JP4838751B2 - Head suspension - Google Patents

Description

  The present invention relates to a mechanical component constituting an information recording apparatus such as a hard disk apparatus for recording information with high density. More specifically, the present invention relates to a mechanism for positioning a recording / reproducing head on a disk surface, and relates to a head suspension that supports a recording / reproducing head at its tip and performs a swinging motion by an actuator.

  The function of the head suspension is to apply a spring elastic force to the head slider provided with the recording / reproducing head at the bending elastic portion, and to support and position the head slider via the flexure. For this reason, it is desirable that the head suspension has a low primary bending vibration. The natural frequency of the higher-order mode is designed to be as high as possible, and the natural frequency of the torsional vibration mode and the sway mode swinging in the radial direction of the disc is designed to be as high as possible.

  However, in the operating state, the air flow generated by the rotation of the disk collides with the head suspension, and induces high frequency vibrations in the head suspension. Among the turbulent excitation vibrations, the sway mode vibrations affect the track deviation. Moreover, since the frequency of the sway mode is as high as 15 kHz or more, it is difficult to suppress it by positioning control. Other vibration modes that cause the track position shift include a primary torsional vibration mode and a secondary torsional vibration mode. The frequency of the primary torsional vibration mode is several kHz and the secondary torsional vibration mode is tens of kHz. Among these, for the first torsional vibration mode, the initial deformation shape of the bending elastic portion of the head suspension is adjusted so that the vibration node line is matched with the head position so as not to affect the head track deviation. However, it is difficult to remove completely due to variation in setting, which often causes a track shift.

  The secondary torsional vibration mode can be adjusted in the same manner, but it is not always easy to suppress it stably and sufficiently. For this reason, in addition to the technology for optimizing the air flow in the disk that does not excite the vibration of the suspension, a technique that attenuates the natural vibration by using a viscoelastic sheet attached to the surface of the load beam is used as a technique to suppress the vibration of the suspension. Has been. The viscoelastic sheet has the effect of suppressing vibrations in a wide frequency range. However, the damping effect has a certain limit, and the mode damping ratio for each vibration mode is 0.01 or less. Therefore, it has been a problem to further suppress these vibration modes in order to realize further higher track density.

  In response to this problem, there has been proposed a structure that further suppresses these vibration modes by applying the principle of the dynamic vibration absorber to the structure of the head suspension. 2 and 3 are examples of a dynamic vibration absorber structure proposed for suppressing the sway mode.

  2A and 2B show a head suspension with a dynamic vibration absorber according to Patent Document 1. FIG. 2A shows the entire structure, and FIG. 2B shows the structure of the dynamic vibration absorber. A part of the front end of the side rail 23 is separated from the load beam to form a vibrating beam 23C which is an L-shaped cantilever beam. The vibration beam 23C constitutes an additional vibration system for the suspension sway mode vibration. The length of the vibration beam 23C is designed to be tuned to the natural frequency of the sway mode of the suspension, and an appropriate amount of viscoelastic sheet 23d is pasted on the plane portion of the vibration beam 23C to thereby sway mode vibration amplitude. It is clarified experimentally and theoretically that can be lowered.

  FIG. 3 shows an example of a dynamic vibration absorber structure for the secondary torsional vibration mode proposed in Patent Document 1. A tongue-shaped beam is hollowed out at a position close to the bending elastic portion 33a of the load beam 33, and the length of the beam is adjusted to synchronize with the natural frequency of the secondary torsional vibration mode. Further, a viscoelastic sheet is attached to one side of the beam to suppress the resonance in the secondary torsional vibration mode.

  FIG. 4 is an example in which a dynamic vibration absorber for the sway mode according to Patent Document 2 is added. A T-shaped portion is hollowed out in the load beam to form a cantilever composed of a mass portion 74 and an elastic portion 72, thereby constituting an additional vibration system as a sway mode dynamic vibration absorber. By adjusting the mass of the mass part and the rigidity of the elastic part, the natural frequency of the dynamic vibration absorber is synchronized with the natural frequency of the suspension. Further, the viscoelastic sheet 88 is affixed by attaching the load beam 68 so that the surfaces 80 and 84 of the dynamic vibration absorber are coupled.

US20060209466 US6,697,225B2

  A conventional head suspension with a dynamic vibration absorber has a viscoelastic sheet attached to an additional vibration system. Since the damping effect of the viscoelastic sheet varies depending on the ambient temperature, even if the viscoelastic sheet is designed to have an optimum value at the standard temperature, it becomes smaller than the optimum value at a high temperature and larger than the optimum value at a low temperature.

  Furthermore, as the head suspension becomes smaller, the viscoelastic sheet also needs to be reduced. However, such a reduced viscoelastic sheet is required due to current processing accuracy, manufacturing errors, and the thickness of the viscoelastic sheet itself. It is very difficult to manufacture.

  Further, in the case of a dynamic vibration absorber having a viscoelastic sheet, if the viscoelastic sheet is peeled off or forgotten to be applied, the effect as a dynamic vibration absorber will not be produced. Further, since a manufacturing process of attaching a viscoelastic sheet is required, the production cost is increased, and the cost is further increased by downsizing.

  The present invention provides a head suspension that can obtain a damping effect that does not vary with temperature and that can further suppress vibration modes with a simple structure.

  The head suspension according to the present invention comprises a main body portion having a head slider at one end and an actuator carriage coupled at the other end, and a beam that vibrates relatively in the width direction of the main body portion. A dynamic vibration absorber with a number satisfying the natural frequency of the head suspension and the tuning condition is composed of the beam and the main body, and the squeeze damping effect of the air film in the gap between the beam and the main body determines the optimum damping condition for the dynamic vibration absorber. Try to meet.

  That is, instead of attaching the viscoelastic sheet, the vibration mode is suppressed using the damping effect of the air in the gap between the beam and the suspension body.

  Since the present invention uses an air film instead of a viscoelastic material, it is possible to eliminate the disadvantage that the damping effect varies with temperature, and to provide a head suspension that exhibits a damping effect that is not affected by the ambient temperature. Can be provided. Moreover, since it can form in a series of processes which manufacture and process a head suspension, the process for sticking a viscoelastic sheet | seat can be deleted.

  FIG. 1 shows an outline of the mechanism configuration of a hard disk device (HDD) 41 to which the present invention is applied first. The hard disk device 41 includes a disk spindle mechanism 42 that rotates a disk that is a recording medium, and a head positioning mechanism 43 that positions a head slider 44 on the disk. The head positioning mechanism 43 is a swing arm type, and performs high track density recording by positioning and controlling a head slider 44 installed at the tip of the head slider 44 in the radial direction of the disk with a track position signal recorded on the disk as a target. The head suspension 10 is a portion made of a stainless steel plate on the front end side of the head positioning mechanism 43.

  Hereinafter, embodiments of the present invention and optimum damping conditions of the dynamic vibration absorber and air film shape conditions at that time will be described in detail.

  FIG. 5A is an overall view of the head suspension according to this embodiment. The head suspension 10 includes a load beam 1 having a narrow surface made of a thin stainless steel plate, a bending elastic portion 2 for applying a load force to the head slider 4 by a flexible spring elastic force, and a longitudinal direction for increasing the bending rigidity of the load beam 1. A side rail 5 provided on the side and a base plate 6 for coupling the head suspension 10 to a swing actuator carriage on the base side of the head positioning mechanism (see FIG. 1). A head slider 4 is installed at the tip of the head suspension 10 via a flexure 3. The function of the head suspension 10 is to apply a load force of 1 to 2 grams to the head slider 4 by the bending elastic portion 2, and to support and position the head slider 4 without deformation in the in-plane direction of the disk (see FIG. 1). It is.

FIG. 5B is an enlarged view of the portion of the dynamic vibration absorber that precedes the broken line 7 at the tip of the load beam 1 in FIG. The additional vibration system as the dynamic vibration absorber according to the present embodiment is a head suspension formed by bending a square beam formed to extend to the tip of the side rail 5 with a side rail 5 and a gap 9 and 180 degrees. This is an L-shaped vibrating beam 8 of a cantilever beam that vibrates in the width direction. A gap 9 between the vibration beam 8 and the side rail 5 has an air film thickness d described later. The vibration beam 8 is designed to have a length a so that the bending natural frequency f ₂ is synchronized with the sway mode natural frequency f ₁ of the head suspension 10. In the present specification, the head suspension 10 excluding the cantilever portion may be referred to as a head suspension body in order to distinguish the cantilever portion from other portions.

Here, the tuning condition will be described. When carrying thereon an additional oscillating system is a natural frequency of f ₂ initially suspension tip at the distal end of the main vibration system of the head suspension having f natural frequency of ₁ (resonance frequency), f ₁ and f ₂ frequency close Resonance frequencies (resonance frequencies) f ₁ ′, f ₂ ′ (f ₂ ′ <f ₂ <f ₁ <f ₁ ′) occur in the response characteristics. The frequency response characteristics of the sway mode vibration of the suspension body with respect to two different dampings are described below with the vertical axis representing the amplitude magnification and the horizontal axis representing the dimensionless excitation frequency ratio (f / f ₁ ). Thus, there are two fixed points P and Q whose amplitude magnification does not change depending on the damping ratio ζ. The amplitude magnifications of these fixed points P and Q change only with the natural frequency ratio f ₂ / f ₁ . Therefore, if f ₂ / f ₁ is appropriately selected, the amplitude magnification at the point P and the amplitude magnification at the point Q can be made equal. Determining the natural frequency f ₂ of the additional vibration system with respect to the resonance frequency f ₁ of the main vibration system so that the amplitude magnifications of these fixed points P and Q are equal to each other is called tuning. A condition where the magnifications are equal to each other is called a tuning condition.

  In general, these fixed points P and Q are not resonance points between the main vibration system and the additional vibration system. However, if the damping ratio of the damping attached to the additional vibration system is appropriately selected, the values of the two resonance points, that is, the horizontal axis (excitation frequency ratio) at which the frequency response characteristic takes the maximum value are close to the fixed points P and Q, respectively. The amplitude magnification substantially coincides with the amplitude magnification of the fixed points P and Q. Designing the damping ratio of the additional vibration system under such conditions is called optimum damping conditions. The principle of the dynamic vibration absorber is a simple method that reduces vibrations against the excitation force in a wide band of the main vibration system by adding an additional vibration system that satisfies the tuning condition and the optimum damping condition to the main vibration system with small damping. The optimum design conditions for dynamic vibration absorbers for linear vibration systems are explained in detail in general textbooks and reference books such as mechanical engineering manuals.

Therefore, the head suspension as the main vibration system and the beam as the additional vibration system according to the present embodiment are modeled into a linear vibration system. Then, the tuning condition of the head suspension and the beam is f ₂ / f ₁ = 1 when the ratio m ₂ / m ₁ of the equivalent mass m ₂ of the primary bending vibration of the beam to the equivalent mass m ₁ of the sway mode of the suspension is μ. / (1 + μ). For example, when μ = 0.05, the tuning condition is satisfied if the design is such that f ₂ = 0.95f ₁ . The vibration beam 8 is designed to have a natural frequency f ₂ that satisfies this tuning condition.

In the case of a cantilever having a uniform thickness and width, the equivalent mass m ₂ of the primary bending vibration is given by m ₂ = 33ρabh / 140. The bending stiffness of the beam is given by k = 4Ebh ³ / a ³ . Therefore, the natural frequency (resonance frequency) f ₂ of the primary bending of the additional vibration system is

It becomes. Here, h is the thickness of the beam, a is the length of the beam, b is the width of the beam, E is the Young's modulus of the beam, and ρ is the density of the material used for the beam. As represented by Expression (1), the natural frequency f ₂ of the additional vibration system does not depend on the beam width b, but is proportional to the thickness h and inversely proportional to the square of the length a.

In the case of the embodiment shown in FIG. 5, since it is not an ideal cantilever, it deviates from the equation (1). Therefore, the natural frequency f ₂ that satisfies the tuning condition in a first order approximation is obtained using the equation (1), the frequency response of the actual suspension is analyzed in the vicinity, and the fixed points P and Q generated in the vicinity of the resonance frequency of the sway mode. The length a of the beam is determined so as to satisfy the tuning condition in which the response magnifications are equal. Since the equivalent mass m ₂ of the sway mode of the actual suspension cannot be easily obtained, the beam shape of the additional vibration system is designed to satisfy the tuning condition, and the relational expression of f ₂ / f ₁ = 1 / (1 + μ) The mass ratio μ is obtained from the relational expression between the size of the fixed points P and Q and the equivalent mass ratio m ₂ . In the case of the suspension having a length of 14 mm and a plate thickness of 38 μm used in this embodiment, the natural frequency f ₁ of the sway mode is about 18 kHz, and the length a of the additional vibration beam that satisfies the optimum tuning condition is 1 0.1 to 1.3 mm.

  In the present invention, the width of the gap 9 between the vibration beam 8 of the additional vibration system and the side rail is designed so that the squeeze damping effect of the air in the gap 9 is optimally attenuated when the vibration beam 8 bends and vibrates. . This optimum attenuation is attenuation such that the resonance peaks come to the fixed points P and Q. In the design of a dynamic vibration absorber for an actual suspension, the optimum damping can be obtained by trial and error by changing the damping acting on the additional vibration system.

Based on the model of the linear vibration system, when the damping ratio of the air film damping to the primary bending vibration of the additional vibration system is ζ, the optimum damping ratio ζ _OP that minimizes the amplitude magnification (resonance magnification) of the sway mode is the additional vibration Using the equivalent mass ratio μ of the system to the main vibration system, [3μ / 8 (1 + μ) ³ ] ^1/2 is given. The fixed points P and Q at this time become resonance, and the maximum amplitude (amplitude magnification) of the frequency response characteristic is the height of these fixed points P and Q. The amplitude magnification at this time is given by [(2 + μ) / μ] ^1/2 . For example, when the ratio μ = m ₂ / m ₁ of the equivalent mass m ₁ of the head suspension to the equivalent mass m ₂ of the additional vibration system is 0.05, the optimum damping ratio is ζ _OP = 0.127, and at this time, the excitation force The amplitude magnification in the sway mode resonance with respect to the static displacement is 6.4.

  Next, a design method for the gap 9 giving the optimum damping ratio, that is, the distance between the beam 8 and the side rail 5, that is, the thickness d of the air film will be described. In general, the damping coefficient c (when the air film having a length (beam length) l, width b (= beam width b = width of the gap 9) and thickness d vibrates slightly in the thickness direction at a frequency f ( Ns / mm) is approximately

Given in. Here δ = b / l, σ = 24πμ a fb 2 / (p a d 2), where, mu _a: viscosity coefficient of air, _{p a:} is the ambient pressure.

The magnitude of the optimum damping of the dynamic vibration absorber can be generally expressed by the damping ratio of the additional vibration system. If the equivalent mass of the additional vibration system is m ₂ and the equivalent stiffness is k ₂ , the damping ratio ζ is

Given in.

FIG. 6 shows a thickness d on the side of a length of 0.8 mm (corresponding to the air film length l) from the tip of a stainless steel cantilever having a thickness of 38 μm and a length a = 1.2 to 1.5 mm. The damping coefficient c when assuming that the air film thickness d is uniformly sinusoidally oscillating with the same vibration amplitude as that of the beam tip is calculated from the equation (2). The damping ratio ζ is obtained from the equation (1) for obtaining the natural frequency of the shaped additional vibration system and the above-described equivalent mass m ₂ using the equation (3), and the beam width b and the gap 9 between the beam 8 and the side rail 5 are obtained. It is shown as a function of the thickness d of the air film present in FIG. According to equation (1), the natural frequency of the primary bending when the beam length is 1.2 mm is 23.5 kHz, and when the beam length a is 1.5 mm, f ₂ = 15.0 kHz. there were. In Figure 6, taking the width b of the natural frequency f ₂ and the beam of the cantilever as a parameter. From this figure, it can be seen that the magnitude of the damping given to the primary bending vibration of the beam varies depending on the beam width b. Further, the higher the frequency, the stronger the attenuation, and the smaller the air film thickness d.

Now, to set the equivalent mass ratio of the additional vibration system to μ = 0.05 and to suppress the maximum amplitude magnification of the suspension vibration to the minimum 6.4, the beam damping ratio by the air film is set to the optimum damping ratio ζ _op = What is necessary is just to design to 0.127. From the tuning conditions and FIG. 6, to obtain this optimum damping ratio, the beam length a is 1.2 mm, the beam natural frequency f ₂ is 23.5 kHz, and the beam width b is 0.1 mm and 0.2 mm. And 0.4 mm, the thickness d of the air film is 2.89 (≈3) μm, 5.48 (≈5.5) μm, and 10.4 (≈10) μm, respectively.

Consider the general relationship between the air film width b and thickness (gap) d that gives the desired damping ratio ζ. In the expression (2), since δ ≦ 0.5 in many cases, δ ² ≈0, and σ is proportional to the frequency and the square of b / d. However, even when the frequency is 23.4 kHz, b When / d <50, σ <1.0, so σ ² ≈0. Therefore, equation (2) is approximately

It is represented by On the other hand, in the expression of the damping ratio ζ in the expression (3), the equivalent mass m ₂ of the additional vibration system is considered to be proportional to the length l of the squeezed air film. )

It becomes. In other words, the damping ratio ζ is proportional to the cube of the air film width / gap ratio (b / d) and inversely proportional to the natural frequency f ₂ of the additional vibration system. As described above, in FIG. 6, the value of the gap d at which the curve at 23.5 kHz gives the optimum value ζ = 0.127 is when b = 0.4, 0.2, 0.1 mm. These are 10.4, 5.48, and 2.89 μm, respectively. When b / d is calculated here, it becomes 38, 36, and 34 at each time, and takes almost the same value. Therefore, the validity of the expression (5) can be verified. Therefore, when designing the air film shape as the optimum attenuation condition, the beam width b may be about 34 times the gap d.

Here, the design conditions of the squeeze air film corresponding to the set region of these damping ratios ζ from which the damping effect can be obtained are b / when f ₂ ≈f ₁ = 23.5 kHz and ζ = 0.127 in FIG. Since d = 34 (d / b = 1/34), when calculated using the simplified formula (5), b / d = 21.5 when ζ = 0.025, and ζ = 0.67. b / d = 64.4. This is the case where the natural frequency of the additional vibration system is 23.5 kHz, but in the case of the natural frequency of 18.0 kHz in the sway mode, ζ = 0.025 (approximately 0.127 of 0.127) from the equation (5). When 1/5), b / d = 20, and when ζ = 0.67 (about 5.3 times 0.127), b / d = 59. From these, it can be said that the dynamic vibration absorber for the sway mode near 23.5 kHz may be designed in the range of b / d = 21.5 to 64.4. That is, the gap d may be in the range of 1 / 64.4 to 1 / 21.5 of the beam width b. Further, it can be said that the dynamic vibration absorber for the sway mode near 18 kHz may be designed in the range of b / d = 20 to 59. That is, the gap d may be in the range of 1/59 to 1/20 of the beam width b.

In order to design the sway mode resonance magnification to another value α, the additional vibration system should be set so that the mass ratio μ _α satisfies α = [(2 + μ) / μ] ^1/2 as the amplitude magnification of the maximum amplitude. design, to design a natural frequency of the additional vibration system so as to satisfy the tuning conditions for this mu _alpha. Next, since the optimum damping ratio ζ _{OP in} this case is given by [3 μ _α / 8 (1 + μ _α ) ³ ] ^1/2 , the width b of the air film at which the damping ratio ζ becomes this value from the characteristics of FIG. And the thickness d may be designed.

However, such a design method is used when the equivalent mass ratio between the main vibration system and the additional vibration system can be calculated by mode analysis or the like. When the equivalent mass ratio cannot be calculated, the size of the additional vibration system (in Example 1 in FIG. 5, the equivalent mass changes when the beam width is changed) is changed, and the fixed points P and Q of the frequency response characteristics are equal. these fixed points P when designing the natural frequency f ₂ such that the height, so that the height of Q is equal to or less than the desired resonant magnification, designing the trial added vibration system shape.

Next, in order to discuss the damping effect of the frequency response by the dynamic vibration absorber and the damping effect when the design state deviates from the tuning condition and the optimum damping condition, the equivalent mass ratio of the additional vibration system μ = m ₂ / m ₁ FIG. 7, FIG. 8, and FIG. 9 show frequency response characteristics depending on the excitation force of the main vibration system when = 0.05. FIG. 7 is a frequency response characteristic when the tuning condition f ₂ / f ₁ = 1 / (1 + μ) = 0.95 is strictly satisfied and the damping ratio ζ of the additional vibration system is changed. It is shown as an amplitude ratio to displacement against force. FIG. 8 shows frequency response characteristics when the natural frequency f ₂ of the additional vibration system is f ₂ / f ₁ = 1.000, which is 5% larger than the tuning condition f ₂ / f ₁ = 0.952. FIG. 9 is a frequency response characteristic when the natural frequency f ₂ of the additional vibration system is 5.4% smaller than the tuning condition f ₂ / f ₁ = 0.952, f ₂ / f ₁ = 0.900. .

First, from FIG. 7, under the tuning condition, when the damping ratio ζ is the optimum value ζ _op = 0.127, the amplitude magnification of the fixed points P and Q becomes the maximum value 6.4. When the currently used viscoelastic sheet is attached to the suspension, the amplitude magnification is 50 or more, so that it can be seen that the vibration absorber of this embodiment is significantly suppressed. Further, when the damping ratio ζ is 0.254 which is twice the optimum value or 0.064 which is 1/2, the maximum amplitude magnification is 10.4 and 9.5, respectively. Further, the amplitude magnification is 20.5 and 17.5 at the maximum even when the damping ratio ζ is 0.032 that is ¼ of the optimum value, or 0.508 that is four times the optimum value. Even when the damping ratio ζ is 0.016, which is 1/8 of the optimum value, or 1.016, which is 8 times, the maximum amplitude magnification is 26.2 and 39.2. It can be seen that there is sufficient vibration suppression effect by the device.

Next, as shown in FIG. 8, when the natural frequency of the additional vibration system increases by 5% with respect to the natural frequency of the main vibration system, the amplitude magnification increases at point P, and the amplitude magnification decreases at point Q. To do. However, at the optimum damping ratio ζ _OP = 0.127, the maximum value of the amplitude magnification hardly deviates from the points P and Q. The maximum value of the amplitude magnification is 8.9 at the point P, which is about 1.3 times that of the optimum tuning condition. In this state, when the damping ratio ζ increases to 0.254, which is twice the optimum value ζ _OP 0.127, or decreases to ½, 0.064, the maximum value of the amplitude magnification is 11.5, 12. 6 Furthermore, even if the damping ratio ζ changes to 0.508, which is four times the optimum value ζ _OP , or 0.032, which is 1/4, it can be seen that the maximum value of the amplitude magnification is 21.9 and 21.0, respectively. . Further, even when the attenuation ratio ζ is 0.016 which is the optimum value ζ _OP 1/8, the amplitude magnification is 35.1. Therefore, it can be seen that when the damping ratio ζ is in the range of 1/8 to 4 times the optimum value ζ _OP , there is a sufficient damping effect by the dynamic vibration absorber of this embodiment. It can be seen that even when the damping ratio ζ is 1.016 which is eight times the optimum value ζ _OP , the amplitude magnification is 45.3, and the vibration is suppressed to the same level or more as the viscoelastic sheet.

As shown in FIG. 9, when the natural frequency of the additional vibration system is reduced by 5% with respect to the natural frequency of the main vibration system, the amplitude magnification is increased at the point Q, and the amplitude magnification is decreased at the point P. To do. However, at the optimum damping ratio ζ _OP = 0.127, the local maximum value hardly deviates from the points P and Q. The maximum value of the amplitude magnification occurs near the Q point and is 9.3, which is approximately 1.5 times the optimum tuning condition. In this state, when the damping ratio ζ increases to 0.254, which is twice the optimum value ζ _OP 0.127, or increases to 0.064, which is 1/2, the maximum value of the amplitude magnification is 10.7, 14 respectively. .7. Furthermore, even if the damping ratio ζ changes to 0.508 which is four times the optimum value ζ _OP or 0.032 which is ¼, the maximum value of the amplitude magnification is 20.0 and 26.9, respectively. Even when the damping ratio ζ is 1.016, which is 8 times ζ _OP, the amplitude magnification is 38.0. Therefore, it can be seen that when the damping ratio ζ is in the range of 1/4 to 4 times the optimum value ζ _OP , there is a sufficient vibration damping effect by the dynamic vibration absorber. It can be seen that even when the damping ratio ζ is 0.016 which is 1/8 of ζ _OP, the amplitude magnification is 51.3, which is the same as that of the viscoelastic sheet.

From the above, if the tuning condition of the additional vibration system is designed to be within ± 5% of the optimum value ζ _OP and the damping ratio ζ is in the range of 1/2 to 2 times the optimum value ζ _OP , the maximum Can be suppressed to a range of 15 or less. Furthermore, if the damping ratio ζ is in the range of 1/4 to 4 times the optimum value ζ _OP, a sufficient effect can be obtained.

FIG. 10 shows frequency response characteristics when the natural frequency f ₂ of the additional vibration system is f ₂ / f ₁ = 1.050, which is 10% larger than the tuning condition f ₂ / f ₁ = 0.952. FIG. 11 shows frequency response characteristics when the natural frequency f ₂ of the additional vibration system is f ₂ / f ₁ = 0.857 which is 10% smaller than the tuning condition f ₂ / f ₁ = 0.952.

As shown in FIG. 10, when f ₂ / f ₁ = 1.050, the amplitude magnification increases at point P, the amplitude magnification decreases at point Q, and when the damping ratio ζ is the optimum value ζ _op = 0.127, the fixed point P It can be seen that the amplitude magnification of is about 14. Even when the damping ratio ζ becomes 0.254, which is twice the optimum value ζ _op , the maximum amplitude magnification is about 13 which is almost the same as ζ _op . In addition, when the damping ratio ζ is 0.064 that is 1/2 of the optimum value ζ _op , or 0.508 that is four times the maximum value, the maximum amplitude magnification is about 19 and 24, respectively. I understand that there is. In addition, when the damping ratio ζ is 0.032, which is 1/4 of the optimum value, or 1.016, which is eight times, the maximum amplitude magnification is about 34 and 42, respectively, which is more than the same level as the viscoelastic sheet. You can see that it is shaken.

As shown in FIG. 11, when f ₂ / f ₁ = 0.857, the amplitude magnification decreases at point P, the amplitude magnification increases at point Q, and when the damping ratio ζ is the optimum value ζ _op = 0.127, The amplitude magnification is about 13. When the damping ratio ζ is 0.254, which is twice the optimum value ζ _op , the maximum amplitude magnification is about 12 which is slightly smaller than ζ _op . In addition, when the damping ratio ζ is 0.508, which is four times the optimum value ζ _op , or 0.064, which is 1/2, the maximum amplitude magnification is about 20 and 21, respectively. I understand that there is. In addition, when the damping ratio ζ is 0.032 that is ¼ of the optimum value, or 1.016 that is 8 times, the maximum amplitude magnification is about 39 and 37, respectively, which is more than the same level as the viscoelastic sheet. You can see that it is shaken.

From the results of FIGS. 10 and 11, the tuning condition of the additional vibration system is designed to be within ± 10% of the optimum value ζ _OP , and the damping ratio ζ is in the range of 1/2 to 4 times the optimum value ζ _OP. It can be seen that sufficient damping effect can be obtained.

Further, from FIGS. 8 to 11, when the damping ratio ζ is displaced in the direction of decreasing from the optimum value ζ _{OP, the} maximum amplitude magnification increases according to the amount of displacement, but is displaced in the direction of increasing from the optimum value ζ _OP . In this case, it was found that the maximum amplitude magnification did not change much.

For confirmation, Figure 12 the frequency response characteristics when the natural frequency _{f 2} of the additional vibration system was tuned condition _f 2 _/ f 1 = 0.952 than 15% larger _f 2 _/ f 1 = 1.095, the frequency response characteristics when the natural frequency _{f 2} of the additional vibration system was tuned condition _f 2 _/ f 1 = 0.952 than 15% smaller _f 2 _/ f 1 = 0.809 shown in FIG. 13. It can be seen that when the damping ratio ζ becomes smaller than the optimum value ζ _OP , the maximum amplitude magnification further increases, but the increase in the maximum amplitude magnification on the side larger than the optimum value ζ _OP is small. Even if the tuning condition of the additional vibration system is within ± 15% of the optimum value ζ _OP , if the damping ratio ζ is in the range of 1/2 to 4 times the optimum value ζ _OP , the maximum amplitude magnification is It is 30 or less, and it can be seen that a sufficient damping effect can be obtained.

Next, a general design condition of the squeezed air film for obtaining the damping ratio ζ for obtaining the damping effect of the dynamic vibration absorber will be considered using the squeezed air film damping equation (2) and FIG. Since the beam forming the additional vibration system as the dynamic vibration absorber does not want to provide a heavy object at the tip of the suspension, it is desirable to make it as light as possible (small if it is made of the same material as the suspension). However, if it is too small, the suppression effect will be reduced, and it will be difficult to create a beam that provides adequate air film attenuation. Therefore, when applied to the suspension of the magnetic disk device, the ratio μ = m ₂ / m ₁ of the equivalent mass m ₂ of the beam to the equivalent mass m ₁ of the suspension sway mode is practically in the range of 0.03 to 0.1. Is. The optimum damping ratio ζ _OP in μ in this range is 0.102 to 0.168 from the formula [3 μ / 8 (1 + μ) ³ ] ^1/2 . Hereinafter, the amplitude magnification when the ratio μ of the equivalent mass m ₂ of the beam to the equivalent mass m ₁ of the sway mode of the suspension is changed from 0.05 will be examined.

FIG. 14 shows the result when μ is doubled to 0.1 and the tuning condition is satisfied. When μ is 0.1, the tuning condition f ₂ / f ₁ = 1 / (1 + μ) = 0.909, and the optimum damping ratio ζ _OP is [3μ / 8 (1 + μ) ³ ] ^1/2 = 0.168. It becomes. FIG. 15 shows frequency response characteristics when the natural frequency f ₂ of the additional vibration system is 10% larger than the tuning condition f ₂ / f ₁ = 0.909 and f ₂ / f ₁ = 1.00. FIG. 16 shows frequency response characteristics when the natural frequency f ₂ of the additional vibration system is f ₂ / f ₁ = 0.82 which is 10% smaller than the tuning condition f ₂ / f ₁ = 0.909.

From the result of FIG. 14, when μ = 0.1 and the damping ratio ζ under the tuning condition is the optimum value ζ _op = 0.168, the amplitude magnification of the fixed points P and Q is 4.6. The maximum amplitude magnification when the damping ratio ζ becomes 0.336, which is twice the optimum value ζ _op , and when the damping ratio ζ becomes 0.084, which is ½ of the optimum value ζ _op , respectively. The values are almost the same as 7.1 and 7.2. Further, the maximum amplitude magnification when the damping ratio ζ becomes 0.671 which is four times the optimum value ζ _op and when the damping ratio ζ becomes 0.042 which is ¼ of the optimum value ζ _op is , 14.1 and 10.2, respectively. When the damping ratio ζ becomes 1.343, which is eight times the optimum value ζ _op , and when the damping ratio ζ becomes 0.021, which is 1/8 of the optimum value ζ _op , , 27.7 and 18.0, respectively, indicating that there is a sufficient damping effect. Furthermore, it can be seen that when the damping ratio ζ is 0.010 which is 1/16 of the optimum value, the maximum amplitude magnification is 44.5, respectively, and the vibration is suppressed to the same level or more as the viscoelastic sheet.

From the result of FIG. 15, when f ₂ / f ₁ = 1.00, when μ = 0.1 and the damping ratio ζ is the optimum value ζ _op = 0.168, the amplitude magnification of the fixed point P is 7.4. Become. The maximum amplitude magnification when the damping ratio ζ becomes 0.336, which is twice the optimum value ζ _op , and when the damping ratio ζ becomes 0.084, which is ½ of the optimum value ζ _op , respectively. 8.6 and 9.2. Further, the maximum amplitude magnification when the damping ratio ζ becomes 0.671 which is four times the optimum value ζ _op and when the damping ratio ζ becomes 0.042 which becomes ¼ of the optimum value ζ _op is 15.7 and 18.6, respectively. When the damping ratio ζ becomes 1.343, which is eight times the optimum value ζ _op , and when the damping ratio ζ becomes 0.021, which is 1/8 of the optimum value ζ _op , It is 30.8 and 32.1 respectively, and it can be seen that there is a sufficient damping effect. Further, when the damping ratio ζ becomes 0.010 which is 1/16 of the optimum value ζ _op , the maximum amplitude magnification is 46.3, respectively, and it can be seen that vibration is suppressed to the same extent as the viscoelastic sheet.

From the results of FIG. 16, when f ₂ / f ₁ = 0.82, when μ = 0.1 and the damping ratio ζ is the optimum value ζ _op = 0.168, the amplitude magnification of the fixed points P and Q is 7. 6 When the damping ratio ζ is 0.336, which is twice the optimum value ζ _op , the maximum amplitude magnification is 7.5. The maximum amplitude magnification when the damping ratio ζ becomes 0.084 which is 1/2 of the optimum value ζ _op and when the damping ratio ζ becomes four times the optimum value ζ _op is 0.671, respectively. , 13.1. The maximum amplitude magnification is 23.4 when the damping ratio ζ becomes 0.042, which is 1/4 of the optimum value ζ _op , and when the damping ratio ζ becomes 1.343, which is eight times the optimum value ζ _op. 24.9, indicating that there is a sufficient damping effect. When the damping ratio ζ is 0.021 which is 1/8 of the optimum value ζ _op , the maximum amplitude magnification is 43.5, and it can be seen that vibration is suppressed to the same extent as the viscoelastic sheet. When the damping ratio ζ becomes 0.010 which is 1/16 of the optimum value ζ _op , the maximum amplitude magnification is 71.2.

14 to 16, μ is set to 1.0, and the tuning condition of the additional vibration system is designed to be within ± 10% of the optimum value ζ _OP , and the damping ratio ζ is ¼ to 8 of the optimum value ζ _OP. It can be seen that a sufficient damping effect can be obtained if the range is doubled. Incidentally, in the case of designing the tuning condition of the additional vibration system within 0 to + 10% of the optimum value zeta _OP it is sufficient if the damping ratio zeta is a 1 / 8-8 times the optimum value zeta _OP It can be seen that a damping effect can be obtained.

FIG. 17 shows the result when μ is set to about half of 0.03 and the tuning condition is satisfied. When μ is 0.03, the tuning condition f ₂ / f ₁ = 1 / (1 + μ) = 0.971, and the optimum damping ratio ζ _OP is [3μ / 8 (1 + μ) ³ ] ^1/2 = 0.101. It becomes. FIG. 18 shows frequency response characteristics when the natural frequency f ₂ of the additional vibration system is f ₂ / f ₁ = 1.07, which is 10% larger than the tuning condition f ₂ / f ₁ = 0.97. FIG. 19 shows frequency response characteristics when the natural frequency f ₂ of the additional vibration system is 10% smaller than the tuning condition f ₂ / f ₁ = 0.97, f ₂ / f ₁ = 0.87.

From the result of FIG. 17, when μ = 0.03 and the damping ratio ζ under the tuning condition is the optimum value ζ _op = 0.101, the amplitude magnification of the fixed points P and Q is 8.2. The maximum amplitude magnification when the damping ratio ζ becomes 0.203, which is twice the optimum value ζ _OP , and when the damping ratio ζ becomes 0.051, which is 1/2 of the optimum value ζ _OP , is 13.7. 11.9. When the damping ratio ζ becomes 0.406, which is four times the optimum value ζ _OP , and when the damping ratio ζ becomes 0.025, which is ¼ of the optimum value ζ _OP , the maximum amplitude magnification is 26.6, respectively. And 21.1. When the damping ratio ζ is 0.012 which is 1/8 of the optimum value ζ _OP , the maximum amplitude magnification is 36.2, which indicates that there is a sufficient vibration damping effect. Furthermore, it can be seen that when the damping ratio ζ becomes 0.811 which is eight times the optimum value ζ _OP , the maximum amplitude magnification is 48.6, respectively, and the vibration is suppressed to the same level or more as the viscoelastic sheet.

From the result of FIG. 18, when f ₂ / f ₁ = 1.07, when μ = 0.03 and the damping ratio ζ is the optimum value ζ _op = 0.101, the amplitude magnification of the fixed point P is 19.3. . Further, the maximum amplitude magnification when the damping ratio ζ becomes 0.203 which is twice the optimum value ζ _OP is also 19.2. The maximum amplitude magnification when the damping ratio ζ becomes 0.406, which is four times the optimum value ζ _OP , and when the damping ratio ζ becomes 0.051, half the optimum value ζ _OP , is 31.8 respectively. 30.1, which shows that there is a sufficient damping effect. When the damping ratio ζ becomes 0.811 which is 8 times the optimum value ζ _OP and when the damping ratio ζ becomes 0.025 which is ¼ of the optimum value ζ _OP , the maximum amplitude magnification is 54. 1 and 57.1. Further, the maximum amplitude magnification when the damping ratio ζ becomes 0.012 which is 1/8 of the optimum value ζ _OP is 111.8.

From the result of FIG. 19, when f ₂ / f ₁ = 0.87, when μ = 0.03 and the damping ratio ζ is the optimum value ζ _op = 0.101, the amplitude magnification of the fixed points P and Q is 19.3. It becomes. When the damping ratio ζ is 0.203, which is twice the optimum value, the maximum amplitude magnification is 17.1 respectively. The maximum amplitude magnification is 26.6 when the damping ratio ζ is 0.406, which is four times the optimum value ζ _op , and when the damping ratio ζ is 0.051, half the optimum value ζ _op. 31.3, which shows that there is a sufficient damping effect. It can be seen that the maximum amplitude magnification when the damping ratio ζ is 0.812 which is eight times the optimum value ζ _op is 42.1, and is damped to the same level or more as the viscoelastic sheet. When the damping ratio ζ becomes 0.025 that is ¼ of the optimum value ζ _op and when the damping ratio ζ becomes 0.012 that is １ ／ of the optimum value ζ _op , the maximum amplitude magnification is 59 respectively. .7, 111.1.

From the results of FIGS. 17 to 19, if μ is set to 0.03 and the tuning condition of the additional vibration system is designed to the optimum value ζ _OP , the damping ratio ζ is in the range of 1/8 to 4 times the optimum value ζ _OP. It can be seen that a sufficient vibration control effect can be obtained. Further, if the tuning condition of the additional vibration system is designed within ± 10% of the optimum value and the damping ratio ζ is in the range of 1/2 to 4 times the optimum value ζ _OP, a sufficient damping effect can be obtained. I understand that.

When the results of changing the mass ratio μ, f ₂ / f ₁ from the tuning condition, and ζ are examined, when the mass ratio μ is increased (= the mass of the additional vibration system is increased), the deviation from the tuning condition is caused. It can be seen that the increase in maximum amplitude magnification is small. Therefore, when μ = 0.1, even when f ₂ / f _{1 (} not shown) is −15% or + 15% from the tuning condition, the maximum amplitude magnification is smaller than μ = 0.05, and the dynamic vibration absorber It can be said that the damping effect is sufficient. In addition, even when μ is larger than 0.1, the same effect can be exhibited within a range in which μ = 0.1 and there is a damping effect. As the mass ratio μ increases, the amount of change in the maximum amplitude magnification when ζ is displaced from the optimum value ζ _OP also decreases.

In addition, when the mass ratio μ is the same, the amplitude magnification increases in principle as f ₂ / f ₁ deviates from the tuning condition, but when f ₂ / f ₁ shifts to the smaller side and ζ is increased The amplitude magnification does not simply increase. As ζ increases, the amplitude magnification becomes smaller than the amplitude magnification at the same ζ in the tuning condition. Specifically, when μ = 0.03, there is a boundary between 0.406, which is 4 times the optimum value ζ _OP , and 0.812, which is 8 times, and when ζ = 0.810, f _{As 2} / f ₁ decreases, the maximum amplitude magnification decreases. The same is true when μ = 0.05, and there is a boundary between 0.508, which is 4 times the optimum value ζ _OP , and 1.016, which is 8 times, and when ζ = 1.016, f _{As 2} / f ₁ decreases, the maximum amplitude magnification decreases.

Further, when μ = 0.1, there is a boundary between 0.336, which is twice the optimum value ζ _OP , and 0.672, which is four times the optimum value ζ _OP , and when ζ = 0.672 or 1.344 , F ₂ / f ₁ becomes smaller, the maximum amplitude magnification becomes smaller. This is because if f ₂ / f ₁ is smaller than the tuning condition, the optimum attenuation in that case shifts to a larger value. When there is a deviation from the tuning condition, there is an optimum attenuation ratio when the deviation occurs, and when the evaluation is made with respect to the attenuation ratio, the range of the attenuation ratio that makes the amplitude magnification below a desired value becomes narrow.

Further, when f ₂ / f ₁ becomes smaller than the tuning condition, the maximum amplitude magnification increases on the fixed point P side as the damping ratio ζ becomes smaller, and when f ₂ / f ₁ becomes larger, It can be seen that it increases on the fixed point Q side. On the other hand, when the damping ratio ζ increases, the maximum amplitude magnification does not increase so much even if f ₂ / f ₁ changes. As the mass ratio μ increases, the maximum amplitude magnification when the damping ratio ζ is large decreases, and the maximum amplitude magnification when the damping ratio ζ is small increases slightly. As μ decreases, the maximum amplitude magnification increases when the damping ratio ζ is large, and the maximum amplitude magnification decreases when the damping ratio ζ is small.

FIG. 20 summarizes the results of changing the mass ratio μ and ζ when f ₂ / f ₁ satisfies the tuning condition. The case where mass ratio becomes smaller than 0.03 from this figure is examined. When the mass of the additional vibration system becomes 0, the amplitude magnification diverges and becomes ∞. However, in the range excluding 0, as long as the tuning condition is satisfied, the mass ratio becomes smaller than 0.03. It can be seen that if the damping ratio ζ is designed within a range of 1/2 to 4 times the optimum value ζ _OP , a sufficient vibration damping effect can be exhibited.

The results when the tuning conditions are changed are summarized as follows. When the mass ratio μ = 0.03 or more, if f ₂ / f ₁ is in the range of −10 to + 10% of the tuning condition, the damping ratio ζ is in the range of 1/2 to 2 times the optimum value ζ _OP . If designed with, sufficient damping effect can be demonstrated.

Further, when the mass ratio μ = 0.05 or more, if f ₂ / f ₁ is set within the range of −15 to + 15% of the tuning condition, the damping ratio ζ is ½ to 2 times the optimum value ζ _OP . If designed within the range, sufficient vibration damping effect by the dynamic vibration absorber can be exhibited. Further, if f ₂ / f ₁ is set within the range of −10 to + 10% of the tuning condition, if ζ is designed within a range of ¼ to 4 times the optimum value ζ _OP , the dynamic vibration absorber is sufficient. The vibration control effect can be demonstrated. Alternatively, if f ₂ / f ₁ is in the range of the tuning condition to −15%, a sufficient damping effect can be obtained by designing the damping ratio ζ within the range of 1/2 to 8 times the optimum value ζ _OP. Can be demonstrated.

When mass ratio μ = 0.10 or more, if f ₂ / f ₁ is in the range of −10 to + 10% of the tuning condition, the damping ratio ζ is ¼ to 8 times the optimum value ζ _OP . If designed within the range, sufficient vibration damping effect by the dynamic vibration absorber can be exhibited. If f ₂ / f ₁ is in the range of the tuning condition to + 10%, a sufficient damping effect can be obtained by designing the damping ratio ζ within a range of 1/8 to 8 times the optimum value ζ _OP. it can.

  From the characteristics of the dynamic vibration absorber described in detail above, an overview of the range of the damping ratio in which the damping effect can be obtained within the range of ± 10 from the tuning condition is as follows. When 60 and μ = 0.05, ζ = 0.016 to 1.016, and when μ = 0.10, ζ = 0.021 to 1.34. The range of the damping ratio including these is ζ = 0.001 to 1.34. Accordingly, when the value of the ratio d / b between the gap and the width of the rectangular air film that gives such an attenuation ratio is calculated using the approximate expression (5), d / b = 1/16 to 1/1 at 18 kHz. 74. From the above results, for example, when b = 0.3 mm, d = 19 μm to 4 μm.

In the above embodiment, the gap 9 does not need to be uniform along the longitudinal direction of the vibration beam 8. The vibration beam 8 is bending vibration, and the effect of air film attenuation is small on the base side of the vibration beam 8 having a small vibration amplitude. Further, the bending of the root side becomes more difficult as the gap 9 becomes smaller. Considering these things, the space near the base is widened to facilitate manufacture, and a region of about 60% to 70% length from the tip side of the vibration beam 8 having a large vibration amplitude is a squeezed air film. If the optimum gap d value that gives the optimum damping ratio ζ _OP has an accuracy of about 25%, the damping ratio of the additional vibration system can be set in a range from twice to half the optimum value ζ _OP. The maximum amplitude magnification can be suppressed to 1.56 times or less of the best value at the worst. Further, although the vibration beam 8 in FIG. 5 is bent and formed on the inner side of the side rail, it can of course be formed by bending the outer side.

  The head suspension with a dynamic vibration absorber shown in FIG. 5 is shown for a head suspension with a side rail 5 as shown in FIG. 5A, but it can also be applied to a head suspension without a side rail 5. In this case, the portion corresponding to the side rail 5 is provided only in the portion ahead of the dotted line 7 at the tip end portion of the suspension as shown in FIG.

  As shown in FIG. 6, if the width b of the air film is increased, the air film thickness d having the same damping ratio ζ can be increased. If the air film thickness d can be increased, the vibration beam 8 can be easily formed. If the width of the beam constituting the dynamic vibration absorber is increased to, for example, 1 mm, the equivalent mass of the additional vibration system also increases, and since the width of the air film is 1 mm, the thickness of the air film can be increased to about 25 μm.

  In the sway mode dynamic vibration absorber of this embodiment having such a structure, the air film thickness d can be increased to the level of the side rail regardless of the thickness of the suspension, so that the dynamic vibration absorber of the suspension is optimally tuned and optimally damped. It can be easily set to a state close to the conditions.

  Further, the equivalent mass of the additional vibration system attached to both sides of the tip of the load beam is 10% or less with respect to the equivalent mass of the sway mode of the main vibration system, so the mass increase relative to the actual head suspension is a fraction of that. Degree.

  FIG. 21 shows another embodiment of the dynamic vibration absorber structure for damping the sway mode. FIG. 21 is a view of the load beam 1 of this embodiment as viewed from above. Since the head suspension shown in FIG. 21 is not provided with a side rail (see FIG. 5A), the thickness of the plate of the load beam 1 is large, for example, about 0.1 mm. For this reason, the bending elastic portion 2 for applying a load force to the slider is formed of a thin or flexible member different from the load beam 1 and is coupled to the load beam 1 and the base plate 6 by welding or the like.

  In the embodiment of FIG. 21, a cantilever beam of an additional vibration system having a dynamic vibration absorber function is installed at the center of the tip of the suspension. 11 is a mass part of the dynamic vibration absorber, 12 is an elastic part, and the bending vibration of the primary mode that vibrates in the suspension plane by adjusting the length a of the dynamic vibration absorber and the width and length of the elastic part 12. The number is designed to tune to the natural frequency of the sway mode of the suspension. The tuning condition design method is the same as in the first embodiment.

As for the damping, the value of the gap, that is, the air film thickness d is designed so that the squeeze damping effect of air generated in the gap 13 between the mass part 11 and the suspension main part becomes the optimum damping ratio ζ _OP of the dynamic vibration absorber. The conditions for optimum attenuation are the same as in the first embodiment. The gap providing the optimum damping varies depending on the area of the surface where the squeeze damping effect occurs, that is, the suspension plate thickness and the length l of the mass portion 11. For example, in order to suppress the resonance magnification of the sway mode to 6.4, the ratio μ = m ₂ / m ₁ of the sway mode equivalent mass m ₁ of the head suspension to the equivalent mass m ₂ of the mass part 11 is 0.05. In this case, the damping ratio ζ of the additional vibration system of the dynamic vibration absorber by the air film may be set to 0.127 which is the optimum value ζ _OP .

  The gap giving this damping ratio ζ is the length of the mass portion 11, that is, the length of the air film 1 = 0.8 mm, and the plate thickness of the load beam 1, that is, the width b of the air film is 0.1 mm. From FIG. 6, d = 2.5 to 3.0 μm. When the plate thickness of the load beam 1 is 0.2 mm, d = 5.0 to 6.0 μm. From equation (5), the damping ratio ζ of an air film having an elongated surface is generally proportional to the cube of b / d. Therefore, if the width b of the narrow surface of the air film is increased, the air film thickness d is also proportional. Can be large.

The gap between the suspension main body and the dynamic vibration absorber other than the air film thickness d can be set to a sufficiently large value that is easy to process. In FIG. 21, the width of the elastic portion 12 is narrowed to reduce the bending vibration frequency in the in-plane direction of the dynamic vibration absorber, so that the gap between the suspension body and the dynamic vibration absorber is large, but this makes the suspension structure weak. When the frequency of the sway mode decreases, for example, the gap can be set to about 50 μm along the constricted shape of the elastic portion 12. However the thickness d of the air film gap 13 facing the mass portion 11, it is important to reduce as close to the optimum damping ratio zeta _OP is obtained. The amount of deviation allowed from the tuning condition and the optimum damping ratio ζ _OP is the same as in the first embodiment.

  Needless to say, the dynamic vibration absorber structure shown in FIG. 21 can also be applied to a suspension with side rails.

  The sway mode dynamic vibration absorber of this embodiment having such a structure can be formed by etching or the like at the time of manufacturing the suspension body, and it is not necessary to bend or attach a viscoelastic sheet. Further, since the molding is easy, the manufacturing error can be reduced, and the suspension dynamic vibration absorber can be manufactured in a state close to the designed tuning condition and optimum damping condition.

For a head suspension having a load beam without side rails, the cantilever structure shown in FIGS. 22A and 22B is also conceivable. In FIG. 22A, reference numerals 14a and 14b denote additional vibration systems that bend and vibrate in the vibration direction of the sway mode. The beam structure extends to the base side of the load beam 1 with the vicinity of both ends of the load beam 1 as a fulcrum. FIG. 22B is an enlarged view of a portion surrounded by a dotted ellipse 18 in FIG. Since it is integrally processed by etching or the like from the same plate material as the load beam 1, the thickness of the additional vibration system is the same as that of the load beam 1, for example, 0.1 mm. This additional vibration system has a cantilever beam structure, and the width of the mass portion 14a ₁ is increased and the width of the elastic portion 14a ₂ is decreased in order to independently design the equivalent mass m ₂ and the natural frequency.

Width of the mass portion 14a ₁ which defines the air film thickness d is large when you want to further suppress the vibration amplitude of the sway mode. Also the condition of the length l of the attenuating air film, the length of the mass portion 14a ₁ is more desirably 0.6 mm. The width of the elastic portion 14a ₂ is designed so that the natural frequency f ₂ of the additional vibration system 14a can be tuned to the natural frequency f ₁ of the main vibration system. The design method satisfying this tuning condition is as in the other embodiments already described. Width of the mass portion 14a ₁ and the elastic portion 14a ₂ when this tuning design condition the entire additional vibration system 14a can also be achieved as a uniform cantilever may be equal.

Next, in order to provide a proper damping by squeeze effect of the air to the additional vibration system, the gap 19 between the mass portion 14a ₁ and the load beam 1 it is processed to have reduced uniformity. This gap (air film thickness d) is determined using FIG. 6 as described in the first and second embodiments. In the embodiment of FIG. 22, since the thickness of the load beam (the width of the air film) is about 0.1 mm, d = 2 to 3 μm even if the length of the mass portion is 0.8 mm. Therefore, the gap between the additional vibration system and the load beam needs to be processed with considerably high precision. The amount of deviation allowed from the tuning condition and the optimum damping ratio ζ _OP is the same as in the first embodiment.

  Since the third embodiment shown in FIG. 22 has the additional vibration system in the same plane as the load beam, the degree of excitation of the vibration of the additional vibration system due to wind disturbance is smaller than that of the embodiment of FIG.

  Any of the sway mode dynamic vibration absorbers of the present embodiment having the structure as described above can be molded at the time of manufacturing the suspension body and is suitable for mass production. Therefore, the dynamic vibration absorber of the suspension can always be set in a state close to optimum tuning and optimum damping conditions.

  Next, an embodiment using a dynamic vibration absorber in the secondary torsional vibration mode is shown in FIGS. FIG. 23 shows a view of only the load beam 1 as seen from above. The structure of the other parts is the same as in the other embodiments. The dynamic vibration absorber in the secondary torsional vibration mode is close to the root of the load beam 1 and is close to a position where the antinode of the secondary torsional vibration mode is located. The tongue-shaped beam 15a is a cantilever beam that bends and vibrates in the thickness direction of the head suspension. , 15b are provided so as to be cut out from the suspension body. Since the natural frequency of the secondary torsional vibration mode is about 15 kHz, the length of the tongue-shaped beams 15a and 15b varies depending on the size and thickness of the suspension, but is about 1.3 mm to 1.5 mm.

  In this embodiment, the AA cross section of FIG. 23 is as shown in FIG. 24, and a thin air damping plate 16 different from the suspension main part is bonded or welded to one side of the load beam 1. The air attenuating plate 16 has a width that covers at least about 60% of the region from the tips of the two tongue beams toward the root, and a region including a portion facing the tongue beams 15a and 15b is air. Removal processing is performed by etching or the like so as to have a predetermined depth of the film thickness d. As described above, the lengths of the tongue beams 15a and 15b are designed so that the natural frequency of the bending vibration of the tongue beams 15a and 15b is synchronized with the natural frequency of the secondary vibration mode of the suspension. Further, the gap between the air attenuating plate and the tongue-shaped beams 15a and 15b is designed to have a thickness d that optimizes the air film attenuation effect as described above.

  Specifically, the length of the tongue beams 15a and 15b synchronized with the natural frequency of the secondary torsional vibration mode is 1.3 mm to 1.5 mm when the plate thickness of the load beam 1 is 40 μm. 1 is longer when the plate thickness is thicker, and shorter when it is thinner. On the other hand, the value of the gap giving the optimum attenuation is shown in FIG. 6 when the length l of the air film forming the gap is 0.8 mm, and when the width b is 0.2 mm and 0.4 mm, d = The ratio of the air film width b to the air film thickness d is about 36 at 5.5 to 6.5 μm and 10 to 12 μm. When the natural frequency of the vibration to be damped is 15 kHz, the range of b / d that gives a damping ratio that provides a damping effect as a dynamic vibration absorber with an amplitude magnification of 20 or less is 19 to 55. That is, the gap d may be in the range of 1/55 to 1/19 of the beam width b. FIG. 6 shows the case where the width of the air film is a maximum of 0.4 mm. However, if the width of the air film can be up to 1.0 mm by design, the thickness of the air film is d = 18 to 53 μm. .

  Another embodiment shown in FIG. 25 is also conceivable as a method for maintaining the gap between the air attenuating plate and the tongue-like beam in a desired size. In this case, the squeeze air attenuating plate 17 has a uniform thickness and has a flat surface and is bonded or welded to one side of the load beam 1. On the other hand, the portions of the tongue-shaped beams 15a and 15b having a dynamic vibration absorber function are removed only by a gap having a desired air film thickness d by etching or the like. This embodiment has an advantage that the processing of the suspension can be concentrated only on the load beam 1. In addition, when the load beam 1 is thick, the length of the tongue-shaped beam that satisfies the synchronization condition with the secondary torsional vibration mode becomes long, which causes a problem that the head suspension structure becomes weak and other natural frequencies are lowered. . Therefore, an advantage that the length of the tongue-shaped beam satisfying the tuning condition can be shortened by making the tongue-shaped beam itself thin.

  The dynamic vibration absorber structure to which the air film damping of the present invention is added can be considered to be modified in accordance with the design conditions of the head suspension. For example, when the load beam 1 is as thick as 0.1 mm, the bending primary natural frequency of the tongue beams 15a and 15b is made to coincide with the secondary torsional frequency of the head suspension, and the tongue beam length is 1.5 mm. In order to achieve the degree, it is possible to reduce the width and thickness of the root portions of the tongue-like beams 15a and 15b that do not oppose the air attenuation plate 16.

  Further, when the dynamic vibration absorber for secondary torsional vibration shown in FIGS. 23, 24 and 25 is applied to a head suspension having a bending elastic part 2 using another part as shown in FIG. 21, squeeze air The damping plates 17 and 18 can be processed as an integral part with the bending elastic part 2, thereby reducing the number of parts.

  In addition, the tongue-shaped beam is cut out from the suspension body and then folded back to form a cantilever, so that the suspension body can be used in place of the air attenuation plate 16 or the air attenuation plate 17. The number of parts can be reduced. Alternatively, the air damping plate may be wider than the suspension body, and the tongue beam may be provided on the side of the suspension body.

  Since the dynamic vibration absorber of this embodiment having such a structure can be easily designed with a cantilever beam shape and an air film thickness d with a degree of freedom, the suspension dynamic vibration absorber is close to optimum tuning and optimum damping conditions. It can be easily set to the state.

  The present invention can also be applied to the case where the primary torsional vibration mode of the head suspension is damped using the principle of a dynamic vibration absorber. FIG. 26 shows an embodiment thereof, and FIG. 26A shows a load beam 1 in which a pair of dynamic vibration absorbers 20 using squeeze air film damping are provided in the vicinity of both edges of the tip. In the case of primary torsional vibration, the position where the vibration amplitude becomes large is the tip of the load beam 1, and therefore, the fixed portion of the cantilever type dynamic vibration absorber is selected near the tip of the load beam 1.

  There are two types of structures and configuration methods of the dynamic vibration absorber of this embodiment. FIGS. 26 (B) and 26 (C) show side views of the respective examples viewed from the direction BB in FIG. 26 (A).

  In FIG. 26B, a narrow beam material is continuously extended from the load beam material in the length direction on both sides of the front end portion of the load beam 1, and the base with the load beam 1 is bent to the load beam 1 side. It is formed by. As described above, the cantilever has the mass portion 20a and the elastic portion 20b, and its bending vibration is designed so that the natural frequency is tuned to the natural frequency of the primary torsional vibration of the head suspension to be damped. The Further, the gap d between the mass portion and the load beam surface is designed to be optimally attenuated as described above.

  FIG. 26 (C) shows a shape like the sway mode dynamic vibration absorber shown in FIG. 22 that is processed in an integrated structure by extending the width direction of the load beam 1 and subsequently loads the connecting portion with the base load beam 1. The base of the cantilever is bent at a right angle so that the gap between the surface of the mass portion 20c of the cantilever and the load beam surface is a desired minute gap d. The cantilever has a mass portion 20c and an elastic portion 20d, and is designed so that the natural frequency of bending vibration of the cantilever is synchronized with the natural frequency of primary torsional vibration of the head suspension.

  Since the natural frequency of the primary torsional vibration is 5 to 7 kHz, the cantilever length that satisfies the tuning condition is considerably long in the uniform beam shape. Therefore, in the application example shown in FIGS. 26 (B) and 26 (C), the thickness of the elastic portions 20b and 20d of the additional vibration system is reduced in advance by etching or the like as compared with the thickness of the mass portion or the load beam 1. The lengths of the portions 20b and 20d are not increased so much that the natural frequency of the cantilever beam satisfies the tuning condition with the natural frequency of the primary torsional vibration of the head suspension. If the cantilever-type dynamic vibration absorber is sufficiently long, the elastic portions 20b and 20d can be designed so as to satisfy the tuning condition without being thinned. Further, the gap d between the mass parts 20a and 20c and the load beam surface is set so that the attenuation of the squeeze air film is optimal.

  When the natural frequency of the vibration to be damped is 6 kHz, b / d may be about 24 from the equation (5). Further, as described above, the ratio of the air film width b to the thickness d when the tuning condition is allowed to be ± 5% and the attenuation ratio is allowed to change from four times the optimum value to ¼. Is in the range of b / d = 14-41. That is, the gap d may be in the range of 1/41 to 1/14 of the beam width b. In the case of b / d = 14, when b = 0.5 and 1.0 mm, d = 36 and 71 μm, and the gap d can be considerably increased.

  Since the dynamic vibration absorber of the present embodiment having such a structure can increase the length and width of the cantilever portion, it is easy to set the dynamic vibration absorber of the suspension to a state close to optimal tuning and optimal damping conditions. it can.

  Here, the secondary torsional vibration mode exists where the antinode of the vibration mode is close to the root of the suspension. Since the tip of the suspension is also an antinode of vibration mode, this embodiment can be applied to secondary torsion. However, it cannot be applied simultaneously with primary torsion and secondary torsion. For simultaneous application, secondary twisting may be as shown in FIG. 23 and primary twisting as shown in FIG. If the range of the ratio between the air film width b and the thickness d is b / d = 19 to 41, that is, the gap d is in the range of 1/41 to 1/19 of the beam width b, the previous embodiment. It can be applied to any of the sway, the primary twist, and the secondary twist described in (1).

  Further, the present invention provides a dynamic vibration absorber structure with air film damping for sway mode, a dynamic vibration damper structure with air film damping for secondary torsional vibration, and air for primary torsional vibration shown in the previous embodiments. It also includes providing a head suspension having any two or three of the dynamic damper structures with membrane damping simultaneously. In general, it is known that adding a dynamic vibration absorber to multiple vibration modes attenuates a wide range of vibration modes sandwiched between those frequencies. The sway mode and the primary and secondary torsional vibration modes A head suspension having a dynamic vibration absorber structure with air film damping can be expected to suppress aerodynamically excited vibration modes in a wide frequency range.

  As can be seen from the above description, the present invention does not require a viscoelastic sheet that has been used for damping in the past to be locally applied, and can be designed with the aid of a computer to design the optimal tuning and optimum damping shape of the dynamic vibration absorber. In addition, a desired damping effect can be given to the dynamic vibration absorber by precisely machining the gap that forms the thickness of the air film. Moreover, since the damping effect of the air film hardly changes depending on the temperature, the damping effect of the dynamic vibration absorber does not change due to environmental conditions and effects with time.

It is an outline of a mechanism configuration of a hard disk device. It is an example of the prior art of the head suspension with a dynamic vibration absorber for sway modes. It is an example of the prior art of the head suspension with a dynamic vibration absorber for secondary torsional vibration modes. It is another example of the prior art of the head suspension with a dynamic vibration absorber for sway modes. It is an example of the Example of the dynamic vibration damper for sway modes with an air film attenuation | damping of this invention. This is the relationship between the damping ratio of the cantilever beam vibration system by the air film and the width and thickness of the air film. It is a frequency response characteristic by an excitation frequency ratio when μ = 0.05 and a tuning condition is satisfied. When μ = 0.05, f ₂ / f ₁ is a frequency response characteristic with an excitation frequency ratio increased by 5% in the tuning condition. When μ = 0.05, f ₂ / f ₁ is a frequency response characteristic by an excitation frequency ratio with a tuning condition reduced by 5%. When μ = 0.05, f ₂ / f ₁ is a frequency response characteristic due to an excitation frequency ratio increased by 10% in the tuning condition. When μ = 0.05, f ₂ / f ₁ is a frequency response characteristic according to an excitation frequency ratio with a tuning condition reduced by 10%. When μ = 0.05, f ₂ / f ₁ is a frequency response characteristic according to an excitation frequency ratio with a tuning condition increased by 15%. When μ = 0.05, f ₂ / f ₁ is a frequency response characteristic by an excitation frequency ratio with a tuning condition reduced by 15%. It is a frequency response characteristic by an excitation frequency ratio when μ = 0.10 and a tuning condition is satisfied. When μ = 0.10, f ₂ / f ₁ is a frequency response characteristic due to an excitation frequency ratio increased by 10% in the tuning condition. When μ = 0.10, f ₂ / f ₁ is a frequency response characteristic by an excitation frequency ratio with a tuning condition reduced by 10%. It is a frequency response characteristic by an excitation frequency ratio when μ = 0.03 satisfies the tuning condition. When μ = 0.03, f ₂ / f ₁ is a frequency response characteristic according to an excitation frequency ratio with an increase of 10% in the tuning condition. When μ = 0.03, f ₂ / f ₁ is a frequency response characteristic according to an excitation frequency ratio with a tuning condition reduced by 10%. It is a figure which shows the relationship between (micro | micron | mu) and amplitude magnification when the damping ratio is displaced from optimal damping ratio (zeta) _OP . It is an example of the other Example of the dynamic vibration damper for sway modes with an air film attenuation | damping. It is an example of the other Example of the dynamic vibration damper for sway modes with an air film attenuation | damping. It is an example of the Example of the dynamic vibration damper for secondary torsional vibration modes with an air film damping | damping. It is an example of the shape of the AA cross section of the air film | membrane attenuation | damping board and suspension in the Example of FIG. FIG. 24 is another example of the AA cross section of the air film damping plate and the suspension in the embodiment of FIG. 23. It is an example of the Example of the dynamic vibration damper for primary torsional vibration modes with an air film damping | damping.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 ... Load beam, 2 ... Bending elastic part, 3 ... Flexure, 4 ... Head slider, 5 ... Side rail, 6 ... Base plate, 7 ... Dotted line which shows a boundary, 8 ... Vibration beam, 9 ... Gap, 10 ... Head suspension, DESCRIPTION OF SYMBOLS 11 ... Mass part, 12 ... Elastic part, 13, 19 ... Gap, 15a, 15b ... Tongue beam, 16, 17 ... Air damping plate, 20 ... Dynamic vibration damper with air film damping

Claims (20)

  A main body portion having a head slider at one end and an actuator carriage coupled at the other end and a beam that vibrates relatively in the width direction of the main body portion. A dynamic vibration absorber having a frequency satisfying the tuning frequency and the natural frequency of the head suspension is configured, and the gap between the beam and the main body portion is optimal as the dynamic vibration absorber because of the squeeze damping effect of the air film in the gap. Head suspension that satisfies the damping conditions.   The head suspension according to claim 1, wherein the beam vibrates in a width direction of the main body.   The head suspension according to claim 1, wherein the beam vibrates in a thickness direction of the main body.   A main body part having a head slider installed at one end and an actuator carriage coupled to the other end, and a beam that vibrates relatively in the width direction of the main body part, and the beam has a gap between the main body part and the beam. A head suspension constituting a dynamic vibration absorber of the main body using an air film.   The beam has a length, width, and thickness that are synchronized with the natural frequency of the head suspension within a range of ± 10% of the natural frequency of the beam, and the gap is set to 1/41 to 1/1 of the width of the beam. The head suspension according to claim 4, wherein the head suspension is 22. When the natural frequency and equivalent mass of the beam and the suspension are f ₂ , m ₂ , f ₁ , and m ₁ , respectively, f ₂ / f ₁ = 1 / (1 + m ₂ / m ₁ ) and the motion The damping ratio ζ of the vibration absorber is 1/2 [3 ( m ₂ / m ₁ ) / 8 ( ₁ + m ₂ / m ₁ ) ³ ] ^1/2 ≦ ζ ≦ 4 [3 ( m ₂ / m ₁ ) / 8 ( The head suspension according to claim 4, wherein 1+ m ₂ / m ₁ ) ³ ] ^1/2 . When the natural frequency and equivalent mass of the beam and the suspension are respectively f ₂ , m ₂ , f ₁ , and m ₁ ,
0.03 ≦ m ₂ / m ₁ ,
f ₂ / f ₁ = 1 / (1 + m ₂ / m ₁ ),
The damping ratio ζ of the dynamic vibration absorber is 1/8 [3 ( m ₂ / m ₁ ) / 8 (1+ m ₂ / m ₁ ) ³ ] ^1/2 ≦ ζ ≦ 4 [3 ( m ₂ / m ₁ ). / 8 ( ₁ + m ₂ / m ₁ ) ³ ] ^1/2
The head suspension according to claim 4. When the natural frequency and equivalent mass of the beam and the suspension are respectively f ₂ , m ₂ , f ₁ , and m ₁ ,
0.03 ≦ m ₂ / m ₁ ,
0.9 / ( ₁ + m ₂ / m ₁ ) ≦ f ₂ / f ₁ ≦ 1.1 / (1+ m ₂ / m ₁ ),
The damping ratio ζ of the dynamic vibration absorber is 1/2 [3 ( m ₂ / m ₁ ) / 8 (1+ m ₂ / m ₁ ) ³ ].
^1/2 ≦ ζ ≦ 4 [3 ( m ₂ / m ₁ ) / 8 (1+ m ₂ / m ₁ ) ³ ] ^1/2
The head suspension according to claim 4. When the natural frequency and equivalent mass of the beam and the suspension are respectively f ₂ , m ₂ , f ₁ , and m ₁ ,
0.05 ≦ m ₂ / m ₁ ,
f ₂ / f ₁ = 1 / (1 + m ₂ / m ₁ ),
The damping ratio ζ of the dynamic vibration absorber is 1/8 [3 ( m ₂ / m ₁ ) / 8 (1+ m ₂ / m ₁ ) ³ ] ^1/2 ≦ ζ ≦ 8 [3 ( m ₂ / m ₁ ). / 8 ( ₁ + m ₂ / m ₁ ) ³ ] ^1/2
The head suspension according to claim 4. When the natural frequency and equivalent mass of the beam and the suspension are respectively f ₂ , m ₂ , f ₁ , and m ₁ ,
0.05 ≦ m ₂ / m ₁ ,
0.9 / ( ₁ + m ₂ / m ₁ ) ≦ f ₂ / f ₁ ≦ 1.1 / (1+ m ₂ / m ₁ ),
The damping ratio ζ of the dynamic vibration absorber is 1/4 [3 ( m ₂ / m ₁ ) / 8 (1+ m ₂ / m ₁ ) ³ ] ^1/2 ≦ ζ ≦ 4 [3 ( m ₂ / m ₁ ). / 8 ( ₁ + m ₂ / m ₁ ) ³ ] ^1/2
The head suspension according to claim 4. When the natural frequency and equivalent mass of the beam and the suspension are respectively f ₂ , m ₂ , f ₁ , and m ₁ ,
0.10 ≦ m ₂ / m ₁ ,
0.9 / ( ₁ + m ₂ / m ₁ ) ≦ f ₂ / f ₁ ≦ 1.1 / (1+ m ₂ / m ₁ ),
The damping ratio ζ of the dynamic vibration absorber is 1/4 [3 ( m ₂ / m ₁ ) / 8 (1+ m ₂ / m ₁ ) ³ ] ^1/2 ≦ ζ ≦ 8 [3 ( m ₂ / m ₁ ). / 8 ( ₁ + m ₂ / m ₁ ) ³ ] ^1/2
The head suspension according to claim 4. When the natural frequency and equivalent mass of the beam and the suspension are respectively f ₂ , m ₂ , f ₁ , and m ₁ ,
0.10 ≦ m ₂ / m ₁ ,
1 / (1 + m ₂ / m ₁ ) ≦ f ₂ / f ₁ ,
The damping ratio ζ of the dynamic vibration absorber is 1/8 [3 ( m ₂ / m ₁ ) / 8 (1+ m ₂ / m ₁ ) ³ ] ^1/2 ≦ ζ ≦ 8 [3 ( m ₂ / m ₁ ). The head suspension according to claim 4, which is / 8 ( ₁ + m ₂ / m ₁ ) ³ ] ^1/2 .   The beam is formed by extending a side rail provided on a side portion in the longitudinal direction of the main body portion in the longitudinal direction and folding the side rail in parallel with the side rail, and the gap is formed between the beam and the beam. The head suspension according to claim 4, wherein the head suspension is a gap with the side rail.   The head suspension according to claim 4, wherein the beam is provided in parallel with a longitudinal side portion of the main body portion in the width direction, and the gap is a gap between the beam and the side portion.   The head suspension according to claim 4, wherein the beam is formed by extending a side portion of the main body portion in the width direction and turning it back.   The head suspension according to claim 4, wherein the beam is formed by extending a side portion of the main body portion in the length direction and turning it back.   The head suspension according to claim 4, wherein the beam is provided by hollowing out a part of the main body.   The head suspension according to claim 17, wherein the beam is formed by folding the hollowed portion.   A main body portion having a head slider at one end and an actuator carriage coupled at the other end, a beam that vibrates relatively in the thickness direction of the main body portion, and a plate member that is opposed to the thickness direction of the main body. A head suspension constituting a dynamic vibration absorber of the main body using an air film in a gap between the plate member and the beam.   The head suspension according to claim 19, wherein the plate material is a plate material constituting a bending elastic portion that applies a load force to the head slider.

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