JP3840742B2 - Dynamic pressure gas bearing structure - Google Patents

Description

【００４５】
図４に示されるように、中心角θで規定される軸体１の外周面における各円周位置に対応して空隙関数ｈ（θ）を式（１）に代入することにより、各円周位置での発生圧力を求めた。
【００４６】
（１） 空隙変化率が動圧に与える影響
空隙一定部分の厚みｈ₀ は直径Ｄ₁ に対する比率で０．０００６２５（軸体１の直径Ｄ₁ と軸受体２の内径Ｄ₂ の差が直径Ｄ₁ に対する比率で０．００１２５）、空隙拡大部分の最大深さｄは直径Ｄ₁ に対する比率で０．００１２５とした。また、空隙拡大部分の幅Ｗは中心角で６０°とした。このとき、軸体１と軸受体２を同心に配置し、軸受体２を２０，０００ｒｐｍで回転させたときに発生する最大圧力を数値計算により求めて、空隙変化率が動圧に与える影響を検討した。形状関数ｇ（θ）は図５に示されるように規定し、空隙変化率αのみを変化させた。
【００４７】
その結果、表１に示すように、軸体が真円形状では大気圧との圧力差が０であるのに対し、空隙変化率αが１．０×１０^-4／°以上、１０．０×１０^-4／°未満の範囲内で０．１×１０⁵ Ｐａ以上の圧力上昇が得られることがわかった。したがって、本発明のラジアル軸受構造においては軸体と軸受体を同心に配置しても動圧が発生するため、高速回転においても回転精度の高い軸受構造を提供することができることがわかった。
【００４８】
【表１】

【００４９】
（２） 空隙拡大部分の最大深さが動圧に与える影響
上記（１）において良好な動圧の発生が得られた空隙変化率を有する空隙変化部分を備えたラジアル軸受構造について、空隙拡大部分の最大深さｄのみを変化させて発生圧力を計算した。すなわち、形状関数ｇ（θ）は図６に示されるように規定し、最大深さｄのみを変化させた。
【００５０】
その結果として、直径Ｄ₁ に対する比率で表わされた空隙拡大部分の最大厚みｄに対して、発生する動圧の変化は表２に示される。
【００５１】
【表２】

【００５２】
軸体の直径に対する空隙拡大部分の最大深さの比率ｄ／Ｄ₁ が０．００７以下であれば、少なくとも、高速回転する軸受体を支持するために必要な動圧が得られる。したがって、高速回転において、回転に伴う変動が小さい動圧気体軸受構造が得られることが推察できる。
【００５３】
（３） 空隙拡大部分の幅が動圧に与える影響
図７に示される形状関数ｇ（θ）を有するラジアル軸受構造において、空隙拡大部分の幅Ｗのみを変化させたときの発生圧力を計算した。その結果は表３に示される。なお、ここでは、空隙拡大部分が空隙変化部分のみを含む場合（Ｗｂ＝０）と、底部分を有する場合（Ｗｂ＞０）のそれぞれについて計算を行なった。
【００５４】
【表３】

【００５５】
表３に示されるように、空隙拡大部分の幅Ｗが広くなると、大気圧との圧力差が大きくなるため、高速回転においてより変動が小さい動圧気体軸受構造が得られることが推察できる。
【００５６】
（４） 空隙一定部分の厚みが動圧に与える影響
図８に示される形状関数ｇ（θ）を有するラジアル軸受構造において、空隙一定部分の厚みｈ₀ （軸体の直径Ｄ₁ と軸受体の内径Ｄ₂ の差２ｈ₀ ）のみを変化させた場合の発生圧力を計算した。その結果は表４に示される。
【００５７】
【表４】

【００５８】
表４に示されるように、直径差２ｈ₀ が直径Ｄ₁ に対する比率で０．００２５以下であれば、動圧の発生が得られる。したがって、直径差がその範囲内であれば、高速回転中で変動が少ない、回転精度の高い動圧気体軸受構造が得られることが推察できる。
【００５９】
実施例２
以上のシミュレーション結果に基づいて、高速回転時に安定となる良好な動圧発生の結果が得られたものについて、実際にラジアル軸受構造を作製し、発生動圧による浮上回転数の測定を行なった。空隙変化率αは３．０×１０^-4／ｄｅｇ、空隙拡大部分の数（加工箇所の数）は３、深さｄ／Ｄ₁ は０．００１２５、空隙拡大部分（加工箇所）の幅Ｗは１０°とした。形状関数ｇ（θ）は図１０に示されるものを採用した。
【００６０】
外径１０ｍｍの軸体を用いて、直径差２ｈ₀ ／Ｄ₁ が０．０００６２５、軸体１の外周面と軸受体２の内周面とが対向する、軸体１の長手方向に沿った長さ（円筒状の空隙の長さ）が２５ｍｍとなるように軸受構造を構成した。
【００６１】
図１０の形状関数を備えた軸体の加工方法としては、図１２に示されるように円環状の砥石６の内周面を研削面６１として用いて、砥石６を矢印Ｐで示される方向に回転させながら、軸体１の外周面を破線で示されるように研削加工することにより行なわれた。
【００６２】
図１３に示すように、軸体１の外周面と軸受体２の内周面とが対向する部分の長さ、すなわち空隙３の長さ（２５ｍｍ）に対して加工部（空隙拡大部分）の長さ、すなわち空隙変化部分３１の長さが１２．５ｍｍ（空隙の長さに対して５０％）となるように軸受構造を構成した。空隙変化部分３１の数や位置は図１３の（Ａ）〜（Ｅ）に示すように５種類とした。
【００６３】
上記のようにして作製された軸体を固定軸とし、軸受体を可動体とした。図１６に示されるように、軸体１と軸受体２を組込んで、モータ１００を作製した。軸受体２の上部には蓋７が設けられた。軸受体２の下部には磁石８が設けられた。軸受体２の下部に設けられた磁石８に対向して、反発力が作用するように、軸体１の外周面上には同様の磁石８が設けられた。さらに、軸受体２の上部外周面上には環状円板９が設けられた。
【００６４】
このように構成されたモータ１００を高速回転させて、すなわち軸受体２を軸体１に対して５０００ｒｐｍの回転数で回転させた後、回転の駆動源を止めて軸受体２が停止するまで放置した。このとき、軸体１と軸受体２とが接触し始める回転数を測定し、これを浮上回転数とした。
【００６５】
浮上回転数の測定結果は表５に示す。
【００６６】
【表５】

【００６７】
軸体１に対する空隙変化部分３１の加工位置が空隙３の両端部にあるもの（図１３のＤ）と比較すると、空隙３の一方端にあるもの（図１３のＡとＢ）、両端にないもの（図１３のＣとＥ）では、浮上のために発生する動圧が高く、浮上回転数は低いことが示された。
【００６８】
実施例３
実施例１のシミュレーション結果に基づいて、良好な動圧発生の結果が得られたものについて、さらにいくつかのラジアル軸受構造を作製し、発生動圧による浮上回転数の測定を行なった。空隙変化率αは３．０×１０^-4／ｄｅｇ、空隙拡大部分の数（加工箇所の数）は３、深さｄ／Ｄ₁ は０．００１２５とした。
【００６９】
形状関数ｇ（θ）は図９と図１０に示されるものを採用した。軸体の実際の作製にあたっては、横断面が真円の軸体を機械加工して目的とする形状関数を得た。
【００７０】
具体的には、軸体の加工方法は以下のようにして行なわれた。
図９に示される形状関数は、図１１に示すように軸体１をＶブロック５の上において研削砥石４を矢印Ｐで示される方向に回転させながら、軸体１を矢印Ｑで示される方向に移動することによって研削加工を行なうことにより、得られた。
【００７１】
図１０の形状関数を備えた軸体の加工方法としては、図１２に示されるように円環状の砥石６の内周面を研削面６１として用いて、砥石６を矢印Ｐで示される方向に回転させながら、軸体１の外周面を破線で示されるように研削加工することにより行なわれた。
【００７２】
また、図１０の形状関数を備えた軸体の加工は、上記２つの方法とは異なる方法でも行なわれた。すなわち、図１１に示される研削加工を終了した後、軸体１を僅かに回転させて再度、図１１に示される研削加工を実施し、加工部分が所定の幅になるまで繰返し加工を行なった。この加工方法によれば、実際の形状関数は、図１４に示されるように、空隙拡大部分の最大厚みを規定する底部分に凹凸が形成される。この場合、図１４に示すように、空隙拡大部分の最大深さとして平均深さｄ_m を採用すれば、図１０と同様の形状関数を備えた軸体と考えられる。
【００７３】
比較例として、図１５に示されるように空隙変化率αが本発明の範囲よりも大きい形状関数（α＝２０×１０^-4／ｄｅｇ）を有する軸体を用いたものについてもラジアル軸受構造を作製し、浮上回転数の測定を行なった。
【００７４】
軸体１に対して空隙拡大部分（空隙変化部分）を加工するに際して、軸体１の端面から加工を施し、空隙拡大部分の長さ（加工長さ）を変化させて数種類の軸体を得た。
【００７５】
上述のようにして作製された軸体を実施例２と同様に固定軸とし、軸受体を可動体とした。図１６に示されるように、軸体１と軸受体２を組込んで、モータ１００を作製した。軸受体２の上部には蓋７が設けられた。軸受体２の下部には磁石８が設けられた。軸受体２の下部に設けられた磁石８に対向して、反発力が作用するように、軸体１の外周面上には同様の磁石８が設けられた。さらに、軸受体２の上部外周面上には環状円板９が設けられた。
【００７６】
なお、上記のラジアル軸受構造の作製に関して外径が１０ｍｍの軸体を用い、軸体と軸受体が対向する長さ、すなわち円筒状の空隙の長さは２５ｍｍとなるように軸体と軸受体とを組立てた。
【００７７】
このように構成されたモータ１００を高速回転させて、すなわち軸受体２を軸体１に対して５０００ｒｐｍの回転数で回転させた後、回転の駆動源を止めて軸受体２の回転が停止するまで放置した。そして、軸体１と軸受体２が接触し始める回転数を測定し、これを浮上回転数とした。
【００７８】
図９の形状関数を備えた軸体を用いた結果は表６に示される。図１０に示される形状関数を備えた軸体で、図１２に示される方法で加工された軸体を用いた結果は表７、図１１で示される加工方法によって作製された軸体を用いた結果は表８に示される。また、比較例として、図１５の形状関数を備えた軸体を用いた結果は表９に示されている。
【００７９】
なお、表６〜表８において、比較例とされるものは、加工長さ（空隙変化部分の長さ）が本発明の範囲外であるものを示す。
【００８０】
【表６】

【００８１】
【表７】

【００８２】
【表８】

【００８３】
【表９】

【００８４】
実施例４
さらに形状関数の条件を変えて軸体を加工し、実施例３と同様にラジアル軸受構造を作製し、浮上回転数の測定を行なった。
【００８５】
実施例３と同一寸法の軸体、軸受体を用い、空隙変化部分αは３．０×１０^-4／ｄｅｇ、軸体の端面からの加工長さ（空隙変化部分の長さ）は７ｍｍとした。
【００８６】
形状関数ｇ（θ）は図９、図１０、図１４で示されるものを採用した。
実施例３に示す加工方法により軸体を作製した。
【００８７】
比較例として、軸体と軸受体の横断面形状が真円のものについてもラジアル軸受構造を作製し、浮上回転数の測定を行なった。
【００８８】
モータの組立および浮上回転数の測定について、実施例３と同一の方法で行なった。
【００８９】
図９の形状関数を備えた軸体を用いた結果は表１０に示される。図１０に示される形状関数を備えた軸体で、図１２に示される方法で加工された軸体を用いた結果は表１１、図１１で示される加工方法によって作製された軸体を用いた結果は表１２に示される。また、比較例として、真円構造のものを用いた結果は表１３に示されている。
【００９０】
なお、表１０〜表１２において、比較例とされるものは、深さｄ／Ｄ₁ 、幅Ｗ、直径差２ｈ₀ ／Ｄ₁ および空隙拡大部分の数（加工箇所の数）のいずれかが本発明の範囲外であるものを示す。
【００９１】
【表１０】

【００９２】
【表１１】

【００９３】
【表１２】

【００９４】
【表１３】

【００９５】
表６〜表８の加工長さ４ｍｍの比較例および表９〜表１３の比較例での浮上回転数が２０００ｒｐｍ以上となるのは、浮上状態での不安定性の影響によるものと考えられる。
【００９６】
以上の結果から明らかなように、本発明例では、軸体と軸受体とが非接触となる回転数、すなわち浮上回転数が極めて低く、回転によって発生する動圧が大きいことが認められる。この結果、軸体と軸受体との接触が起こるのは極めて低い回転数の領域であり、接触による衝撃は小さく、破損や焼付きは非常に起こり難い。また、非接触で高速で回転している状態での外部からの振動等の負荷に対しても、発生動圧が高いため高い安定性を示し、高い回転精度を得ることができる。
【００９７】
一方、比較例においては十分な動圧の発生が認められず、高い回転数で軸体と軸受体とが接触するため、損傷や焼付きを引き起こす可能性が高いことがわかる。また、高速回転時にも発生動圧が低いため、不安定であり、高い回転精度を得ることができない。
【００９８】
以上に開示された実施例はすべての点で例示であって制限的なものではないと考慮されるべきである。本発明の範囲は、以上の実施例ではなく、特許請求の範囲によって示され、特許請求の範囲と均等の意味および範囲内でのすべての修正や変形を含むものと解釈されるべきである。
【００９９】
【発明の効果】
以上のように、本発明に従って規定された空隙と空隙変化部分を備えることにより、高速回転下で高い回転精度を有するとともに、より高い寿命を有する動圧気体軸受構造を提供することができる。
【図面の簡単な説明】
【図１】軸体の外周面と軸受体の内周面とが対向した領域に設けられる円筒状の空隙を示す図である。
【図２】この発明に従って形成される空隙変化部分の一例を示す図である。
【図３】この発明に従った空隙を定義するために用いられる軸体の外周面と軸受体の内周面を示す横断面図である。
【図４】本発明の動圧気体軸受構造のシミュレーション計算のモデルとして軸体の外周面と軸受体の内周面を示す図である。
【図５】空隙変化率が動圧に与える影響を検討するために用いられた形状関数を示す図である。
【図６】空隙拡大部分の最大深さが動圧に与える影響を検討するために用いられた形状関数を示す図である。
【図７】空隙拡大部分の幅が動圧に与える影響を検討するために用いられた形状関数を示す図である。
【図８】空隙一定部分の厚み（直径差）が動圧に与える影響を検討するために用いられた形状関数を示す図である。
【図９】実施例３と４で用いられた形状関数を示す図である。
【図１０】実施例２〜４で用いられた形状関数を示す図である。
【図１１】軸体の加工方法の一例を示す斜視図である。
【図１２】軸体の加工方法の別の例を示す横断面図である。
【図１３】実施例２で採用された空隙変化部分の種々の位置Ａ〜Ｅを示す軸体と軸受体の概略的な断面を示す縦断面図である。
【図１４】実施例３と４で用いられた形状関数を示す図である。
【図１５】比較例で用いられた形状関数を示す図である。
【図１６】本発明の動圧気体軸受構造が適用されたモータの概略構成を示す図である。
【符号の説明】
１ 軸体
２ 軸受体
３ 空隙
３１ 空隙変化部分
５０ 中心軸[0001]
BACKGROUND OF THE INVENTION
The present invention relates generally to a dynamic pressure gas bearing structure, and more specifically to a dynamic pressure gas bearing structure that supports a rotating body that rotates at a high speed.
[0002]
[Background Art and Problems to be Solved by the Invention]
In recent years, with the increase in storage capacity and the reduction in access time of a rotational drive unit of a magnetic recording device, for example, a hard disk driver (hereinafter referred to as “HDD”), the spindle motor for driving an HDD has a high rotational speed corresponding thereto. Therefore, high rotational accuracy has been demanded. In order to rotate a precision motor that requires such a high rotational speed and high rotational accuracy at a higher speed, it has been proposed to use an air bearing (dynamic pressure gas bearing) as a rotational drive unit. In the rotary drive unit using this air bearing, when the rotating body rotates, air is forcibly introduced into at least the gap between the radial gas bearing body and the rotating body. Thereby, the air pressure in the gap is increased, and the rotating body rotates at a high speed via the air bearing. Thus, the use of the air bearing is expected to maintain the rotational accuracy even during high-speed rotation.
[0003]
In the radial type gas bearing as described above, a wedge-shaped gap is formed by the eccentricity of the shaft body in the bearing body, as shown in, for example, Junichi Togai, “Gas Bearing” Kyoritsu Shuppan (1984). The Pressure is generated because air is compressed as it passes through the wedge-shaped gap. As a result, the shaft body and the bearing body can be supported in a non-contact manner.
[0004]
However, according to Satoshi Mori, “About the Whirl of Gas Bearing”, pp. 481 to 488 and “Lubrication”, Vol. 20, No. 7 (1975), the cylindrical journal bearing is not suitable for supporting a vertical shaft. When placed under load, an instability phenomenon called “Woile” is observed. This phenomenon swings around the bearing so that the shaft is pressed against the bearing surface by centrifugal force at any rotational speed. In the cylindrical journal bearing, the bearing center and the rotation center are deviated by a static load, so that pressure is generated at one place to bring about stable rotation. However, when a cylindrical journal bearing is used in a vertical structure, that is, a structure that supports a vertical shaft, the bearing is placed in a no-load state, so that the pressure generation point changes due to disturbance and the rotation becomes unstable. .
[0005]
When applying such a dynamic pressure gas bearing to the HDD, the positional accuracy of the rotating body is regarded as important, and it is necessary to eliminate the instability factors as described above.
[0006]
In view of this, a dynamic pressure gas bearing structure capable of maintaining high rotational accuracy at high speed rotation is proposed in Japanese Patent Application Laid-Open No. 8-312639.
[0007]
However, in the proposed dynamic pressure gas bearing structure, the shaft body and the bearing body can be stably rotated without contact with each other due to the generated dynamic pressure at high speed rotation. The number of rotations when rotating while in contact with the body is relatively large. Therefore, there has been a problem that there is a high possibility that the shaft body and the bearing body come into contact with each other at a high rotational speed at the time of starting and stopping the rotation and cause damage and seizure. Therefore, the dynamic pressure gas bearing structure proposed in the above publication may have a short life.
[0008]
Accordingly, an object of the present invention is to provide a dynamic pressure gas bearing structure capable of maintaining high rotational accuracy at high speed rotation and increasing the bearing life.
[0009]
[Means for Solving the Problems]
A dynamic pressure gas bearing structure according to the present invention includes a columnar shaft body and a hollow cylindrical bearing body facing the shaft body with a gap in the radial direction. When the shaft body and the bearing body are arranged so that the respective central axes coincide with each other, a substantially cylindrical gap is formed by the outer peripheral surface of the shaft body and the inner peripheral surface of the bearing body. In the cross-sectional shape perpendicular to the shaft of the shaft body and the bearing body, the point where the radial straight line passing through the central axis intersects the contour line corresponding to the outer peripheral surface of the shaft body, and the outer shape corresponding to the inner peripheral surface of the bearing body The gap thickness (h) is defined by the distance between the points intersecting the line.
[0010]
The gap thus defined has at least one gap changing portion. In the gap changing portion, the thickness of the gap changes with respect to the central angle corresponding to the circumference along the outer peripheral surface of the shaft body. The diameter of the shaft body is D₁ When the change amount of the gap thickness is Δh and the change amount of the central angle is Δθ, the change rate α of the gap is expressed by the following equation.
[0011]
α = (Δh / D₁ ) / Δθ [/ °]
The void change portion has a void change rate within the following range.
[0012]
  1.0 × 10^-Four≦ α <10.0 × 10^-Four
  The length of the gap changing portion along the longitudinal direction of the shaft body is 20% or more and 90% or less of the length of the cylindrical gap along the longitudinal direction of the shaft body.The air gap changing portion is formed at one end in the longitudinal direction of the shaft body in the cylindrical air gap.
[0014]
Preferably, the gap includes a gap constant portion having a substantially constant thickness and a gap enlarged portion having a thickness larger than the constant thickness. The void enlarged portion includes a void changing portion.
[0015]
Preferably, the shaft diameter (D₁ ) Thickness of the constant gap portion expressed as a ratio (h)₀ ) And the maximum thickness (h_max) (H)_max-H₀ ) / D₁ Is 0.007 or less.
[0016]
More preferably, the void enlarged portion has a width (W) corresponding to 5 ° or more in central angle.
[0017]
Diameter of shaft (D₁ ) Thickness of the constant gap portion expressed as a ratio (h)₀ / D₁ ) Is preferably 0.00125 or less.
[0018]
By increasing the cross-sectional shape of at least one of the shaft body and the bearing body to a non-circular shape, the gap enlarged portion is formed.
[0019]
It is preferable that three or more void enlarged portions are arranged along the outer peripheral surface of the shaft body.
[0020]
In the present invention, when the shaft body and the bearing body are arranged so that the respective central axes coincide with each other, a gap changing portion having a predetermined gap change rate is formed between the outer peripheral surface of the shaft body and the inner peripheral surface of the bearing body. It is formed between. Therefore, the wedge-shaped gap can be formed without the shaft body being eccentric in the bearing body. Therefore, when air or lubricating oil or the like flows in the gap formed by the shaft body and the bearing body, dynamic pressure is generated due to the wedge-type effect.
[0021]
Specifically, as shown in detail in part A of FIG. 3, in the expanded portion of the gap changing portion, the flow path of the fluid expands, and thus the density of streamlines per unit cross-sectional area decreases. As a result, a negative pressure portion is formed. On the other hand, in the reduced portion of the gap changing portion, the streamline density per unit cross-sectional area increases, so the pressure rises. The radial load is supported by the pressure thus generated.
[0022]
In order to perform such a pressure change efficiently, the air gap change rate α is 1.0 × 10^-Four/ ° or more, 10.0 × 10^-FourMust be within the range of less than / °. This is because when the gap change rate is out of the above range, the shape effect of the gap change portion is not sufficiently obtained, and the pressure increase due to the entrainment of the fluid having viscosity becomes insufficient.
[0023]
For the stability in a state where the shaft body and the bearing body are rotating in a non-contact state by the generated dynamic pressure, the gap changing portion having the above-described shape is required. In the present invention, as a result of paying attention to the process from the rotation state to the rotation in the non-contact state (referred to as the floating state), a bearing structure having a longer life can be obtained. it can.
[0024]
When the rotation of the shaft body or the bearing body starts, the shaft body and the bearing body rotate relatively while being in contact with each other. When the rotational speed increases, the generated dynamic pressure increases. When the rotational speed reaches a certain rotational speed, the shaft body and the bearing body are brought into a non-contact state, that is, a floating state. The rotation speed at this time is referred to as a floating rotation speed. This means that the shaft body or the bearing body is rotating while the shaft body and the bearing body are in contact with each other until the floating rotational speed is reached.
[0025]
In consideration of the influence of the contact on the shaft body and the bearing body, it is preferable that the rotational speed in the contact state is small. In other words, the smaller the levitation speed, the less likely damage will occur due to rotation stoppage.
[0026]
This ascent speed decreases when the dynamic pressure required for ascent is large. In addition to the wedge effect, the dynamic pressure required for levitation is more prominent as the gap is smaller. Therefore, the air gap changing portion is a portion in which the air gap is widened, and thus works in a direction to reduce the dynamic pressure.
[0027]
In order to effectively generate the dynamic pressure for levitation, the length of the gap changing portion along the longitudinal direction of the shaft body is the length of the cylindrical gap formed between the shaft body and the bearing body. If it is 90% or less, dynamic pressure is effectively generated at a location where there is no gap change portion, and the flying speed is reduced. Further, when the length of the gap change portion is less than 20% of the length of the cylindrical gap, the wedge effect is reduced, and the shape effect of the gap change portion described above cannot be obtained sufficiently, and the floating rotation state is stable. I can't get sex.
[0028]
The cylindrical gap portion means a cylindrical gap 3 formed by the outer peripheral surface of the shaft body 1 and the inner peripheral surface of the bearing body 2 shown in FIG. As shown in FIG. 2, the length of the cylindrical gap 3 is the length along the longitudinal direction of the shaft body 1 at the portion where the outer peripheral surface of the shaft body 1 and the inner peripheral surface of the bearing body 2 face each other. Means. The gap changing portion 31 is formed in a part of the cylindrical gap 3. The length of the gap changing portion 31 is defined as the length along the longitudinal direction of the shaft body 1. The ends of the cylindrical gap 3 mean one end and the other end along the longitudinal direction of the shaft body 1 at a portion where the outer peripheral surface of the shaft body 1 and the inner peripheral surface of the bearing body 2 face each other.
[0030]
The gap changing portion tends to reduce the dynamic pressure for rising as described above, but when the gap changing portion exists at the end of the cylindrical gap, the dynamic pressure is more significantly reduced. When a gap changing portion is present at the end of the cylindrical gap, the fluid inside the gap changing portion and the fluid outside the bearing portion are communicated. That is, the fluid inside the gap changing portion can enter and leave the space outside the bearing structure. This means that the pressure (dynamic pressure) generated in the gap is released to the external space through the gap changing portion. As a result, the generated dynamic pressure tends to decrease and the levitation rotational speed tends to increase.
[0031]
Thickness of constant gap (h₀ ) And the maximum thickness (h_max) (H)_max-H₀ ), That is, the maximum depth (d) of the void enlarged portion is the diameter of the shaft (D₁ ) To 0.00025 or more and 0.007 or less. When the maximum depth of the void enlarged portion exceeds the upper limit value, the shape effect of the wall surface of the void enlarged portion does not act, and the dynamic pressure action due to the wedge effect cannot be obtained. Further, the lower limit value of the maximum depth (d) of the gap enlarged portion depends on required accuracy in implementation.
[0032]
The void-enlarged portion preferably has a width (W) corresponding to 5 ° or more in central angle. This is considered to be because if the width (W) is less than 5 °, the fluid flow does not sufficiently develop with respect to the change in the shape of the air gap, and the desired effect cannot be obtained. The upper limit of the width (W) is determined by the number of gap enlarged portions arranged on the outer peripheral surface of the shaft body, but the ratio of the gap enlarged portion to the circumference of the cross-sectional shape of the shaft body is 2 minutes. It is preferable that it is 1 or less. This is thought to be due to the fact that the volume of fluid flowing between the shaft body and the bearing body increases when the thickened gap enlargement part increases with respect to the constant gap part, so the efficiency of dynamic pressure generation decreases. It is done.
[0033]
Thickness of constant gap (h₀ ) Is the diameter of the shaft (D₁ ) Is preferably 0.00125 or less. In other words, the diameter of the shaft (D₁ ) And the inner diameter of the bearing body (D₂ ) (D)₂ -D₁ ) Is the diameter of the shaft (D₁ ) Is preferably 0.0025 or less. The reason for this is that, similar to the action due to the maximum depth of the gap expansion portion, if the thickness of the gap constant portion increases, the shape effect of the gap does not work sufficiently, and the dynamic pressure generation effect due to the wedge-shaped gap does not function. It is done.
[0034]
The gap enlarged portion is obtained by making at least one of the outer peripheral surface of the shaft body and the inner peripheral surface of the bearing body into a non-circular shape in a cross-sectional shape. Note that the cross-sectional shapes of the shaft body and the bearing body can be determined by the shape of the gap defined above.
[0035]
When the present invention is applied to an actual bearing structure, it is necessary to support a load in the radial direction, so that it is necessary to achieve a mechanical balance in at least two places. Therefore, specifically, it is necessary to arrange | position two above-mentioned space | gap enlarged parts along the outer peripheral surface of a shaft body. However, if only two gap expansion portions are provided, there is a risk that mechanical fluctuations easily occur due to disturbance in the direction orthogonal to the balance direction. Therefore, more preferably, a structure that achieves a mechanical balance at three or more locations, that is, three or more locations along the outer peripheral surface of the shaft body are provided with the gap expansion portion.
[0036]
In order to obtain a pressure increase more efficiently, it is preferable that the gap change rate of the expanded portion and the reduced portion is different in the gap changed portion. Further, it is preferable to gradually reduce the pressure in the expanded portion accompanied by the pressure drop to suppress the flow loss as much as possible. In the reduced portion, it is considered effective to suppress the loss of fluid energy due to the friction of the wall surface by rapidly increasing the pressure. Even when the gap change rates of the expanded portion and the reduced portion are made different as described above, the gap change rate needs to be within the above-described range.
[0037]
Further, the dynamic pressure gas bearing structure of the present invention can obtain excellent rotational accuracy even when applied to a horizontal structure, that is, a bearing structure for supporting a horizontal shaft. Even in a state where the shaft body is eccentric in the bearing body due to the static load in the horizontal structure, pressure is generated by the gap changing portion of the present invention in addition to the wedge-shaped gap formed between the shaft body and the bearing body due to the eccentricity. . Therefore, also in the horizontal structure, it is considered that the rotational stability is increased by the same effect as in the vertical structure.
[0038]
【Example】
Example 1
The inventors first logically simulated the influence of the shape of the gap formed between the shaft body and the bearing body on the generated dynamic pressure by numerical analysis.
[0039]
FIG. 3 is a cross-sectional view showing the outer peripheral surface of the shaft body and the inner peripheral surface of the bearing body used to define the shape of the air gap in the dynamic pressure gas bearing structure of the present invention. As shown in FIG. 3, a gap 3 is formed between the outer peripheral surface of the shaft body 1 and the inner peripheral surface of the bearing body 2. Shaft body 1 has diameter D₁ Have Bearing body 2 has an inner diameter D₂ Have The thickness of the gap 3 is represented by h.
[0040]
As shown in detail in part A of FIG. 3, an enlarged gap portion and a fixed gap portion are defined. The void enlarged portion includes a void changing portion. The void enlarged portion has a width W (represented by a central angle corresponding to the circumference which is the cross-sectional shape of the shaft body). The gap enlarged portion has a base width Wb. The gap changing portion has an expanded portion and a reduced portion. The thickness of the constant gap is h₀ It is represented by The maximum thickness of the gap expansion part is h_maxIt is represented by Thickness h of constant gap₀ Is the diameter D of the shaft₁ And inner diameter D of bearing body₂ Is one half of the difference (diameter difference). The maximum depth d of the void enlarged portion is the maximum thickness h of the void enlarged portion._maxAnd the thickness h of the gap₀ It is expressed by the difference between The air gap change rate is represented by a slope α. Note that the shape of the gap defined as described above is determined when the shaft body 1 and the bearing body 2 are arranged so that the central shaft 50 coincides.
[0041]
FIG. 4 is a cross-sectional view showing the outer peripheral surface of the shaft body 1 and the inner peripheral surface of the bearing body 2 shown in order to define the shape function and the gap function of the radial bearing structure as shown in FIG. The gap 3 shown in FIG. 4 is defined by the gap function h (θ). The shape of the outer peripheral surface of the shaft body 1 is defined by a shape function g (θ). In this way, the dynamic pressure generated during rotation by changing the shape of the air gap defined by the function was obtained by numerical calculation.
[0042]
The numerical calculation was performed under the following assumptions.
(I) The fluid (here, air) is an incompressible fluid, and the flow is laminar.
[0043]
(Ii) The fluid satisfies the following equation with respect to the gap function h (θ) in the circumferential direction.
[0044]
[Expression 1]

[0045]
As shown in FIG. 4, by substituting the gap function h (θ) into the equation (1) corresponding to each circumferential position on the outer circumferential surface of the shaft body 1 defined by the central angle θ, The generated pressure at the position was determined.
[0046]
(1) Effect of air gap change rate on dynamic pressure
Thickness h of constant gap₀ Is the diameter D₁ To 0.000625 (diameter D of shaft 1₁ And the inner diameter D of the bearing body 2₂ The difference in diameter D₁ 0.00125), and the maximum depth d of the void enlarged portion is the diameter D₁ The ratio was 0.00125. In addition, the width W of the enlarged gap portion was set to 60 ° in the central angle. At this time, the shaft body 1 and the bearing body 2 are arranged concentrically, the maximum pressure generated when the bearing body 2 is rotated at 20,000 rpm is obtained by numerical calculation, and the influence of the air gap change rate on the dynamic pressure is determined. investigated. The shape function g (θ) was defined as shown in FIG. 5, and only the air gap change rate α was changed.
[0047]
As a result, as shown in Table 1, when the shaft body is a perfect circle, the pressure difference from the atmospheric pressure is 0, whereas the void change rate α is 1.0 × 10^-Four/ ° or more, 10.0 × 10^-Four0.1 × 10 within the range of less than / °^Five It was found that a pressure increase of Pa or more was obtained. Therefore, in the radial bearing structure of the present invention, dynamic pressure is generated even if the shaft body and the bearing body are arranged concentrically, and it has been found that a bearing structure with high rotational accuracy can be provided even at high speed rotation.
[0048]
[Table 1]

[0049]
(2) Effect of the maximum depth of the void expansion part on the dynamic pressure
With respect to the radial bearing structure including the gap changing portion having the gap change rate at which good dynamic pressure was generated in the above (1), the generated pressure was calculated by changing only the maximum depth d of the gap expanding portion. That is, the shape function g (θ) is defined as shown in FIG. 6, and only the maximum depth d is changed.
[0050]
As a result, the diameter D₁ Table 2 shows the change in the dynamic pressure generated with respect to the maximum thickness d of the gap enlarged portion expressed as a ratio to the above.
[0051]
[Table 2]

[0052]
Ratio of maximum depth of void expansion portion to shaft body diameter d / D₁ If it is 0.007 or less, at least the dynamic pressure required to support the bearing body rotating at high speed can be obtained. Therefore, it can be inferred that a dynamic pressure gas bearing structure with small fluctuations associated with rotation can be obtained at high speed rotation.
[0053]
(3) Effect of the width of the gap expansion part on the dynamic pressure
In the radial bearing structure having the shape function g (θ) shown in FIG. 7, the generated pressure was calculated when only the width W of the gap enlarged portion was changed. The results are shown in Table 3. Here, the calculation was performed for each of the case where the gap enlarged portion includes only the gap changing portion (Wb = 0) and the case where the gap enlarged portion has a bottom portion (Wb> 0).
[0054]
[Table 3]

[0055]
As shown in Table 3, it can be inferred that, when the width W of the gap enlarged portion is increased, the pressure difference from the atmospheric pressure is increased, so that a dynamic pressure gas bearing structure with less fluctuation at high speed rotation can be obtained.
[0056]
(4) Influence of thickness of constant gap on dynamic pressure
In the radial bearing structure having the shape function g (θ) shown in FIG.₀ (Shaft diameter D₁ And inner diameter D of bearing body₂ Difference of 2h₀ ), The pressure generated when only the pressure was changed was calculated. The results are shown in Table 4.
[0057]
[Table 4]

[0058]
As shown in Table 4, the diameter difference is 2h.₀ Is diameter D₁ If the ratio is 0.0025 or less, the generation of dynamic pressure can be obtained. Therefore, if the diameter difference is within the range, it can be inferred that a dynamic pressure gas bearing structure with high rotational accuracy and little fluctuation during high-speed rotation can be obtained.
[0059]
Example 2
Based on the above simulation results, a radial bearing structure was actually fabricated for the one that obtained the result of good dynamic pressure generation that became stable during high-speed rotation, and the flying rotation speed was measured by the generated dynamic pressure. Void change rate α is 3.0 × 10^-Four/ Deg, the number of void enlarged portions (number of machining points) is 3, depth d / D₁Was 0.00125, and the width W of the void enlarged portion (processed portion) was 10 °. The shape function g (θ) shown in FIG. 10 was adopted.
[0060]
Using a shaft body with an outer diameter of 10 mm, the diameter difference is 2 h.₀/ D₁Is 0.000625, and the length along the longitudinal direction of the shaft body 1 (the length of the cylindrical gap) in which the outer peripheral surface of the shaft body 1 and the inner peripheral surface of the bearing body 2 face each other is 25 mm. A bearing structure was constructed.
[0061]
As a processing method of the shaft body having the shape function of FIG. 10, the inner peripheral surface of the annular grindstone 6 is used as the grinding surface 61 as shown in FIG. 12, and the grindstone 6 is moved in the direction indicated by the arrow P. The rotation was performed by grinding the outer peripheral surface of the shaft body 1 as indicated by the broken line while rotating.
[0062]
As shown in FIG. 13, the length of the portion where the outer peripheral surface of the shaft body 1 and the inner peripheral surface of the bearing body 2 face each other, that is, the length of the gap 3 (25 mm), The bearing structure was configured such that the length, that is, the length of the gap changing portion 31 was 12.5 mm (50% with respect to the length of the gap). As shown in FIGS. 13A to 13E, the number and positions of the gap changing portions 31 are five types.
[0063]
The shaft body manufactured as described above was a fixed shaft, and the bearing body was a movable body. As shown in FIG. 16, the shaft body 1 and the bearing body 2 were assembled to produce the motor 100. A lid 7 was provided on the upper part of the bearing body 2. A magnet 8 is provided below the bearing body 2. A similar magnet 8 is provided on the outer peripheral surface of the shaft body 1 so that a repulsive force acts against the magnet 8 provided in the lower part of the bearing body 2. Further, an annular disk 9 was provided on the upper outer peripheral surface of the bearing body 2.
[0064]
The motor 100 thus configured is rotated at a high speed, that is, after the bearing body 2 is rotated at a rotational speed of 5000 rpm with respect to the shaft body 1, the rotation drive source is stopped and the bearing body 2 is left until it stops. did. At this time, the rotational speed at which the shaft body 1 and the bearing body 2 started to contact each other was measured, and this was defined as the floating rotational speed.
[0065]
Table 5 shows the measurement results of the flying speed.
[0066]
[Table 5]

[0067]
Compared to the processing position of the gap changing portion 31 with respect to the shaft body 1 at both ends of the gap 3 (D in FIG. 13), the one at one end of the gap 3 (A and B in FIG. 13), not at both ends The ones (C and E in FIG. 13) showed that the dynamic pressure generated for levitation was high and the levitation speed was low.
[0068]
Example 3
Based on the simulation results of Example 1, several radial bearing structures were produced for those for which good dynamic pressure generation results were obtained, and the levitation rotational speed was measured by the generated dynamic pressure. Void change rate α is 3.0 × 10^-Four/ Deg, the number of void enlarged portions (number of machining points) is 3, depth d / D₁Was 0.00125.
[0069]
The shape function g (θ) shown in FIGS. 9 and 10 was adopted. In the actual production of the shaft body, a shaft body having a perfectly circular cross section was machined to obtain a desired shape function.
[0070]
Specifically, the shaft body was processed as follows.
The shape function shown in FIG. 9 is the direction indicated by the arrow Q while rotating the grinding wheel 4 in the direction indicated by the arrow P on the V block 5 as shown in FIG. It was obtained by carrying out grinding by moving to.
[0071]
As a processing method of the shaft body having the shape function of FIG. 10, the inner peripheral surface of the annular grindstone 6 is used as the grinding surface 61 as shown in FIG. 12, and the grindstone 6 is moved in the direction indicated by the arrow P. The rotation was performed by grinding the outer peripheral surface of the shaft body 1 as indicated by the broken line while rotating.
[0072]
Further, the shaft body having the shape function of FIG. 10 was processed by a method different from the above two methods. That is, after the grinding process shown in FIG. 11 is completed, the shaft body 1 is slightly rotated and the grinding process shown in FIG. 11 is performed again, and the machining process is repeated until the processed part has a predetermined width. . According to this processing method, in the actual shape function, as shown in FIG. 14, irregularities are formed in the bottom portion that defines the maximum thickness of the void enlarged portion. In this case, as shown in FIG. 14, the average depth d is defined as the maximum depth of the gap enlarged portion._mIs considered to be a shaft body having the same shape function as in FIG.
[0073]
As a comparative example, as shown in FIG. 15, a shape function (α = 20 × 10 × 10) in which the air gap change rate α is larger than the range of the present invention.^-FourA radial bearing structure was also prepared for a shaft body having a / deg), and the flying speed was measured.
[0074]
When machining the gap enlarged portion (gap changing portion) for the shaft body 1, machining is performed from the end face of the shaft body 1, and the length (working length) of the gap enlarged portion is changed to obtain several types of shaft bodies. It was.
[0075]
The shaft body manufactured as described above was a fixed shaft in the same manner as in Example 2, and the bearing body was a movable body. As shown in FIG. 16, the shaft body 1 and the bearing body 2 were assembled to produce the motor 100. A lid 7 was provided on the upper part of the bearing body 2. A magnet 8 is provided below the bearing body 2. A similar magnet 8 is provided on the outer peripheral surface of the shaft body 1 so that a repulsive force acts against the magnet 8 provided in the lower part of the bearing body 2. Further, an annular disk 9 was provided on the upper outer peripheral surface of the bearing body 2.
[0076]
Regarding the production of the radial bearing structure described above, a shaft body having an outer diameter of 10 mm is used, and the length of the shaft and the bearing body, that is, the length of the cylindrical gap is 25 mm. And assembled.
[0077]
The motor 100 thus configured is rotated at a high speed, that is, the bearing body 2 is rotated with respect to the shaft body 1 at a rotational speed of 5000 rpm, and then the rotation drive source is stopped to stop the rotation of the bearing body 2. Left until. And the rotation speed which the shaft body 1 and the bearing body 2 begin to contact was measured, and this was made into the floating rotation speed.
[0078]
The results using the shaft body with the shape function of FIG. The shaft body having the shape function shown in FIG. 10 and the shaft body machined by the method shown in FIG. 12 was used, and the shaft body produced by the machining method shown in Table 7 and FIG. 11 was used. The results are shown in Table 8. As a comparative example, Table 9 shows the results of using a shaft body having the shape function of FIG.
[0079]
In Tables 6 to 8, a comparative example indicates that the processing length (the length of the gap changing portion) is outside the scope of the present invention.
[0080]
[Table 6]

[0081]
[Table 7]

[0082]
[Table 8]

[0083]
[Table 9]

[0084]
Example 4
Further, the shaft body was processed under different conditions of the shape function, a radial bearing structure was produced in the same manner as in Example 3, and the flying speed was measured.
[0085]
A shaft body and a bearing body having the same dimensions as in Example 3 were used, and the air gap changing portion α was 3.0 × 10.^-Four/ Deg, the processing length from the end face of the shaft body (length of the gap changing portion) was 7 mm.
[0086]
The shape function g (θ) shown in FIGS. 9, 10, and 14 was adopted.
A shaft was produced by the processing method shown in Example 3.
[0087]
As a comparative example, a radial bearing structure was also prepared for a shaft and a bearing having a perfect cross-sectional shape, and the flying speed was measured.
[0088]
The assembly of the motor and the measurement of the flying speed were performed in the same manner as in Example 3.
[0089]
The results using the shaft with the shape function of FIG. The shaft body having the shape function shown in FIG. 10 and the shaft body machined by the method shown in FIG. 12 was used, and the shaft body produced by the machining method shown in Table 11 and FIG. 11 was used. The results are shown in Table 12. As a comparative example, Table 13 shows the results of using a perfect circle structure.
[0090]
In Tables 10 to 12, a comparative example is a depth d / D.₁, Width W, diameter difference 2h₀/ D₁And the number of void-enlarged portions (the number of processed portions) is outside the scope of the present invention.
[0091]
[Table 10]

[0092]
[Table 11]

[0093]
[Table 12]

[0094]
[Table 13]

[0095]
The reason why the flying speed in the comparative example with a processing length of 4 mm in Tables 6 to 8 and the comparative example in Tables 9 to 13 is 2000 rpm or more is considered to be due to the influence of instability in the flying state.
[0096]
As is clear from the above results, in the present invention example, it is recognized that the rotational speed at which the shaft body and the bearing body are not in contact, that is, the floating rotational speed is extremely low, and the dynamic pressure generated by the rotation is large. As a result, the contact between the shaft body and the bearing body occurs in an extremely low rotational speed region, the impact due to the contact is small, and breakage and seizure are unlikely to occur. In addition, since the generated dynamic pressure is high even with respect to a load such as external vibration in a state of rotating at high speed without contact, high stability can be obtained and high rotation accuracy can be obtained.
[0097]
On the other hand, in the comparative example, the generation of sufficient dynamic pressure is not recognized, and the shaft body and the bearing body come into contact with each other at a high rotational speed, so that it is highly likely to cause damage and seizure. In addition, since the generated dynamic pressure is low even during high-speed rotation, it is unstable and high rotational accuracy cannot be obtained.
[0098]
The embodiments disclosed above are to be considered in all respects as illustrative and not restrictive. The scope of the present invention is shown not by the above embodiments but by the scope of claims, and should be interpreted to include all modifications and variations within the meaning and scope equivalent to the scope of claims.
[0099]
【The invention's effect】
As described above, by providing the gap and the gap changing portion defined according to the present invention, it is possible to provide a dynamic pressure gas bearing structure having high rotational accuracy under a high speed rotation and having a longer life.
[Brief description of the drawings]
FIG. 1 is a view showing a cylindrical gap provided in a region where an outer peripheral surface of a shaft body and an inner peripheral surface of a bearing body face each other.
FIG. 2 is a diagram showing an example of a gap changing portion formed according to the present invention.
FIG. 3 is a cross-sectional view showing an outer peripheral surface of a shaft body and an inner peripheral surface of a bearing body used for defining a gap according to the present invention.
FIG. 4 is a diagram showing an outer peripheral surface of a shaft body and an inner peripheral surface of a bearing body as models for simulation calculation of the dynamic pressure gas bearing structure of the present invention.
FIG. 5 is a diagram showing a shape function used to examine the influence of the air gap change rate on the dynamic pressure.
FIG. 6 is a diagram showing a shape function used for studying the influence of the maximum depth of the gap expansion portion on the dynamic pressure.
FIG. 7 is a diagram showing a shape function used for studying the influence of the width of the gap enlarged portion on the dynamic pressure.
FIG. 8 is a diagram showing a shape function used to study the influence of the thickness (diameter difference) of a constant gap portion on dynamic pressure.
FIG. 9 is a diagram showing shape functions used in Examples 3 and 4. FIG.
FIG. 10 is a diagram showing shape functions used in Examples 2 to 4.
FIG. 11 is a perspective view showing an example of a shaft body machining method.
FIG. 12 is a cross-sectional view showing another example of a shaft body machining method.
FIG. 13 is a longitudinal sectional view showing a schematic cross section of a shaft body and a bearing body showing various positions A to E of a gap changing portion employed in Example 2;
14 is a diagram showing shape functions used in Examples 3 and 4. FIG.
FIG. 15 is a diagram illustrating a shape function used in a comparative example.
FIG. 16 is a diagram showing a schematic configuration of a motor to which the dynamic pressure gas bearing structure of the present invention is applied.
[Explanation of symbols]
1 shaft body
2 Bearing body
3 gap
31 Void change part
50 central axis

Claims (7)

A cylindrical shaft body, and a hollow cylindrical bearing body facing the shaft body with a gap in the radial direction,
When the shaft body and the bearing body are arranged so that the respective central axes coincide with each other, a substantially cylindrical gap is formed by the outer peripheral surface of the shaft body and the inner peripheral surface of the bearing body,
In a cross-sectional shape perpendicular to the shaft and the shaft of the bearing body, a point where a radial straight line passing through the central axis intersects an outline corresponding to the outer peripheral surface of the shaft body, and an inner circumference of the bearing body The thickness (h) of the gap is defined by the distance between the points intersecting the outline corresponding to the surface,
The gap has at least one gap changing portion in which the thickness of the gap changes with respect to a central angle corresponding to the circumference along the outer peripheral surface of the shaft body,
When the diameter of the shaft body is D ₁ , the change amount of the gap is Δh, and the change amount of the central angle is Δθ, the change rate α of the gap is
α = (Δh / D ₁ ) / Δθ [/ °]
Represented by
The gap changing portion is 1.0 × 10 ⁻⁴ ≦ α <10.0 × 10 ^−4.
With a void change rate in the range of
The length of the gap changing portion along the longitudinal direction of the shaft body, Ri cylindrical length 90% der than 20% or more of the gap along the longitudinal direction of the shaft body,
The dynamic pressure gas bearing structure , wherein the gap changing portion is formed at one end in the longitudinal direction of the shaft body in the cylindrical gap . The air gap, and the gap constant portion having a substantially constant thickness, and a gap enlarged portion having a greater thickness than the predetermined thickness, the gap enlarged portion comprises said gap change portion of claim 1 Dynamic pressure gas bearing structure. Difference (h _max −h ₀ ) / D between the thickness (h ₀ ) of the gap constant portion expressed as a ratio to the diameter (D ₁ ) of the shaft body and the maximum thickness (h _max ) of the gap enlarged portion The dynamic pressure gas bearing structure according to claim 2 , wherein ₁ is 0.007 or less. The gap enlarged portion has a width (W) which corresponds to less than 5 ° with the central angle, the dynamic pressure gas bearing structure according to claim 2 or 3. The shaft having a diameter (D ₁₎ the gap constant part thickness expressed as a percentage of (h ₀ / D ₁₎ is 0.00125 or less, according to any of claims 2 to 4 Dynamic pressure gas bearing structure. The hydrodynamic gas bearing structure according to any one of claims 2 to 5 , wherein the gap enlarged portion is formed by making the cross-sectional shape of at least one of the shaft body and the bearing body into a non-circular shape. . The dynamic pressure gas bearing structure according to any one of claims 2 to 6 , wherein the gap expanding portion is arranged at three or more locations along the outer peripheral surface of the shaft body.

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