuration in which load beam resonance does not occur in the second conventional example. It is a figure which shows the positioning frequency response characteristic of the head element in the head assembly of a 2nd prior art example.

  Hereinafter, preferred embodiments of the present invention will be described with reference to the drawings. In addition, this invention is not limited to the following embodiment. The constituent elements described below include those that can be easily assumed by those skilled in the art and those that are substantially the same. Furthermore, the constituent elements described below can be appropriately combined. In addition, various omissions, replacements, or changes of components can be made without departing from the scope of the present invention.

  FIG. 1 schematically shows the overall configuration of a load / unload type magnetic disk device (HDD device) on which a head assembly according to a preferred embodiment of the present invention is mounted. As shown in FIG. 1, a magnetic disk device 1 includes a housing 4, a magnetic disk 6 that is driven to rotate about a shaft 5 by a spindle motor, a head assembly 2 in which a slider 3 having a head element 7 is mounted at the tip, It comprises a support arm 8 that supports the head assembly 2 at its tip.

  A coil portion of a voice coil motor (VCM) is attached to the rear end portion of the support arm 8. The support arm 8 is rotatable in parallel with the surface of the magnetic disk 6 about the horizontal rotation shaft 9. The VCM is composed of a coil part (not shown) and a magnet part 10 covering the coil part. A ramp mechanism 11 is provided from the outer side of the data area of the magnetic disk 6 to the outer side of the magnetic disk 6, and the slider 3 is moved by a tab 12 provided at the forefront of the head assembly 2 on the inclined surface. Then, it is separated from the magnetic disk 6 to be in an unloaded state.

  During operation of the magnetic disk device 1 (during high-speed rotation of the magnetic disk), the slider 3 floats with a slight flying height with respect to the surface of the magnetic disk 6 and is in a loaded state. On the other hand, at the time of non-operation (when the magnetic disk is stopped or during low-speed rotation when starting and stopping), the tab 12 at the tip of the head assembly 2 is lifted onto the ramp mechanism 11, so that the slider 3 is in an unloaded state. is there.

  FIG. 2 is a perspective view schematically showing the overall configuration of the head assembly in a preferred embodiment of the present invention. Hereinafter, for convenience of explanation, the Z-axis positive direction in the figure is referred to as the upper surface side of the head assembly 2, and the Z-axis negative direction is referred to as the back surface side or the lower surface side of the head assembly 2. The slider 3 includes a head element 7 composed of an inductive write head element and an MR read thin film magnetic head such as a giant magnetoresistive (GMR) read head element or a tunnel magnetoresistive (TMR) read head element. The trailing edge is provided on the Y-axis positive side in FIG.

  In FIG. 2, a head assembly 2 includes a base plate 13, a load beam 14, a flexure 15, a first thin film piezoelectric element as a first driving means 16a, and a second driving means 16b as main components. A second thin film piezoelectric element and a slider 3 are provided. A dynamic vibration absorber 33 is formed in the flexure 15. Further, the base plate 13 is configured to be attached to the distal end portion of the support arm 8.

  The load beam 14 is fixed to the base plate 13 by a plurality of beam welding points 17a. The leaf spring 18 is formed on the load beam 14 and applies a predetermined pressing force to the slider 3 against the magnetic disk. Furthermore, the load beam 14 has a structure in which bending portions 19 are provided on both sides to increase the strength. The flexure 15 that is a wiring board is fixed to the load beam 14 by a beam welding point 17b. In FIG. 2, the attitude angle of the slider 3 is indicated by Dp for the pitch direction, Dr for the roll direction, and Dy for the yaw direction. The base plate 13 and the load beam 14 are line symmetric with respect to the central axis parallel to the Y-axis direction in each figure.

  FIG. 3 is an exploded perspective view schematically showing a head assembly according to a preferred embodiment of the present invention. That is, FIG. 3 shows a state in which the head assembly 2 is disassembled into a load beam 14, a flexure 15, a base plate 13, first and second driving means 16 a and 16 b, and a slider 3. The flexure 15 is a wiring board in which an insulating layer is generally coated on a flexure substrate 24 which is a thin stainless steel plate having a thickness of about 18 μm, and a copper foil is plated thereon. The stainless steel plate, the insulating layer, and the copper foil are arbitrarily selected. It is etched into the shape of and is precisely processed into a predetermined shape.

  As shown in FIG. 3, the slider 3 is bonded and fixed on a slider support plate 20 formed on the flexure 15. The fulcrum projection 21 is integrally formed on the center line near the tip of the load beam 14. The slider support plate 20 is supported by the first outrigger 22a and the second outrigger 22b, and is supported by the fulcrum projection 21 in a point-contact manner from the back to the center position of the slider so as to be rotatable. Therefore, the slider support plate 20 is supported by the load beam 14 with a pivot structure. Further, the first and second outriggers 22a and 22b hold the posture of the slider 3 flexibly. As a result, the slider 3 smoothly follows the change in posture due to the waviness of the disk surface. A pressing force generated by the leaf spring 18 of the load beam 14 acts between the fulcrum protrusion 21 and the slider support plate 20. Therefore, the slider support plate 20 is held by the frictional force generated by this pressing force in the X direction and the Y direction.

  The first driving means 16a and the second driving means 16b are bonded on the first piezoelectric support portion 23a and the second piezoelectric support portion 23b of the flexure 15. The first and second driving means 16a and 16b are alternately expanded and contracted to give the slider support plate 20 a rotational force in the yaw direction along the plane, and the slider support plate 20 is rotated about the fulcrum protrusion 21. Let Note that the first piezoelectric support portion 23 a and the second piezoelectric support portion 23 b are formed of an insulating layer 41 constituting the flexure 15. Further, the dynamic vibration absorber 33 is provided on the slider support plate 20. The dynamic vibration absorber 33 is located on the rear end side of the load beam with respect to the fulcrum protrusion 21.

  FIG. 4 is an exploded perspective view showing the configuration of the flexure provided in the head assembly in the preferred embodiment of the present invention. In FIG. 4, the flexure is originally an integrated flexure, but the flexure substrate 24 and the head element wiring 25 (wiring portion) are shown separately for easy understanding. The dynamic vibration absorber 33 includes a first weight portion 33a, a second weight portion 33e, a spring portion 33b, and a damping portion 33c, and is provided by etching a flexure laminate. Specifically, the first weight portion 33 a and the spring portion 33 b of the dynamic vibration absorber 33 are made of the same copper foil as that of the head element wiring 25. The damping part 33 c is formed of a polyimide insulating layer 41. The second weight 33e of the dynamic vibration absorber 33 is formed by etching from the flexure substrate 24. Moreover, the wiring material of copper foil is exposed on the upper surface of the first weight part 33a, and it is possible to add a mass (mass adjusting part) such as a solder ball to the surface.

  FIG. 5a is a plan view of the first driving means provided in the head assembly according to the preferred embodiment of the present invention. Moreover, FIG. 5b represents the AA cross section in FIG. 5a, and FIG. 5c represents the BB cross section in FIG. 5a. Since the first driving unit 16a and the second driving unit 16b have the same structure, only the structure of the first driving unit 16a is illustrated here. An upper electrode 27a is formed on the upper surface side of the thin film piezoelectric body 26, and a lower electrode 27b is formed on the lower surface side. Since the first driving means 16a is very thin and easily damaged, a base 28 is provided as a flexible reinforcing material.

  The first driving means 16a is entirely covered with an insulating cover 30 made of polyimide to protect the thin film piezoelectric body 26. Note that a part of the insulating cover 30 is removed from the C part and the D part in FIG. In part C, the lower electrode 27b is exposed and is electrically connected to the first electrode pad 29a. In the portion D, the upper electrode 27a is exposed and is electrically connected to the second electrode pad 29b. Thereby, the thin film piezoelectric body 26 of the first driving means 16a can be expanded and contracted by applying a voltage to the first electrode pad 29a and the second electrode pad 29b. The direction of polarization of the thin film piezoelectric body 26 is indicated by an arrow in FIG. When an electric field is applied in the polarization direction (a negative voltage is applied to the first electrode pad 29a and a positive voltage is applied to the second electrode pad 29b), the thin film piezoelectric body 26 is moved in the in-plane direction of the piezoelectric film by the piezoelectric constant d31. Will contract. In contrast to the direction of polarization, it extends when an electric field is applied. If a negative voltage is applied to the third electrode pad 29c corresponding to the first electrode pad 29a and a positive voltage is applied to the fourth electrode pad 29d corresponding to the second electrode pad 29b, the second driving means The thin film piezoelectric body 16b is contracted in the in-plane direction of the piezoelectric film by the piezoelectric constant d31.

  FIG. 6 is a plan view of the main part of the tip of the head assembly according to the preferred embodiment of the present invention as viewed from the upper surface side (slider side). FIG. 7 is a plan view of the main portion of the tip of the head assembly according to the preferred embodiment of the present invention as viewed from the lower surface side (plan view of the head assembly of FIG. 6 as viewed from the rear surface side). For convenience of explanation, the load beam 14 is not shown.

In FIG. 6, the head element wiring 25 (wiring portion) is arranged so as to surround the slider 3, and the head electrode terminal 31 of the slider 3 is connected to the corresponding head element wiring 25 (wiring portion) by a solder ball. A first bent portion 32a and a second bent portion 32b are formed on the first and second outriggers 22a and 22b arranged on both sides of the slider support plate 20, respectively. Further, the intersections of the extension lines L1 and L2 of the first and second bent portions 32a and 32b are configured to coincide with the fulcrum protrusion 21. As a result, the first bent portion 32 a and the second bent portion 32 b are easily bent, so that the slider support plate 20 is slightly rotated around the fulcrum protrusion 21.

  The head element wiring 25 (wiring portion) is partially fixed to the first and second outriggers 22a and 22b (CC portion in FIG. 6), and the first extending from the slider support plate 20 is provided. The drive rib 36a and the second drive rib 36b are also fixed in the same manner (F-F portion in FIG. 6).

  The first and second driving means 16a and 16b are driven by applying a voltage to the first, second, third and fourth electrode pads 29a, 29b, 29c and 29d. The drive wiring 37a is arranged to input a voltage to the first electrode pad 29a and the fourth electrode pad 29d, and the ground wiring 37b connects the second electrode pad 29b and the third electrode pad 29c. Yes. Thus, when an alternating drive signal is input to the drive wiring 37a, the first drive means 16a and the second drive means 16b expand and contract in opposite directions.

  The slider support plate 20 is formed with a T-type limiter 34 (FIGS. 6 and 7) for lifting the slider 3 from the magnetic disk surface when the slider 3 is unloaded from the magnetic disk 6. The T-type limiter portion 34 is bent at the bent portion 34a on the side opposite to the slider 3, and is engaged with a hole portion 35 (shown in FIG. 3) formed in the load beam 14. It should be noted that the T-type limiter 34 and the hole 35 are not in contact during normal operation other than load / unload.

  FIG. 7 is a view of FIG. 6 viewed from the back side. In FIG. 7, the first link 39a having high rigidity is formed between the first joint 40a and the second joint 40b that are easily deformed. The first joint 40a is connected to the first drive rib 36a, and the second joint 40b is connected to the first fixing portion 24a that is a part of the flexure 15. Similarly, the rigid second link 39b is formed between the easily deformable third joint 40c and the fourth joint 40d, and the third joint 40c is connected to the second drive rib 36b, The fourth joint 40 d is supported by a second fixing portion 24 b that is a part of the flexure 15.

  A first separation groove 44a for separating the first driving means 16a from the second joint 40b and the flexure substrate 24 is provided. The first separation groove 44 a is formed along a range corresponding to the length of the thin film piezoelectric body 26 in the longitudinal direction (Y-axis direction). The head assembly 2 has an axisymmetric shape with respect to an axis of symmetry parallel to the Y axis, and the same applies to the second separation groove 44b.

  8a to 8f are cross-sectional views of the main parts in FIG. In the flexure 15, an insulating layer 41 such as polyimide is formed on a stainless flexure substrate 24 having a thickness of 18 μm, and a head element wiring 25 (wiring portion) is formed thereon. The head element wiring 25 is used for insulation or protection. And covered with a wiring cover layer 42 of polyimide or the like. Further, the flexure 15 secures a necessary mechanical function by etching the flexure substrate 24 into an arbitrary shape. FIG. 8a is a cross-sectional view taken along the line CC in FIG. 8b is a DD cross section in FIG. 6, FIG. 8c is a FF cross section in FIG. 6, FIG. 8d is a GG cross section in FIG. 6, FIG. 8e is a HH cross section in FIG. It is sectional drawing which shows EE cross section.

  In the CC portion shown in FIG. 8 a, the first outrigger 22 a is composed of a flexure substrate 24 and is connected to the slider support plate 20. An insulating layer 41 is formed on a part of the first outrigger 22a, and a head element wiring 25 (wiring portion) made of copper foil is formed thereon, and a wiring cover is provided so as to cover the head element wiring 25 (wiring portion). 42 is formed. 8B, the flexure substrate 24 on the back surface side of the head element wiring 25 (wiring portion) is removed by etching, and the slider support plate 20, the first outrigger 22a, and the head element wiring 25 ( The wiring part) is separated. In the FF portion shown in FIG. 8c, the first drive rib 36a which is the flexure substrate 24 extending from the slider support plate 20 and a part of the head element wiring 25 (wiring portion) are fixed, and the head element wiring 25 ( The wiring portion) and the first outrigger 22a are separated.

  In the GG portion shown in FIG. 8D, the first joint 40a has the same cross-sectional shape as the DD cross section, and the head element wiring 25 (wiring portion) obtained by etching away the flexure substrate 24 of the flexure 15 and the insulating layer 41. The wiring cover 42 is formed. 8H, the second joint 40b is formed by a head element wiring 25 (wiring portion) obtained by etching away the flexure substrate 24 of the flexure 15, an insulating layer 41, and a wiring cover. Since the first and second joints 40a and 40b are more flexible than the first link 39a, when the first driving means 16a expands and contracts, the first link 39a is the second link 39a. A slight rotational movement is performed around the joint 40b. Similarly, when the second drive unit 16b expands and contracts, the second link 39b performs a fine rotation about the fourth joint 40d. In conjunction with this, the slider support plate 20 rotates around the fulcrum protrusion 21.

  As shown in FIG. 8f, the first driving means 16a is bonded onto the first piezoelectric support 23a at a position overlapping the reinforcing plate 43a of the first link 39a (part indicated by a broken line P in FIG. 8f). Has been. Further, the first driving means 16a is bonded to the first piezoelectric material support portion 23a so that the other end portion also overlaps the flexure substrate 24 of the flexure 15 (part indicated by a broken line Q in FIG. 8f). . Similarly, the second driving means 16b is bonded to the reinforcing plate 43b of the second link 39b on the second piezoelectric support portion 23b at a position where one tip end portion overlaps. Further, the other tip of the second driving means 16b is also bonded onto the second piezoelectric support 23b at a position overlapping the flexure substrate 24 of the flexure 15. Thereby, the displacement of the thin film piezoelectric body 26 can be reliably transmitted to the first link 39a (or the second link 39b).

  In the present embodiment, the dynamic vibration absorber 33 is located on the rear end side of the load beam 14 with respect to the fulcrum protrusion 21 and is between the first driving means 16a and the second driving means 16b, and the first driving means 16a. And the second drive means 16b. The dynamic vibration absorber 33 includes a first weight portion 33a, a second weight 33e, a spring portion 33b that connects the first weight portion 33a to the slider support plate 20, and a first weight portion 33a that is a slider support plate. 20 includes a damping portion 33c that suppresses an amplitude that vibrates with respect to 20, a spring portion 33b, and a frame portion 33d that supports the damping portion 33c. The first weight portion 33 a and the spring portion 33 b are formed by etching on the insulating layer 41 of the flexure 15 with the same copper foil as the head element wiring 25. The damping portion 33c is formed by etching with an insulating layer 41 made of polyimide. Further, the second weight portion 33 e is formed by etching on the flexure substrate 24. Thus, since the structure of the dynamic vibration absorber 33 can be processed by the etching process of the flexure 15, a new processing process is not required and the processing cost is not required at all.

  In the present embodiment, the first weight portion 33a and the second weight portion 33e are substantially rectangular and are provided along the X direction in the figure. Can be set arbitrarily. These first and second weight portions 33 a and 33 e serve as mass portions in the dynamic vibration absorber 33.

  The frame portion 33d is provided in a frame shape so as to surround the first weight portion 33a, the spring portion 33b, and the second weight portion 33e. The spring portion 33b is provided so as to extend along the negative direction of the Y axis, and one end in the longitudinal direction is connected to the frame portion 33d. The spring portion 33b is connected to the first weight portion 33a in the vicinity of the center in the longitudinal direction.

  Thus, the first weight part 33a, the second weight part 33e, the spring part 33b, and the damping part 33c have a substantially H-shape as a whole. With this configuration, the dynamic vibration absorber 33 has a degree of freedom of vibration with respect to the rotation direction of the slider support plate 20. Here, the damping portion 33 c plays a role of suppressing vibration in the rotation direction of the slider support plate 20.

  FIG. 9 is a diagram illustrating an operation state of the dynamic vibration absorber. When the slider 3 rotates around the fulcrum protrusion 21, the first weight portion 33a (second weight portion 33e) reciprocates in the direction of the arrow shown in the figure. The arrow direction coincides with the direction in which the head element 7 of the slider 3 crosses the recording track. In the resonance mode in which the slider 3 resonates greatly, the first and second weight portions 33a and 33e of the dynamic vibration absorber 33 function to absorb the vibration of the slider 3 and suppress the resonance.

  10a and 10b are views showing a cross section of the dynamic vibration absorber 33, and FIG. 10a is a cross sectional view showing a JJ cross section in FIG. 10b is a cross-sectional view showing a cross section taken along line II in FIG. The spring part 33 b is provided on the upper surface side of the damping part 33 c formed of the insulating layer 41. The first weight portion 33 a is provided on the upper surface side of the insulating layer 41, and the second weight portion 33 e is provided on the back surface side of the insulating layer 41. In order to match the resonance frequency of the dynamic vibration absorber 33 to the optimum value, the shapes of the first weight 33a and the second weight 33e can be arbitrarily determined. Moreover, the first weight part 33a may have a mass adjusting part. Specifically, it is possible to finely adjust the resonance frequency by adding a solder ball 33f on the first weight portion 33a. The position where the solder ball 33f is added may be provided at one place on the axis of symmetry (Y axis) passing through the fulcrum protrusion 21, and a plurality of the central axes may be provided symmetrically. In addition, the wiring cover layer 42 is not provided in the position which adds the solder ball 33f on the 1st weight part 33a.

  Hereinafter, a head assembly having a frequency response characteristic that does not resonate in the track direction of the head element will be specifically described according to this embodiment.

(Conventional example)
First, configurations of the first conventional example and the second conventional example will be described. FIG. 18 is a view showing a head assembly of a first conventional example. FIG. 20 is a diagram showing a head assembly of a second conventional example. The head assembly according to the first conventional configuration shown in FIG. 18 is obtained by removing the dynamic vibration absorber 33 from the configuration of the head assembly 2 of the present embodiment shown in FIG. 6, and the other configurations are shown in FIG. Same as the configuration. Further, the head assembly according to the second conventional example shown in FIG. 20 corresponds to Japanese Patent Application Laid-Open No. 2003-228561, and is replaced by the center of gravity instead of the dynamic vibration absorber 33 in the configuration of the head assembly 2 of the present embodiment shown in FIG. The adjustment counter balance 60 is attached to the slider support plate 20, and the center of gravity Gr of the rotating portion is aligned with the fulcrum protrusion 21. Other configurations are the same as those shown in FIG.

FIG. 19a is a simplified model diagram of the first conventional example in FIG. FIG. 19B is a diagram showing a state in which the slider reciprocates around the fulcrum protrusion by applying an alternating voltage to the first driving means and the second driving means in the first conventional example. It is. The center of gravity Gr of the rotating portion including the slider 3 and the slider support plate 20 is located at a distance S ₁ from the fulcrum protrusion 21 toward the head element 7. Therefore, the center of gravity Gr moves in the X direction when the slider 3 rotates. The reaction due to the movement of the center of gravity Gr is transmitted to the load beam 14 through the fulcrum protrusion 21 and shakes the load beam 14 in the X direction. The rolling behavior of the load beam 14 is shown in FIG. 19c.

  FIG. 19d is a diagram showing the response characteristic of the displacement in the X direction of the head element 7 with respect to the input voltage applied to the driving means in the first conventional example. In this response characteristic, a large peak with a gain of 20 dB appeared at a frequency of 25 kHz. The center of gravity Gr of the rotating portion including the slider 3 and the slider support plate 20 is separated from the fulcrum protrusion 21, and the reaction force in the X direction due to this rotating motion is transmitted to the load beam 14 through the fulcrum protrusion 21. When the frequency of this rotational motion coincides with the roll mode (Sway mode) of the load beam 14, the roll resonance (Sway mode) mode of the load beam 14 is excited.

Next, a second conventional example will be described. FIG. 21a is a simplified model diagram of the second conventional example in FIG. FIG. 21 b is a diagram showing a state in which the slider is reciprocatingly rotated around the fulcrum protrusion by applying an alternating voltage to the first driving means and the second driving means in the second conventional example. It is. FIG. 21c is a model diagram in which the slider rotation portion and the counter balance are further simplified in the second conventional example. Here, M corresponds to the weight of the rotating part including the slider 3 and the slider support plate 20 of the first conventional example, and m ₃ corresponds to the mass of the counterbalance 60. S ₁ is the distance between the center of mass of M and the fulcrum protrusion 21, and S 3 is the distance between the center of mass of the counterbalance 60 and the fulcrum protrusion 21. If the center of gravity of the rotating part including the slider 3 and the slider support plate 20 and the counter balance 60 is located at the position of the fulcrum protrusion 21, no reaction force is generated in the fulcrum protrusion 21 when the whole reciprocally rotates. . When the condition that the center of gravity is at the position of the fulcrum protrusion 21 is expressed by a simple expression, the following expression (1) is obtained.

  When the angular velocity of the rotational amount θ of the rotational motion is applied to both sides of the formula (1), the following formula (2) is obtained.

Here, the angular velocity × distance, the angular velocity × distance S ₁ the speed V1 of the rotation amount θ so is the speed and the angular speed × distance S3 of the rotation amount θ and the speed V3, can be represented by the following formula (3) It becomes.

Therefore, it can be understood that the reaction force does not act on the fulcrum protrusion 21 when the momentum of M and m ₃ is balanced around the fulcrum protrusion 21. That is, since no reaction force acts on the fulcrum protrusion 21, the load beam 14 is stationary even in the roll mode. This state is shown in FIG. 21d.

  FIG. 21e is a diagram showing the response characteristic of the displacement of the head element in the X direction with respect to the input voltage applied to the driving means in the second conventional example. By adding the counter balance 60 to reduce the influence of the reaction force at the fulcrum protrusion 21, the roll resonance of the 25 kHz load beam 14 is greatly reduced. However, a new peak of resonance appeared at 30 kHz. This resonance is determined by the mass of the slider 3 and the slider support plate 20 and the spring constant of the driving means 16, and is the resonance of the slider 3 in the yaw direction. Thus, in the second conventional example, the resonance of the first conventional example 25 kHz was improved, but a new resonance of 30 kHz occurred. This becomes a problem when the control band for head positioning is further expanded.

(Example)
Hereinafter, an embodiment of the present invention will be described. FIG. 11a is a model diagram of one embodiment of the present invention in which the configuration of FIG. 6 is simplified. A first weight portion 33a (second weight portion 33e) is attached to the slider support plate 20 via a spring portion 33b and a damping portion 33c. The first driving means 16 a has one end fixed to the L-shaped first link 39 a and the other end fixed to the flexure substrate 24. A first joint 40a and a second joint 40b are disposed at both ends of the first link 39a. Similarly, the second driving means 16 b is fixed at one end to the L-shaped second link 39 b and the other end is fixed to the flexure substrate 24. A third joint 40c and a fourth joint 40d are disposed at both ends of the second link 39b. In FIG. 11a, the first line segment L1 connecting the first joint 40a and the second joint 40b, and the second line segment L2 connecting the third joint 40c and the fourth joint 40d are the load beam 14 The fulcrum protrusions 21 intersect each other.

  Here, the function of the dynamic vibration absorber 33 according to the embodiment will be described first. FIG. 14 is a simplified model diagram for explaining the action of the dynamic vibration absorber by the first weight part (second weight part), the spring part, and the damping part. In FIG. 14, the main weight portion 50 corresponds to the inertial mass of the rotating portion including the slider 3 and the slider support plate 20. The main spring portion 51 mainly corresponds to the elastic coefficient of the first and second driving means 16a and 16b. The main damping part 52 corresponds to the sum of the damping coefficients of the insulating cover 30, the base 28, and the first and second piezoelectric support parts 23 a and 23 b of the driving means 16. The sub-weight part 53 is the sum of the inertial masses centered on the fulcrum protrusion 21 due to the mass of the first weight part 33a and the second weight part 33e. The auxiliary spring portion 54 corresponds to the spring portion 33b. The auxiliary damping unit 55 corresponds to the damping unit 33c. The base surface 56 corresponds to the flexure substrate 24 in FIG.

  FIG. 15a is a diagram showing frequency response characteristics when the external weight f is applied to the main weight portion in the case where the dynamic vibration absorber in FIG. 14 is not provided. In this case, a resonance frequency ω0 guided from the main weight portion 50 and the main spring portion 51 appears, and ω0 at this time is expressed by the following equation (4). Here, M is the mass (inertial mass) of the main weight 50, and K is the spring constant of the main spring.

  FIG. 15b is a diagram showing frequency response characteristics when there is no sub-damping portion of the dynamic vibration absorber in FIG. In this frequency response characteristic, two resonance peaks ω1 and ω3 appear, and an anti-resonance peak ω2 appears between the resonance peaks. Here, when ω3 <ω2 <ω1 and the main damping unit 52 is ignored, the resonance peak ω1 and the resonance peak ω3 are expressed by the following equations (5) and (6). Further, the anti-resonance peak ω <b> 2 is expressed by the following formula (7). Here, M is a mass (inertial mass) of the main weight portion 50, K is a spring constant of the main spring portion, m is a mass of the sub weight portion, and k is a spring constant of the sub weight portion.

  In order to suppress the resonance of ω0 by the dynamic vibration absorber 33, k and m of the dynamic vibration absorber 33 are set so that the antiresonance ω2 substantially matches ω0.

  FIG. 15c is a diagram showing the frequency response characteristics in FIG. Due to the action of the sub-damping portion 55, the resonance peak appearing in FIG. 15b is eliminated, and a flat characteristic can be obtained uniformly from a low frequency to a high frequency. In this way, by adding the dynamic vibration absorber 33 to the system that originally has a resonance of ω0, the resonance ω0 of the main weight 50 can be eliminated and a flat frequency response can be obtained.

  Next, the operation of an embodiment of the present invention in which the dynamic vibration absorber 33 is applied to the head assembly 2 will be described. FIG. 11c is a model diagram showing, in a simplified manner, a state in which the slider rotates about the fulcrum protrusion by the first and second driving means in the embodiment of the present invention. First, when the first driving means 16a contracts and the second driving means 16b extends, the slider 3 rotates counterclockwise around the fulcrum protrusion 21 by the action of the first link 39a and the second link 39b. It will turn slightly. FIG. 16 shows the response characteristics of the displacement of the head element 7 in the X direction with respect to the input voltage applied to the driving means 16 at this time. This response characteristic shows a gentle characteristic in which the resonance peak is suppressed. If a resonance peak exists in the response characteristic from the viewpoint of control system design, it is difficult to ensure a gain margin for the control characteristic. Therefore, it is important to obtain a flat frequency response characteristic over a wide frequency band in order to achieve precise positioning of the head element.

  In the case of the first conventional example shown in FIG. 19d, a large resonance peak occurs at 25 kHz. This resonance peak corresponds to the roll mode (Sway mode) of the load beam. In this case, as a control characteristic, only a control band of about 3 kHz can be secured at most. In the case of the second conventional example shown in FIG. 21e, the counterbalance 60 causes the center of gravity Gr of the rotating portion of the slider 3 to coincide with the fulcrum projection 21, so that the 25 kHz load beam roll mode (Sway mode) appears. Absent. However, a slider rotation mode (Yaw mode) appears at 30 kHz. 30 kHz corresponds to ω0 in FIG. In this case, the control band is improved as compared with the case of the first conventional example, but it is difficult to secure a control band of 10 kHz. On the other hand, in the embodiment of the present invention shown in FIG. 16, since the resonance peak can be made gentle, it is possible to sufficiently secure control characteristics of 10 kHz or more. Here, the response characteristic when the damping coefficient is ignored in the characteristic of FIG. 16 is shown by a solid line in FIG. In the figure, the broken line is the same characteristic diagram as in FIG. From FIG. 17, it can be confirmed that the characteristics of FIG. 16 can be obtained by optimizing the damping coefficient.

  The optimization of the damping coefficient will be described with reference to FIGS. FIG. 12 shows a control block of a servo controller using the first drive means and the second drive means. The microactuator constituted by the first drive means 16a and the second drive means 16b and the VCM are connected in parallel, and the head positioning amount is the sum of the outputs of the microactuator and the VCM. The gain Cma of the microactuator controller and the gain Cvcm of the VCM controller are connected in parallel. The control stability of this system can be roughly determined by a round transfer function.

  The round transfer function G is expressed by the following equation (8).

  Here, Gma is a transfer function of the head support mechanism, and Gvcm is a transfer function of the VCM. A Bode diagram of the loop transfer function G is shown in FIG. In this case, the control band is 5 kHz where the gain = 0 dB. FIG. 13a shows a case where the damping part of the dynamic vibration absorber is insufficient. At this time, since the peak value of the first resonance point of the dynamic vibration absorber 33 is 0 dB, there is a problem in control stability. FIG. 13b shows the case where the gain margin is 10 dB, and FIG. 13c shows the case where the gain margin is 20 dB. Generally, it is necessary to secure a gain margin of 10 dB. In the case of FIG. 13c, the control band can be set to 10 kHz.

  Next, a mechanism for suppressing the roll mode (Sway mode) of the load beam 14 with the dynamic vibration absorber 33 will be described. FIG. 11b is a diagram showing the behavior of the first and second weight portions of the dynamic vibration absorber when the driving means is operated in a frequency range lower than the first resonance frequency (ω3) of the dynamic vibration absorber. FIG. 11c shows the behavior of the first and second weight portions of the dynamic vibration absorber when the driving means is operated in a frequency range higher than the first resonance frequency (ω3) of the dynamic vibration absorber. In FIG. 11 b, when the slider 3 is rotated counterclockwise, the first and second weight portions 33 a and 33 e are also largely swung counterclockwise around the fulcrum protrusion 21. Conversely, in FIG. 11 c, when the slider 3 rotates counterclockwise, the first and second weight portions 33 a and 33 e swing clockwise around the fulcrum protrusion 21. In this way, the operation mode of the dynamic vibration absorber 33 changes with the first resonance frequency ω3 as a boundary. Here, the first resonance frequency ω <b> 3 of the dynamic vibration absorber 33 is a higher frequency than the roll resonance mode of the load beam 14. Note that the mode in which ω0 of 30 kHz is suppressed corresponds to FIG.

  FIG. 11d is a model diagram in which the slider, slider support plate, and dynamic vibration absorber in FIG. 11a are further simplified. FIG. 11e is a simplified model diagram illustrating the operating state of FIG. 11b. In FIG. 11e, the mass M (hereinafter simply referred to as “mass M”) of the slider 3 and the slider support plate 20 is rotated about the fulcrum protrusion 21 by a rotation amount θ1, and is arranged at a distance S2 from the fulcrum protrusion 21. The mass m (hereinafter simply referred to as “mass m”) of the first weight portion 33a (second weight portion 33e) of the dynamic vibration absorber 33 thus moved is moved by a rotation amount θ2 around the fulcrum protrusion. Since the rolling mode of the load beam 14 is lower than the first resonance frequency ω3, the vibration period of the reciprocating motion of the rotation amount θ1 of the mass M and the rotation amount θ2 of the mass m is the same, and the phase delay is small. When the swing width (rotation amount) θ1 and θ2 are compared, the swing width θ2 of the first weight portion 33a (second weight portion 33e) of the dynamic vibration absorber 33 is larger than the swing width θ1 of the slider 3. That is, θ1 <θ2. Since the vibration periods of the mass M and the mass m are the same, the angular velocity of the first weight portion 33a (second weight portion 33e) of the dynamic vibration absorber 33 (hereinafter simply referred to as “the angular velocity of the dynamic vibration absorber 33”). Greatly increases. That is, the angular velocity of the swing width (rotation amount) θ1 and the angular velocity of the swing width θ2 have the relationship of the following formula (9).

  Here, if the momentum of the mass M and the momentum of the mass m are equal, the reaction force can be prevented from acting on the fulcrum protrusion 21. Therefore, when the velocity of the mass M is V1 and the velocity of the mass m is V2, the following equation (10 ), No reaction force is generated on the fulcrum protrusion.

In the present embodiment, since the swing width (rotation amount) of the mass M of the slider 3 and the slider support plate 20 is θ1, the velocity V1 of the mass M is the angular velocity of the swing width θ1 × distance S ₁ , mass m. Since the velocity V2 of the angular velocity is the angular velocity of the deflection width θ2 × the distance S2, the equation (10) can be expressed as the following equation (11).

  When the momentum of the mass m is matched with the momentum of the mass M, the equation (11) indicates that the mass m, the angular velocity of the swing width θ2 of the dynamic vibration absorber 33, and the distance S2 can be set freely. That is, even if the mass m is small, the angular velocity of the deflection width θ2 or the distance S2 may be increased. In the frequency response characteristics of FIG. 17, the roll mode of the load beam 14 is 25 kHz, and the first resonance frequency (ω3) of the dynamic vibration absorber 33 is 27.5 kHz. By setting the first resonance frequency (ω3) of the dynamic vibration absorber 33 to be a slightly higher frequency and balancing the momentum of the mass M and the momentum of the mass m, the dynamics of the slider 3 and the slider support plate 20 are increased. The center of gravity can be adjusted to the fulcrum protrusion 21. In other words, the reaction of the rotating operation of the slider 3 is not transmitted to the load beam 14. Thereby, the rolling mode of the load beam 14 can be suppressed. That is, the weight portion has a first resonance point (ω3) that resonates by vibration in the same direction as the direction in which the head element 7 of the slider 3 rotates with respect to the fulcrum protrusion 21, and the first resonance point (ω3 ) Is higher than the resonance frequency of the roll mode of the load beam, so that the roll resonance of the load beam 14 can be suppressed. Further, when the setting of the resonance frequency of the dynamic vibration absorber 33 is changed in a direction that brings the first resonance frequency (ω3) closer to the frequency of the roll mode, the amplitude of the mass m increases and the angular velocity of the dynamic vibration absorber 33 increases rapidly. Become bigger. Therefore, since the mass m can be reduced, it is possible to reduce the weight of the rotating portion including the slider 3 and the slider support plate 20.

  As described above, according to the present embodiment, by arranging the dynamic vibration absorber 33 instead of the counter balance, the yaw mode of the slider 3 can be suppressed, and the roll mode of the load beam 14 can also be suppressed. The control bandwidth can be greatly expanded. Further, the mass m of the first weight portion 33a (second weight portion 33e) of the dynamic vibration absorber 33 can be reduced, and the head and the disk can be prevented from crashing when the slider 3 floats on the disk. Is possible. Therefore, stable flying characteristics of the slider 3 can be obtained.

DESCRIPTION OF SYMBOLS 1 Magnetic disk apparatus 2 Head assembly 3 Slider 4 Housing 5 Spindle motor shaft 6 Magnetic disk 7 Head element 8 Support arm 9 Horizontal rotation shaft 10 of VCM Magnet part 11 Lamp mechanism 12 Tab 13 Base plate 14 Load beam 15 Flexure 16a 1st Drive means (thin film piezoelectric element)
16b Second driving means (thin film piezoelectric element)
17a, 17b Beam welding point 18 Leaf spring 19 Bending portion 20 Slider support plate 21 Support point projection 22a First outrigger 22b Second outrigger 23a First piezoelectric body support portion 23b Second piezoelectric body support portion 24 Flexure substrate 24a First fixing portion 24b Second fixing portion 25 Head element wiring 26 Thin film piezoelectric body 27a Upper electrode 27b Lower electrode 28 Base 29a First electrode pad 29b Second electrode pad 29c Third electrode pad 29d Fourth Electrode pad 30 Insulating cover 31 Head electrode terminal 32a First bent portion 32b Second bent portion 33 Dynamic vibration absorber 33a First weight portion 33b of dynamic vibration absorber Spring portion 33c of dynamic vibration absorber Damping portion 33d of dynamic vibration absorber Dynamic vibration absorber frame portion 33e Dynamic vibration absorber second weight portion 33f Solder ball 34 T-type limiter 34a bending portion 35 load beam hole 36a first drive rib 36b second drive rib 37a drive wiring 37b ground wiring 39a first link 39b second link 40a first joint 40b second joint 40c third Joint 40d Fourth joint 41 Insulating layer 42 Wiring cover layers 43a, 43b Reinforcement plate 44a First separation groove 44b Second separation groove 50 Main weight part 51 Main spring part 52 Main damping part 53 Sub weight part 54 Sub spring Part 55 Sub-damping part 56 Base surface 60 Counter balance

Claims (4)

A slider having a head element;
A slider support plate for holding the slider;
A load beam for holding the slider support plate;
A fulcrum protrusion provided at the tip of the load beam and rotatably supporting the slider support plate;
Driving means for rotating the slider support plate around the fulcrum protrusion;
A dynamic vibration absorber provided on the slider support plate,
The dynamic vibration absorber is located on the rear end side of the load beam with respect to the fulcrum protrusion, and has a degree of vibration freedom in the rotation direction of the slider support plate, and includes a weight portion and the weight portion. A spring portion connected to the slider support plate, and a damping portion for suppressing the amplitude of the weight portion vibrating relative to the slider support plate,
The weight portion has a first resonance point that resonates due to vibration in the same direction as the direction in which the head element of the slider rotates with respect to the fulcrum protrusion,
The head assembly according to claim 1, wherein the first resonance point has a frequency higher than a resonance frequency of a rolling mode of the load beam . A slider having a head element;
A slider support plate for holding the slider;
A load beam for holding the slider support plate;
A fulcrum protrusion provided at the tip of the load beam and rotatably supporting the slider support plate;
Driving means for rotating the slider support plate around the fulcrum protrusion;
A dynamic vibration absorber provided on the slider support plate,
The dynamic vibration absorber is located on the rear end side of the load beam with respect to the fulcrum protrusion, and has a degree of vibration freedom in the rotation direction of the slider support plate, and includes a weight portion and the weight portion. A spring portion connected to the slider support plate, and a damping portion for suppressing the amplitude of the weight portion vibrating relative to the slider support plate,
The head assembly according to claim 1, wherein the weight portion, the spring portion, and the damping portion are provided by etching a flexure laminate. A slider having a head element;
A slider support plate for holding the slider;
A load beam for holding the slider support plate;
A fulcrum protrusion provided at the tip of the load beam and rotatably supporting the slider support plate;
Driving means for rotating the slider support plate around the fulcrum protrusion;
A dynamic vibration absorber provided on the slider support plate,
The dynamic vibration absorber is located on the rear end side of the load beam with respect to the fulcrum protrusion, and has a degree of vibration freedom in the rotation direction of the slider support plate, and includes a weight portion and the weight portion. A spring portion connected to the slider support plate, and a damping portion for suppressing the amplitude of the weight portion vibrating relative to the slider support plate,
The weight assembly includes a mass adjusting unit. Magnetic disk device characterized by head assembly according is mounted to one of claims 1 to 3.

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