JP2004053004A - Dynamic pressure bearing device and manufacturing method - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To improve shock resistance, to reduce A-RRO (an oscillation motion of a shaft), to reduce changes in bearing rigidity for a temperature change, and to suppress current loss. <P>SOLUTION: This dynamic pressure bearing device is made as a shaft body by integrally forming a circular thrust plate 3 having a diameter larger than that of a shaft 2 and the shaft 2 by stainless steel. A sleeve 11 is formed by an aluminum silicon alloy. A linear expansion coefficient of the shaft body is made -5% not less than 17×10<SP>-6</SP>not more than +5%. A linear expansion coefficient of the sleeve 11 is made -5% not less than of 14×10<SP>-6</SP>not more than +5% in a measurement of a range of 0°C-100°C. Shaft direction thickness of the thrust plate 3 is made as 0.1 to 0.7mm. Hardness of the sleeve 11 is made smaller than that of the shaft 2. Dynamic pressure grooves 21 and 22 for radial bearings are provided in parts facing to the shaft 2. A first thrust dynamic pressure groove 23 is provided at a position facing to the thrust plate 3 of a counter plate 13. A second thrust dynamic pressure groove 24 is provided at the sleeve 11. <P>COPYRIGHT: (C)2004,JPO

Description

【００９４】
この数式１において、離心率ｅが０のとき真円となる。なお、軸受け剛性とは、Ｆをロータに作用する力の総和（外力の総和）とし、ｘをロータの変化量とすると、ｋ＝Ｆ／ｘで表されるｋを指し、ここではその単位を〔ｋｇ・ｆ／μｍ〕として示す。
【００９５】
図６に示される２０℃のように温度が低い時は、離心率は比較的余裕があるが、高温では粘度が低下し、軸受け剛性が低くなるので、隙間は２〜３μｍにしなけらえば使用に耐えることはできない。すなわち、図８の６０℃や図９の８０℃の場合のグラフに示されるように、高温となると上ラジアル動圧溝２７（図１では、上側の第１のラジアル動圧溝２１に相当）の部分での離心率ｅが３μｍ程度の隙間のときに０．０１を超えてしまう。一方、隙間が過小の場合は、軸損トルクが増加したり、最悪の場合には、シャフト２とスリーブ１１とが接触して回転不能となる場合も有り得る。ここで軸損トルクとは、軸（シャフト）が回転する際に、油の粘度等によって失われる軸トルクであり、その値は小さい程良い。ここでは、その単位を〔ｇ・ｃｍ〕で示す。
【００９６】
次に、動圧軸受装置の基本動作の一つである温度変化による剛性の変化について説明する。
【００９７】
隙間は、シャフトやスリーブの熱膨張によっても変化する。シャフトとスリーブの材質の組み合わせによる温度変化に対する隙間の変化を図１０に示す。動圧軸受装置は、回転により油に圧力を発生させるが、隙間が小さいほど、また、油の粘度が高いほど、高い圧力を発生させることができる。油の粘度は、低温では高く、高温では低くなる。したがって、シャフトおよびスリーブの材質は、選定時に、シャフトの線膨張係数がスリーブのそれよりも大きくすれば高温にて隙間が減少するので、粘度低下による圧力低下を補うことができる。
【００９８】
現在のスピンドルモータに採用されている材料と本発明で使用されるアルミシリコン合金をそれぞれシャフトとし、スリーブとした場合の組み合わせと、その特性は次のように整理される。まず、シャフトとスリーブとを同材質にすると、隙間は温度によって変化しない。隙間の長さが変化しないということは、シャフトの外径もスリーブの内径も共に大きくなりながら隙間が同じということであり、これは隙間部分の体積が増加することを示す。高温になると油の粘性は低下する。このため、隙間部分の体積の増加と油の粘性低下が生じ、軸受け剛性は悪化する。また、シャフトとスリーブとが同材質であると、シャフトとスリーブとが衝突した際にシャフトに傷がつき軸受け特性が悪化する。シャフトとスリーブとを同材質とすることは、このように欠点が多く採用するのは好ましくない。
【００９９】
シャフトをＳＵＳ−４００系とし、スリーブを真鍮（ＢｓＢｎ）あるいは青銅（Ｂｒｏｎｚｅ）系とする組み合わせは加工性に関しては比較的良好となる。しかし、熱膨張的には高温になると隙間が増加し、油の粘性低下とともに、軸受け剛性は悪化に向かう。特に、１．８インチや２．５インチなどのように小型のＨＤＤに使用されるスピンドルモータの動圧軸受装置は軸受け剛性に余裕が無いため、各動圧溝部の隙間を極めて微小に設計しなければならない。しかし、要求される厳しい公差をクリアすることは困難であり、また、温度（低温）によってはシャフトとスリーブが干渉することが生じ、軸受けとしては機能しなくなる。また、スリーブの材質は、摩耗性に対して十分では無く、表面にニッケル／クロムなどのメッキが必要となる。このため、この組み合わせも問題ありと言える。
【０１００】
シャフトをＳＵＳ−３００系とし、スリーブをＳＵＳ−４００系とする組み合わせは熱膨張だけを考えると好ましい。すなわち、温度上昇に伴って隙間が減少するので、油の粘度低下による剛性の低下を補う効果を有する。しかし、従来の技術で詳述したように、スリーブをＳＵＳ−４００系とすることは、その硬度が高いことから種々の問題が発生する。すなわち、高温になると、隙間が小さくなり、シャフトがスリーブに接触したり、ロックしたりする。また、内径の寸法、真円度、面相度、円筒度などの高精度加工や動圧溝加工が困難となる。内径寸法のバラツキが大きくなると、シャフトとの適正な隙間を得ることができず、測定による選別組み合わせをしなければならない。動圧溝加工においても溝深さの均一性や対称性が悪くなり、軸受け特性に悪い影響を及ぼす。このように、シャフトをＳＵＳ−３００系とし、スリーブをＳＵＳ−４００系とする組み合わせは、性能面、生産性、価格面などで問題を生ずる。
【０１０１】
シャフトをＳＵＳ−３００系とし、スリーブを本発明の第２の一体部材で用いられているアルミシリコン合金とすると、図１０に示す表に示されるように、温度が高くなると、隙間が少しずつ狭くなる。具体的に示せば、その隙間は、温度０℃で、０．２μｍ広くなり、２０℃で変化０、４０℃で０．２μｍ狭くなり、６０℃で０．４μｍ狭くなり、８０℃で０．６μｍ狭くなり、１００℃で０．８μｍ狭くなる。この各値は、油の粘度低下による剛性の低下を補う上で最適な値となる。
【０１０２】
この図１０で示す値は、シャフトの径を３ｍｍとしたものであり、シャフトの径を他の値、たとえば、２．５ｍｍや２ｍｍとすると、シャフト側の膨張による増加量の絶対値が小さくなるため、隙間の変化量も小さくなる。しかし、一般的に、シャフトの径が小さい場合、軸受け剛性との関係では、隙間も小さくする必要があり、元々の隙間に対する隙間の変化量の割合は、シャフトの径が小さい場合もそれほどの変化はない。ただし、厳密に言えば、その割合は変化するので、シャフトをＳＵＳ−３００系とし、スリーブを本発明のアルミシリコン合金とした場合であっても、両者の線膨張係数の差をシャフトの径によって異ならせる必要がある。上述の実施の形態では、両者の線膨張係数の差を、１．５×１０^−６〜４．５×１０^−６程度としているが、シャフトの径を２ｍｍ程度とすると、隙間は２μｍ程度となり、両者の差は７×１０^−６以下は必要とされ、逆に、シャフトの径を４ｍｍ程度とすると、隙間は４μｍ程度となり、両者の差は１×１０^−６以上あれば良いものとなる。
【０１０３】
次に、動圧軸受装置の基本動作の説明の最後として、消費電流と剛性の関係について説明する。
【０１０４】
軸受けの剛性と電流は、相反する条件となる。つまり、軸受けの剛性を大きくするためには、隙間を小さくしたり、油の粘度を高くしなければならない。一方、電流を小さくするため、すなわち損失トルクを小さくするためには、隙間を大きくしたり、油の粘度を低くしなければならない。同じ軸受剛性を維持しながら軸損トルクを小さくさせるためにはシャフトの径を小さくして隙間も小さくすれば良い。しかしながら、スリーブの内径が小さくなり、かつ公差もより厳しくなるので、従来から使用されている材質のように固い部材、たとえば真鍮やＳＵＳ−３００系を従来よりさらに精度良く加工することは極めて困難となる。従来の材質を使用した場合は、シャフト径とスリーブ内径を全数測定し選別組み合わせをする必要が生ずるが、この方法は生産コストが大きく上昇する。しかも、スリーブ内径が小径となるため、その真円度も精度が悪くなって歩留まりも悪くなる。
【０１０５】
以上の考察を図１１および図１２を参照しながら説明する。なお、図１１は、シャフトをＳＵＳ−４００系とし、スリーブを真鍮とした場合を示し、図１２は、本発明の構成、すなわち、シャフト２をＳＵＳ−３０４とし、スリーブ１１を本発明に使用されるアルミシリコン合金とした場合を示す。
【０１０６】
図１１に示すように、シャフトとしてＳＵＳ−４００系を採用し、スリーブとして真鍮を採用した場合、温度が５８℃当たりを越すと、上ラジアル動圧溝２７における離心率ｅが許容限界値である０．０１を越す。また、傾斜角が温度の上昇に伴い急激に大きくなる。ここで傾斜角とは、シャフトのベースに対する傾斜を指し、ベースに対して９０度（垂直）の状態を０度としたものである。
【０１０７】
これに対して、本発明の組み合わせ（シャフト２に相当するシャフト２６をＳＵＳ−３００系とし、スリーブ１１に相当する上ＢＲＧ２７および下ＢＲＧ２８を本発明のアルミシリコン合金としたもの）では、図１２に示すように、温度が２０℃から８０℃の間において、離心率ｅが許容限界値である０．０１内に入ることとなる。
【０１０８】
このように、本実施の形態において使用するアルミシリコン合金は、アルミニウムの基本特性としての加工性、耐食性、軽量、安価、高強度などを悪化させないで、線膨張係数がシャフト２の値よりも小さい値となる。すなわち、アルミニウムに珪素（Ｓｉ）やその他の添加物を入れていくことにより、線膨張係数が低下していく性質を利用し、線膨張係数が０℃〜１００℃の範囲の測定で、１４×１０^−６プラス／マイナス５％以内のアルミシリコン合金としたものである。なお、このアルミシリコン合金の線膨張係数は、ＳＵＳ−３００系とＳＵＳ−４００系の各線膨張係数の略中間の値、すなわち１４×１０^−６となっている。加えて、このアルミシリコン合金は、０℃〜１００℃の温度範囲で、約１３．５〜約１５．３×１０^−６となっており、かつその値は温度が高くなる程大きい値となる。なお、０℃〜１００℃の温度範囲は、ハードディスク用モータとして必要とされる耐熱条件であり、使用温度環境でもある。
【０１０９】
本実施の形態で使用されるアルミシリコン合金（Ｓｉが３０重量％で、Ｃｕその他が２〜３重量％で残りがＡｌ）の他の特性は、次のとおりである。引っ張り強度は４６．６ｋｇｆ／ｍｍ^２、耐力は４０．０ｋｇｆ／ｍｍ^２、ヤング率は９７００ｋｇｆ／ｍｍ^２、硬度は１５０Ｈｖ（ビッカース硬さ）、密度は２．６ｇ／ｃｍ^３である。
【０１１０】
上述の実施の形態では、シャフト２とスラスト板３とをＳＵＳ−３００系の一体部材とし、また、スリーブ１１とハブ１２とを上述のアルミシリコン合金の一体部材としているので、ＲＲＯ（特にＡ−ＲＲＯ）とＮＲＲＯが共に非常に小さなものとなる。これはスリーブとハブとのはめ合いが無くなると共にシャフトとスラスト板とのはめ合いが無くなるためである。これによってスリーブ１１部分とハブ１２部分の直角度は完全に維持される。すなわち、スリーブとハブを一体材質とし、ハブ１２部分のディスク載置部１２ｃやその他の部分を同時加工すれば各部分の垂直度や水平度を高精度なものとすることができる。シャフト２とスラスト板３の一体部材についても同様である。
【０１１１】
また、両一体部材は、共にはめ込み構造ではないので、耐衝撃性が向上すると共に、不均一な応力も働かず、時間的（経時的）にも温度的にも安定した精度を維持することが可能となる。また、この実施の形態では、スリーブ１１部分とハブ１２部分とが一体部材で構成されているので、ハブ１２部分のスリーブ１１部分に対する固定を考慮する必要がなくなり、コア４やコイル巻線５のための収納空間を十分大きくすることができ、電流ロスを小さくすることができる。また、硬度がよりやわらかとなるスリーブ１１側にラジアル軸受用の動圧溝２１，２２を設けているので、従来の真鍮のスリーブと同様にボール転造でスリーブ１１の内部に動圧溝を簡単に、かつひび割れが生じないように形成することができる。
【０１１２】
また、上述の実施の形態の動圧軸受装置やこの動圧軸受装置が使用されるスピンドルモータ１０では、回転部品やシャフト２の垂直度を高精度に維持できるものとなっているので、安定した動圧力が得られ、寿命も長くなる。また、両一体部材は、共にはめ込み構造でないので、耐熱衝撃性も向上する。
【０１１３】
また、スラスト軸受用の動圧溝２３，２４をスラスト板３に対向するカウンタプレート１３とスリーブ１１とに設けているので、スラスト板３という面積が小さくなりがちなものに動圧溝を設置する場合に比べ、動圧溝の設置位置の自由度が増し、より適切な動圧効果が得やすいものとなると共に動圧溝の中心を出しやすいものとなる。
【０１１４】
また、この実施の形態では、その一体部材を線膨張係数が０℃〜１００℃の範囲の測定で、１４×１０^−６プラス／マイナス５％以内のアルミシリコン合金としているので、スリーブ１１部分内に動圧溝を形成しやすくなると共に、シャフト２部分としてＳＵＳ３００系を採用することができる。しかも、ハブ１２の根元部分の厚さ（従来のモータのスリーブとハブとの結合部分におけるハブ側）を、載置されるディスク１４の厚さの１．３倍としているので、ディスク１４との一体回転がスムーズとなると共に、コア４やコイル巻線５の収納空間Ｓを大きくすることができる。収納空間Ｓを大きくすることで、電流ロスを抑えることができる。
【０１１５】
また、シャフト２のスラスト板３部分の軸方向厚さを０．３ｍｍとしているので、すなわち従来に比べ１／３〜１／５としているので、スラスト軸受けの機能を十分満足させつつ、モータの軸方向長さを小さくでき、薄型化が可能になると共に電流ロスが大幅に減少する。
【０１１６】
また、このディスク駆動装置では、Ａ−ＲＲＯが小さくなり、軸受け剛性の変化が小さくなり、しかも電流ロスを抑えることができるので、ディスク１４からの情報の読み取りやディスク１４への情報の書き込みがミスなく行えるものとなると共に電池使用の際の長時間稼働を達成することができる。なお、Ａ−ＲＲＯの他のＲＲＯやＮＲＲＯも非常に小さくなる。
【０１１７】
また、スラスト軸受用の動圧溝２３，２４をスラスト板３部分に対向するカウンタプレート１３とスリーブ１１とに設けているので、スラスト板３という面積が小さくなりがちなものに動圧溝を設置する場合に比べ、スラスト軸受け用の動圧溝の設置位置や設計の自由度が増し、より適切な動圧効果が得やすいものとなる。このため、ディスク駆動装置として、情報の書き込みミスや読み取りミスが無い高品質な装置とすることができる。さらに、このように、このディスク駆動装置は、内部のスピンドルモータの各種のＲＲＯやＮＲＲＯ（これらの中で少なくともＡ−ＲＲＯ）が小さくなり、軸方向長さが小さくなるので、情報の読み取りや書き込みが安定すると共に、モータの薄型化に伴い装置の薄型化が可能となる。
【０１１８】
上述の実施の形態は、本発明の好適な実施の形態の例であるが、本発明の要旨を逸脱しない範囲で種々変更可能である。たとえば、上述の動圧軸受装置を使用したスピンドルモータとしては、図１３に示すスピンドルモータ３０のように、ハブ１２部分の外径の軸方向長さを延伸させ、ディスク１４の搭載数を増すようにしても良い。なお、他の構成は、基本的に図１に示すスピンドルモータ１０と同様となっている。
【０１１９】
また、上述の動圧軸受装置を使用したスピンドルモータとしては、図１４に示すスピンドルモータ４０のように、ハブ１２部分の外形の軸方向長さを延伸させ、その収納空間Ｓにコア４、コイル巻線５、ヨーク１５、磁石１６を配するようにしても良い。このスピンドルモータ４０もディスク１４の搭載数を増すことができる。なお、他の構成は、基本的に図１や図１３に示すスピンドルモータ１０，３０と同様となっている。
【０１２０】
さらに、図１５のスピンドルモータ５０のように、２．５インチ以下のＨＤＤに搭載されるのに好ましい形状としても良い。この小型のスピンドルモータ５０は、ハブ１２部分の上部のカウンタプレート１３付近を少し上方に延伸させて図１５のようなボス形状部５１とし、このボス形状部５１の外径に溝５２を設けて、ここにクランパ１８の内径部をはめ込むことによって、２枚のディスク１４，１４を固定させるものである。なお、２枚のディスク１４，１４の間には、ディスク１４，１４の間隔を形成すると共にその間隔を維持するための円筒形の間隔保持部材５３が設けられている。
【０１２１】
このスピンドルモータ５０は、クランパ１８の厚さを薄くし、かつディスク１４，１４を固定するねじがないのでハブ１２部分の軸方向高さを短くできる。また、従来のものと同一の高さとした場合には、第１のラジアル動圧溝２１を上方に移動することができるので、モーメント剛性を大きくすることが可能となる。また、ハブ１２の根本部分５４（従来のモータのスリーブとハブとの結合部分におけるハブ側）の厚さをディスク１４の厚さと略同一としているため、ディスク１４との一体回転がスムーズになると共にコア４、コイル巻線５のための収納空間Ｓを薄型にも拘わらず大きなものとすることができる。収納空間Ｓが大きくなるため、電流ロスを抑えることができる。
【０１２２】
上述の実施の形態の動圧軸受装置を使用した各スピンドルモータ１０，３０，４０，５０は、軸固定型の動圧軸受装置を採用しているため、いわゆる軸固定型のスピンドルモータとなっているが、このように軸固定型とすることで、相対回転部分（スリーブ１１とシャフト２部分、ハブ１２とスラスト板３部分）の垂直度を高精度に維持できるという利点を有する。すなわち、スピンドルモータでは、相対回転部分が２ヶ所あるが、垂直度を得るには、そららの部分を一体部材とするのが好ましい。このような要請に対して、軸固定型のスピンドルモータの場合、相対回転部分を、スラスト板３とシャフト２との一体部材と、ハブ１２とスリーブ１１の一体部材の両一体部材に分離でき、かつ組み込みが可能となる。これに対し、軸回転型のスピンドルモータの場合、シャフトとスラスト板とを一体部材とすると、ハブ部分を含めたロータ部分（シャフト、スラスト板、ハブが含まれる）を組み込めなくなる。このように、垂直度の精度を上げるためには、軸固定型のスピンドルモータが好ましい。
【０１２３】
また、上述の各スピンドルモータ１０，３０，４０，５０では、回転部品やシャフト２の垂直度を高精度に維持できるものとなっているので、安定した動圧力が得られ、寿命も長くなる。また、両一体部材は、共にはめ込み構造でないので、耐衝撃性や耐熱衝撃性も向上する。このような性質を一部犠牲にしても、従来以上の利点を有するスピンドルモータとすることができる。たとえば、上述の実施の形態では、シャフト２部分とスラスト板３部分を一体部材とすると共にスリーブ１１部分とハブ１２部分を一体部材とした動圧軸受装置として、２つの一体部材を有するスピンドルモータ１０，３０，４０，５０としているが、いずれか一方の一体部材のみを有する動圧軸受装置を備えるスピンドルモータとしても良い。その場合でも従来のスピンドルモータや動圧軸受装置に比べ、Ａ−ＲＲＯ等の面で有利な効果を有するものとなる。また、本発明で示したアルミシリコン合金をスリーブ１１のみに使用し、スリーブ１１とハブ１２とを一体部材とせず、２部品で構成する場合も、各種の軸受け特性は向上する。
【０１２４】
また、上述した実施の形態では、いわゆる軸固定型のスピンドルモータについて説明したが、軸回転型のスピンドルモータのスリーブ部分に、またはスリーブ部分とベース部分を一体化し、その一体部材に、本発明の第２の一体部材の材料となるアルミシリコン合金を使用し、シャフトにＳＵＳ−３００系を使用するようにしても良い。この場合、シャフトとハブとを一体部材とすると、さらに好ましいものとなる。
【０１２５】
また、第２の一体部材の線膨張係数を、０℃〜１００℃で１４×１０^−６プラス／マイナス５％以内としたが、第１の一体部材であるシャフト体の線膨張係数が１７×１０^−６プラス／マイナス５％以内である場合、第２の一体部材の線膨張係数は、０℃〜１００℃の温度範囲で、１１×１０^−６〜１５×１０^−６の範囲であれば、従来に比べ、相当な効果を有するものとなる。また、線膨張係数が０℃〜１００℃の温度範囲で、１７×１０^−６程度のものとしては、ＳＵＳ−３００系が好ましいが、他の金属部材としたり、表面処理によって表面のみこの値とした金属部材を採用しても良い。
【０１２６】
さらに、第２の一体部材を所定のアルミ合金とすることで、かなりの硬度と所定の線膨張係数を有するものとし、その線膨張係数をシャフト２の線膨張係数に対して、０℃〜１００℃の温度範囲で、１×１０^−６〜７×１０^−６だけ小さくすれば、隙間が数ミクロンのものにおける軸受け剛性が安定したものとなる。また、上述の実施の形態では、第２の一体部材は、その線膨張係数が、温度が高くなるほど、その値が高くなる（１０℃当たり、約０．１５〜０．２×１０^−６の割合）ような性質を有するものとしたが、その高くなる割合を他の値としたり、温度が高くなるほどその変化値が大きくなるものとしても良い。また、第２の一体部材は、０℃〜１００℃の範囲で、線膨張係数が一定値（約１１×１０^−６〜１５×１０^−６の範囲の特定値）となるものとしても良い。
【０１２７】
さらに、ハブ１２の根本部分（従来のモータのスリーブとハブとの結合部分におけるハブ側）の厚さを、ディスク１４の厚さと略同一としたり、１．３倍程度としているが、この関係は、載置されるディスク１４の厚さの０．５倍以上で２倍以下の範囲であれば、ディスク１４との一体回転がスムーズになると共にコア４、コイル巻線５のための収納空間Ｓを薄型にも拘わらず大きなものとすることができる。収納空間Ｓを大きくすることで、電流ロスも抑えることができる。
【０１２８】
また、シャフト２のスラスト板３部分の軸方向厚さを、０．１〜０．７ｍｍ、より好ましくは０．１５〜０．３５ｍｍとするのが良い。このような構成とすると、スラスト軸受けの機能を十分満足させつつ、モータの軸方向長さを小さくでき、薄型化が可能になると共に電流ロスが大幅に減少する。また、上述の実施の形態では、第１の一体部材となるシャフト側を０℃〜１００℃の範囲で、線膨張係数が一定となるものにしているが、ここで一定とは、全く変化しないもののみならず、アルミシリコン合金に比べ、その変化が５分の１程度以下のものを含むものとする。
【０１２９】
また、上述の実施の形態では、ＨＤＤ用のスピンドルモータを示したが、光走査（スキャナ）装置用のモータ等、他の装置用のスピンドルモータとしても良い。そのような場合、スラスト板３やカウンタプレート１３のいずれか一方または両者が不要となることがある。また、ハブ１２に相当する部分がなくなる場合もあり得る。また、動圧軸受装置としては、油の代わりに、他の液体を利用したり、空気などの気体を利用するものとしても良い。
【０１３０】
【発明の効果】
本発明では、耐衝撃性を向上させ得、Ａ−ＲＲＯ（いわゆる軸の揺動運動）を小さくでき、温度変化に対する軸受け剛性の変化を小さくでき、しかも電流ロスを押さえることができると共に小型化、薄型化が可能となる動圧軸受装置を得ることができる。また、他の発明では、温度変化に対する軸受け剛性の変化を小さくでき、安定した性能を長期に渡って維持することができる動圧軸受装置を得ることができる。
【０１３１】
また、他の発明では、電流ロスを押さえることができると共にスラスト側の動圧溝の設置位置の自由度が増し、動圧効果を適切なものとすることができる動圧軸受装置を得ることができる。さらに、他の発明では、温度変化に対する軸受け剛性の変化を小さくでき、しかもシャフトがスリーブに衝突してしまうのを防止できる動圧軸受装置を得ることができる。
【０１３２】
さらに、他の発明では、Ａ−ＲＲＯ（いわゆる軸の揺動運動）を小さくできると共に、薄型化への対応を簡単に行うことができる動圧軸受装置の製造方法を得ることができる。すなわち、薄型化しても、ベースに対するシャフトの垂直度を維持できるスピンドルモータを製造することができる。
【図面の簡単な説明】
【図１】本発明の実施の形態に係る動圧軸受装置と、その動圧軸受装置を使用したスピンドルモータの構造を示す断面図である。
【図２】図１の動圧軸受装置を使用したスピンドルモータのカウンタプレートおよびその周辺の拡大図で、第１の一体部材を省略した図である。
【図３】図１の動圧軸受装置を使用したスピンドルモータのスリーブ部分とハブ部分を一体化した一体部材の材料となるアルミシリコン合金を得るための製造方法のステップを示す図である。
【図４】図１の動圧軸受装置を使用したスピンドルモータの動圧溝に使用される油の特性（温度に対する油の粘度の変化）を示すグラフである。
【図５】図１の動圧軸受装置を使用したスピンドルモータや従来のスピンドルモータの各種性能を評価するための評価用の動圧軸受装置を示す図である。
【図６】図４に示す特性の油と図５に示す評価用の動圧軸受装置を使用して、シャフトの外径とスリーブの内径との隙間を０．５μｍ単位で変化させたときの軸受け特性（軸損トルクと離心率）を示すグラフで、温度が２０℃のときのグラフである。
【図７】図４に示す特性の油と図５に示す評価用の動圧軸受装置を使用して、シャフトの外径とスリーブの内径との隙間を０．５μｍ単位で変化させたときの軸受け特性（軸損トルクと離心率）を示すグラフで、温度が４０℃のときのグラフである。
【図８】図４に示す特性の油と図５に示す評価用の動圧軸受装置を使用して、シャフトの外径とスリーブの内径との隙間を０．５μｍ単位で変化させたときの軸受け特性（軸損トルクと離心率）を示すグラフで、温度が６０℃のときのグラフである。
【図９】図４に示す特性の油と図５に示す評価用の動圧軸受装置を使用して、シャフトの外径とスリーブの内径との隙間を０．５μｍ単位で変化させたときの軸受け特性（軸損トルクと離心率）を示すグラフで、温度が８０℃のときのグラフである。
【図１０】図１の動圧軸受装置を使用したスピンドルモータや従来のスピンドルモータのシャフトやスリーブに使用されている材質の組み合わせによる温度変化に対する隙間の変化を示す表である。
【図１１】従来のスピンドルモータに採用されている材質の組み合わせによる温度に対する軸受け特性（軸損トルク、傾斜角、離心率、軸受け剛性）を示すグラフである。
【図１２】図１に示す本実施の形態に係る動圧軸受装置を使用したスピンドルモータに採用されている材質の組み合わせによる温度に対する軸受け特性（軸損トルク、傾斜角、離心率、軸受け剛性）を示すグラフである。
【図１３】本実施の形態に係る動圧軸受装置を使用したスピンドルモータの第１の変形例の構造を示す断面図である。
【図１４】本実施の形態に係る動圧軸受装置を使用したスピンドルモータの第２の変形例の構造を示す断面図である。
【図１５】本実施の形態に係る動圧軸受装置を使用したスピンドルモータの第３の変形例の構造を示す断面図である。
【図１６】従来の軸固定型の動圧軸受装置を使用したスピンドルモータの構造を示す断面図である。
【図１７】従来の軸回転型の動圧軸受装置を使用したスピンドルモータの構造を示す断面図である。
【符号の説明】
１　ベース（ステータの一部）
２　シャフト（第１の一体部材の一部でシャフト体の一部）
３　スラスト板（第１の一体部材の一部でシャフト体の一部）
４　コア（ステータの一部）
５　コア巻線（ステータの一部）
１０　スピンドルモータ
１１　スリーブ（第２の一体部材の一部、ロータの一部）
１２　ハブ（第２の一体部材の一部、ロータの一部）
１２ａ　ディスク載置用の段部
１２ｃ　ディスク載置部
１３　カウンタプレート（ロータの一部）
１４　ディスク
１５　ヨーク（ロータの一部）
１６　磁石（ロータの一部）
２１　第１のラジアル動圧溝（ラジアル軸受用の動圧溝）
２２　第２のラジアル動圧溝（ラジアル軸受用の動圧溝）
２３　第１のスラスト動圧溝
２４　第２のスラスト動圧溝[0001]
TECHNICAL FIELD OF THE INVENTION
The present invention relates to a dynamic bearing device and a method of manufacturing the dynamic bearing device.
[0002]
[Prior art]
The hydrodynamic bearing device is mounted on a hard disk drive mechanism (hereinafter, referred to as HDD), a digital versatile disk mechanism (hereinafter, referred to as DVD), a scanner device, or the like. Used when rotating. Bearing devices used for HDDs, DVDs, and the like are required to have high-speed rotation, high rotation accuracy, low noise, high reliability, and the like.
[0003]
Conventionally, ball bearings have been used as bearing devices, but in recent years, dynamic pressure bearing devices have been used. The dynamic pressure bearing device basically has characteristics that meet the above-mentioned requirements in principle, and is suitable for an HDD or the like. However, in order to exhibit its excellent characteristics, various conditions must be considered in the structure and application to equipment. That is, in the dynamic pressure bearing device, the shaft and the shaft supporting portion are supported in a non-contact state by oil or air pressure generated by relative rotation. At this time, in order to maintain the predetermined pressure, it is necessary to very strictly control the gap, the viscosity of oil and air, and the shape accuracy such as a right angle, etc., according to the theory of fluid dynamics, and also to change the temperature and the time Special measures must also be taken to prevent small changes to changes.
[0004]
Hereinafter, the structure of a spindle motor employing a conventional hydrodynamic bearing device with various ideas will be described. In addition, a spindle motor adopting a moving body bearing device is shown in FIG. 4 of Patent Document 1 or a shaft fixed type in which a shaft is fixed as shown in Patent Document 2, and shown in FIG. 1 and Patent Document 3 of Patent Document 1. There are known two types of shaft rotation types in which such a shaft rotates together with the rotor. First, a spindle motor 100 using a conventional shaft-fixed type dynamic pressure bearing device will be described with reference to FIG.
[0005]
In the spindle motor 100, a shaft (shaft) 102 is press-fitted and fixed to a base 101, and a thrust plate 103 is press-fitted and fixed to a distal end side of the shaft 102. On the outer periphery of the central cylindrical portion of the base 101, there are arranged cores 104 made of a plurality of thin magnetic metal plates, each of which has a central hole fitted into the cylindrical portion. A plurality of salient poles extending in the radial direction are arranged on the core 104 at predetermined intervals in the circumferential direction, and the salient poles are provided with coil windings 105. The base 101, the core 104, and the coil winding 105 constitute a stator of the spindle motor 100.
[0006]
A cylindrical sleeve 111 serving as a bearing is arranged around the shaft 102, and a counter plate 112 is fixed to the sleeve 111 by swaging so as to close a center hole (upper portion in FIG. 16) of the sleeve 111. A cylindrical hub 114 for mounting a disk 113 such as a hard disk is fixed to the outer periphery of the sleeve 111 by press-fitting, shrink fitting, bonding or the like.
[0007]
A cylindrical magnet 116 is fixed to an inner peripheral portion of the cylindrical portion of the hub 114 via a metal cylindrical yoke 115 serving as a magnetic material. The hub 114 is provided with a screw hole through which a screw 117 for attaching and detaching the disk 113 is inserted. Then, after inserting the disk 113 between the clamper 118 and the hub 114, the disk 113 can be fixed to the hub 114 by tightening the screw 117 while inserting the screw 117 into the screw hole. The sleeve 111, the counter plate 112, the hub 114, the yoke 115, and the magnet 116 constitute a rotor of the spindle motor 100.
[0008]
Oil is filled in a slight gap between the shaft 102 (including the thrust plate 103) and the sleeve 111, and the oil is held in a state where the oil does not leak out of the sleeve 111. A radial dynamic pressure groove 121 for restricting the movement of the shaft 102 in the radial direction is provided at one end of the inner peripheral surface of the sleeve 111, and a radial dynamic pressure groove 122 is provided at the other end. Here, the shaft 102, the sleeve 111, and the radial dynamic pressure grooves 121, 122 constitute a first dynamic pressure fluid bearing.
[0009]
A thrust dynamic pressure groove for restricting the movement of the rotor in the thrust direction is provided on a surface of the thrust plate 103 facing the counter plate 112, and a similar thrust is formed on a surface of the thrust plate 103 facing the step portion 123 of the sleeve 111. A dynamic pressure groove is provided. Here, the thrust plate 103, the sleeve 111, the counter plate 112, and two thrust dynamic pressure grooves provided on the thrust plate 103 constitute a second hydrodynamic bearing.
[0010]
Next, based on FIG. 17, the structure of a spindle motor 200 using a conventional shaft rotating type dynamic pressure bearing device will be described.
[0011]
In this spindle motor 200, a cylindrical sleeve 202 is press-fitted and fixed at the center of a base 201, and a counter plate 203 is swaged and fixed to the sleeve 202 so as to close a central hole of the sleeve 202. A core 204 on which a thin magnetic metal plate is laminated is arranged around the outer periphery of the sleeve 202, and a coil winding 205 is applied to a salient pole portion of the core 204. The base 201, the sleeve 202 (including the counter plate 203), the core 204, and the coil winding 205 constitute the stator of the spindle motor 200.
[0012]
A shaft 212 bonded and fixed to the hub 211 is inserted into the sleeve 202, and a thrust plate 213 is press-fitted and fixed to the tip of the shaft 212. A cylindrical magnet 215 is fixed to the hub 211 via a cylindrical yoke 214. The shaft 212 is provided with a threaded recess serving as a female thread portion for threadingly engaging a screw 217 to which a disc 216 such as a versatile disc can be attached.
[0013]
Then, after inserting the disk 216 between the clamper 218 and the hub 211, the screw 217 is inserted into the screw recess, and the disk 216 can be attached to the hub 211 by tightening the screw 217. Here, the hub 211, the shaft 212, the thrust plate 213, the yoke 214, and the magnet 215 constitute a rotor of the spindle motor 200.
[0014]
Oil is filled in a slight gap between the sleeve 202 and the shaft 212, and the oil is kept in a state where the oil does not leak out of the sleeve 202. A pair of radial dynamic pressure grooves 221 and 222 are provided on one end and the other end of the inner peripheral surface of the sleeve 202, respectively. Here, a first dynamic pressure fluid bearing is constituted by the sleeve 202, the shaft 212, and the radial dynamic pressure grooves 221 and 222. Thrust dynamic pressure grooves are provided on both sides of the surface of the thrust plate 213 facing the counter plate 203 and the surface of the sleeve 202 facing the step portion 223. Here, the second hydrodynamic bearing is constituted by the thrust plate 213, the sleeve 202 and the two thrust dynamic pressure grooves.
[0015]
[Patent Document 1] Japanese Patent Application Laid-Open No. 8-335366 (FIGS. 1 and 4)
[Patent Document 2] JP-A-2000-41359
[Patent Document 3] JP-A-2000-3244753
[0016]
[Problems to be solved by the invention]
The fixed spindle motor 100 using the conventional dynamic pressure bearing device and the spindle motor 200 using the shaft rotation type dynamic pressure bearing device have the following five major drawbacks. A first drawback is that the spindle motors 100 and 200 using the conventional hydrodynamic bearing device are vulnerable to impact because the parts that rotate relative to each other are composed of a combination of different members.
[0017]
A second drawback is that in the spindle motors 100 and 200 employing the conventional hydrodynamic bearing device, the verticality between the shafts 102 and 212 and the rotor tends to be insufficient, and the rotational runout at the hubs 114 and 211 parts. Is to increase. The runout includes a repeatable runout (Repeatable Runout: hereinafter referred to as RRO) and a non-repeatable runout (Non-Repeatable Runout: hereinafter referred to as RRO). The types include shaft runout and surface runout, and the conventional spindle motors 100 and 200 have a large shaft runout RRO (hereinafter, referred to as A-RRO). When the A-RRO of the hubs 114 and 211 increases, the RRO of the runout on the surfaces of the disks 113 and 216 increases, and the recording of information on the disks 113 and 216 and the reading of information from the disks 113 and 216 become difficult. An error occurs.
[0018]
In the conventional spindle motors 100 and 200, the reason why the A-RRO becomes large is considered to be as follows. That is, in the case of the spindle motor 100, the fixing portion 100a of the sleeve 111 and the hub 114 is fixed by press-fitting, shrink fitting, bonding, or the like. Inevitably occurs. Although this displacement also occurs at the time of assembling, a stress in the attachment remains in the fixing portion 100a, and a precision displacement occurs due to a change in temperature, a change in diameter, or the like. This deviation in accuracy leads to a deterioration in verticality, which directly leads to a deterioration in A-RRO.
[0019]
In the case of the spindle motor 200, since the diameter of the shaft 212 is small and the contact width between the shaft 212 and the fixed portion 200a of the hub 211 is small, it is extremely difficult to manage the squareness of the fixed portion 200a with high accuracy. The squareness (perpendicularity) cannot be maintained. Further, even if the right angle comes out at the time of assembling, the fixing of the shaft 212 and the hub 211 is performed by press-fitting, bonding, or the like. Stress remains, and the squareness is not maintained due to changes in temperature and time. The state where the right angle cannot be obtained is directly related to the deterioration of A-RRO.
[0020]
In any of the spindle motors 100 and 200, when the shaft 102 and the thrust plate 103 are assembled or the shaft 212 and the thrust plate 213 are assembled, the thrust plates 103 and 213 are slightly inclined with respect to the shafts 102 and 212. I do. This inclination (deterioration of verticality) causes deterioration of A-RRO of the hubs 114 and 211.
[0021]
A third disadvantage is that in the conventional spindle motors 100 and 200, the change in bearing stiffness with respect to the temperature change is large. If the bearing stiffness is too low, the repeatable runout (RRO) and the non-repeatable runout (NRRO) become large and cannot be used as a motor.
[0022]
In the conventional spindle motors 100 and 200, a large change (decrease) in the bearing rigidity with respect to a temperature change is considered to be due to the following points. That is, in the conventional spindle motors 100 and 200, the shafts 102 and 212 are made of stainless steel 400 series (SUS-400 series), for example, SUS430 containing 18% Cr, and the sleeves 111 and 202 are made of brass or stainless steel. The steel is 300 series (SUS-300 series), for example, SUS304 containing 18% of Cr and 8% of Ni. When these materials are employed, the gap between the shaft 102 and the sleeve 111 and the gap between the shaft 212 and the sleeve 202 increase when the temperature rises due to the influence of the difference in linear expansion coefficient, and the dynamic pressure effect decreases. Therefore, the bearing rigidity in the radial direction is reduced. In addition, when the temperature rises, the viscosity of the dynamic pressure oil decreases, and the bearing rigidity further decreases.
[0023]
In the case of SUS430 and SUS304 shown above, the linear expansion coefficient of the harder SUS430 is 10.4 × 10 ^-6 The softer SUS304 has a linear expansion coefficient of 16.4 × 10 ^-6 When the temperature rises, the sleeves 111 and 202 expand more, and the gap between the shafts 102 and 212 becomes wider. The problem of bearing stiffness is caused not only in the radial direction but also in the thrust direction by the same cause.
[0024]
The problem of bearing stiffness can be improved if the material of the sleeves 111 and 202 is the same as the material of the shafts 102 and 212, ie, SUS-400 (there is no influence due to the difference in linear expansion coefficient). 111 and 202 are hardened, and it becomes difficult to perform groove processing, shape processing, and the like on them, thereby greatly reducing productivity and increasing production costs. Further, the change in the viscosity of the oil accompanying the rise in the temperature of the bearing portion is not yet improved, and the problem remains.
[0025]
It should be noted that there is theoretically a possibility that the material of the shafts 102 and 212 is the same SUS-300 as the material of the sleeves 111 and 202. However, when the shafts 102 and 212 are made of the same material as the sleeves 111 and 202 and are made of a soft material, the shafts 102 and 212 are easily damaged, and rotation stability is easily deteriorated. Locking between the sleeves 111 and 202 becomes easy, and this configuration cannot be adopted.
[0026]
Further, the shafts 102 and 212 may be made of SUS-300 system, and the sleeves 111 and 202 may be made of SUS-400 system. However, if the sleeves 111 and 202 are made of SUS-400 system, the difference in thermal expansion is too large, so , The shafts 102 and 212 are locked. That is, since the SUS-400 system has a small degree of thermal expansion, when the temperature becomes high, the gap between the shafts 102 and 212 and the sleeves 111 and 202 becomes zero and the SUS-400 is locked. In addition, sometimes the gaps become small, so that the shafts 102 and 212 hit the rigid SUS-400-based sleeves 111 and 202 and damage the shafts 102 and 212.
[0027]
Further, when the shafts 102 and 212 are made of SUS-300 and the sleeves 111 and 202 are made of SUS-400, the SUS-400 has high hardness. It is difficult to perform fine processing, high precision processing such as high cylindricity, and dynamic pressure groove processing. Variations in the inner diameter result in an inability to obtain an appropriate gap, and require a time-consuming step of sorting and combining by measurement. This poses a problem in terms of productivity and cost. The fact that the dynamic pressure groove processing becomes difficult is directly related to the deterioration of the uniformity and symmetry of the groove depth, and the bearing characteristics are deteriorated. It also has an adverse effect on productivity and price.
[0028]
For these reasons, a combination in which the shafts 102 and 212 are made of SUS-300 and the sleeves 111 and 202 are made of SUS-400 cannot be practically adopted.
[0029]
Further, a fourth disadvantage is that current loss is large in the conventional spindle motors 100 and 200. There are two possible causes for this. One is a structural problem. That is, in order to obtain strength and efficient workability at the fixed portion 100a between the sleeve 111 and the hub 114 of the fixed shaft type spindle motor 100 shown in FIG. 16, the axial length of the fixed portion 100a may be increased. However, the radial thickness of each component must be made sufficiently large, but for that purpose, the space for accommodating the core 104 and the coil winding 105 becomes narrow, and as a result, the electromagnetic characteristics (Kt) become small. . This tendency that the storage space becomes narrow and the electromagnetic characteristics become small also occurs in the spindle motor 200 of the rotary shaft type.
[0030]
Another cause of the large current loss is the problem of the bearing structure. That is, the shaft loss torque of the bearing portion can be reduced without reducing the bearing rigidity by reducing the diameters of the shafts 102 and 212 and reducing the radial gap. However, in the case of the spindle motor 100, when the diameter of the shaft 102 is reduced, it becomes difficult to obtain a coupling force and a squareness at the fixed portion 100b between the shaft 102 and the base 101. In the case of the spindle motor 200, when the diameter of the shaft 212 is reduced, the contact area of the fixed portion 200a between the shaft 212 and the hub 211 is further reduced, and the A-RRO is further deteriorated as described above. In addition, when the radial gap is reduced, the effect of the difference in the coefficient of linear expansion becomes large.
[0031]
One of the causes of the large current loss is the thickness of the thrust plates 103 and 213. At present, in order to obtain the verticality of the thrust plates 103 and 213 and the fixing strength at the fixing portions 100c and 200b, the thrust plates 103 and 213 have to be thickened. Specifically, the thickness of the thrust plates 103 and 213 is set to about 1 to 1.5 mm. For this reason, the current loss is considerably large.
[0032]
As described above, it is extremely difficult to reduce the diameters of the shafts 102 and 212 and to reduce the thickness of the thrust plates 103 and 213. As a result, the current increases and the current loss increases.
[0033]
A fifth disadvantage is that the conventional spindle motors 100 and 200 can easily cope with a 3.5-inch hard disk, but are more compact and thinner for a 2.5-inch or 1.8-inch hard disk. It is difficult to deal with. That is, in the conventional spindle motors 100 and 200, since the relative rotation part is an assembled part obtained by assembling separate members, it is difficult to manage the gap, which is the gap of the bearing part, so that a large gap cannot be avoided. This makes it difficult to reduce the size and thickness. In particular, in consideration of the strength and accuracy of assembling, the thrust plates 103 and 213 and the hubs 114 and 211 have to be thickened, and the axial thickness tends to be large.
[0034]
The conventional hydrodynamic bearing device accounts for a large part of the five disadvantages of the conventional spindle motors 100 and 200. That is, the second problem is the deterioration of the verticality, the third problem is the change in bearing rigidity, the fourth problem is the problem of the shaft structure which is one of the problems of the current loss, and the thickness of the thrust plate. The problem of the thickness of the thrust plate, which is one of the factors that make it difficult to reduce the size and thickness of the fifth disadvantage, is also a problem of the conventional hydrodynamic bearing device itself.
[0035]
The present invention has been made in order to solve the problems of the above-described hydrodynamic bearing device, and can improve impact resistance, can reduce A-RRO (so-called rocking motion of a shaft), and can prevent temperature change. It is an object of the present invention to provide a hydrodynamic bearing device that can reduce a change in bearing stiffness, suppress current loss, and can be reduced in size and thickness. Another object of the present invention is to provide a dynamic pressure bearing device capable of reducing a change in bearing rigidity with respect to a temperature change and maintaining stable performance for a long period of time.
[0036]
Another invention provides a dynamic pressure bearing device that can suppress current loss, increase the degree of freedom in the installation position of a thrust-side dynamic pressure groove, and make the dynamic pressure effect appropriate. The purpose is to: Still another object of the present invention is to provide a hydrodynamic bearing device capable of reducing a change in bearing stiffness with respect to a temperature change and preventing a shaft from colliding with a sleeve.
[0037]
Still another object of the present invention is to provide a method of manufacturing a dynamic pressure bearing device capable of reducing A-RRO (so-called rocking motion of a shaft) and easily responding to a reduction in thickness.
[0038]
[Means for Solving the Problems]
In order to achieve the above-described object, a dynamic pressure bearing device according to the present invention includes a dynamic pressure bearing device including a base, a shaft fixed to the base, and a sleeve rotatably disposed around the shaft. A shaft body is formed by integrally forming a circular thrust plate having a diameter larger than the diameter of the shaft portion and a shaft, which is provided at an end opposite to the base of the shaft and having a diameter larger than that of the shaft portion, to form a shaft body. It is made of an aluminum silicon alloy having silicon as a component, and has a linear expansion coefficient of 17 × 10 ^-6 Within ± 5%, the coefficient of linear expansion of the sleeve is 14 × 10 ^-6 It is within ± 5%, the axial thickness of the thrust plate is 0.15 to 0.35 mm, and the hardness of the sleeve is smaller than the hardness of the shaft. A dynamic pressure groove for a radial bearing is provided at the position where the thrust plate is opposed to the thrust plate at a position facing the thrust plate and covering the thrust plate and closing the center hole of the sleeve. A first thrust dynamic pressure groove for a thrust bearing is provided, and a second thrust dynamic pressure groove is formed on a sleeve portion facing a surface of the thrust plate opposite to the surface facing the first thrust dynamic pressure groove. Provided.
[0039]
In the hydrodynamic bearing device of the present invention, the shaft fixed to the base and the thrust plate are integrated into an integral member, so that the impact resistance is improved and the verticality is easily obtained. For this reason, A-RRO of the sleeve that rotates relative to the shaft can be reduced. In addition, since the respective linear expansion coefficients of the sleeve and the shaft body are set to be larger on the shaft side, a change in bearing rigidity with respect to a temperature change is reduced. In addition, since the sleeve on the rotating side is made of an aluminum silicon alloy having aluminum and silicon, the weight is approximately the same as aluminum, and current loss and inertia are small.
[0040]
Further, since this dynamic pressure bearing device is of a fixed shaft type, it is possible to incorporate the shaft body even if the thrust plate has a larger diameter than the shaft. In addition, since the dynamic pressure groove for the radial bearing is provided on the sleeve side where the hardness becomes softer, the dynamic pressure groove is easily formed inside the sleeve by ball rolling like a conventional brass sleeve, and cracks are not generated. It can be formed so as not to occur, and a protrusion provided in the groove for dynamic pressure (a step which projects to the center side on the surface of the sleeve before forming the dynamic pressure groove is often provided. Portion) prevents the surface of the shaft from being damaged. Furthermore, since the sleeve is made of a softer material than the shaft body, the shaft side is not damaged even when both are in contact when the shaft body is inserted into the sleeve, and stable performance can be obtained over a long period of time. .
[0041]
Further, since the axial thickness of the thrust plate portion is set to 0.15 to 0.35 mm, current loss is reduced, and the size and thickness can be reduced. In addition, since the dynamic pressure grooves for the thrust bearing are provided on the counter plate and the sleeve facing the thrust plate, the dynamic pressure grooves are smaller than those in the case where the area of the thrust plate tends to be small. This increases the degree of freedom in the installation position and design of the device, making it easier to obtain a more appropriate dynamic pressure effect, and at the same time, reducing the cost.
[0042]
A hydrodynamic bearing device according to another aspect of the present invention is a hydrodynamic bearing device having a shaft, a sleeve arranged to be relatively rotatable on the shaft, and a base for fixing either the shaft or the sleeve. The sleeve is formed of an aluminum silicon alloy having aluminum and silicon as components, and the linear expansion coefficient of the shaft is 17 × 10 ^-6 Within ± 5%, the coefficient of linear expansion of the sleeve is 11 × 10 when measured in the range of 0 ° C. to 100 ° C. ^-6 ~ 15 × 10 ^-6 The hardness of the sleeve is smaller than the hardness of the shaft, and a dynamic pressure groove for a radial bearing is provided on a portion of the sleeve which is a softer member and facing the shaft.
[0043]
In the hydrodynamic bearing device according to the present invention, since the respective linear expansion coefficients of the sleeve and the shaft body are set to be larger on the shaft side, the change in the bearing rigidity with respect to the temperature change is reduced. Also, this dynamic pressure bearing device has a dynamic pressure groove for the radial bearing on the sleeve side where the hardness becomes softer, so that the dynamic pressure groove is formed inside the sleeve by ball rolling like a conventional brass sleeve. Can be formed easily and without cracks, and a protrusion provided on the groove for dynamic pressure (a step which protrudes to the center side on the surface of the sleeve before forming the dynamic pressure groove is provided. In many cases, this step prevents the surface of the shaft from being damaged. Furthermore, since the sleeve is made of a material softer than the shaft, the shaft side is not damaged even when both are in contact with each other when the shaft is inserted into the sleeve, and stable performance can be maintained for a long period of time.
[0044]
Further, a dynamic pressure bearing device of another invention is a dynamic pressure bearing device having a shaft, a sleeve disposed relatively rotatable on the shaft, and a base for fixing either the shaft or the sleeve. The sleeve is formed of an aluminum silicon alloy composed of 65 to 69% by weight of Al, 28 to 32% by weight of Si, and 1 to 5% by weight of Cu and the like, and has a linear expansion coefficient of 0 to 100 ° C. 1 × 10 with respect to the coefficient of linear expansion of the shaft ^-6 ~ 7 × 10 ^-6 Only, the hardness of the sleeve is made smaller than the hardness of the shaft, and a dynamic pressure groove for a radial bearing is provided on a portion of the sleeve which is a softer member and facing the shaft.
[0045]
In the hydrodynamic bearing device according to the present invention, since the linear expansion coefficients of the sleeve and the shaft are set to be larger on the shaft side, a change in bearing rigidity with respect to a temperature change is reduced. Since the sleeve is made of aluminum-silicon alloy, which is mostly made of aluminum and silicon, the weight is about the same as or slightly lighter than aluminum. When used, current loss and inertia are reduced. In addition, this dynamic pressure bearing device has a dynamic pressure groove for the radial bearing on the sleeve side where the hardness becomes softer, so that the dynamic pressure groove is formed inside the sleeve by ball rolling like a conventional brass sleeve. Can be formed easily and without cracks, and the projection of the groove for dynamic pressure prevents the surface of the shaft from being damaged. Furthermore, since the sleeve is made of a material softer than the shaft, the shaft side is not damaged even when both are in contact with each other when the shaft is inserted into the sleeve, and stable performance can be maintained for a long period of time.
[0046]
According to another aspect of the present invention, there is provided a hydrodynamic bearing device including a shaft, a sleeve disposed relatively rotatable on the shaft, and a base for fixing either the shaft or the sleeve. A thrust plate is provided at an end opposite to the above, a sleeve is formed of an aluminum silicon alloy having aluminum and silicon as components, and a shaft is composed of 10.5 to 32% by weight of Cr and 4 to 13% by weight of Ni. % Of a ferrous alloy, and a dynamic pressure groove for a radial bearing is provided on an inner surface of the sleeve, which is a portion facing the shaft of the sleeve. The thrust plate is located at a position facing the thrust plate. And a thrust bearing for the thrust bearing at a position facing the thrust plate of the counter plate provided so as to cover the center hole of the sleeve. Of providing thrust dynamic pressure groove, opposite the surface from facing the sleeve portion and the first thrust dynamic pressure groove and opposing surfaces of the thrust plate is provided with a second thrust dynamic pressure groove.
[0047]
Since the hydrodynamic bearing device of the present invention uses an aluminum silicon alloy containing silicon lighter than aluminum as the sleeve, if the sleeve portion is the rotating side, the rotating portion is considerably lighter than stainless steel and slightly lighter than aluminum. . Therefore, the current loss is reduced and the inertia is reduced. Further, when the sleeve is fixed, the weight of the dynamic pressure bearing device is reduced, which can contribute to the weight reduction of the device incorporating the dynamic pressure bearing device. Furthermore, since the dynamic pressure groove for the thrust bearing is provided on the counter plate and the sleeve facing the thrust plate, the dynamic pressure groove is set in comparison with the case where the dynamic pressure groove is installed on the thrust plate where the area tends to be small. This increases the degree of freedom in the installation position and design of the device, making it easier to obtain a more appropriate dynamic pressure effect, and at the same time, reducing the cost.
[0048]
Further, a dynamic pressure bearing device of another invention is a dynamic pressure bearing device having a shaft, a sleeve disposed relatively rotatable on the shaft, and a base for fixing either the shaft or the sleeve. At least in the temperature range of 0 ° C. to 100 ° C., the coefficient of linear expansion increases as the temperature increases, and is formed of an aluminum-silicon alloy having aluminum and silicon as components. The coefficient of linear expansion of the shaft is larger than the coefficient of linear expansion of the sleeve. The hardness of the sleeve is made smaller than the hardness of the shaft while being kept constant in the temperature range, and a dynamic pressure groove for a radial bearing is provided on a portion of the sleeve which is a softer member and facing the shaft.
[0049]
The dynamic pressure bearing device according to the present invention is characterized in that the sleeve is made of an aluminum-silicon alloy having aluminum and silicon as components, the linear expansion coefficient of which increases as the temperature increases at least in a temperature range of 0 to 100 ° C. Since the coefficient is larger than the linear expansion coefficient of the integral member and is constant in the above temperature range, when the temperature increases, the gap between the dynamic pressure grooves in the sleeve becomes narrower, the rigidity of the bearing is maintained, and the The risk of contact with the shaft can be greatly reduced. Here, the term “constant linear expansion coefficient” includes not only the coefficient that does not change at all but also the coefficient whose change is about one-fifth or less as compared with the aluminum silicon alloy.
[0050]
Further, in another invention, in addition to the above-described hydrodynamic bearing device, the shaft is fixed to the base and the sleeve is rotatably arranged around the shaft. Since this dynamic pressure bearing device is of a fixed shaft type, it is possible to incorporate this shaft body even if the shaft and the thrust plate are formed as an integral member and the thrust plate has a larger diameter than the diameter of the shaft. . As described above, according to the present invention, the degree of freedom in design is increased.
[0051]
In addition, in another invention, in addition to the dynamic pressure bearing device of the above-described invention, a disk holding hub that rotates integrally with the sleeve is formed integrally with the sleeve from an aluminum silicon alloy. In this hydrodynamic bearing device, since the sleeve and the hub are integrated into an integral member, the impact resistance of the hub portion is improved, and the verticality of the hub can be easily obtained. For this reason, A-RRO as a bearing device can be made very small. Moreover, miniaturization and thinning are also possible.
[0052]
According to another aspect of the present invention, in addition to the above-described hydrodynamic bearing device, a thrust plate having a diameter larger than the diameter of the shaft is provided at an end of the shaft opposite to the base. The plate and the shaft are integrally formed of stainless steel to form a shaft body, and the axial thickness of the thrust plate portion is 0.1 to 0.7 mm.
[0053]
In the hydrodynamic bearing device of the present invention, the shaft and the thrust plate are integrated to form a shaft body, so that the impact resistance of the shaft body is improved and the verticality of the shaft portion is easily obtained. For this reason, A-RRO as a bearing device can be reduced. Moreover, miniaturization and thinning are also possible. In addition, since the axial thickness of the thrust plate portion is 0.1 to 0.7 mm, the axial length can be further reduced while sufficiently satisfying the function of the thrust bearing, thereby further reducing the size and thickness. It becomes possible. In addition, current loss can be reduced by making the thrust plate thinner.
[0054]
According to another aspect of the invention, in addition to the above-described hydrodynamic bearing device, the rotor integrally rotating with the sleeve is formed integrally with the sleeve from an aluminum silicon alloy. In this hydrodynamic bearing device, since the sleeve and the rotor are integrated and formed as an integral member, the impact resistance of the rotor portion is improved, and the verticality of the rotor can be easily obtained. Therefore, the A-RRO of the rotor can be made very small. Moreover, miniaturization and thinning are also possible.
[0055]
A method of manufacturing a dynamic bearing device according to the present invention includes a method of manufacturing a dynamic bearing device including a base, a shaft fixed to the base, and a sleeve rotatably disposed around the shaft. After the shaft is assembled into the base, the part of the bottom of the shaft and the part of the bottom of the base are cut at the same time. It is cut out so as to have a constant plane, and has a thickness in a completed state.
[0056]
When this manufacturing method is adopted, even if the shaft-fixed type dynamic pressure bearing device has a small thickness in the axial direction, the shaft can be sufficiently perpendicular to the base. That is, it is possible to reduce the A-PRO of the sleeve and to easily cope with the reduction in thickness.
[0057]
[Embodiment of the present invention]
Hereinafter, a hydrodynamic bearing device according to an embodiment of the present invention will be described with reference to the drawings. In this embodiment, a dynamic pressure bearing device used for a spindle motor will be described. For a disk drive equipped with a hydrodynamic bearing device, only the spindle motor and its periphery will be described, and a head portion for reading and writing disk information, a control circuit for controlling the spindle motor and its head portion, etc. The circuit part and other mechanical parts are the same as those of the conventional disk device, and the description thereof will be omitted.
[0058]
The dynamic pressure bearing device according to this embodiment is incorporated in a spindle motor 10. The spindle motor 10 is a spindle motor having a fixed shaft type dynamic pressure bearing device, and is mounted on an HDD. In this spindle motor 10, as shown in FIG. 1, a shaft 2 is press-fitted into a base 1 made of a hard resin material such as a metal such as aluminum or a functional resin, etc. One or more are fixed. A thrust plate 3 is provided integrally with the shaft 2 on the tip side (the upper side in FIG. 1) of the shaft 2 and serves as a thrust plate portion. That is, the shaft 2 and the thrust plate 3 are formed as a single-piece shaft body composed of one part. In the following, this integral member may be referred to as a first integral member instead of the shaft body in relation to a second integral member described later.
[0059]
This first integral member (shaft body) is made of stainless steel (SUS-304) which is an iron-based alloy containing 18% by weight of Cr and 8% by weight of Ni. The first integral member cuts the outer periphery of the rod-shaped SUS-304 to form a cylindrical shaft 2 portion and a circular (specifically flat cylindrical) thrust plate 3 portion. ing. The axial thickness of the thrust plate 3 is set to 0.3 mm, but if it is in the range of 0.1 to 0.7 mm, it is preferable in terms of strength and current loss. In addition, this range becomes more preferable if it is set to 0.15 to 0.35 mm.
[0060]
The linear expansion coefficient of this first integral member is 16.4 × 10 ^-6 Vickers (hardness) is about 196. The first integrated member may be any other SUS-300-based member. That is, an iron-based alloy containing 10.5 to 32% by weight of Cr and 4 to 13% by weight of Ni may be used. However, the linear expansion coefficient is 17 × 10 ^-6 It is preferably within plus / minus 5%. The coefficient of linear expansion is a coefficient of linear expansion relating to a change in length in a solid. When 1 is a length, lo is a length at 0 ° C., and θ is a temperature, the value of (dl / dθ) / lo is expressed as Point. The linear expansion coefficient of the shaft (SUS-304) serving as the first integral member is substantially constant in the range of 0 ° C to 100 ° C. Note that the term “constant” described herein includes not only a state that does not change at all but also a state that the change is about one-fifth or less compared to a second integrated member described later.
[0061]
A cylindrical cylindrical portion 1a is provided at the center of the base 1, and a thin magnetic metal plate whose central hole portion fits into the cylindrical portion 1a is arranged on the outer periphery of the cylindrical portion 1a in a stacked state. , The core 4. The fixing of the core 4 is performed by combining one or more of adhesion, press fitting, caulking, and the like. A plurality of salient poles extending in the radial direction are formed in the core 4 at predetermined intervals in the circumferential direction, and a coil winding 5 is applied to each of the salient poles. The storage space S in which the core 4 and the coil winding 5 are stored can be made wider than that of a conventional spindle motor for the reasons described below. The base 1, the core 4, the coil winding 5, and the like constitute a stator of the spindle motor 10.
[0062]
Around the first integrated member, a second integrated member in which a cylindrical sleeve 11 portion serving as a bearing and a disk-shaped hub 12 portion are integrally formed is arranged. The base 1, the shaft 2 or the shaft body, and the sleeve 11 form a basic component of the hydrodynamic bearing device.
[0063]
The second integrated member is made of an aluminum silicon alloy containing 67 to 68% by weight of Al, 30% by weight of Si, and 2 to 3% by weight of Cu and the like, and subjected to heat treatment. Its coefficient of linear expansion is 14 × 10 ^-6 Of which 14.7 × 10 4 was measured in the range of 40 ° C. to 100 ° C. ^-6 And 16.0 × 10 6 in the measurement in the range of 100 ° C. to 200 ° C. ^-6 And 17.6 × 10 6 in the measurement in the range of 200 ° C. to 300 ° C. ^-6 And 19.0 × 10 3 in the measurement in the range of 300 ° C. to 400 ° C. ^-6 It becomes. During this measurement, in the range of 40 ° C. to 400 ° C., the differential expansion measurement was performed in the range of room temperature to 400 ° C., at a heating rate of 10 ° C. per minute, and in a nitrogen stream.
[0064]
As described above, the first integral member has a coefficient of linear expansion that increases as the temperature increases (in the range of 40 ° C. to 400 ° C., on average, about 0.15 × per 10 ° C.). 10 ^-6 ).
[0065]
In consideration of the measurement error and the slight change of the contained metal ratio, the linear expansion coefficient is 14 × 10 4 in the range of 0 ° C. to 100 ° C. in relation to the first integrated member side. ^-6 It is sufficient if the value is within plus / minus 5% of this value. ^-6 It should be within plus / minus 5%. The material distribution may be an aluminum silicon alloy in the range of 65 to 69% by weight of Al, 28 to 32% by weight of Si, and 1 to 5% by weight of Cu and others. As described above, the linear expansion coefficient of the aluminum silicon alloy was determined to be 14 × 10 ^-6 Can be adjusted also by the content of silicon (Si) so as to be within ± 5%.
[0066]
The second integral member is a newly developed aluminum silicon alloy having the above-described coefficient of linear expansion. Its hardness is softer than that of the first integral member, and its hardness is about 114 to 165 [kg / mm] in Vickers hardness. If the hardness is 80 or more in terms of Vickers hardness, the second integrated member does not dent when the two integrated members collide with each other when the rotation of the second integrated member starts and stops. The hardness varies depending on the method and conditions of the heat treatment, and the value required for the apparatus can be appropriately obtained from this range (Vickers hardness 114 to 165 [kg / mm]). When the extruded member is not subjected to the heat treatment, the Vickers hardness is 114 [kg / mm], and the second integral member obtained by the heat treatment has a Vickers hardness of 165 [kg / mm]. ]. By changing the method of heat treatment, the hardness can be further increased slightly. In this second integral member, the extruded member is formed into a predetermined shape by hot forging, and then heat-treated.
[0067]
A disc-shaped metal counter plate 13 is fixed to the sleeve 11 of the second integral member so as to close one end (upper side in FIG. 1) of the center hole. As shown in FIG. 2, the counter plate 13 is placed on the upper flat portion 11a on the upper end side of the sleeve 11 of the second integral member, and then the sleeve 11 is swaged to fix the counter. This is performed by a method of holding down the plate 13. The swaged portion is shown as 11b. After the caulking operation, an adhesive may be applied to the caulked portion 11b.
[0068]
The sleeve 11 portion of the second integral member includes a large-diameter sleeve inner diameter portion 11c with which the outer periphery of the counter plate 13 is engaged, and a thrust plate 3 portion other than the upper step flat portion 11a and the caulked portion 11b described above. A small-diameter sleeve inner diameter portion 11d whose outer periphery is opposed, a lower step flat portion 11e whose lower surface in FIG. 1 is opposed to a portion protruding to the outer periphery of the thrust plate 3 portion, and a shaft 2 portion are opposed to the smallest diameter. And a shaft-facing inner surface 11f.
[0069]
A disk mounting step 12a for mounting a disk 14 as a hard disk, a yoke 15 made of a cylindrical magnetic metal, A large-diameter cylindrical portion 12b for concentrically fixing and holding a cylindrical magnet 16 fixed to the yoke 15 by bonding, press fitting, caulking, or the like, a disk mounting portion 12c for mounting the disk 14, and a disk 14 A screw hole 12d for inserting a screw 17 for fixing the device on the mounting portion 12c, and a clamper for mounting a clamper 18 for inserting the disk 14 between the disk mounting portion 12c. A mounting portion 12e, a clamper fitting step portion 12f into which the center hole of the clamper 18 fits, and a coil concaved in a concave shape to secure a sufficient space for the coil winding 5. A countercurrent recess 12g is formed. The rotor of the spindle motor 10 is constituted by the second integral member, the counter plate 13, the yoke 15, the magnet 16, and the like.
[0070]
The aluminum-silicon alloy used as the material of the second integral member is distributed as described above. The manufacturing method is as shown in FIG. That is, first, a rapidly solidified powder of an aluminum alloy is manufactured by an atomizing method (step S51). As the atomizing method, a gas atomizing method, an ultrasonic gas atomizing method, or the like is employed.
[0071]
This rapidly solidified powder causes a molten alloy of aluminum and silicon to flow out of a tundish (a container having a hole at the bottom), and at the same time, causes a jet of a spray medium (gas or liquid) to impinge on the molten metal stream. After the molten metal is scattered into fine droplets, heat is deprived and solidified to form. The powder particles have a diameter of about 100 μm including a large number of silicon having a diameter of about 2 μm, for example.
[0072]
Then, an additive consisting of ceramics and a special alloy powder is added (step S52), and the final material is 67-68% by weight of Al, 30% by weight of Si, and 2-3% by weight of Cu (copper) and others. And rapidly solidified powder. Then, a billet-shaped green compact is obtained by hot pressing (step S53). Thereafter, a degassing process is performed in which the powder is heated in a vacuum or in a non-oxidizing atmosphere to remove the oxide adsorbed on each surface of the powder and the water adsorbing the non-oxide, and the like (step S54). Next, an extruded material is obtained by performing hot extrusion (step S55). The extruded material is processed to form a second integral member. Thereafter, heat treatment is appropriately performed.
[0073]
The linear expansion coefficient of the obtained aluminum silicon alloy changes depending on the content of silicon (Si). In this embodiment, 30% by weight of silicon is contained, and about 14 × 10 ^-6 Has a linear expansion coefficient of As the silicon content is further increased, the linear expansion coefficient decreases proportionally, and when the silicon content is about 36%, about 13 × 10 ^-6 When the silicon content becomes about 44 to 48%, 10 × 10 ^-6 (This is about the same as the linear expansion coefficient of the SUS-400 series). Instead of changing only silicon, nickel may be mixed with silicon, and the linear expansion coefficient may be lowered by gradually increasing the total amount of silicon and nickel from 30% by weight.
[0074]
The disk 14 is mounted when the spindle motor 10 is mounted on a disk drive. This mounting is performed by mounting the disk 14 on the disk mounting portion 12c, mounting the clamper 18 on the clamper mounting portion 12e, and tightening the screw 17 while inserting the screw 17 into the screw hole 12d. You. The disk 14 of this embodiment is a disk for a 2.5-inch hard disk.
[0075]
This is a small gap between a first integral member in which the shaft 2 and the thrust plate 3 are integrally formed and a second integral member in which the sleeve 11 and the hub 12 are integrally formed. Oil for dynamic pressure is put in the portion where the groove is formed, and the oil is held so as not to leak out from the sleeve 11 portion.
[0076]
The gap between the outer peripheral surface of the shaft 2 and the shaft-facing inner surface 11f of the sleeve 11 is 2 to 4 μm. In general, when the diameter of the shaft 2 is 4 mm, the gap is 4 μm, and when the diameter is 3 mm, the gap is 3 μm. The size of the gap is 1/1000 of the diameter of the shaft 2. ing. In this embodiment, the diameter of the shaft 2 portion is 2.5 mm, and the gap between the shaft 2 portion and the shaft-facing inner surface portion 11f is 2.5 μm. The gap between the thrust plate 3 and the counter plate 13 and the gap between the thrust plate 3 and the sleeve 11 are also 2.5 μm.
[0077]
Two dynamic pressure grooves for radial bearings are provided on the shaft-facing inner surface 11f. That is, a first radial dynamic pressure groove 21 for restricting the radial movement of the second integrated member is provided on one end side (upper side in FIG. 1) of the shaft-facing inner surface portion 11f. 1 (on the lower side), a second radial dynamic pressure groove 22 having the same function is provided.
[0078]
The thrust direction of the second integral member is set at a position facing the thrust plate 3 of the counter plate 13 covering the thrust plate 3 and facing a portion of the thrust plate 3 protruding in the outer peripheral direction from the shaft 2 portion. A first thrust dynamic pressure groove 23 for restricting the movement of the first thrust dynamic pressure groove 23 is provided. A second thrust dynamic pressure groove having a similar function is provided at a position facing a lower step flat portion 11e of the sleeve 11, that is, a flat portion of a portion of the thrust plate 3 that protrudes in the outer peripheral direction from the shaft 12 portion. 24 are provided.
[0079]
As shown in FIG. 1 and FIG. 2, the first and second radial dynamic pressure grooves 21 and 22 are provided with a plurality of> -shaped grooves arranged in parallel in the circumferential direction, and make a round in the shaft-facing inner surface portion 11f. It is formed as follows. Each of the first and second thrust dynamic pressure grooves 23 and 24 has a large number of> -shaped grooves arranged in a circular shape. Each of the dynamic pressure grooves 21, 22, 23, and 24 may have another shape. For example, the thrust dynamic pressure grooves may be formed by concentrically arranging circles having different diameters, or may be formed in a spiral shape. May be adopted.
[0080]
Each of the dynamic pressure grooves 21, 22, 23, and 24 is formed by pressing, etching, cutting, rolling such as ball rolling, electric discharge machining, or the like. The oil for dynamic pressure is applied to each of the above-described first radial dynamic pressure groove 21, the second radial dynamic pressure groove 22, the first thrust dynamic pressure groove 23, and the second thrust dynamic pressure groove 24. , Each groove 21, 22, 23, 24 is filled with oil.
[0081]
The order of assembling the hydrodynamic bearing device and the spindle motor 10 configured as described above will be described below.
[0082]
First, a shaft (first integrated member) integrally formed with the thrust plate 3 is inserted into a second integrated member having the hub 12. Next, the counter plate 13 is arranged so that the outer periphery thereof is opposed to the large-diameter sleeve inner diameter portion 11c and is in contact with the upper step flat portion 11a, and then the counter plate is swaged and fixed to the sleeve 11 portion. I do. Then, if necessary, an adhesive is applied to the caulked portion 11b. Thereafter, a gap (about 2.5 μm in this example) between the first integrated member and the second integrated member is filled with dynamic pressure oil using a vacuum device. Note that the yoke 15 and the magnet 16 are concentrically fixed to the second integral member by press-fitting, bonding, caulking, or the like in advance.
[0083]
On the other hand, the core 4 is fixed to the base 1 by bonding, press-fitting, caulking, or the like, and the coil winding 5 is wound around salient poles to form a stator. After applying the coil winding 5 to the core 4, the core 4 may be fixed to the base 1. The shaft 2 is fixed to the center hole of the base 1 constituting the stator by a method combining one or more of press fitting, shrink fitting, bonding, welding and the like.
[0084]
In this assembled state, the length of the shaft 2 and the axial thickness of the portion of the base 1 into which the shaft 2 is press-fitted are larger than in the completed state. Therefore, after assembling, a part of the bottom side of the base 1 and the shaft 2 is cut off by cutting to have a size in a completed state. Thus, the spindle motor 10 is completed. This incorporation method is adopted in order to obtain accurate verticality. In order to obtain an accurate verticality, a press-fitting amount approximately equal to the diameter of the shaft 2 is required. Since the diameter of the shaft 2 of the spindle motor 10 for a 2.5-inch hard disk is about 2.5 mm, the thickness of the press-fit portion of the shaft 2 of the base 1 is once set to 2.5 mm, and the verticality is sufficiently increased. After being taken out, about 1.5 mm is cut out, the thickness of the bottom portion of the base 1 in the final form is 1 mm, and the base 1 is thinned.
[0085]
By adopting such an assembling method, the verticality of the shaft 2 with respect to the base 1 is maintained, and the thickness is reduced. Further, since the bottom side of the base 1 is cut off, the work of chamfering the outer end of the center hole of the base 1 is not required. It is preferable that this assembling method (a method of press-fitting the shaft 2 and then cutting it out) is adopted for a spindle motor for a hard disk having a very small size such as 2.5 inches or 1.8 inches.
[0086]
When the spindle motor 10 is mounted on the disk drive, the disk 14 is mounted on the disk mounting portion 12c, and the disk 14 is sandwiched between the disk mounting portion 12c and the clamper 18 is mounted on the disk mounting portion 12e. Put on. Thereafter, the clamper 18 is tightened with the screw 17 to fix the disk 14. The disk drive is completed by incorporating the spindle motor 10 in this state inside the disk drive and providing other mechanisms, other circuits, and the like.
[0087]
Next, the basic operation of the hydrodynamic bearing device will be described, and the characteristics of the spindle motor 10 using the hydrodynamic bearing device and the characteristics of the disk drive device equipped with the spindle motor 10 and the characteristics of the conventional device will be described together with the description. Will be described.
[0088]
First, the basic operation of the hydrodynamic bearing device will be described. First, the prerequisites for evaluation, that is, the change in oil viscosity, the bearing device for evaluation, and the relationship between the change in clearance and the characteristics of the bearing are shown in FIG. This will be described with reference to FIG.
[0089]
FIG. 4 shows a change in kinematic viscosity of oil (common oil used in a dynamic pressure bearing device) with respect to temperature. The kinematic viscosity is a value obtained by dividing the viscosity (absolute viscosity) by the density. For example, the kinematic viscosity is a viscosity of 1 mPa · s (milli-pascal second) and the density is 1 g / cm. ² Then 1mm ² / S (square millimeters per second). The unit of the conventional centistokes (cSt) is the same value. As for the kinematic viscosity, the larger the numerical value is, the higher the value is.
[0090]
As shown in FIG. 4, the kinematic viscosity of the oil rapidly decreases as the temperature increases. A decrease in viscosity results in a decrease in bearing stiffness, ie, dynamic pressure. For this reason, when the temperature rises, it is necessary to reduce the gap between the shaft and the shaft support in order to maintain the bearing rigidity. However, if the clearance is too small, the bearing stiffness becomes too high, and as described later, the shaft loss torque increases, or in the worst case, the shaft and the shaft support come into contact with each other to lock (unrotatable). In some cases, this may occur.
[0091]
FIG. 5 shows an evaluation dynamic pressure bearing device. In this device, the diameter L1 of the shaft 26 is 3 mm, the width L2 of the upper radial dynamic pressure groove (upper BRG) 27 is 3 mm, the width L3 of the lower radial dynamic pressure groove (lower BRG) 28 is 2 mm, In the position and direction shown in FIG. 5, 0.01 kg is applied as a lateral external force F1. That is, the distance L4 between the upper BRG 27 and the lower BRG 28 is 1 mm, and a lateral external force F1 is applied to a position 1.5 mm away from the intermediate position toward the upper BRG 27. A thrust bearing having a diameter larger than that of the shaft 26 is provided below. Further, the rotation speed of the shaft 26 is set to 7200 RPM. Using such an oil and a dynamic pressure bearing device for evaluation, a gap between the shaft 2 (corresponding to the shaft 26) and the shaft-facing inner surface portion 11f of the sleeve 11 (corresponding to the upper BRG 27 and the lower BRG 28) is reduced to 0. FIGS. 6 to 9 show the results of analyzing the bearing characteristics (eccentricity and shaft loss torque) when changed in units of 0.5 μm.
[0092]
In the case of HDD, the most important characteristic is NRRO, which is largely affected by the bearing stiffness. As the bearing stiffness becomes smaller (lower), the gap changes and the eccentricity increases, but the allowable value of the eccentricity is experimentally and empirically about 0.01. The eccentricity e is an index indicating how flat the ellipse is. When the major axis from the center of the ellipse to the outer periphery in the longitudinal direction is a, and the minor axis from the outer periphery in the transverse direction to b, the following is given. This is defined by Expression 1.
[0093]
[Formula 1]

[0094]
In Equation 1, when the eccentricity e is 0, the circle becomes a perfect circle. The bearing stiffness refers to k represented by k = F / x, where F is the sum of forces acting on the rotor (sum of external forces) and x is the amount of change in the rotor. [Kg · f / μm].
[0095]
When the temperature is low, such as 20 ° C. shown in FIG. 6, the eccentricity has a relatively large margin, but at high temperatures, the viscosity decreases and the bearing rigidity decreases. Can not withstand. That is, as shown in the graphs at 60 ° C. in FIG. 8 and at 80 ° C. in FIG. 9, when the temperature becomes high, the upper radial dynamic pressure groove 27 (corresponding to the upper first radial dynamic pressure groove 21 in FIG. 1). When the eccentricity e at the portion is a gap of about 3 μm, it exceeds 0.01. On the other hand, if the gap is too small, the shaft loss torque may increase, or in the worst case, the shaft 2 and the sleeve 11 may come into contact with each other and become unable to rotate. Here, the shaft loss torque is a shaft torque lost due to oil viscosity or the like when the shaft rotates, and the smaller the value, the better. Here, the unit is shown in [g · cm].
[0096]
Next, a change in rigidity due to a temperature change, which is one of the basic operations of the hydrodynamic bearing device, will be described.
[0097]
The gap also changes due to thermal expansion of the shaft or sleeve. FIG. 10 shows the change in the gap with respect to the temperature change due to the combination of the shaft and sleeve materials. The dynamic pressure bearing device generates pressure in the oil by rotation. However, the smaller the gap and the higher the viscosity of the oil, the higher the pressure can be generated. The viscosity of the oil is high at low temperatures and low at high temperatures. Therefore, when the material of the shaft and the sleeve is selected, if the linear expansion coefficient of the shaft is larger than that of the sleeve, the gap decreases at a high temperature, so that the pressure drop due to the viscosity decrease can be compensated.
[0098]
The combination of the material used for the current spindle motor and the aluminum-silicon alloy used in the present invention as a shaft and a sleeve are summarized as follows. First, if the shaft and the sleeve are made of the same material, the gap does not change with temperature. The fact that the length of the gap does not change means that the gap is the same while the outer diameter of the shaft and the inner diameter of the sleeve both increase, which indicates that the volume of the gap increases. At higher temperatures, the viscosity of the oil decreases. For this reason, the volume of the gap increases and the viscosity of the oil decreases, and the bearing rigidity deteriorates. Further, if the shaft and the sleeve are made of the same material, the shaft is damaged when the shaft and the sleeve collide, and the bearing characteristics deteriorate. It is not preferable to use the same material for the shaft and the sleeve because such disadvantages are often employed.
[0099]
A combination of a shaft made of SUS-400 and a sleeve made of brass (BsBn) or bronze (Bronze) provides relatively good workability. However, in terms of thermal expansion, when the temperature becomes high, the gap increases, and as the viscosity of the oil decreases, the bearing rigidity tends to deteriorate. In particular, the dynamic pressure bearing device of a spindle motor used for a small HDD such as 1.8 inches or 2.5 inches has no margin in bearing rigidity. Therefore, the gap between each dynamic pressure groove is designed to be extremely small. There must be. However, it is difficult to meet the required tight tolerances, and depending on the temperature (low temperature), the shaft and the sleeve may interfere with each other, and the bearing will not function. Further, the material of the sleeve is not sufficient for abrasion, and requires plating of nickel / chromium on the surface. Therefore, it can be said that this combination is also problematic.
[0100]
A combination in which the shaft is made of SUS-300 and the sleeve is made of SUS-400 is preferable in consideration of only thermal expansion. That is, since the gap decreases as the temperature rises, it has an effect of compensating for a decrease in rigidity due to a decrease in oil viscosity. However, as described in detail in the related art, when the sleeve is made of SUS-400, various problems occur due to its high hardness. That is, when the temperature becomes high, the gap becomes small, and the shaft comes into contact with the sleeve or locks. In addition, it is difficult to perform high-precision processing such as inner diameter, roundness, surface roughness, cylindricity, and dynamic pressure groove processing. If the variation in the inner diameter becomes large, it is not possible to obtain an appropriate gap with the shaft, and it is necessary to perform a selective combination by measurement. Even in the dynamic pressure groove machining, the uniformity and symmetry of the groove depth are deteriorated, and the bearing characteristics are adversely affected. As described above, the combination of the SUS-300 shaft and the SUS-400 sleeve causes problems in performance, productivity, and price.
[0101]
If the shaft is made of SUS-300 and the sleeve is made of the aluminum silicon alloy used in the second integral member of the present invention, as shown in the table of FIG. 10, when the temperature becomes high, the gap gradually narrows. Become. Specifically, the gap increases by 0.2 μm at a temperature of 0 ° C., changes by 0 ° C. at 20 ° C., narrows by 0.2 μm at 40 ° C., narrows by 0.4 μm at 60 ° C., and decreases by 0.4 μm at 80 ° C. It becomes 6 μm narrower and 0.8 μm narrower at 100 ° C. These values are optimal in compensating for a decrease in rigidity due to a decrease in oil viscosity.
[0102]
The values shown in FIG. 10 are obtained by setting the diameter of the shaft to 3 mm. If the diameter of the shaft is set to another value, for example, 2.5 mm or 2 mm, the absolute value of the increase due to the expansion on the shaft side decreases. Therefore, the amount of change in the gap also becomes small. However, in general, when the diameter of the shaft is small, it is necessary to reduce the gap in relation to the bearing rigidity, and the ratio of the change amount of the gap to the original gap is not so large even when the shaft diameter is small. There is no. However, strictly speaking, the ratio changes, so even if the shaft is made of SUS-300 and the sleeve is made of the aluminum silicon alloy of the present invention, the difference between the linear expansion coefficients of the two is determined by the diameter of the shaft. It needs to be different. In the above embodiment, the difference between the two linear expansion coefficients is 1.5 × 10 ^-6 ~ 4.5 × 10 ^-6 However, if the diameter of the shaft is about 2 mm, the gap is about 2 μm, and the difference between the two is 7 × 10 ^-6 The following is required. Conversely, if the diameter of the shaft is about 4 mm, the gap is about 4 μm, and the difference between them is 1 × 10 ^-6 Anything above would be good.
[0103]
Next, as the last of the description of the basic operation of the hydrodynamic bearing device, the relationship between the current consumption and the rigidity will be described.
[0104]
Bearing stiffness and current are in conflicting conditions. That is, in order to increase the rigidity of the bearing, it is necessary to reduce the gap or increase the viscosity of the oil. On the other hand, in order to reduce the current, that is, to reduce the torque loss, it is necessary to increase the gap or reduce the viscosity of the oil. In order to reduce the shaft loss torque while maintaining the same bearing rigidity, the diameter of the shaft may be reduced and the clearance may be reduced. However, since the inner diameter of the sleeve becomes smaller and the tolerance becomes tighter, it is extremely difficult to process a hard member such as a conventionally used material, for example, brass or SUS-300 system with higher accuracy than before. Become. When the conventional material is used, it is necessary to measure all the shaft diameters and the sleeve inner diameters and select and combine them. However, this method greatly increases the production cost. In addition, since the inner diameter of the sleeve is small, the roundness and accuracy are poor, and the yield is low.
[0105]
The above consideration will be described with reference to FIGS. 11 shows the case where the shaft is made of SUS-400 and the sleeve is made of brass. FIG. 12 shows the configuration of the present invention, that is, the shaft 2 is made of SUS-304, and the sleeve 11 is used in the present invention. This shows the case where an aluminum silicon alloy is used.
[0106]
As shown in FIG. 11, when SUS-400 is used as the shaft and brass is used as the sleeve, when the temperature exceeds about 58 ° C., the eccentricity e in the upper radial dynamic pressure groove 27 is an allowable limit value. Exceeding 0.01. Further, the inclination angle increases sharply as the temperature rises. Here, the inclination angle refers to the inclination of the shaft with respect to the base, and the angle of 90 degrees (vertical) with respect to the base is defined as 0 degree.
[0107]
On the other hand, in the combination of the present invention (the shaft 26 corresponding to the shaft 2 is a SUS-300 type, and the upper BRG 27 and the lower BRG 28 corresponding to the sleeve 11 are the aluminum silicon alloy of the present invention), as shown in FIG. As shown, when the temperature is between 20 ° C. and 80 ° C., the eccentricity e falls within the allowable limit of 0.01.
[0108]
As described above, the aluminum silicon alloy used in the present embodiment has a linear expansion coefficient smaller than that of the shaft 2 without deteriorating the workability, corrosion resistance, light weight, low cost, high strength, and the like as basic characteristics of aluminum. Value. That is, by using silicon (Si) and other additives in aluminum, the property of decreasing the coefficient of linear expansion is used. 10 ^-6 Aluminum / silicon alloy within plus / minus 5%. The coefficient of linear expansion of this aluminum silicon alloy is a value approximately intermediate between the coefficients of linear expansion of the SUS-300 series and the SUS-400 series, that is, 14 × 10. ^-6 It has become. In addition, this aluminum silicon alloy has a temperature range of about 13.5 to about 15.3 × 10 3 at a temperature range of 0 ° C. to 100 ° C. ^-6 And the value increases as the temperature increases. Note that the temperature range of 0 ° C. to 100 ° C. is a heat-resistant condition required for a hard disk motor, and is also an operating temperature environment.
[0109]
Other characteristics of the aluminum silicon alloy (30% by weight of Si, 2 to 3% by weight of Cu and others and the balance of Al) used in the present embodiment are as follows. Tensile strength is 46.6kgf / mm ² , Proof stress is 40.0kgf / mm ² , Young's modulus is 9700kgf / mm ² , Hardness is 150 Hv (Vickers hardness), density is 2.6 g / cm ³ It is.
[0110]
In the above-described embodiment, the shaft 2 and the thrust plate 3 are formed as an integral member of the SUS-300 system, and the sleeve 11 and the hub 12 are formed as an integral member of the above-mentioned aluminum silicon alloy. RRO) and NRRO are both very small. This is because the fit between the sleeve and the hub is lost, and the fit between the shaft and the thrust plate is lost. As a result, the perpendicularity between the sleeve 11 and the hub 12 is completely maintained. That is, if the sleeve and the hub are made of an integral material and the disk mounting portion 12c and other portions of the hub 12 are simultaneously processed, the verticality and the horizontality of each portion can be made high precision. The same applies to the integral member of the shaft 2 and the thrust plate 3.
[0111]
In addition, since the two integrated members are not of a fitting structure, the impact resistance is improved, and non-uniform stress is not exerted, so that stable accuracy can be maintained both temporally (temporarily) and thermally. It becomes possible. Further, in this embodiment, since the sleeve 11 and the hub 12 are formed as an integral member, it is not necessary to consider fixing the hub 12 to the sleeve 11, and the core 4 and the coil winding 5 are not required to be fixed. Can be made sufficiently large, and the current loss can be reduced. In addition, since the dynamic pressure grooves 21 and 22 for the radial bearing are provided on the sleeve 11 side where the hardness becomes softer, the dynamic pressure grooves are easily formed inside the sleeve 11 by ball rolling similarly to the conventional brass sleeve. And it can be formed so as not to cause cracks.
[0112]
Further, in the dynamic bearing device of the above-described embodiment and the spindle motor 10 in which the dynamic bearing device is used, the verticality of the rotating parts and the shaft 2 can be maintained with high accuracy, so that the stable. Dynamic pressure is obtained, and the life is prolonged. Further, since both of the integral members are not of a fitting structure, the thermal shock resistance is also improved.
[0113]
Further, since the dynamic pressure grooves 23 and 24 for the thrust bearing are provided on the counter plate 13 and the sleeve 11 facing the thrust plate 3, the dynamic pressure grooves are provided on the thrust plate 3 where the area tends to be small. As compared with the case, the degree of freedom of the installation position of the dynamic pressure groove is increased, so that a more appropriate dynamic pressure effect can be easily obtained and the center of the dynamic pressure groove can be easily set.
[0114]
Further, in this embodiment, the integral member was measured at a linear expansion coefficient of 0 ° C. ^-6 Since it is made of an aluminum silicon alloy within plus / minus 5%, a dynamic pressure groove can be easily formed in the sleeve 11 portion, and SUS300 can be adopted as the shaft 2 portion. In addition, the thickness of the root portion of the hub 12 (the hub side at the joint portion between the conventional motor sleeve and the hub) is 1.3 times the thickness of the disk 14 to be mounted. The integral rotation becomes smooth and the storage space S for the core 4 and the coil winding 5 can be increased. The current loss can be suppressed by enlarging the storage space S.
[0115]
Further, since the axial thickness of the thrust plate 3 portion of the shaft 2 is set to 0.3 mm, that is, 1/3 to 1/5 of the conventional one, the function of the thrust bearing can be sufficiently satisfied and the shaft of the motor can be sufficiently satisfied. The length in the direction can be reduced, the thickness can be reduced, and the current loss is significantly reduced.
[0116]
In addition, in this disk drive device, A-RRO is reduced, the change in bearing stiffness is reduced, and current loss can be suppressed, so that reading of information from the disk 14 and writing of information to the disk 14 are not mistaken. And can be operated for a long time when using a battery. In addition, other RRO and NRRO of A-RRO also become very small.
[0117]
Further, since the dynamic pressure grooves 23 and 24 for the thrust bearing are provided on the counter plate 13 and the sleeve 11 facing the thrust plate 3, the dynamic pressure grooves are provided on the thrust plate 3 where the area tends to be small. As compared with the case where the dynamic pressure groove for the thrust bearing is installed, the degree of freedom in the installation position and design of the dynamic pressure groove for the thrust bearing is increased, and a more appropriate dynamic pressure effect is easily obtained. For this reason, it is possible to provide a high-quality disk drive device without any information writing or reading errors. Further, as described above, in this disk drive device, since various RROs and NRROs (at least A-RROs among them) of the internal spindle motor are reduced and the axial length is reduced, information reading and writing are performed. And the device can be made thinner as the motor becomes thinner.
[0118]
The above-described embodiment is an example of a preferred embodiment of the present invention, but can be variously modified without departing from the gist of the present invention. For example, as a spindle motor using the above-described hydrodynamic bearing device, as in a spindle motor 30 shown in FIG. 13, the axial length of the outer diameter of the hub 12 is extended to increase the number of disks 14 mounted. You may do it. The other configuration is basically the same as that of the spindle motor 10 shown in FIG.
[0119]
In addition, as a spindle motor using the above-described hydrodynamic bearing device, as in a spindle motor 40 shown in FIG. The winding 5, the yoke 15, and the magnet 16 may be provided. This spindle motor 40 can also increase the number of disks 14 mounted. Other configurations are basically the same as those of the spindle motors 10 and 30 shown in FIG. 1 and FIG.
[0120]
Further, as in the case of the spindle motor 50 in FIG. 15, the shape may be a shape suitable for being mounted on an HDD of 2.5 inches or less. This small spindle motor 50 has a boss-shaped portion 51 as shown in FIG. 15 extending slightly upward in the vicinity of the counter plate 13 above the hub 12, and a groove 52 is provided on the outer diameter of the boss-shaped portion 51. The two disks 14 are fixed by fitting the inner diameter of the clamper 18 here. In addition, between the two disks 14, 14, there is provided a cylindrical spacing member 53 for forming an interval between the disks 14, 14 and maintaining the interval.
[0121]
In the spindle motor 50, the thickness of the clamper 18 is reduced, and the axial height of the hub 12 can be shortened because there is no screw for fixing the disks 14, 14. If the height is the same as that of the conventional one, the first radial dynamic pressure groove 21 can be moved upward, so that the moment rigidity can be increased. Further, since the thickness of the root portion 54 of the hub 12 (the hub side at the joint portion between the sleeve and the hub of the conventional motor and the hub) is substantially the same as the thickness of the disk 14, the integral rotation with the disk 14 becomes smoother. The storage space S for the core 4 and the coil winding 5 can be made large even though it is thin. Since the storage space S becomes large, current loss can be suppressed.
[0122]
Each of the spindle motors 10, 30, 40, and 50 using the dynamic pressure bearing device of the above-described embodiment employs a fixed-shaft type dynamic pressure bearing device, and thus is a so-called fixed-shaft type spindle motor. However, by using the fixed shaft as described above, there is an advantage that the verticality of the relative rotation portion (the sleeve 11 and the shaft 2 portion, the hub 12 and the thrust plate 3 portion) can be maintained with high accuracy. That is, in the spindle motor, there are two relative rotation portions, but in order to obtain verticality, it is preferable that those portions be formed as an integral member. In response to such a demand, in the case of a fixed shaft type spindle motor, the relative rotation portion can be separated into an integral member of the thrust plate 3 and the shaft 2 and an integral member of the hub 12 and the sleeve 11, And it can be incorporated. On the other hand, in the case of the shaft rotation type spindle motor, if the shaft and the thrust plate are formed as an integral member, the rotor portion (including the shaft, the thrust plate, and the hub) including the hub portion cannot be incorporated. Thus, in order to increase the accuracy of the verticality, a fixed shaft type spindle motor is preferable.
[0123]
Further, in each of the spindle motors 10, 30, 40, and 50 described above, since the verticality of the rotating parts and the shaft 2 can be maintained with high accuracy, a stable dynamic pressure can be obtained and the life is prolonged. Further, since both of the integrated members are not of a fitting structure, impact resistance and thermal shock resistance are also improved. Even if such a property is partially sacrificed, a spindle motor having more advantages than before can be obtained. For example, in the above-described embodiment, the spindle motor 10 having two integral members is used as a hydrodynamic bearing device in which the shaft 2 and the thrust plate 3 are integral members and the sleeve 11 and the hub 12 are integral members. , 30, 40, and 50, but may be a spindle motor including a hydrodynamic bearing device having only one of the integrated members. Even in such a case, compared to the conventional spindle motor and the dynamic pressure bearing device, it has an advantageous effect in terms of A-RRO and the like. Also, when the aluminum silicon alloy shown in the present invention is used only for the sleeve 11 and the sleeve 11 and the hub 12 are not formed as an integral member but are formed of two parts, various bearing characteristics are improved.
[0124]
Further, in the above-described embodiment, a so-called fixed shaft type spindle motor has been described. However, the sleeve portion of the shaft rotation type spindle motor, or the sleeve portion and the base portion are integrated, and the integrated member of the present invention is provided. An aluminum silicon alloy as a material of the second integral member may be used, and the shaft may be made of SUS-300. In this case, it is more preferable that the shaft and the hub be an integral member.
[0125]
The coefficient of linear expansion of the second integral member is 14 × 10 at 0 ° C. to 100 ° C. ^-6 Within ± 5%, the linear expansion coefficient of the shaft body as the first integral member is 17 × 10 ^-6 When it is within plus / minus 5%, the linear expansion coefficient of the second integral member is 11 × 10 in a temperature range of 0 ° C. to 100 ° C. ^-6 ~ 15 × 10 ^-6 Within this range, a considerable effect can be obtained as compared with the related art. In a temperature range where the linear expansion coefficient is 0 ° C to 100 ° C, 17 × 10 ^-6 As the degree, SUS-300 is preferable, but another metal member or a metal member having a surface only with this value by surface treatment may be used.
[0126]
Further, by forming the second integral member from a predetermined aluminum alloy, the second integral member has a considerable hardness and a predetermined linear expansion coefficient. 1 × 10 in the temperature range of ℃ ^-6 ~ 7 × 10 ^-6 If it is made smaller, the bearing rigidity in the case where the gap is several microns becomes stable. Further, in the above-described embodiment, the linear expansion coefficient of the second integral member increases as the temperature increases (about 0.15 to 0.2 × 10 5 per 10 ° C.). ^-6 However, the rate of increase may be set to another value, or the change value may increase as the temperature increases. In addition, the second integral member has a constant linear expansion coefficient (about 11 × 10 ^-6 ~ 15 × 10 ^-6 (A specific value in the range of).
[0127]
Further, the thickness of the root portion of the hub 12 (the hub side at the joint portion between the sleeve and the hub of the conventional motor) is substantially the same as the thickness of the disk 14 or about 1.3 times. If the thickness is in the range of 0.5 times to 2 times the thickness of the disk 14 to be mounted, the rotation with the disk 14 becomes smooth and the storage space S for the core 4 and the coil winding 5 is provided. Can be made large despite its thinness. By making the storage space S large, current loss can also be suppressed.
[0128]
The axial thickness of the thrust plate 3 of the shaft 2 is preferably 0.1 to 0.7 mm, more preferably 0.15 to 0.35 mm. With such a configuration, the axial length of the motor can be reduced while sufficiently satisfying the function of the thrust bearing, so that the motor can be made thinner and the current loss is greatly reduced. Further, in the above-described embodiment, the linear expansion coefficient is made constant in the range of 0 ° C. to 100 ° C. on the shaft side serving as the first integral member, but here, the constant does not change at all. Not only those but also those whose change is about 1/5 or less as compared with aluminum silicon alloy.
[0129]
Further, in the above-described embodiment, the spindle motor for the HDD is shown, but a spindle motor for another device such as a motor for an optical scanning (scanner) device may be used. In such a case, one or both of the thrust plate 3 and the counter plate 13 may be unnecessary. Further, there may be a case where a portion corresponding to the hub 12 is omitted. Further, as the dynamic pressure bearing device, instead of oil, another liquid or a gas such as air may be used.
[0130]
【The invention's effect】
In the present invention, impact resistance can be improved, A-RRO (so-called rocking motion of the shaft) can be reduced, change in bearing rigidity with respect to temperature change can be reduced, current loss can be suppressed, and downsizing can be achieved. A dynamic pressure bearing device that can be made thin can be obtained. Further, according to another aspect of the invention, it is possible to obtain a dynamic pressure bearing device capable of reducing a change in bearing rigidity with respect to a temperature change and maintaining stable performance for a long period of time.
[0131]
Further, in another invention, it is possible to obtain a dynamic pressure bearing device that can suppress current loss and increase the degree of freedom of the installation position of a thrust-side dynamic pressure groove, and can make a dynamic pressure effect appropriate. it can. Further, according to another aspect of the present invention, it is possible to obtain a dynamic pressure bearing device that can reduce a change in bearing rigidity with respect to a temperature change and that can prevent a shaft from colliding with a sleeve.
[0132]
Further, in another aspect of the invention, it is possible to obtain a method of manufacturing a dynamic pressure bearing device that can reduce A-RRO (so-called swinging motion of a shaft) and can easily cope with a reduction in thickness. That is, it is possible to manufacture a spindle motor that can maintain the verticality of the shaft with respect to the base even when the thickness is reduced.
[Brief description of the drawings]
FIG. 1 is a cross-sectional view showing a structure of a dynamic bearing device according to an embodiment of the present invention and a spindle motor using the dynamic bearing device.
FIG. 2 is an enlarged view of a counter plate of a spindle motor using the hydrodynamic bearing device of FIG. 1 and a periphery thereof, in which a first integral member is omitted.
3 is a diagram showing steps of a manufacturing method for obtaining an aluminum silicon alloy as a material of an integral member integrating a sleeve portion and a hub portion of a spindle motor using the dynamic pressure bearing device of FIG. 1;
4 is a graph showing characteristics of oil (change in oil viscosity with respect to temperature) used in a dynamic pressure groove of a spindle motor using the dynamic pressure bearing device of FIG.
5 is a diagram showing an evaluation dynamic pressure bearing device for evaluating various performances of a spindle motor using the dynamic pressure bearing device of FIG. 1 and a conventional spindle motor.
FIG. 6 shows a case where the gap between the outer diameter of the shaft and the inner diameter of the sleeve is changed in units of 0.5 μm using the oil having the characteristics shown in FIG. 4 and the dynamic pressure bearing device for evaluation shown in FIG. FIG. 7 is a graph showing bearing characteristics (shaft loss torque and eccentricity) when the temperature is 20 ° C. FIG.
FIG. 7 shows a case where the gap between the outer diameter of the shaft and the inner diameter of the sleeve is changed in units of 0.5 μm using the oil having the characteristics shown in FIG. 4 and the dynamic pressure bearing device for evaluation shown in FIG. 7 is a graph showing bearing characteristics (shaft loss torque and eccentricity) when the temperature is 40 ° C.
8 shows a case where the gap between the outer diameter of the shaft and the inner diameter of the sleeve is changed in units of 0.5 μm using the oil having the characteristics shown in FIG. 4 and the hydrodynamic bearing device for evaluation shown in FIG. 7 is a graph showing bearing characteristics (shaft loss torque and eccentricity) when the temperature is 60 ° C.
FIG. 9 shows a case where the gap between the outer diameter of the shaft and the inner diameter of the sleeve is changed in units of 0.5 μm using the oil having the characteristics shown in FIG. 4 and the hydrodynamic bearing device for evaluation shown in FIG. 6 is a graph showing bearing characteristics (shaft loss torque and eccentricity) when the temperature is 80 ° C.
10 is a table showing a change in a gap with respect to a temperature change due to a combination of materials used for a shaft or a sleeve of a spindle motor using the hydrodynamic bearing device of FIG. 1 or a conventional spindle motor.
FIG. 11 is a graph showing bearing characteristics (shaft loss torque, inclination angle, eccentricity, bearing stiffness) with respect to temperature by a combination of materials used in a conventional spindle motor.
FIG. 12 shows bearing characteristics with respect to temperature (shaft loss torque, inclination angle, eccentricity, bearing stiffness) due to a combination of materials used in a spindle motor using the hydrodynamic bearing device according to the present embodiment shown in FIG. FIG.
FIG. 13 is a sectional view showing a structure of a first modified example of the spindle motor using the dynamic pressure bearing device according to the present embodiment.
FIG. 14 is a sectional view showing a structure of a second modified example of the spindle motor using the hydrodynamic bearing device according to the present embodiment.
FIG. 15 is a sectional view showing a structure of a third modified example of the spindle motor using the dynamic pressure bearing device according to the present embodiment.
FIG. 16 is a sectional view showing a structure of a spindle motor using a conventional shaft-fixed type dynamic pressure bearing device.
FIG. 17 is a cross-sectional view showing a structure of a spindle motor using a conventional shaft rotation type dynamic pressure bearing device.
[Explanation of symbols]
1 Base (part of stator)
2 Shaft (part of the first integral member and part of the shaft body)
3 Thrust plate (part of the first integral member and part of the shaft body)
4 cores (part of stator)
5 core winding (part of stator)
10 Spindle motor
11 sleeve (part of the second integral member, part of the rotor)
12 Hub (part of the second integral member, part of the rotor)
12a Step portion for mounting disk
12c disk mounting part
13 Counter plate (part of rotor)
14 disks
15 Yoke (part of rotor)
16 magnets (part of rotor)
21 1st radial dynamic pressure groove (dynamic pressure groove for radial bearing)
22 Second radial dynamic pressure groove (dynamic pressure groove for radial bearing)
23 1st thrust dynamic pressure groove
24 Second Thrust Dynamic Pressure Groove

Claims (10)

In a hydrodynamic bearing device having a base, a shaft fixed to the base, and a sleeve rotatably arranged around the shaft,
A shaft body formed integrally with a circular thrust plate having a diameter greater than the diameter of the shaft portion and the shaft is provided integrally with stainless steel, which is provided at an end of the shaft opposite to the base.
The sleeve is formed of an aluminum silicon alloy having aluminum and silicon as components,
The linear expansion coefficient of the shaft body is within 17 × 10 ⁻⁶ plus / minus 5%,
The coefficient of linear expansion of the sleeve is set within 14 × 10 ⁻⁶ plus / minus 5% in a measurement in a range of 0 ° C. to 100 ° C.,
The axial thickness of the thrust plate portion is 0.15 to 0.35 mm,
The hardness of the sleeve is smaller than the hardness of the shaft, and a dynamic pressure groove for a radial bearing is provided in a portion of the sleeve that is a softer member and that faces the shaft.
A first thrust movement for a thrust bearing is provided at a position facing the thrust plate and facing the thrust plate on a counter plate provided to cover the thrust plate and close the center hole of the sleeve. A pressure groove, and a second thrust dynamic pressure groove is provided on the sleeve portion facing a surface of the thrust plate opposite to a surface facing the first thrust dynamic pressure groove,
A dynamic pressure bearing device characterized by the above-mentioned. In a hydrodynamic bearing device having a shaft, a sleeve arranged to be rotatable relative to the shaft, and a base for fixing either the shaft or the sleeve,
The shaft is made of stainless steel,
The sleeve is formed of an aluminum silicon alloy having aluminum and silicon as components,
The linear expansion coefficient of the shaft is within 17 × 10 ⁻⁶ plus / minus 5%,
The linear expansion coefficient of the sleeve is set to 11 × 10 ⁻⁶ to 15 × 10 ^{−6 in} a measurement in a range of 0 ° C. to 100 ° C.,
A dynamic pressure bearing device wherein the hardness of the sleeve is smaller than the hardness of the shaft, and a dynamic pressure groove for a radial bearing is provided on a portion of the sleeve which is a softer member and facing the shaft. . In a hydrodynamic bearing device having a shaft, a sleeve arranged to be rotatable relative to the shaft, and a base for fixing either the shaft or the sleeve,
The sleeve is made of an aluminum-silicon alloy composed of 65 to 69% by weight of Al, 28 to 32% by weight of Si, and 1 to 5% by weight of Cu and the like. In the measurement in the range of 100 ° C., the linear expansion coefficient of the shaft is reduced by 1 × 10 ^{−6 to} 7 × 10 ⁻⁶ , the hardness of the sleeve is smaller than the hardness of the shaft, and a softer member is obtained. A dynamic pressure bearing device, wherein a dynamic pressure groove for a radial bearing is provided on a portion of the sleeve facing the shaft. In a hydrodynamic bearing device having a shaft, a sleeve arranged to be rotatable relative to the shaft, and a base for fixing either the shaft or the sleeve,
A thrust plate is provided at the end of the shaft opposite to the base,
The sleeve is formed of an aluminum silicon alloy having aluminum and silicon as components,
The shaft is made of stainless steel which is an iron-based alloy containing 10.5 to 32% by weight of Cr and 4 to 13% by weight of Ni,
A dynamic pressure groove for a radial bearing is provided on an inner surface of the sleeve, which is a portion of the sleeve facing the shaft, and covers the thrust plate at a position facing the thrust plate and closes a center hole of the sleeve. A first thrust dynamic pressure groove for a thrust bearing is provided at a position facing the thrust plate of the counter plate provided as described above, and a surface of the thrust plate facing the first thrust dynamic pressure groove is A dynamic pressure bearing device, wherein a second thrust dynamic pressure groove is provided in the sleeve portion facing the opposite surface. In a hydrodynamic bearing device having a shaft, a sleeve arranged to be rotatable relative to the shaft, and a base for fixing either the shaft or the sleeve,
The sleeve is formed of an aluminum-silicon alloy having aluminum and silicon as components, the linear expansion coefficient of which increases as the temperature increases, at least in a temperature range of 0 ° C. to 100 ° C. The coefficient of linear expansion is larger than the coefficient of linear expansion and is constant in the temperature range, the hardness of the sleeve is smaller than the hardness of the shaft, and a portion of the sleeve that is a softer member and is opposed to the shaft is a radial bearing. A hydrodynamic bearing device comprising a hydrodynamic groove. The dynamic pressure bearing device according to any one of claims 2 to 5, wherein the shaft is fixed to the base and the sleeve is rotatably arranged around the shaft. 7. The hydrodynamic bearing device according to claim 6, wherein a disk holding hub that rotates integrally with the sleeve is formed integrally with the sleeve from the aluminum silicon alloy. At the end of the shaft opposite to the base, a thrust plate having a diameter larger than the diameter of the shaft is provided, and the thrust plate and the shaft are integrally formed of stainless steel to form a shaft body. 8. The hydrodynamic bearing device according to claim 6, wherein the plate portion has an axial thickness of 0.1 to 0.7 mm. 9. The hydrodynamic bearing device according to claim 6, wherein the rotor that rotates integrally with the sleeve is formed integrally with the sleeve from the aluminum silicon alloy. In a method for manufacturing a hydrodynamic bearing device having a base, a shaft fixed to the base, and a sleeve rotatably arranged around the shaft,
The axial thickness of the shaft and the base is made thicker than the completed state, and after the shaft is incorporated into the base, a part of the bottom side of the shaft and a part of the bottom side of the base are A method of manufacturing a hydrodynamic bearing device, characterized in that both the cut surfaces are cut at the same time so that the cut surfaces are constant, and the cut surface is in a completed state.

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