JP2004316816A - Ball bearing and main shaft device for machine tool - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a ball bearing and a main shaft device for a machine tool in which heat generation accompanied by high speed rotation is suppressed , and seizure is prevented with high rigidity, whose rolling fatigue lifetime can be lengthened, and which are advantageous with respect to environment and costs. <P>SOLUTION: Supposing that the curvature radius of the widthwise direction of an inside track groove is Ri, the curvature radius of the widthwise direction section of an outside track groove is Ro, the diameter of a ball is Da, and one-half the value which is an inside diameter of the inner ring subtracted from the outside diameter of the outer ring is bearing section height H, the curvature radius of the widthwise direction of the inside track groove is set to Ri=0.535Da-0.545Da, that of the outside track groove is set to Ro=0.515Da-0.545Da, and the contact angle of the ball with respect to the inside track groove and the outside track groove is set to 17°-26°, the diameter of the ball is set to Da=0.4H-0.56H, the ball bearing is lubricated with grease supplied from outside. <P>COPYRIGHT: (C)2005,JPO&NCIPI

Description

【００２６】
表１に番号１〜２５で示される各玉軸受の回転時（２５０００ｍｉｎ^−１）のアキシアルばね定数を図１に示す。同図に示すように、番号１〜２５の中で回転時のアキシアルばね定数が最も高い玉軸受は番号１の玉軸受であり、逆に回転時のアキシアルばね定数が最も低い玉軸受は番号２５の玉軸受である。
番号１の玉軸受と番号２５の玉軸受とを比較すると、番号１の玉軸受は玉の直径をＤａとすると内側軌道溝および外側軌道溝の幅方向断面の曲率半径Ｒｉ，Ｒｏが０．５２Ｄａであるのに対し、番号２５の玉軸受は内側軌道溝および外側軌道溝の幅方向断面の曲率半径Ｒｉ，Ｒｏが０．６０Ｄａであることから、内側軌道溝および外側軌道溝の幅方向断面の曲率半径Ｒｉ，Ｒｏが大きいほど回転時のアキシアルばね定数が低下し、回転時のアキシアル剛性が低くなることがわかる。
【００２７】
表１に番号１〜２５で示される各玉軸受の回転時（２５０００ｍｉｎ^−１）の内輪ＰＶ値を図２に示す。同図に示すように、番号１〜２５の中で回転時の内輪ＰＶ値が最も高い玉軸受は番号１の玉軸受であり、逆に回転時の内輪ＰＶ値が最も低い玉軸受は番号２０及び２５の玉軸受であり、番号５、１０及び１５の玉軸受も同様である。
【００２８】
番号５、１０、１５、２０及び２５の玉軸受は玉の直径をＤａとすると内側軌道溝の幅方向断面の曲率半径Ｒｉが０．６０Ｄａであり、このことから、外側軌道溝の幅方向断面の曲率半径Ｒｏが変わっても内側軌道溝の幅方向断面の曲率半径Ｒｉが同じ値であれば内輪ＰＶ値は変化しないことがわかり、さらに内側軌道溝の幅方向断面の曲率半径Ｒｉが大きいほど内輪ＰＶ値を低く抑えられることがわかる。
【００２９】
表１の番号１〜２５で示される各玉軸受の内輪を２５０００ｍｉｎ^−１で回転させた時の内輪ＰＶ値と外輪ＰＶ値を図３に示す。同図に示すように、玉軸受の内輪を２５０００ｍｉｎ^−１の高速で回転させた場合には、外輪ＰＶ値よりも内輪ＰＶ値のほうが大きくなる。これは、高速回転になると外輪コントロール（外輪と玉とが純転がりに近い状態）になり、玉の自転軸が外輪側の接触角を基準になるため、内輪側の滑りが大きくなり、ＰＶ値が増大するからである。したがって、焼付きを判断するための指標としては、内輪ＰＶ値を採用することが望ましい。
【００３０】
表１の番号１〜２５で示される各玉軸受の回転時（２５０００ｍｉｎ^−１）の発熱量を図４に示す。同図に示すように、番号１〜２５の中で発熱量が最も高い玉軸受は番号１の玉軸受であり、逆に発熱量が最も低い玉軸受は番号１０，１５，２０，２５の玉軸受である。番号１０，１５，２０，２５の玉軸受は内側軌道溝の幅方向断面の曲率半径ＲｉがＲｉ＝０．６０Ｄａであり、このことから、内側軌道溝の幅方向断面の曲率半径Ｒｉが大きいほど高速回転時の発熱量が低下し、焼付きを抑制できることがわかる。
【００３１】
表１に番号１〜２５で示される各玉軸受の回転時（２５０００ｍｉｎ^−１）の転がり疲労寿命を図５に示す。同図に示すように、番号１〜２５の中で転がり疲労寿命が最も長い玉軸受は番号６の玉軸受であり、逆に転がり疲労寿命が最も短い玉軸受は番号２５の玉軸受である。
番号６の玉軸受と番号２５の玉軸受とを比較すると、番号６の玉軸受は内側軌道溝の幅方向断面の曲率半径ＲｉがＲｉ＝０．５２Ｄａ、外側軌道溝の幅方向断面の曲率半径ＲｏがＲｏ＝０．５４Ｄａであるのに対し、番号２５の玉軸受は内側軌道溝の幅方向断面の曲率半径ＲｉがＲｉ＝０．６０Ｄａ、外側軌道溝の幅方向断面の曲率半径ＲｏがＲｏ＝０．６０Ｄａであることから、内側軌道溝および外側軌道溝の幅方向断面の曲率半径Ｒｉ，Ｒｏが大きいほど転がり疲労寿命が低下することがわかる。
【００３２】
表１に番号１〜２５で示される各玉軸受の回転時（２５０００ｍｉｎ^−１）の静的許容アキシアル荷重を図６に示す。同図に示すように、番号１〜２５の中で静的許容アキシアル荷重が最も高い玉軸受は番号１の玉軸受であり、逆に静的許容アキシアル荷重が最も低い玉軸受は番号５，１０，１５，２０，２５の玉軸受である。
【００３３】
番号５，１０，１５，２０，２５の玉軸受は内側軌道溝の幅方向断面の曲率半径ＲｉがＲｉ＝０．６０Ｄａであり、このことから、内側軌道溝の幅方向断面の曲率半径Ｒｉが大きいほど静的許容アキシアル荷重が低下してしまうことがわかる。
したがって、玉軸受の発熱量やＰＶ値を低減するためには、内側軌道溝および外側軌道溝の幅方向断面の曲率半径を大きくすれば良いが、内側軌道溝および外側軌道溝の幅方向断面の曲率半径を大きくすると軸受の剛性や寿命が低下すると共に静的許容アキシアル荷重が低下してしまう。これと反対に、内側軌道溝および外側軌道溝の幅方向断面の曲率半径を小さくすると発熱量やＰＶ値が上昇してしまう。
【００３４】
工作機械主軸支持用としては、焼付き指標、発熱量、アキシアル剛性、転がり疲労寿命、静的許容アキシアル荷重の全てにおいてバランスの良好な軸受が求められており、このような観点から、本発明者らは、番号１〜２５で示される各玉軸受のアキシアルばね定数、内輪ＰＶ値、発熱量、転がり疲労寿命、静的許容アキシアル荷重の計算結果を絶対値評価で比較した。ここでの絶対値評価は、各条件において最も優れる値を１００、最も劣る値を０とし、その差分を指数化して絶対値評価を行った。また、高速条件下で最も重要となる焼付き指標−内輪ＰＶ値は重要度を上げ、ランクを２倍として計算した。
【００３５】
上記の絶対値評価で使用した計算式を下記に示す。
（１）剛性（Ｋａ）
ＫａＲａｎｋ＝１００−（Ｋａ_Ｍａｘ−Ｋａ）／（（Ｋａ_Ｍａｘ−Ｋａ_Ｍｉｎ）
／１００）
Ｋａ_Ｍａｘ；図１の中で最大値を示すアキシアルばね定数
Ｋａ_Ｍｉｎ；図１の中で最小値を示すアキシアルばね定数
（２）内輪ＰＶ値（Ｐｖ）
ＰｖＲａｎｋ＝（Ｐｖ_Ｍａｘ−Ｐｖ）／（（Ｐｖ_Ｍａｘ−Ｐｖ_Ｍｉｎ）／１００）
Ｐｖ_Ｍａｘ；図２の中で最大値を示す内輪ＰＶ値
Ｐｖ_Ｍｉｎ；図２の中で最小値を示す内輪ＰＶ値
（３）発熱量（Ｗ）
ＷＲａｎｋ＝（Ｗ_Ｍａｘ−Ｗ）／（（Ｗ_Ｍａｘ−Ｗ_Ｍｉｎ）／１００）
Ｗ_Ｍａｘ；図４の中で最大値を示す発熱量
Ｗ_Ｍｉｎ；図４の中で最小値を示す発熱量
（４）転がり疲れ寿命（Ｌ）
ＬＲａｎｋ＝１００−（Ｌ_Ｍａｘ−Ｌ）／（（Ｌ_Ｍａｘ−Ｌ_Ｍｉｎ）／１００）
Ｌ_Ｍａｘ；図５の中で最大値を示す転がり疲れ寿命
Ｌ_Ｍｉｎ；図５の中で最小値を示す転がり疲れ寿命
（５）静的許容アキシアル荷重（Ｆａ）
ＦａＲａｎｋ＝１００−（Ｆａ_Ｍａｘ−Ｆａ）／（（Ｆａ_Ｍａｘ−Ｆａ_Ｍｉｎ）／１００）
Ｆａ_Ｍａｘ；図６の中で最大値を示す静的許容アキシアル荷重
Ｆａ_Ｍｉｎ；図６の中で最小値を示す静的許容アキシアル荷重
（６）ＴＯＴＡＬ−Ｒａｎｋ
ＫａＲａｎｋ＋（ＰｖＲａｎｋ×２）＋ＷＲａｎｋ＋ＬＲａｎｋ＋ＦａＲａｎｋ
上述したＫａＲａｎｋ、ＰｖＲａｎｋ、ＷＲａｎｋ、ＬＲａｎｋ、ＦａＲａｎｋ、ＴＯＴＡＬ−Ｒａｎｋの計算結果を表２に示す。
【００３６】
【表２】

【００３７】
ＴＯＴＡＬ−Ｒａｎｋの評価点は、番号２及び７の玉軸受が最も優れており、最適な仕様と言える。番号２及び７のＴＯＴＡＬ−Ｒａｎｋは、ほぼ同じ値である。
表２のＲａｎｋ係数からいずれかの指標が１００点を取るような軸受仕様は、いずれかの値が０もしくは０に近い値になり、両極端となる。このため、工作機械主軸用としては、あまり良い仕様とは言えない。
【００３８】
表２に示される計算結果から、工作機械主軸支持用として最も適している玉軸受は玉の直径をＤａとすると内側軌道溝の幅方向断面の曲率半径ＲｉがＲｉ＝０．５４Ｄａであり、また外側軌道溝の幅方向断面の曲率半径ＲｏがＲｏ＝０．５２Ｄａもしくは０．５４Ｄａであることがわかる。しかし、軸受を実際に製作する際には製作誤差も含まれるため、内側軌道溝の幅方向断面の曲率半径ＲｉをＲｉ＝０．５３５Ｄａ〜０．５４５Ｄａ、外側軌道溝の幅方向断面の曲率半径ＲｏをＲｏ＝０．５１５Ｄａ〜０．５４５Ｄａの範囲とするのが望ましい。
【００３９】
次に、本発明者らは、玉の接触角の影響を確認するために、内輪内径を７０ｍｍ、外輪外径を１１０ｍｍ、玉径をＤａ＝８．７３１ｍｍ、玉数を２５個、内側軌道溝および外側軌道溝の幅方向断面の曲率半径をＲｉ＝Ｒｏ＝０．５４Ｄａ、内外輪材質をＳＵＪ２、玉材質をＳｉ_３Ｎ_４とし、玉の接触角をα＝１２°、１５°、１８°、２０°、２２°、２５°、３０°とした場合の内輪ＰＶ値、外輪ＰＶ値、発熱量について計算した。なお、ここでの計算は、回転数；２５０００ｍｉｎ^−１、ＤＢ組合せ（間座幅２０ｍｍ）、軸嵌め合い；Ｔ２１（しまり嵌め、嵌め合いタイト量；２１μｍ）、ハウジング嵌め合い；Ｌ８（すきま嵌め、嵌め合いルーズ量；８μｍ）、組込み時予圧荷重；アキシアルばね定数１００Ｎ／μｍになる予圧荷重（この組込み時のアキシアルばね定数１００Ｎ／μｍは切削加工可能な剛性目安として使用）、温度条件；外輪温度上昇２４℃、内外輪温度差１６℃の条件で計算した。その計算結果を図７乃至図９に示す。
【００４０】
図７に示すように、内側軌道溝および外側軌道溝に対する玉の接触角がα＝３０°の場合には、ジャイロモーメントの影響により玉の滑りが大きくなるため、内輪ＰＶ値が急激に上昇する。したがって、内輪ＰＶ値を低く抑えるためには、内側軌道溝および外側軌道溝に対する玉の接触角を３０°より小さい角度にする必要があるが、内側軌道溝および外側軌道溝に対する玉の接触角を例えば１２°にするとアキシアル剛性が小さくなり、組込み時のアキシアルばね定数を１００Ｎに設定しようとした場合に過大な予圧荷重をかけなくてはならず、回転時の予圧も大きくなる。このため、図７に示すように、内側軌道溝および外側軌道溝に対する玉の接触角を１５°とした玉軸受よりも内輪ＰＶ値が増大するため、内側軌道溝および外側軌道溝の幅方向断面の曲率半径を表１に示す番号７のようにＲｉ，Ｒｏ＝０．５４Ｄａと設定した場合には、内側軌道溝および外側軌道溝に対する玉の接触角を１５°〜２５°とすることが望ましい。
【００４１】
また、図８に示すように、内側軌道溝および外側軌道溝に対する玉の接触角がα＝３０°の場合には、ジャイロモーメントの影響により玉の滑りが大きくなるため、発熱量が急激に上昇する。したがって、発熱量を低く抑えるためには、内側軌道溝および外側軌道溝に対する玉の接触角を３０°より小さい角度にする必要があるが、内側軌道溝および外側軌道溝に対する玉の接触角が１５°の場合には、内側軌道溝および外側軌道溝に対する玉の接触角が１８°の場合よりも発熱量が増大する。このため、内側軌道溝および外側軌道溝の幅方向断面の曲率半径を表１に示す番号７のようにＲｉ，Ｒｏ＝０．５４Ｄａと設定した場合には、内側軌道溝および外側軌道溝に対する玉の接触角を１８°〜２５°とすることが望ましい。
【００４２】
さらに、図９に示すように、内側軌道溝および外側軌道溝に対する玉の接触角がα＝１５°以下の場合には、上記と同様の理由により、外輪ＰＶ値が増大する。したがって、外輪ＰＶ値を低く抑えるためには、内側軌道溝および外側軌道溝に対する玉の接触角を１５°より大きい角度にする必要があるが、内側軌道溝および外側軌道溝に対する玉の接触角が３０°の場合には、ジャイロモーメントの影響により玉の滑りが大きくなり、内側軌道溝および外側軌道溝に対する玉の接触角が２５°の場合よりも外輪ＰＶ値が上昇する。このため、内側軌道溝および外側軌道溝の幅方向断面の曲率半径を表１に示す番号７のようにＲｉ，Ｒｏ＝０．５４ａと設定した場合には、内側軌道溝および外側軌道溝に対する玉の接触角を１８°〜２５°とすることが望ましい。
【００４３】
以上、図７乃至図９の結果より内側軌道溝及び外側軌道溝に対する玉の接触角は１８°〜２５°とすることが望ましいと言えるが、軸受を実際に製作する際には製作誤差も含まれるため、内側軌道溝及び外側軌道溝に対する玉の接触角は１７°〜２６°とすることが望ましい。
次に、本発明者らは、玉径の確認するために、内輪内径を７０ｍｍ、外輪外径を１１０ｍｍ、内側軌道溝および外側軌道溝の幅方向断面の曲率半径をＲｉ＝Ｒｏ＝０．５４Ｄａ、内側軌道溝および外側軌道溝に対する玉の接触角を１８°、内外輪材質をＳＵＪ２、玉材質をＳｉ_３Ｎ_４とし、玉径、玉数および断面比（＝玉径／（外輪外径−内輪内径）／２）×１００）を表３に示す組合せとした場合の内輪ＰＶ値を計算した。なお、ここでの計算は、回転数；２５０００ｍｉｎ^−１、ＤＢ組合せ（間座幅２０ｍｍ）、軸嵌め合い；Ｔ２１（しまり嵌め、嵌め合いタイト量；２１μｍ）、ハウジング嵌め合い；Ｌ８（すきま嵌め、嵌め合いルーズ量；８μｍ）、組込み時予圧荷重；アキシアルばね定数１００Ｎ／μｍになる予圧荷重（この組込み時のアキシアルばね定数１００Ｎ／μｍは切削加工可能な剛性目安として使用）、温度条件；外輪温度上昇２４℃の条件で計算した。その計算結果を図１０に示す。
【００４４】
【表３】

【００４５】
図１０から明らかなように、玉の直径を７．１４４ｍｍ、７．９３８ｍｍとした場合には、内輪と外輪との温度差が大きくなると内輪ＰＶ値が急激に上昇し、内輪と外輪との温度差が３０°付近では内輪ＰＶ値が１９００〜２１００ＭＰａ・ｍ／ｓ程度となる。また、玉の直径を１１．９０６ｍｍ、１２．７ｍｍとした場合は、内輪ＰＶ値の変化は緩やかであるが、元々のＰＶ値自体が高く、内輪と外輪の温度差が３０°付近は、前述の７．１４４ｍｍ、７．９３８ｍｍと同じく内輪ＰＶ値が１９００〜２１００ＭＰａ・ｍ／ｓ程度となる。これに対し、玉の直径を８．７３１ｍｍ、９．５２５ｍｍ、１０．３１９ｍｍ、１１．１１２ｍｍとした場合には、内輪と外輪との温度差が大きくなると内輪ＰＶ値も上昇するが、内輪と外輪との温度差が３０°付近では内輪ＰＶ値が１８００ＭＰａ・ｍ／ｓ程度となる。したがって、軸受断面高さ（外輪の外径から内輪の内径を差し引いた値の１／２）をＨとすると玉の直径ＤａがＤａ＝０．４３７Ｈ〜０．５５６Ｈである場合には、内輪と外輪との温度差に対する内輪ＰＶ値の変化率がほぼ同等となるので、玉の直径ＤａをＤａ＝０．４３７Ｈ〜０．５５６Ｈに設定することにより、内輪と外輪との間に温度差があってもＰＶ値を低くすることができる。
【００４６】
図１１は、本発明の第１実施形態に係るグリース潤滑される玉軸受の部分断面図である。本実施形態の玉軸受（アンギュラ玉軸受）１０は、工作機械の主軸支持用の外輪カウンタボア軸受である。同図に示すように、本発明の一実施形態に係る玉軸受１０は、リング状に形成された内輪１１と、この内輪１１の外周に同心円状に配置された外輪１２とを備えており、内輪１１及び外輪１２のうち少なくとも一方は、重量比でＣを０．２〜１．２％、Ｓｉを０．４〜１．５％、Ｍｏを０．５〜１．５％、Ｃｒを０．５〜２．０％、残部Ｆｅ及び不可避的不純物元素を含有する鉄鋼材料からなり、表面炭素濃度が０．３〜１．３％、表面窒素濃度が０．２〜０．８％となるように、浸炭窒化処理が施されている。また、玉軸受１０は内輪１１と外輪１２との間に転動自在に組み込まれた複数の玉１３と、これらの玉１３を保持する保持器１４とを備えており、内輪１１の外周面には内側軌道溝１１ａが形成されている。この内側軌道溝１１ａは外輪１２の内周面に形成された外側軌道溝１２ａと対向しており、これらの軌道溝１１ａ，１２ａは幅方向断面が円弧状に形成されている。
【００４７】
また図１２は、本発明の一実施形態に係る玉軸受の玉と軌道溝との関係を示す図である。図に示すように、内輪１１の外径面には、その円周方向に沿って所定曲率半径の内側軌道溝１１ａが形成され、外輪１２の内径面には、その円周方向に沿い且つ上記内側軌道溝１１ａと径方向で対向した所定曲率半径の外側軌道溝１２ａが形成され、その内側軌道溝１１ａと外側軌道溝１２ａとによって形成される空間内に、複数の玉１３が円周方向に並んで介挿されている。そして、各玉１３は、上記軌道溝１１ａ，１２ａにそれぞれ転がり接触する。そして、本実施の形態では、上記玉１３の直径をＤａとし、上記外側軌道溝１２ａの曲率半径をＲｏ、上記内側軌道溝１１ａの曲率半径をＲｉとした場合に、Ｒｉ＝０．５３５Ｄａ〜０．５４５Ｄａ、Ｒｏ＝０．５１５Ｄａ〜０．５４５Ｄａを満足するように、転動体である玉１３の直径Ｄａや各軌道溝１１ａ，１２ａの曲率半径Ｒｉ，Ｒｏを決定したものである。
【００４８】
また、本実施形態においては、外輪１２のカウンタボア側（図では右側）に、外輪１２を径方向に貫通する補給要素としての補給孔１５が設けられている。補給孔１５は、直径０．１〜５ｍｍの円形断面を有している。補給孔１５は、外輪１２の内径面の、外輪軌道１２ａに隣接する箇所に開口している。
補給孔１５は、外輪１２の周方向に間隔をあけた複数箇所に設けられてもよい。なお、補給孔１５は、接触部１２ｂのある側に設けられてもよく、接触部１２ｂ以外の部分に設ければよい。
【００４９】
アンギュラ玉軸受１０の軸受空間には、軸受空間容積の１０〜２０％の量のグリースが初期封入される。そして、軸受使用時には、次のようなグリース補給方法が適用される。すなわち、適宜なタイミングで（間欠的、定期的に）、補給孔１５を介して、一回の補給量が軸受空間容積の４％以下となるようにグリースショットさせる。なお、一回の補給量は、０．００４ｃｃ〜０．１ｃｃであることが望ましい。ここで、軸受空間容積とは、外輪内径と内輪外径との間にできる空間から、転動体の体積及び保持器の体積を差し引いた容積を意味する。
【００５０】
（実施例）
図１１に示した構成に基づく図１３に示された背面組合せの一対のアンギュラ玉軸受を用いて、本発明の実施例と比較例を表４に示す。また、表４の比較例Ａ、Ｂおよび実施例Ａ、Ｂの内輪を下記の評価条件で５０００ｍｉｎ^−１〜３１０００ｍｉｎ^−１の速度で回転させたときの外輪の温度変化を図１４に示す。
【００５１】
【表４】

【００５２】
［評価条件］
組合せ ２列背面組合せ（背面組合せ定位置予圧）
潤滑 オイルエア潤滑 給油量 ０．０３ｃｃ／４ｍｉｎ／１Ｂｒｇ
組込時アキシアルぱね定数 １００Ｎ／μｍ
（予圧の影響を受けないように同じばね定数に設定）
駆動方法 ベルト
冷却方法 外筒冷却 冷却油温室温＋３℃（室温２２℃）、６ＮＬ／ｍｉｎ
評価方法 温度上昇、焼き付き限界
図１４に示すように、玉径を７．９３８ｍｍ、玉数を２６個、内側軌道溝および外側軌道溝の幅方向断面の曲率半径をＲｉ，Ｒｏ＝０．５４Ｄａ、内側軌道溝および外側軌道溝に対する玉の接触角を１５°、内外輪材質をＳＵＪ２、玉材質をセラミックとした玉軸受（比較例Ａ）の場合には、内輪の回転速度が２４０００ｍｉｎ^−１を超えると外輪温度が急激に上昇し、内輪の回転速度が２７０００ｍｉｎ^−１に達すると焼付きを起こすことがわかる。これに対し、玉径を８．７３１ｍｍ、玉数を２５個、内側軌道溝の幅方向断面の曲率半径をＲｉ＝０．５８Ｄａ、外側軌道溝の幅方向断面の曲率半径をＲｏ＝０．５４Ｄａ、内側軌道溝および外側軌道溝に対する玉の接触角を１８°、内外輪材質をＳＨＸ（日本精工株式会社商品名、Ｃ；０．２〜１．２重量％、Ｓｉ；０．７〜１．５重量％、Ｍｏ；０．５〜１．５重量％、Ｃｒ；０．５〜２．０重量％、残部Ｆｅ及び不可避的不純物元素を含有する鉄鋼材料、表面炭素濃度：０．８〜１．３％、表面窒素濃度：０．２〜０．８％）、玉材質をセラミックとした玉軸受（比較例Ｂ）の場合には、内輪の回転速度が３１０００ｍｉｎ^−１に達しても焼付きを起さないことがわかる。
【００５３】
また、玉径を８．７３１ｍｍ、玉数を２５個、内側軌道溝および外側軌道溝の幅方向断面の曲率半径をＲｉ，Ｒｏ＝０．５４Ｄａ、内側軌道溝および外側軌道溝に対する玉の接触角を１８°、内外輪材質をＳＵＪ２、玉材質をセラミックとした玉軸受（実施例Ａ）の場合や、玉径を８．７３１ｍｍ、玉数を２５個、内側軌道溝および外側軌道溝の幅方向断面の曲率半径をＲｉ，Ｒｏ＝０．５４Ｄａ、内側軌道溝および外側軌道溝に対する玉の接触角を１８°、内外輪材質をＳＨＸ、玉材質をセラミックとした玉軸受（実施例Ｂ）の場合も比較例Ｂと同様に、内輪の回転速度が３１０００ｍｉｎ^−１に達しても焼付きを起さないことがわかる。
【００５４】
以上のことから、内側軌道溝の幅方向断面の曲率半径をＲｉ＝０．５３５Ｄａ〜０．５４５Ｄａ、外側軌道溝の幅方向断面の曲率半径をＲｏ＝０．５１５Ｄａ〜０．５４５Ｄａ、内側軌道溝および外側軌道溝に対する玉の接触角を１７°〜２６°、玉の直径を軸受断面高さＨ（外輪の外径から内輪の内径を差し引いた値の１／２）に対して０．４Ｈ〜０．５６Ｈに設定することにより、高速回転に伴う発熱量を抑制して焼付きを防止できると共に転がり疲労寿命を長くすることができ、かつ低コスト化を図ることができる。
【００５５】
さらに、図１１に示した第１実施形態のアンギュラ玉軸受を用いて、以下のような実験を行った。
内径６５ｍｍ、外径１００ｍｍ、幅１８ｍｍ、接触角１８°、玉径７．１４４ｍｍ、使用グリースはイソフレックスＮＢＵ１５（ＮＯＫクリューバー（株）製）、グリース初期封入量は軸受空間容積の１５％（２．３ｃｃ）である、一対のアンギュラ玉軸受により、試験機の回転主軸を支持した。一対のアンギュラ玉軸受は背面組合せとし、軸受間隔は１００ｍｍとした。
【００５６】
（実験１）
試験機の回転主軸をいくつかの設定速度で回転させ、各設定速度における軸受耐久時間を観測した。結果を表１に示す。
【表５】

グリース補給がない状態では、ｄｍｎ１２０万で１００００時間を越える耐久時間を達成できなかった。これは、グリース寿命が１００００時間であったことを示している。ｄｍｎ１５０万ではグリース寿命が１０００時間であった。ｄｍｎ１８０万ではグリース寿命が１００時間であった。
表５の結果をグラフ化したものを図１５に示す。図１５から、ｄｍｎの上昇に伴って、グリース寿命時間が指数関数的に短くなることがわかる。
【００５７】
（実験２）
実験１のときと同様な条件で、試験機の回転主軸を一対のアンギュラ玉軸受で支持し、適宜な慣らし運転を行った。慣らし運転後のｄｍｎ１５０万における外輪温度は５５℃であった。なお、ｄｍｎ１５０万におけるグリース寿命は、前述のように１０００時間である。
本実験では、ｄｍｎ１５０万で１０００時間運転を行った時点で、いくつかの設定補給量のグリース補給を行い、その後０からｄｍｎ１５０万に５秒で立ち上げて運転を再開するとともに軸受を観察することで、各設定補給量における耐久時間及び軸受の昇温を測定した。結果を表６に示す。
【表６】

補給量が４％以下のときは、初期封入のグリース寿命の半分である５００時間を越えると、焼付が生じた。
一方、補給量が４％を超えると、グリース補給後に軸受が異常昇温（外輪が７５℃以上に急上昇）した。
以上の結果から、グリースの補給間隔は、初期封入のグリース寿命時間の半分以下とし、補給量は軸受空間容積の４％以下とするのが良いことがわかった。
【００５８】
なお、２％のグリース補給を行った後、慣らし運転をした時点で、軸受内部の接触部の脇にかき出されたグリースを採取し、量を測定したところ、２％の量の補給に対し、１．９％の量のグリースがかき出されたことがわかった。これは、転動体（玉）、内外輪の軌道溝及び保持器間に必要最低限のグリースが潤滑膜として保持され、余分なグリースがかき出されたことを示す。
このことから、グリース補給量の下限は軸受空間容積の０．１％であることがわかった。
【００５９】
図１６は、本発明の第２実施形態であるアンギュラ玉軸受２０を示す。なお、内輪軌道溝と外輪軌道溝の幅方向断面の曲率半径、玉の接触角度、玉の直径、グリース潤滑の一回の補給量は第１実施形態と同様である。
アンギュラ玉軸受２０は、内輪２１、外輪２２、内外輪２１の内輪軌道２１ａと外輪２２の外輪軌道２２ａとの間に複数配置された玉２３及び玉２３を円周方向等間隔に保持する保持器２４を備えている。
本実施形態においては、外輪２２のカウンタボア側（図では右側）に、外輪２２を径方向に貫通する補給要素としての補給孔２５が設けられている。補給孔２５の外輪内径面側は、グリースだまり２５ａになっている。グリースだまり２５ａの断面積は、補給孔２５の他の部分の断面積より大きい。グリースだまり２５ａを有しているので、補給孔２５は、段付き円柱状空間になっている。グリースだまり２５ａは、外輪２２の内径面の、外輪軌道２２ａに隣接する箇所に位置している。以下に説明する実施形態においても、補給孔がグリースだまりを有してもよい。
【００６０】
図１７は、本発明の第３実施形態に係るアンギュラ玉軸受３０を示す。なお、内輪軌道溝と外輪軌道溝の幅方向断面の曲率半径、玉の接触角度、玉の直径、グリース潤滑の一回の補給量は第１実施形態と同様である。
アンギュラ玉軸受３０は、内輪３１、外輪３２、内輪３１の内輪軌道３１ａと外輪３２の外輪軌道３２ａとの間に複数配置された玉３３及び玉３３を円周方向等間隔に保持する保持器３４を備えている。本実施形態は、内輪カウンタボア軸受である。
本実施形態においては、外輪３２の外輪軌道３２ａの、接触部３２ｂのある側（図では右側）の反対側に、外輪３２を径方向に貫通する補給要素としての補給孔３５が開口している。尚、補給孔３５は、接触部３２ｂのある側であってもよく、接触部３２ｂ以外の部分に設ければよい。
【００６１】
図１８は、本発明の第４実施形態に係るアンギュラ玉軸受４０を示す。なお、内輪軌道溝と外輪軌道溝の幅方向断面の曲率半径、玉の接触角度、玉の直径、グリース潤滑の一回の補給量は第１実施形態と同様である。
アンギュラ玉軸受４０は、内輪４１、外輪４２、内輪４１の内輪軌道４１ａと外輪４２の外輪軌道４２ａとの間に複数配置された玉４３及び外輪案内の保持器４４を備えている。本実施形態は、外輪カウンタボア軸受である。
本実施形態においては、外輪４２のカウンタボア側（図では右側）に、外輪４２を径方向に貫通する補給要素としての補給孔４５が設けられている。補給孔４５は、保持器４４の片側（図では右側）の案内面４４ａに向けて開口している。
【００６２】
図１９は、本発明の第５実施形態に係るアンギュラ玉軸受５０を示す。なお、内輪軌道溝と外輪軌道溝の幅方向断面の曲率半径、玉の接触角度、玉の直径、グリース潤滑の一回の補給量は第１実施形態と同様である。
アンギュラ玉軸受５０は、内輪５１、外輪５２、内輪５１の内輪軌道５１ａと外輪５２の外輪軌道５２ａとの間に複数配置された玉５３及び外輪案内の保持器５４を備えている。本実施形態は、外輪カウンタボア軸受である。
本実施形態においては、外輪５２の反カウンタボア側（図では左側）に、外輪５２を径方向に貫通する補給要素としての補給孔５５が設けられている。補給孔５５は、保持器５４の片側（図では左側）の案内面５４ａに向けて開口している。
【００６３】
図２０は、本発明の第１〜５実施形態に記載のアンギュラ玉軸受を用いて構成される工作機械用主軸装置としてのスピンドル装置を示す図である。ここでは、例として第１実施形態のアンギュラ玉軸受１０を用いている。
【００６４】
軸受１０は、主軸１に外嵌し、そしてハウジング７に内嵌している。主軸１は、軸受１０を介して、ハウジング７に対し回転可能である。軸受１０の各内輪及び外輪間には、それぞれ主軸１及びハウジング７に沿って配置された内輪間座５及び外輪間座６が配置されている。内輪間座５及び外輪間座６の軸方向両端には、それぞれ内輪押さえ部材８及び外輪押さえ部材９が配置され、各間座を介して各軸受に予圧を与えている。内輪押さえ部材８及び外輪押さえ部材９の間には、図示せぬ間隙が形成されており、両押さえ部材間にラビリンスを形成している。
【００６５】
ハウジング７には、ハウジング７を貫通し、各軸受１０の外輪に形成された補給孔に追加グリースを補給するノズル（グリース供給こま）４が固定されている。グリースは、グリース補給器２から補給パイプ３を介してノズル４に供給され、そして径方向に軸受内部に補給される。グリース補給器２は、適宜なタイミングで（間欠的、定期的に）、一回の補給量が軸受空間容積の０．１〜４％となるようにグリースショットする。
【００６６】
図２０では、第１実施形態の軸受１０を例として挙げたが、勿論その他の実施形態２〜５、又はそれらの任意の組合せを代わりに用いてもよい。
また、その他の軸受の外輪に同様の補給孔を設けても同様の効果が期待されることは言うまでもない。
【００６７】
図２１は、以下に説明する第６実施形態に係る転がり軸受２００を用いて構成される工作機械用主軸装置としてのスピンドル装置を示す図である。
【００６８】
軸受２００は、主軸１に外嵌し、ハウジング７に内嵌している。主軸１は、軸受２００を介して、ハウジング７に対し回転可能である。軸受２００の各内輪及び外輪間には、それぞれ主軸１及びハウジング７に沿って配置された内輪間座５００ａ，５００ｂ，５００ｃ，５００ｄ，５００ｅ及び外輪間座６００ａ，６００ｂ，６００ｃ，６００ｄ，６００ｅが図視左から順に配置されている。
【００６９】
内輪間座５００ａ及び５００ｅ並びに外輪間座６００ａ及び６００ｅの軸方向両端には、それぞれ内輪押さえ部材８ａ，８ｂ及び外輪押さえ部材９ａ，９ｂが配置され、各間座を介して各軸受に予圧を与えている。内輪押さえ部材８ａ及び外輪押さえ部材９ａ並びに内輪押さえ部材８ｂ及び外輪押さえ部材９ｂの間には、図示せぬ間隙が形成されており、両押さえ部材間にラビリンスを形成している。
【００７０】
図２２は、図２０に示すスピンドル装置の拡大断面図である。ここでは、本発明の第６実施形態に係るアンギュラ玉軸受２００並びにその周辺構造について説明する。
【００７１】
図２２に示す各アンギュラ玉軸受２００は、内輪２０１、外輪２０２、内輪２０１の内輪軌道２０１ａと外輪２０２の外輪軌道２０２ａとの間に複数配置された玉２０３、及び、玉２０３を円周方向等間隔に保持する保持器２０４を備えている。外輪２０２は、玉２０３を接触角を持って保持するためのテーパ部２０２ｂを軸方向片側に有している。以下、テーパ部が形成された軸方向一方を正面側、他方を背面側と呼ぶこととする。
【００７２】
本実施形態においては、各アンギュラ玉軸受２００間には、グリース補給用外輪間座６００ｂが配置されている。グリース補給用外輪間座６００ｂには、ハウジング７を貫通した二つのグリース補給用ノズル４が、グリース補給用外輪間座６００ｂに差し込み固定されている。グリース補給用ノズル４には、外部のグリース供給器２から補給パイプ３を介して追加グリースが供給される。
【００７３】
グリース補給用外輪間座６００ｂは、ノズル４の先端から追加グリースをアンギュラ玉軸受２００内部に補給する補給要素としての補給孔２０５を有している。補給孔２０５は、直径０．１〜５ｍｍの円形断面を有しており、軸受２００の内側（保持器２０４よりも内径側）に向けて軸方向に開口している。補給孔２０５は、内輪２０１及び外輪２０２間に背面側から軸方向に追加グリースを供給する。供給されるグリースは、主に保持器２０４よりも内径側に供給される。
【００７４】
なお、補給孔２０５は、径方向に間隔をあけてグリース補給用外輪間座６００ｂの複数箇所に設けられてもよい。また、供給されるグリースは、主に保持器２０４よりも内径側に供給されるほうが好ましいが、外径側に供給してもよい。
【００７５】
各アンギュラ玉軸受２００の軸受空間には、軸受空間容積の１０〜２０％の量のグリースが初期封入される。そして、軸受使用開始後、グリース供給器２は、適宜なタイミングで（間欠的、定期的に）、補給孔２０５を介して、一回の補給量が軸受空間容積の０．１〜４％となるようにグリースショットする。軸受内部に追加補給されたグリースは、玉２０３の転動に伴い、軸受内部全体に馴染み、不足したグリースを補う。これにより、グリースの欠如による軸受の破損を防止し、軸受の長寿命化が図られる。
なお、グリース補給用外輪間座６００ｄにも、ハウジング７を貫通したグリース補給用ノズル４が差し込み固定されており、上記と同様に構成されている。
【００７６】
図２３は、第６実施形態の第１の変形例に係るスピンドル装置の拡大断面図を示す。
本変形例に用いられているアンギュラ玉軸受２２０は、軸に外嵌する内輪２２１、ハウジング１０００に内嵌する外輪２２２、内輪２２１の内輪軌道２２１ａと外輪２２２の外輪軌道２２２ａとの間に転動自在に配置された玉２２３、並びに玉２２３を保持する保持器２２４から構成される。
【００７７】
ハウジング１０００は、径方向内側に突出した凸部１０００ａを有する。軸受２２０の外輪２２２は、軸方向背面側で凸部１０００ａに接している。内輪２２１の軸方向背面側には、凸部１０００ａと軸方向に対向する内輪間座５１０ａが配置されている。
【００７８】
一方、外輪２２２の軸方向正面側には、グリース補給用外輪間座６１０が設けられている。グリース補給用外輪間座６１０は、内輪間座５１０ｂと軸方向に対向している。ハウジング１０００におけるグリース補給用外輪間座６１０の外径面に対応する位置には、グリース補給用ノズル４００をグリース補給用外輪間座６１０に差し込むための開口１０００ｂが形成されている。グリース補給用ノズル４００の基部４００ａは、ねじ等の固定部材４００ｂによりハウジング１０００の外径面上に固定されており、基部４００ａから延出した先端部４００ｃがグリース補給用外輪間座６１０内部に差し込まれている。
【００７９】
グリース補給用外輪間座６１０は、グリース補給用ノズル４００の先端部４００ｃから追加グリースを軸受２２０内部に補給する補給要素としての補給孔２２５を有している。補給孔２２５は、直径０．１〜５ｍｍの円形断面を有している。補給孔２２５は、内輪２２１及び外輪２２２間に正面側から軸方向に追加グリースを供給する。これにより、グリースの欠如による軸受の破損を防止し、軸受の長寿命化が図られる。
なお、補給孔２２５は、径方向に間隔をあけてグリース補給用外輪間座６１０の複数箇所に設けられてもよい。
【００８０】
図２４は、第６実施形態の第２の変形例に係るスピンドル装置の拡大断面図を示す。
本変形例に用いられているアンギュラ玉軸受２３０は、軸に外嵌する内輪２３１、ハウジング１１００に内嵌する外輪２３２、内輪２３１の内輪軌道２３１ａと外輪２３２の外輪軌道２３２ａとの間に転動自在に配置された玉２３３、並びに玉２３３を保持する保持器２３４から構成される。
【００８１】
ハウジング１１００は、径方向内側に突出した凸部１１００ａを有する。軸受２３０の外輪２３２は、軸方向正面側で凸部１１００ａに接している。内輪２３１の正面側には、凸部１１００ａに軸方向に対向する内輪間座５２０ｂが配置されている。一方、外輪２２２の軸方向背面側には、内輪間座５２０ａ及び外輪間座６２０がそれぞれに対向して配置されている。
【００８２】
凸部１１００ａの反対側となるハウジング１０００の外径面には、グリース補給用ノズル４００を凸部１１００ａ内に差し込むための開口１１００ｂが形成されている。グリース補給用ノズル４００の基部４００ａは、ねじ等の固定部材４００ｂによりハウジング１１００の外径面上に固定されており、基部４００ａから延出した先端部４００ｃが凸部１１００ａ内部に差し込まれている。
【００８３】
凸部１１００ａは、グリース補給用ノズル４００の先端部４００ｃから追加グリースを軸受２３０内部に補給する補給要素としての補給孔２３５を有している。補給孔２３５は、直径０．１〜５ｍｍの円形断面を有している。補給孔２３５は、内輪２３１及び外輪２３２間に正面側から軸方向に追加グリースを供給する。これにより、グリースの欠如による軸受の破損を防止し、軸受の長寿命化が図られる。
なお、補給孔２３５は、径方向に間隔をあけて凸部１１００の複数箇所に設けられてもよい。
【００８４】
また、図２５に第６実施形態の第３の変形例に係るスピンドル装置の拡大断面図を示す。
本変形例は、第２変形例のアンギュラ玉軸受２３０の正面側と背面側を入れ替えたものであり、ハウジング１１００の凸部１１００ａがアンギュラ玉軸受２３０の軸方向背面側に設けられている。その他の構成は、図２４に示したものと同様である。
【００８５】
本変形例において、追加グリースは、凸部１１００ａに形成された補給孔２３５から、内輪２３１及び外輪２３２間に背面側から軸方向に供給される。これにより、グリースの欠如による軸受の破損を防止し、軸受の長寿命化が図られる。
【００８６】
図２６は、第６実施形態の第４の変形例に係るスピンドル装置の拡大断面図を示す。
本変形例に用いられているアンギュラ玉軸受２４０は、軸に外嵌する内輪２４１、ハウジング１２００に内嵌する外輪２４２、内輪２４１の内輪軌道２４１ａと外輪２４２の外輪軌道２４２ａとの間に転動自在に配置された玉２４３、並びに玉２４３を保持する保持器２４４から構成される。外輪１２００の正面側端部には、テーパ部から径方向内側に突出した凸部２４２ｂが形成されている。
【００８７】
軸受２３０の外輪２３２は、軸方向正面側で、即ち、凸部２４２ｂが、外輪間座６３０ｂと接しており、軸方向背面側で、外輪間座６３０ａと接している。内輪２３１の背面側及び正面側には、それぞれ外輪間座６３０ａ及び６３０ｂと径方向に対向する内輪間座５３０ａ及び５３０ｂが配置されている。
【００８８】
ハウジング１２００は、外輪２４２の凸部２４２ｂの反対側となる外径面に、グリース補給用ノズル４００を凸部２４２ｂ内に差し込むための開口１２００ｂを有している。グリース補給用ノズル４００の基部４００ａは、ねじ等の固定部材４００ｂによりハウジング１２００の外径面上に固定されており、基部４００ａから延出した先端部４００ｃが、開口１２００ｂを介して外輪２４２の凸部２４２ｂ内部に差し込まれている。
【００８９】
凸部２４２ｂは、グリース補給用ノズル４００の先端部４００ｃから追加グリースを軸受２４０内部に補給する補給要素としての補給孔２４５を有している。補給孔２４５は、直径０．１〜５ｍｍの円形断面を有している。補給孔２４５は、内輪２４１及び外輪２４２間に正面側から軸方向に追加グリースを供給する。これにより、グリースの欠如による軸受の破損を防止し、軸受の長寿命化が図られる。
なお、補給孔２４５は、径方向に間隔をあけて凸部２４２ｂの複数箇所に設けられてもよい。
【００９０】
また、図２７に第６実施形態の第５の変形例に係るスピンドル装置の拡大断面図を示す。
本変形例は、第４変形例のアンギュラ玉軸受２４０の外輪２４２の変形例であり、外輪２４２の凸部２４２ｂが、アンギュラ玉軸受２４０の軸方向背面側に形成されている。その他の構成は、図２６に示したものと同様である。
【００９１】
本変形例において、追加グリースは、凸部２４２ｂに形成された補給孔２４５から、内輪２４１及び外輪２４２間に背面側から軸方向に供給される。これにより、グリースの欠如による軸受の破損を防止し、軸受の長寿命化が図られる。
【００９２】
上記した第６実施形態および第６実施形態の変形例１〜５のように構成することにより、軸方向に追加グリースを軸受内部に補給することが可能となる。
また、その他の軸受において、同様の補給孔を設けても同様の効果が期待されることは言うまでもない。
【００９３】
（実験３）
次に、図２８に示す回転試験装置（６５ミリ・アンギュラ玉軸受用中間Ｓ／Ｐ）を用いて軸方向からグリース追加供給を行う場合の耐久試験を行った。図２８において、軸１は背面組合せ（ＤＢ）のアンギュラ玉軸受２５０，２５０によりハウジング１５００に対して回転可能に構成されている。
【００９４】
図２９は、図２８のアンギュラ玉軸受２５０及びその周辺部材を示す拡大断面図である。アンギュラ玉軸受２５０は、軸１に固定された内輪２５１、ハウジングに固定された外輪２５２、内輪２５１の内輪軌道２５１ａと外輪２５２の外輪軌道２５２ａとの転動自在に配置された転動体２５３、及び転動体２５３を保持する保持器２５４から構成されている。
【００９５】
各アンギュラ玉軸受２５０の正面側には、内輪間座５５０ａ及び外輪間座６５０ａが、そして背面側には、内輪間座５５０及びグリース供給用外輪間座６５０ｂが、それぞれ内輪２５１及び外輪２５２に隣接配置されている。グリース供給用外輪間座６５０ｂ間には、外輪間座６６０が配置されている。
【００９６】
グリース供給用外輪間座６５０ｂには、ハウジング１５００を介してグリース供給用ノズル４５０が差込固定されている。グリース供給用ノズル４５０には、図示せぬグリース補給器から補給パイプ３を介してグリースが供給される。グリース供給用外輪間座６５０は、アンギュラ玉軸受２５０の正面側に開口した補給孔２５５（φ２．０）を有している。グリース供給用ノズル４５０に供給されたグリースは、先端部４５０ａから補給孔２５５に供給され、軸方向に軸受の軸受空間内に供給される。
【００９７】
実験では、本発明に係る玉軸受の構成である、内径６５ｍｍ、外径１００ｍｍ、幅１８ｍｍ、玉径７．１４４ｍｍ、及び接触角１８°のアンギュラ玉軸受を用いた。潤滑に用いたグリースは、イソフレックスＮＢＵ１５であり、グリース初期封入量は、軸受空間容積の１５％とした。両軸受２５０，２５０は、２０，０００ｍｉｎ^−１（ｄｍｎ１８０万）で回転させた。
【００９８】
実験では、比較のため、グリースの追加供給を行わない試験１とグリースの追加供給を行う試験２を行った。試験２では、回転開始後５０時間経過する毎に、０．３ｃｃ（軸受空間容積の１．５％に相当）のグリースを補給孔２５５を介して軸受空間内に追加ショットした。
【００９９】
実験を行った結果、試験１では、実験開始後５００時間後に軸受が焼き付いてしまったため、実験を中止した。一方、試験２では、実験開始後３０００時間が経過しても異常が発生することは無く、無事に実験を終了した。この実験結果から、グリースを補給した場合には、３０００時間以上経過しても何ら軸受に問題は発生せず、軸受の寿命が大幅に延びることが確認された。
【０１００】
次に、本発明の第１〜６実施形態に記載のアンギュラ玉軸受において、一回のグリース補給量を変更した本発明の第７実施形態について説明する。
上記第１〜６実施形態では、焼付きの発生を防止するため、一回のグリース補給量が軸受空間容積の０．１〜４％となるように設定されていたが、さらに、金型用途向けの工作機械等、精度が厳しく要求される装置の主軸に用いられる軸受においては、補給した瞬間の温度の脈動が生じないようにすることが望まれる。このため、第７実施形態においては、一回のグリース補給量を０．００４ｃｃ〜０．１ｃｃとなるように補給する。
【０１０１】
アンギュラ玉軸受の軸受空間には、軸受空間容積の１０〜２０％の量のグリースが初期封入される。ここで、軸受空間容積とは、外輪内径と内輪外径との間にできる空間から、転動体の体積及び保持器の体積を差し引いた容積を意味する。そして、軸受使用時には、次のようなグリース補給方法が適用される。すなわち、適宜なタイミングで（間欠的、定期的に）、補給孔１５を介して、一回の補給量が０．００４ｃｃ〜０．１ｃｃとなるようにグリースショットさせる。グリースショットのバラツキを考慮すると、一回の補給量の上限は、０．１２ｃｃである。また、一回のグリース補給量は、温度の脈動の発生防止を考慮すると、０．０１ｃｃ〜０．０３ｃｃであることが好ましい。以上に示す範囲内のグリースショットを行うことで、グリースの劣化または油膜形成不足による異常昇温の発生および軸受の破損を防止するとともに、グリース補給時の温度の脈動を抑制し、アンギュラ玉軸受が取り付けられる工作機械の軸精度の劣化を防ぐことができる。なお、上記第７実施形態は、上記第１〜第６実施形態に示されたアンギュラ玉軸受の構成のいずれにも適用可能である。
【０１０２】
次に、本発明の第７実施形態について、 内輪内径６５ｍｍ、外輪外径１００ｍｍ、幅１８ｍｍ、玉径７．１４４ｍｍ、接触角１８°のアンギュラ玉軸受７４０，７５０を用いて、グリース補給量と温度の脈動の関係を確認する温度脈動確認試験を行った。温度脈動確認試験に用いられたグリースは、イソフレックスＮＢＵ１５であり、グリースの初期封入量は、軸受空間容積の１５％であった。また、試験条件は、ｄｍｎ＝１８０万であった。
【０１０３】
本脈動確認試験は、図３０に示す試験用主軸装置７００を用いて行われた。試験用主軸装置７００は、支持台７０１上に配置されたハウジングブロック７０２によってハウジング本体７０３が支持される構成を有している。ハウジング本外７０３には、アンギュラ玉軸受７４０，７５０が互いに背面配置で内嵌している。このアンギュラ玉軸受主軸７４０は、主軸７１０に外嵌しており、主軸７１０を回転自在に支承している。
【０１０４】
アンギュラ玉軸受７４０，７５０間には、各アンギュラ玉軸受７４０，７５０の内輪間に内輪間座７１１が、そして各アンギュラ玉軸受７４０，７５０の外輪間に外輪間座７１２がそれぞれ設けられている。また、アンギュラ玉軸受７５０の軸方向後端側（図中右側）には、後端外輪押さえ７１３が設けられている。
また、アンギュラ玉軸受７４０の軸方向前端側（図中左側）には、外輪押さえ部材７１４と内輪押さえ部材７１５が設けられている。各アンギュラ玉軸受７４０，７５０は、外輪押さえ部材７１４及び内輪押さえ部材７１５によって後端外輪押さえ７１３側に軸方向に沿って押圧されている。
【０１０５】
ハウジング７０３には、ノズル部材７３０，７３０が各アンギュラ玉軸受７４０，７５０に対応して取り付けられている。ノズル部材７３０、７３０は、各アンギュラ玉軸受７４０，７５０の外輪に設けられた孔側からグリースを軸受空間内に供給する。本温度脈動確認試験では、グリースは、試験開始後１時間おきに供給するような構成とした。一回のグリース補給時に各軸受に供給されるグリースの量は、０．０３５ｃｃ、０．１０ｃｃ、０．１５ｃｃ、０．３０ｃｃ、０．６０ｃｃとして計５回実験を行った。図３１は、この温度脈動確認試験の結果を示すグラフであり、（ａ）はグリース補給量が０．０３５ｃｃの場合、（ｂ）はグリース補給量が０．１０ｃｃの場合、（ｃ）はグリース補給量が０．１５ｃｃの場合、（ｄ）はグリース補給量が０．３０ｃｃの場合、そして（ｅ）はグリース補給量が０．６０ｃｃの場合を示す。
【０１０６】
図３１（ａ），（ｂ）に示すように、グリース補給量が０．０３５ｃｃ及び０．１０ｃｃの場合には、グリース補給をおこなっても、アンギュラ玉軸受７４０，７５０の軸受温度に全く変化はない。しかしながら、図３１（ｃ）に示すように、グリース補給量が０．１５ｃｃの場合には、グリース補給する毎にアンギュラ玉軸受７４０，７５０の温度が１℃程度上昇しているのがわかる。同様に、図３１（ｄ），（ｅ）に示すように、グリース補給量が０．３０ｃｃ及び０．６０ｃｃの場合にも、グリース補給する毎にアンギュラ玉軸受７４０，７５０の温度が１〜２℃程度上昇しているのがわかる。
【０１０７】
ここで、アンギュラ玉軸受７４０と軸受７５０では、グリース補給前の定常状態での温度が異なっている。この定常状態での温度の差異は、アンギュラ玉軸受７４０と、アンギュラ玉軸受７５０の周辺構造の差異、例えば、ハウジングブロック７０２からの距離の差異や図示せぬ冷却装置との位置関係等により発生した熱の抑制率が異なっているため、定常状態での温度が異なっているものと考えられる。
【０１０８】
いずれにせよ、アンギュラ玉軸受７４０においても、またアンギュラ玉軸受７５０においても、図３１（ｃ）〜図３１（ｅ）に示すように、グリース補給量が０．１５ｃｃ以上の場合には、補給グリースの攪拌抵抗等により軸受が発熱を起こし、軸受の昇温、すなわち温度の脈動が発生していると考えられる。したがって、一回のグリース補給量を０．１ｃｃ以下と設定することにより、温度の脈動が発生しない安定したグリース供給を行うことが可能であることがわかった。
【０１０９】
【発明の効果】
以上説明したように、本発明によれば、高速回転に伴う発熱量を抑制しかつ高剛性で焼付きを防止し、転がり疲労寿命を長くすることができると共に工作機械主軸用軸受として用いる場合の静的アキシアル負荷容量を確保し、さらに環境面、コスト面で有利なグリース潤滑の玉軸受を提供できる。また、この転がり軸受を用いて、高い信頼性を有する工作機械用主軸装置を実現することができる。
【図面の簡単な説明】
【図１】玉軸受の回転時のアキシアルばね定数を示す図である。
【図２】玉軸受の回転時の内輪ＰＶ値を示す図である。
【図３】玉軸受の回転時の内輪ＰＶ値と外輪ＰＶ値を示す図である。
【図４】玉軸受の発熱量を示す図である。
【図５】玉軸受の転がり疲労寿命を示す図である。
【図６】玉軸受の静的許容アキシアル荷重を示す図である。
【図７】玉軸受の内輪ＰＶ値と接触角との関係を示す図である。
【図８】玉軸受の発熱量と接触角との関係を示す図である。
【図９】玉軸受の外輪ＰＶ値と接触角との関係を示す図である。
【図１０】玉軸受の内輪ＰＶ値と内外輪温度差との関係を示す図である。
【図１１】本発明の第１実施形態を示す断面図である。
【図１２】本発明の第１実施形態に係る玉軸受の玉と軌道溝との関係を示す図である。
【図１３】本発明の一実施形態に係る玉軸受が適用された実施例を示す図である。
【図１４】玉軸受の外輪温度と内輪回転数との関係を示す図である。
【図１５】本発明の第１実施形態を用いた実験結果を示すグラフである。
【図１６】本発明の第２実施形態を示す断面図である。
【図１７】本発明の第３実施形態を示す断面図である。
【図１８】本発明の第４実施形態を示す断面図である。
【図１９】本発明の第５実施形態を示す断面図である。
【図２０】本発明の第１〜５実施形態に記載の転がり軸受を用いて構成されるスピンドル装置を示す断面図である。
【図２１】本発明の第６実施形態に記載の転がり軸受を用いて構成されるスピンドル装置を示す断面図である。
【図２２】図２１に示すスピンドル装置の拡大断面図であり、本発明の第６実施形態を示す図である。
【図２３】本発明の第６実施形態の第１変形例を示す断面図である。
【図２４】本発明の第６実施形態の第２変形例を示す断面図である。
【図２５】本発明の第６実施形態の第３変形例を示す断面図である。
【図２６】本発明の第６実施形態の第４変形例を示す断面図である。
【図２７】本発明の第６実施形態の第５変形例を示す断面図である。
【図２８】回転試験装置を示す図である。
【図２９】図２８の回転試験装置の拡大図である。
【図３０】本発明の第７実施形態における温度脈動試験に用いられた試験用主軸装置を示す図である。
【図３１】本発明の第７実施形態における脈動確認試験の結果を示すグラフであり、（ａ）はグリース補給量が０．０３５ｃｃの場合、（ｂ）はグリース補給量が０．１０ｃｃの場合、（ｃ）はグリース補給量が０．１５ｃｃの場合、（ｄ）はグリース補給量が０．３０ｃｃの場合、そして（ｅ）はグリース補給量が０．６０ｃｃの場合を示す。
【図３２】玉軸受の静的許容アキシアル荷重と内側軌道溝曲率半径との関係を示す図である。
【符号の説明】
１０，２０，３０，４０，５０ アンギュラ玉軸受（玉軸受）
１１，２１，３１，４１，５１ 内輪
１２，２２，３２，４２，５２ 外輪
１３，２３，３３，４３，５３ 玉
１４，２４，３４，４４，５４ 保持器
１５，２５，３５，４５，５５ 補給孔[0001]
TECHNICAL FIELD OF THE INVENTION
The present invention relates to a ball bearing used for various machine tools and the like, and a machine tool spindle device.
[0002]
[Prior art]
2. Description of the Related Art In recent years, various types of machine tools have been accelerated in rotation of a spindle and feed of peripheral devices centering on a machining center in order to improve productivity. Along with this, the ball bearings used as bearings for supporting the main shaft and the rotating shaft of peripheral devices described above are also required to be operated at higher speeds. Suppression, seizure prevention, and high rigidity are required for ball bearings.
[0003]
Conventionally, this type of ball bearing includes an inner ring formed in a ring shape, an outer ring concentrically arranged on the outer periphery of the inner ring, and a plurality of balls rotatably incorporated between the outer ring and the inner ring. And an inner raceway groove is formed on the outer peripheral surface of the inner race. The inner raceway groove is opposed to the outer raceway groove formed on the inner peripheral surface of the outer race, and these raceway grooves have an arc-shaped cross section in the width direction.
[0004]
In such a ball bearing, when the ball rolls on the raceway groove with the rotation of the inner ring or the outer ring, frictional heat is generated between the ball and the raceway groove. The calorific value due to the rotation of such a ball bearing shows a lower value as the product of the contact pressure at the contact surface between the ball and the raceway groove of the inner ring and the outer ring and the sliding speed at the aforementioned contact surface between the ball and the ball are smaller. It is clear that
[0005]
In the above-described conventional ball bearing, the radius of curvature of the cross-section in the width direction of the inner raceway groove and the outer raceway groove is set to about 0.52 Da when the diameter of the ball is Da, and the bearing cross-section height (the outer diameter of the outer race) If H is １ ／ of the value obtained by subtracting the inner diameter of the inner ring from), the diameter of the ball is set to about Da = 0.57H.
Further, in order to reduce the centrifugal force acting on the ball, a ball bearing whose diameter is set to Da = 0.40H to 0.50H or an inner raceway groove and an outer raceway groove to reduce the gyro moment. A ball bearing having a contact angle of the ball with respect to the angle of 15 ° is commercialized and used.
[0006]
Ball bearings for supporting machine tool spindles are often used at high speeds under a slight amount of oil lubrication. For this reason, oil film rupture is likely to occur, and the contact surface pressure between the ball and the raceway groove may increase due to the effect of the temperature rise of the bearing and the centrifugal force.
On the other hand, due to the above-mentioned high speed, seizure is more likely to occur in the ball bearing. In order to prevent this seizure, grease lubrication, which is easy to handle and is advantageous in terms of environment and cost, has conventionally been employed for lubrication of ball bearings.
[0007]
Grease-lubricated rolling bearings used for machine tool spindles are usually lubricated only with grease initially sealed to prevent heat generation. If the grease is sealed at the initial stage and rotated at high speed without performing the grease break-in operation, abnormal heat generation may occur due to the grease getting stuck or stirring resistance. I have to. As grease lubricated rolling bearings, there are disclosed a technology of providing a grease reservoir on the inner ring side and continuously supplying grease by centrifugal force, and a technology of effectively utilizing grease sealed in a bearing space by using air (Patent Documents) 1-9).
[0008]
In addition, rolling bearings such as ball bearings are sometimes used coaxially side by side. At this time, in order to prevent the relative position in the axial direction of each bearing from changing even during use, a gap between the bearings is conventionally prevented by a fixed position preload using a spacer or the like.
However, when the preload due to the fixed-position preload becomes excessive due to the centrifugal force and the temperature rise due to the high-speed rotation of the bearing, the contact surface pressure between the ball and the raceway groove increases, and seizure easily occurs. Therefore, even if the temperature changes, a constant pressure preload that keeps the load between the bearings substantially constant by using a spring, pneumatic pressure, or hydraulic pressure has been used to suppress the seizure of the bearing and extend the life.
[0009]
Further, a preload switching spindle is also used to secure the rigidity of the ball bearings and to reduce an increase in preload due to a centrifugal force and a temperature rise (effect of a temperature difference between the inner ring and the outer ring) due to high-speed rotation.
In terms of bearing specifications, for example,
{Circle around (1)} In order to suppress the PV value (P; contact surface pressure, V; sliding speed) and heat generation, the radius of curvature of the width direction cross section of the inner raceway groove is set larger than the radius of curvature of the width direction cross section of the outer raceway groove. (See Patent Document 10),
{Circle around (2)} In addition to the above, at least one of the inner ring and the outer ring has a weight ratio of C of 0.2 to 1.2%, Si of 0.7 to 1.5%, and Mo of 0.5 to 1. 5%, 0.5 to 2.0% of Cr, the balance being Fe and steel materials containing unavoidable impurity elements, and a carbonitriding treatment followed by a quenching and tempering treatment to reduce the surface carbon concentration to 0.8 to 2.0%. 1.3% and a low nitrogen material having a surface nitrogen concentration of 0.2 to 0.8% to further reduce heat generation (see Patent Document 11).
(3) In order to suppress the PV value and heat generation, the radius of curvature of the cross section of the inner raceway groove in the width direction is set to be larger than the radius of curvature of the cross section of the outer raceway groove in the width direction. The one that derives optimal parameters of heat generation and PV value (see Patent Document 12),
(4) In order to improve the heat dissipation of the inner ring and reduce the temperature difference from the outer ring, the width of the inner ring is set larger than the width of the outer ring, and the pitch diameter PCD of the ball ((inner ring inner diameter + outer ring outer diameter) / 2) is set smaller than that of the conventional bearing to reduce the rotation and revolving speed of the ball and reduce heat generation of the bearing (see Patent Document 13).
(5) In order to reduce the temperature rise, at least one of the inner ring and the outer ring is formed of powdered high-speed steel having a hardness of HRC65 or more, and the radius of curvature of the raceway groove when formed of the powdered high-speed steel is used as the bearing steel. 1.0 to 1.12 times that of the case formed by the above, the spin slip and gyro slip generated in the Hertz elastic contact surface are reduced, the temperature rise of the bearing is reduced, and the radius of curvature of the raceway groove is increased. (See Patent Document 14)
(6) In order to secure rigidity at low speeds and reduce torque heat generation at high speeds to satisfy both high speed and high rigidity, raceway grooves formed on the outer peripheral surface of the inner ring and the inner peripheral surface of the outer ring One in which the radius of curvature is continuously changed (see Patent Document 15),
(7) In order to reduce the centrifugal force and gyro moment acting on the ball to suppress heat generation and increase the speed, the diameter of the ball is equal to the bearing cross-section height (the value obtained by subtracting the inner diameter of the inner ring from the outer diameter of the outer ring is 1). / 2) set to 32-35% (see Patent Document 16),
(8) In order to reduce the internal preload and achieve high speed, the outer ring groove bottom thickness is reduced, and the dmn value (dm: ball pitch circle diameter, n: rotation speed) exceeds 2,000,000. There has been proposed a device in which the outer ring expands and deforms when the internal preload of the bearing increases (see Patent Document 17).
[0010]
[Patent Document 1]
Japanese Utility Model Publication No. 1-67331
[Patent Document 2]
Japanese Utility Model Laid-Open No. 4-132220
[Patent Document 3]
Published Japanese Utility Model Application No. 6-35659
[Patent Document 4]
JP-A-6-35653
[Patent Document 5]
Japanese Utility Model Publication No. 5-94531
[Patent Document 6]
Japanese Utility Model Laid-Open No. 5-94532
[Patent Document 7]
JP-A-6-35555
[Patent Document 8]
JP-A-6-35657
[Patent Document 9]
JP-A-5-86029
[Patent Document 10]
JP-A-11-270564
[Patent Document 11]
Re-published patent WO00 / 37813
[Patent Document 12]
JP 2000-145794 A
[Patent Document 13]
JP-A-10-26132
[Patent Document 14]
Japanese Utility Model Registration No. 2526579
[Patent Document 15]
JP-A-9-177795
[Patent Document 16]
JP-A-10-274244
[Patent Document 17]
JP-A-11-82499
[0011]
[Problems to be solved by the invention]
However, for ball bearings used in machine tool spindles and the like, in addition to demands for high speed and low heat generation, cutting of difficult-to-cut materials has been increasing in recent years, and demands for high speed and high rigidity have been increasing. In addition, the demand for cost reduction is becoming stronger. In particular, at higher speeds, it is not uncommon for bearings supporting a main shaft to be used in an environment of dmn (= (bearing inner diameter + bearing outer diameter) ÷ 2 × rotational speed (rpm)) of 1,000,000 or more.
[0012]
Grease lubricated rolling bearings are advantageous in terms of environment and cost, but tend to have a short life at high speed rotation. In the case of grease lubrication, the bearing is seized due to deterioration of the grease before the rolling fatigue life of the bearing. If the number of rotations is extremely high, the grease deteriorates in a short time and seizure occurs early. Even the technology of grease lubricated rolling bearings disclosed in Patent Documents 1-9 has not been sufficiently improved.
[0013]
As for the preloading method when a plurality of bearings are juxtaposed along the axis, a fixed position preload that can simplify the preload mechanism is more desirable than a constant pressure preload from the viewpoint of rigidity and cost reduction.
In order to increase the speed and rigidity of ball bearings, if the bearings themselves are highly rigid and resistant to seizure and the amount of heat generated by rotation is low, even if the above-mentioned low-cost lubrication method or preloading method is used, further high-speed Can be achieved.
[0014]
As described in Patent Literature 10 and Patent Literature 11 described above, in order to reduce the heat generation of the bearing and reduce the PV value of the inner ring, the radius of curvature of the cross section in the width direction of the inner raceway groove may be increased. However, when actually used for supporting the main shaft of a machine tool, the performance required of a ball bearing not only suppresses heat generation and speeds up, but also includes rigidity and durability life. A load capacity against an excessive static axial load acting on the bearing as an unclamping force when the bearing is removed from the main shaft is also required. Therefore, the radius of curvature of the cross section in the width direction of the inner raceway groove is set to 52.5 to the diameter of the ball. In actually setting the range within a wide range of 60% or less, it is necessary to consider the above-mentioned use conditions for the upper limit of the radius of curvature.
[0015]
The same applies to the case of the ball bearing disclosed in Patent Document 12. The concept of the permissible static axial load of a ball bearing is based on the basic load that the ball rides on the shoulder of either the inner raceway groove or the outer raceway groove and the contact surface pressure when an axial load is applied to the ball bearing. There are two types, the surface pressure limit corresponding to the rated load, and the lower value is usually set as the limit axial load. However, in the case of a running load, 70% of the limit axial load is set as the allowable axial load in consideration of the manufacturing error of the bearing, the variation in the tolerance range, and the safety factor.
[0016]
As for the surface pressure limit, a static allowable load coefficient fs expressed by the following equation is used as a general guide.
fs = Co / Po
Co; Basic static load rating (N)
Po; static equivalent load (N)
Here, when the ball bearing is used for supporting the main shaft of the machine tool, it is desirable to set the static allowable load coefficient fs to 1.5 or more. The sum of the amount and the permanent deformation amount of the raceway is 85% of the maximum stress 4200 MPa (Basic static rated load: Coa because it is a static allowable axial load) that is approximately 0.0001 times the diameter of the rolling element, or 3570 MPa. Since fs can be set to about 1.7 to 1.8, the maximum stress 3570 MPa is set as the static allowable surface pressure.
[0017]
FIG. 32 shows the specifications of the ball bearings as inner ring inner diameter; Di = 70 mm, outer ring outer diameter; Do = 110 mm, ball diameter; Da = 8.731 mm, number of balls; 25, radius of curvature of the outer raceway groove in the width direction; Ro = 0.52 Da, contact angle of the ball with the inner raceway groove and the outer raceway groove; 15 °, inner and outer race material; SUJ2, ball material; Si ₃ N ₄ The relationship between the radius of curvature Ri of the cross section in the width direction of the inner raceway groove and the permissible static axial load in the case of is shown.
[0018]
The use of ceramic balls for high-speed applications is now within the common sense.However, when a load is applied, the modulus of longitudinal elasticity of the ceramic is larger than that of bearing steel. Also, the contact stress between the ball and the raceway groove increases, and the static allowable axial load also decreases.
Static permissible axial loads are limited by permissible stresses prior to riding loads. As shown in FIG. 32, when the radius of curvature Ri of the cross-section in the width direction of the inner raceway groove is compared with the conventional ball bearing in which Ri = 0.52 Da is set, when the radius of curvature of the cross-section in the widthwise direction of the inner raceway groove is increased. It can be seen that the ultimate allowable axial load is greatly reduced.
[0019]
When the radius of curvature Ri of the cross section in the width direction of the inner raceway groove is set to 0.58 Da, the static allowable axial load is reduced by 65% as compared with the conventional ball bearing, and the capacity of the machine tool spindle becomes insufficient. Occurs.
When the radius of curvature of the cross-section in the width direction of the inner raceway groove and the outer raceway groove is increased, the contact portion of the ball (hereinafter referred to as “contact ellipse”) contacting the raceway groove is reduced during rotation of the bearing. This makes it possible to suppress the sliding of the ball and to reduce the PV value and the heat generation, but naturally the contact ellipse is reduced. For this reason, since the stress of the contact portion becomes larger than that of the conventional ball bearing, the rolling fatigue life naturally decreases.
[0020]
In addition, as important parameters for heat generation, high speed, and rigidity, (1) the radius of curvature of the cross section in the width direction of the width cross section of the raceway groove, (2) the ball diameter, and (3) the contact angle are mentioned. The content does not take the contact angle parameter into consideration in all cases.
In addition, the recent trend of machine tool spindles is that not only high speed but also variation / variable production is performed, so that machining is performed in various rotation ranges, and the frequency of tool conversion is increasing. In order to realize high-efficiency machining under such machining conditions, it is indispensable to be able to rotate at high speed and to shorten the start / stop time (sudden acceleration / deceleration time) of the spindle. During this transition period of rapid rotation fluctuation, changes in the surrounding environment (heat generation of the motor, cooling of the outer cylinder, etc.) are remarkable, and the bearing for the main shaft is exposed to severe thermal load conditions. To put it simply, a temperature difference between the inner and outer rings may occur, and the internal preload of the bearing may increase rapidly to cause seizure. It is necessary to maintain stable performance even under such conditions, and it is necessary to design the bearing insensitively with respect to the PV value.
[0021]
The present invention has been made in view of the above-described problems, and has as its object the purpose of suppressing heat generation due to high-speed rotation with grease lubrication and preventing seizure with high rigidity, extending the rolling fatigue life. It is an object of the present invention to provide a ball bearing and a spindle device for a machine tool, which can secure static axial load capacity when used as a bearing for a machine tool spindle, and are advantageous in terms of environment and cost.
[0022]
[Means for Solving the Problems]
The object of the present invention is achieved by the following configurations.
(1) An inner race having an inner raceway groove on the outer peripheral surface, an outer race having an outer raceway groove facing the inner raceway groove on the inner peripheral surface, and the inner raceway groove and the outer raceway groove are disposed between the inner raceway groove and the outer raceway groove. A ball bearing comprising a plurality of balls, wherein the widthwise cross section of the inner raceway groove and the outer raceway groove is formed in an arc shape.
The radius of curvature of the cross section in the width direction of the inner raceway groove is Ri, the radius of curvature of the cross section in the width direction of the outer raceway groove is Ro, the diameter of the ball is Da, and the value obtained by subtracting the inner diameter of the inner race from the outer diameter of the outer race. When 1/2 is the bearing section height H, the radius of curvature of the width direction cross section of the inner raceway groove is set to Ri = 0.535 Da to 0.545 Da, and the radius of curvature of the width direction cross section of the outer raceway groove is Ri. Is set to Ro = 0.515 Da to 0.545 Da, the contact angle of the ball to the inner raceway groove and the outer raceway groove is set to 17 ° to 26 °, and the diameter of the ball is set to Da = 0. 4H to 0.56H, and
A ball bearing lubricated by grease supplied from the outside.
(2) The ball bearing according to (1), wherein the ball is made of ceramic.
(3) At least one of the inner ring and the outer ring has a weight ratio C of 0.2 to 1.2%, Si of 0.4 to 1.5%, Mo of 0.5 to 1.5%, Cr Is made of a steel material containing 0.5 to 2.0%, balance Fe and unavoidable impurity elements, and is subjected to carbonitriding to set the surface carbon concentration to a range of 0.8 to 1.3%. The ball bearing according to (1) or (2), wherein the surface nitrogen concentration is set in a range of 0.2 to 0.8%.
(4) Additional grease is supplied through at least one of a supply hole penetrating the outer ring in the radial direction, a supply hole penetrating the outer ring in the axial direction, or a supply hole provided in the outer ring spacer. The ball bearing according to any one of claims (1) to (3).
(5) A spindle device for a machine tool in which a ball bearing for supporting a spindle is mounted in a housing,
A spindle device for a machine tool, wherein the ball bearing is the ball bearing according to any one of (1) to (4).
[0023]
Here, the effective range of each of the above-described component elements will be described.
(I) Si: 0.4 to 1.5% by weight
Si is an element having an effect on temper softening resistance, and has an effect of improving high-temperature strength and delaying decomposition of retained austenite, which is effective in preventing indentation-type delamination in a high-temperature environment. If the Si content is less than 0.4% by weight, the high-temperature strength becomes insufficient and the indentation origin type peeling occurs. Therefore, the lower limit is set to 0.4% by weight. On the other hand, if the Si content exceeds 1.5% by weight, the mechanical strength is reduced and carburization is impeded. Therefore, the upper limit is set to 1.5% by weight.
(II) Mo; 0.5 to 1.5% by weight
Mo is an element having an effect on temper softening resistance like Si, and has an effect of improving high-temperature strength. Mo acts as a carbide forming element for forming fine carbides on the carbonitrided surface. If the Mo content is less than 0.5% by weight, the high-temperature strength becomes insufficient and the amount of carbide deposited on the surface becomes insufficient. Therefore, the lower limit is set to 0.5% by weight. On the other hand, if the Mo content exceeds 1.5% by weight, giant carbides are generated at the stage of the material, leading to the loss of carbides and shortening the rolling fatigue life of the bearing. And
(III) Cr; 0.5 to 2.0% by weight
Cr is an additive element having the same function and effect as Mo. If the Cr content is less than 0.5% by weight, the high-temperature strength becomes insufficient and the amount of carbide precipitated on the surface becomes insufficient. Therefore, the lower limit is set to 0.5% by weight. On the other hand, if the Cr content exceeds 2.0% by weight, giant carbides are generated at the stage of the raw material, causing the carbides to fall off and reducing the rolling fatigue life of the bearing. Therefore, the upper limit is set to 2.0% by weight. And (IV) C: 0.2 to 1.2% by weight
As described above, if the amount of retained austenite is too large, the retained austenite is decomposed and the shape changes with time, and the dimensional stability of the bearing is impaired. On the other hand, the presence of retained austenite on the inner ring surface and the outer ring surface is effective in preventing indentation origin type peeling. Therefore, it is necessary to suppress the amount of retained austenite in the entire bearing in a state where the retained austenite is present on the surface. From such a viewpoint, it is preferable that the amount of the average retained austenite in the steel including the surface and the core is 5% by volume or less, and for that purpose, the carbon concentration on which the retained austenite depends is reduced to 1.2% by weight or less. Therefore, the upper limit was set to 1.2% by weight. On the other hand, if the carbon concentration is less than 0.2% by weight, it takes a long time to obtain a desired carburizing depth by the carbonitriding treatment, which causes an increase in the overall cost. 0.2% by weight.
(V) Surface carbon concentration: 0.8 to 1.3% by weight
When carbon is added to the surface by carbonitriding, the martensitic structure serving as the matrix can be solid-solution strengthened, and a large amount of retained austenite effective in preventing indentation origin type exfoliation can be formed in the very surface layer. If the surface carbon concentration is less than 0.8% by weight, the surface hardness becomes insufficient and the rolling fatigue life and wear resistance are reduced. Therefore, the lower limit is set to 0.8% by weight. On the other hand, if the surface carbon concentration exceeds 1.3% by weight, giant carbides precipitate during carbonitriding treatment and the rolling fatigue life is reduced. Therefore, the upper limit was set to 1.3% by weight.
(VI) Surface nitrogen concentration: 0.2 to 0.8% by weight
When nitrogen is added to the surface by carbonitriding, the tempering resistance is improved, the high-temperature strength is increased, the wear resistance is improved, and a large amount of retained austenite is present in the very surface layer, which is effective in preventing indentation type delamination. Can be. If the surface nitrogen concentration is less than 0.2% by weight, the high-temperature strength is reduced and the wear resistance is improved, so the lower limit was set to 0.2% by weight. On the other hand, if the surface nitrogen concentration exceeds 0.8% by weight, it becomes difficult to finish the grinding during the production of the bearing, and the productivity of the bearing is reduced due to difficult grinding. Therefore, the upper limit value is set to 0.8% by weight. .
(VII) Other component elements
It is preferable to add a small amount of Ti as another component element. This is because, when Ti is added, fine titanium carbide (TiC) or carbonitride (Ti (C + N)) precipitates and disperses in the matrix and improves wear resistance and seizure resistance. In this case, the Ti content is desirably 0.1 to 0.3% by weight. If the Ti content is less than 0.1% by weight, the effect of carbide precipitation cannot be obtained, so the lower limit is set to 0.1% by weight. On the other hand, if the Ti content exceeds 0.3% by weight, a huge precipitate is likely to be formed, and this may become a defect to reduce the rolling fatigue life, so that the upper limit is set to 0.3% by weight. %. Incidentally, when the size of the titanium precipitate (TiC, Ti (C + N)) is 0.1 μm or less, it contributes to improvement of wear resistance and seizure resistance.
[0024]
BEST MODE FOR CARRYING OUT THE INVENTION
Hereinafter, the grease-lubricated ball bearing of the present invention will be described with reference to the drawings. First, the process of selecting the ball bearing according to the present invention will be described.
In order to find the specifications of the most suitable ball bearing for supporting the main shaft of a machine tool, the present inventors set the inner ring inner diameter to 70 mm, the outer ring outer diameter to 110 mm, the ball diameter to 8.731 mm, the number of balls to 25, and the contact angle to 25. 18 °, inner and outer ring material is SUJ2, ball material is Si ₃ N ₄ Axial rigidity, inner ring PV value, outer ring PV value, and heat value when the curvature radii Ri and Ro of the cross-section in the width direction of the inner raceway groove and the outer raceway groove are combined with the ball diameter Da as shown in Table 1. , Rolling fatigue life and static allowable axial load were calculated. Note that the calculation here is performed at a rotation speed of 25000 min. ^-1 , Back combination (spacing width 20 mm), shaft fitting; T21 (tight fit, tight fit amount: 21 μm), housing fit; L8 (clearance fit, loose fit amount: 8 μm), preload during assembly; axial Preload with a spring constant of 100 N / μm (Axial spring constant of 100 N / μm when this is incorporated is used as a measure of rigidity that can be machined), temperature conditions: Calculated under conditions of outer ring temperature rise of 24 ° C and inner and outer ring temperature difference of 16 ° C did.
[0025]
[Table 1]

[0026]
At the time of rotation of each ball bearing indicated by number 1 to 25 in Table 1 (25000 min ^-1 FIG. 1 shows the axial spring constant of (1). As shown in the figure, the ball bearing having the highest axial spring constant during rotation among the numbers 1 to 25 is the ball bearing of number 1, and the ball bearing having the lowest axial spring constant during rotation is number 25. Ball bearings.
Comparing the ball bearing of No. 1 with the ball bearing of No. 25, the ball bearing of No. 1 has a radius of curvature Ri, Ro of 0.52 Da in the width direction cross section of the inner raceway groove and the outer raceway groove when the diameter of the ball is Da. On the other hand, in the ball bearing of No. 25, since the curvature radii Ri and Ro of the width direction cross sections of the inner raceway groove and the outer raceway groove are 0.60 Da, the widthwise cross-sections of the inner raceway groove and the outer raceway groove are 0.65 Da. It can be seen that the larger the radii of curvature Ri and Ro, the lower the axial spring constant during rotation and the lower the axial rigidity during rotation.
[0027]
At the time of rotation of each ball bearing indicated by number 1 to 25 in Table 1 (25000 min ^-1 ) Are shown in FIG. As shown in the figure, the ball bearing with the highest inner ring PV value during rotation among the numbers 1 to 25 is the ball bearing with the number 1 and the ball bearing with the lowest inner ring PV value during the rotation is number 20. And 25 ball bearings, as well as ball bearings with numbers 5, 10 and 15.
[0028]
In the ball bearings of Nos. 5, 10, 15, 20, and 25, the radius of curvature Ri of the cross section in the width direction of the inner raceway groove is 0.60 Da, assuming that the diameter of the ball is Da. It can be seen that the inner ring PV value does not change if the radius of curvature Ri of the width direction cross section of the inner raceway groove is the same value even if the radius of curvature Ro of the inner raceway changes. It can be seen that the inner ring PV value can be kept low.
[0029]
The inner ring of each ball bearing indicated by the numbers 1 to 25 in Table 1 is 25000 min. ^-1 FIG. 3 shows the inner wheel PV value and the outer wheel PV value when rotating at. As shown in FIG. ^-1 , The inner wheel PV value is larger than the outer wheel PV value. The reason for this is that at high speeds, the outer ring is controlled (the outer ring and the ball are close to pure rolling), and the rotation axis of the ball is based on the contact angle on the outer ring side. Is increased. Therefore, it is desirable to use the inner ring PV value as an index for determining image sticking.
[0030]
When each of the ball bearings indicated by numbers 1 to 25 in Table 1 rotates (25,000 min. ^-1 4) shows the calorific value. As shown in the figure, the ball bearing with the highest heat value among the numbers 1 to 25 is the ball bearing with the number 1 and the ball bearing with the lowest heat value with the ball bearings with the numbers 10, 15, 20, and 25. It is a bearing. In the ball bearings of Nos. 10, 15, 20, and 25, the radius of curvature Ri of the cross-section in the width direction of the inner raceway groove is Ri = 0.60 Da. Therefore, the larger the radius of curvature Ri of the cross-section of the inner raceway groove in the width direction is, the larger the radius. It can be seen that the amount of heat generated during high-speed rotation is reduced, and that seizure can be suppressed.
[0031]
At the time of rotation of each ball bearing indicated by number 1 to 25 in Table 1 (25000 min ^-1 5) shows the rolling fatigue life. As shown in the figure, the ball bearing with the longest rolling fatigue life among the numbers 1 to 25 is the ball bearing with the number 6 and the ball bearing with the shortest rolling fatigue life is the ball bearing with the number 25.
Comparing the ball bearing No. 6 with the ball bearing No. 25, the ball bearing No. 6 has a radius of curvature Ri = 0.52 Da in the width direction cross section of the inner raceway groove, and a radius of curvature in the width direction cross section of the outer raceway groove. While Ro is Ro = 0.54 Da, the ball bearing of No. 25 has a radius of curvature Ri = 0.60 Da in the width direction cross section of the inner raceway groove, and the radius of curvature Ro of the width direction cross section of the outer raceway groove is Ro. = 0.60 Da, it can be seen that the larger the radius of curvature Ri, Ro of the cross-section in the width direction of the inner raceway groove and the outer raceway groove, the lower the rolling fatigue life.
[0032]
At the time of rotation of each ball bearing indicated by number 1 to 25 in Table 1 (25000 min ^-1 6) shows the permissible static axial load of FIG. As shown in the drawing, the ball bearing having the highest static allowable axial load among the numbers 1 to 25 is the ball bearing of the number 1, and the ball bearings having the lowest static allowable axial load are the numbers 5 and 10 on the contrary. , 15, 20, and 25 ball bearings.
[0033]
In the ball bearings of Nos. 5, 10, 15, 20, and 25, the radius of curvature Ri of the cross section in the width direction of the inner raceway groove is Ri = 0.60 Da. Therefore, the radius of curvature Ri of the cross section in the widthwise direction of the inner raceway groove is smaller. It is understood that the larger the value, the lower the allowable static axial load.
Therefore, in order to reduce the calorific value and the PV value of the ball bearing, the radius of curvature of the widthwise cross section of the inner raceway groove and the outer raceway groove may be increased, but the widthwise cross section of the inner raceway groove and the outer raceway groove may be reduced. When the radius of curvature is increased, the rigidity and life of the bearing are reduced, and the allowable static load is reduced. Conversely, if the radius of curvature of the cross-section in the width direction of the inner raceway groove and the outer raceway groove is reduced, the calorific value and the PV value increase.
[0034]
For machine tool spindle support, a well-balanced bearing is required for all of the seizure index, calorific value, axial rigidity, rolling fatigue life, and static allowable axial load. Compared the calculation results of the axial spring constant, the inner ring PV value, the calorific value, the rolling fatigue life, and the static allowable axial load of each ball bearing indicated by numbers 1 to 25 by absolute value evaluation. In the absolute value evaluation, the best value under each condition was set to 100 and the worst value was set to 0, and the absolute value was evaluated by indexing the difference. The burn-in index—the inner ring PV value, which is the most important under high-speed conditions, was calculated with the importance increased and the rank doubled.
[0035]
The calculation formula used in the above absolute value evaluation is shown below.
(1) Rigidity (Ka)
KaRank = 100− (Ka _Max −Ka) / ((Ka _Max -Ka _Min )
/ 100)
Ka _Max The axial spring constant showing the maximum value in FIG.
Ka _Min The axial spring constant showing the minimum value in FIG.
(2) Inner ring PV value (Pv)
PvRank = (Pv _Max −Pv) / ((Pv _Max -Pv _Min ) / 100)
Pv _Max The inner ring PV value showing the maximum value in FIG.
Pv _Min The PV of the inner ring showing the minimum value in FIG.
(3) Calorific value (W)
WRank = (W _Max −W) / ((W _Max -W _Min ) / 100)
W _Max The calorific value showing the maximum value in FIG.
W _Min The calorific value showing the minimum value in FIG.
(4) Rolling fatigue life (L)
LRank = 100− (L _Max −L) / ((L _Max -L _Min ) / 100)
L _Max A rolling fatigue life showing the maximum value in FIG.
L _Min Rolling fatigue life showing the minimum value in FIG.
(5) Allowable static load (Fa)
FaRank = 100− (Fa _Max −Fa) / ((Fa _Max -Fa _Min ) / 100)
Fa _Max A static allowable axial load showing the maximum value in FIG.
Fa _Min A static allowable axial load showing the minimum value in FIG.
(6) TOTAL-Rank
KaRank + (PvRank × 2) + WRank + LRank + FaRank
Table 2 shows the calculation results of the above-described KaRank, PvRank, WRank, LRank, FaRank, and TOTAL-Rank.
[0036]
[Table 2]

[0037]
Regarding the evaluation points of the TOTAL-Rank, the ball bearings of Nos. 2 and 7 are the most excellent, and can be said to be the optimum specifications. The TOTAL-Rank of the numbers 2 and 7 has almost the same value.
In the bearing specification in which any index takes 100 points from the Rank coefficient in Table 2, any value is 0 or a value close to 0, which is the extreme. For this reason, it cannot be said that it is a very good specification for a machine tool spindle.
[0038]
From the calculation results shown in Table 2, the ball bearing most suitable for supporting the main shaft of the machine tool has a radius of curvature Ri of the widthwise cross section of the inner raceway groove of Ri = 0.54 Da, where Da is the diameter of the ball, and It can be seen that the radius of curvature Ro of the cross section in the width direction of the outer raceway groove is Ro = 0.52 Da or 0.54 Da. However, when the bearing is actually manufactured, a manufacturing error is also included. Therefore, the radius of curvature Ri of the cross section in the width direction of the inner raceway groove is Ri = 0.535 Da to 0.545 Da, and the radius of curvature of the cross section in the width direction of the outer raceway groove is Ri. It is desirable to set Ro to be in a range of Ro = 0.515 Da to 0.545 Da.
[0039]
Next, to confirm the influence of the contact angle of the balls, the present inventors set the inner ring inner diameter to 70 mm, the outer ring outer diameter to 110 mm, the ball diameter to Da = 8.731 mm, the number of balls to 25, the inner raceway groove. The curvature radius of the cross section of the outer raceway groove in the width direction is Ri = Ro = 0.54 Da, the inner and outer race materials are SUJ2, and the ball material is Si. ₃ N ₄ And the inner ring PV value, the outer ring PV value, and the calorific value were calculated when the contact angles of the balls were α = 12 °, 15 °, 18 °, 20 °, 22 °, 25 °, and 30 °. Note that the calculation here is performed at a rotation speed of 25000 min. ^-1 , DB combination (spacing width 20 mm), shaft fitting; T21 (tight fit, tight fit amount: 21 μm), housing fit; L8 (clearance fit, loose fit amount: 8 μm), preload at assembly; axial Preload with a spring constant of 100 N / μm (Axial spring constant of 100 N / μm when this is incorporated is used as a measure of rigidity that can be machined), temperature conditions: Calculated under conditions of outer ring temperature rise of 24 ° C and inner and outer ring temperature difference of 16 ° C did. 7 to 9 show the calculation results.
[0040]
As shown in FIG. 7, when the contact angle of the ball with respect to the inner raceway groove and the outer raceway groove is α = 30 °, the slip of the ball increases due to the effect of the gyro moment, and the inner ring PV value sharply increases. . Therefore, in order to keep the inner ring PV value low, the contact angle of the ball with respect to the inner raceway groove and the outer raceway groove needs to be smaller than 30 °. For example, when the angle is set to 12 °, the axial rigidity decreases, and when the axial spring constant at the time of installation is set to 100 N, an excessive preload must be applied, and the preload during rotation increases. For this reason, as shown in FIG. 7, since the inner ring PV value is larger than that of a ball bearing in which the contact angle of the ball with the inner raceway groove and the outer raceway groove is 15 °, the widthwise cross section of the inner raceway groove and the outer raceway groove is increased. Is set to Ri, Ro = 0.54 Da as shown in Table 1, the contact angle of the ball to the inner raceway groove and the outer raceway groove is preferably set to 15 ° to 25 °. .
[0041]
Further, as shown in FIG. 8, when the contact angle of the ball with the inner raceway groove and the outer raceway groove is α = 30 °, the ball slides more due to the effect of the gyro moment, and the heat generation amount rises sharply. I do. Therefore, in order to keep the calorific value low, the contact angle of the ball with the inner raceway groove and the outer raceway groove must be smaller than 30 °, but the contact angle of the ball with the inner raceway groove and the outer raceway groove is 15 °. In the case of °, the calorific value increases more than when the contact angle of the ball with the inner raceway groove and the outer raceway groove is 18 °. For this reason, when the radius of curvature of the cross-section in the width direction of the inner raceway groove and the outer raceway groove is set as Ri, Ro = 0.54 Da as shown in Table 1, the ball for the inner raceway groove and the outer raceway groove is Is desirably 18 ° to 25 °.
[0042]
Further, as shown in FIG. 9, when the contact angle of the ball with the inner raceway groove and the outer raceway groove is α = 15 ° or less, the outer ring PV value increases for the same reason as described above. Therefore, in order to keep the outer ring PV value low, the contact angle of the ball with respect to the inner raceway groove and the outer raceway groove needs to be larger than 15 °. In the case of 30 °, the sliding of the ball becomes large due to the influence of the gyro moment, and the outer ring PV value increases as compared with the case where the contact angle of the ball with the inner raceway groove and the outer raceway groove is 25 °. For this reason, when the radius of curvature of the cross-section in the width direction of the inner raceway groove and the outer raceway groove is set to Ri, Ro = 0.54a as shown in Table 7, the ball for the inner raceway groove and the outer raceway groove is Is desirably 18 ° to 25 °.
[0043]
As described above, from the results of FIGS. 7 to 9, it can be said that the contact angle of the ball with respect to the inner raceway groove and the outer raceway groove is desirably set to 18 ° to 25 °. However, when the bearing is actually manufactured, a manufacturing error is included. Therefore, it is desirable that the contact angle of the ball with respect to the inner raceway groove and the outer raceway groove is 17 ° to 26 °.
Next, in order to confirm the ball diameter, the present inventors set the inner ring inner diameter to 70 mm, the outer ring outer diameter to 110 mm, and set the curvature radius of the inner raceway groove and the outer raceway groove in the width direction cross section to Ri = Ro = 0.54 Da. The contact angle of the ball with the inner raceway groove and the outer raceway groove is 18 °, the material of the inner and outer rings is SUJ2, and the material of the ball is Si. ₃ N ₄ And the ball diameter, the number of balls, and the cross-sectional ratio (= ball diameter / (outer ring outer diameter−inner ring inner diameter) / 2) × 100) were calculated as the combinations shown in Table 3 to calculate the inner ring PV value. Note that the calculation here is performed at a rotation speed of 25000 min. ^-1 , DB combination (spacing width 20 mm), shaft fitting; T21 (tight fit, tight fit amount: 21 μm), housing fit; L8 (clearance fit, loose fit amount: 8 μm), preload at assembly; axial The preload was set to a spring constant of 100 N / μm (the axial spring constant of 100 N / μm at the time of assembling was used as a measure of rigidity that can be machined), and temperature conditions were calculated under conditions of an outer ring temperature rise of 24 ° C. FIG. 10 shows the calculation results.
[0044]
[Table 3]

[0045]
As is clear from FIG. 10, when the diameter of the ball is 7.144 mm and 7.938 mm, when the temperature difference between the inner ring and the outer ring increases, the inner ring PV value sharply increases, and the temperature between the inner ring and the outer ring increases. When the difference is around 30 °, the inner ring PV value is about 1900 to 2100 MPa · m / s. When the diameters of the balls are 11.906 mm and 12.7 mm, the change in the inner ring PV value is gradual, but the original PV value itself is high, and the temperature difference between the inner ring and the outer ring is about 30 °. The inner ring PV value is about 1900 to 2100 MPa · m / s, as is the case with 7.144 mm and 7.938 mm. On the other hand, when the diameter of the ball is 8.731 mm, 9.525 mm, 10.319 mm, and 11.112 mm, the inner ring PV value increases as the temperature difference between the inner ring and the outer ring increases. When the temperature difference with the inner ring is around 30 °, the inner ring PV value is about 1800 MPa · m / s. Therefore, if the bearing cross-section height (1/2 of the value obtained by subtracting the inner diameter of the inner ring from the outer diameter of the outer ring) is H, if the diameter Da of the ball is from 0.437H to 0.556H, the inner ring is Since the rate of change of the inner ring PV value with respect to the temperature difference between the outer ring and the outer ring is substantially the same, the temperature difference between the inner ring and the outer ring can be reduced by setting the diameter Da of the ball to 0.437H to 0.556H. However, the PV value can be reduced.
[0046]
FIG. 11 is a partial sectional view of a grease-lubricated ball bearing according to the first embodiment of the present invention. The ball bearing (angular ball bearing) 10 of the present embodiment is an outer ring counterbore bearing for supporting a main shaft of a machine tool. As shown in FIG. 1, a ball bearing 10 according to an embodiment of the present invention includes an inner ring 11 formed in a ring shape, and an outer ring 12 arranged concentrically around the outer periphery of the inner ring 11. At least one of the inner ring 11 and the outer ring 12 has a weight ratio of C of 0.2 to 1.2%, Si of 0.4 to 1.5%, Mo of 0.5 to 1.5%, and Cr of 0%. 0.5 to 2.0%, steel material containing balance Fe and unavoidable impurity elements, the surface carbon concentration is 0.3 to 1.3%, and the surface nitrogen concentration is 0.2 to 0.8%. Thus, the carbonitriding process is performed. Further, the ball bearing 10 includes a plurality of balls 13 rotatably incorporated between the inner ring 11 and the outer ring 12 and a retainer 14 for holding these balls 13. Has an inner raceway groove 11a. The inner raceway groove 11a is opposed to the outer raceway groove 12a formed on the inner peripheral surface of the outer race 12, and these raceway grooves 11a, 12a are formed in a circular cross section in the width direction.
[0047]
FIG. 12 is a view showing a relationship between a ball and a raceway groove of a ball bearing according to an embodiment of the present invention. As shown in the figure, an inner raceway groove 11a having a predetermined radius of curvature is formed on the outer diameter surface of the inner ring 11 along the circumferential direction, and the inner raceway surface of the outer ring 12 is formed along the circumferential direction on the inner diameter surface. An outer raceway groove 12a having a predetermined radius of curvature radially opposed to the inner raceway groove 11a is formed, and a plurality of balls 13 are arranged circumferentially in a space formed by the inner raceway groove 11a and the outer raceway groove 12a. They are inserted side by side. Then, each ball 13 comes into rolling contact with the track grooves 11a and 12a. In this embodiment, when the diameter of the ball 13 is Da, the radius of curvature of the outer raceway groove 12a is Ro, and the radius of curvature of the inner raceway groove 11a is Ri, 0.5 = 035 to 0. The diameter Da of the ball 13 as a rolling element and the radii of curvature Ri and Ro of the raceway grooves 11a and 12a are determined so as to satisfy 0.545 Da and Ro = 0.515 Da to 0.545 Da.
[0048]
In the present embodiment, a supply hole 15 is provided on the counter bore side (the right side in the figure) of the outer ring 12 as a supply element that penetrates the outer ring 12 in the radial direction. The supply hole 15 has a circular cross section having a diameter of 0.1 to 5 mm. The supply hole 15 is open at a location on the inner diameter surface of the outer ring 12 adjacent to the outer ring raceway 12a.
The supply holes 15 may be provided at a plurality of locations spaced apart in the circumferential direction of the outer ring 12. The supply hole 15 may be provided on the side where the contact portion 12b is located, or may be provided on a portion other than the contact portion 12b.
[0049]
In the bearing space of the angular contact ball bearing 10, grease in an amount of 10 to 20% of the bearing space volume is initially sealed. When the bearing is used, the following grease replenishing method is applied. That is, the grease is shot at an appropriate timing (intermittently or periodically) through the supply hole 15 such that the amount of one replenishment is 4% or less of the bearing space volume. It is desirable that the replenishment amount at one time is 0.004 cc to 0.1 cc. Here, the bearing space volume means a volume obtained by subtracting the volume of the rolling element and the volume of the cage from the space formed between the inner diameter of the outer ring and the outer diameter of the inner ring.
[0050]
(Example)
Table 4 shows examples and comparative examples of the present invention using a pair of angular ball bearings of the back combination shown in FIG. 13 based on the configuration shown in FIG. The inner rings of Comparative Examples A and B and Examples A and B in Table 4 were subjected to the following evaluation conditions for 5000 minutes. ^-1 ~ 31000min ^-1 FIG. 14 shows the temperature change of the outer ring when the outer ring is rotated at the speed shown in FIG.
[0051]
[Table 4]

[0052]
[Evaluation conditions]
Combination 2-row back-to-back combination (back-to-back combination fixed position preload)
Lubrication Oil-air lubrication Lubrication amount 0.03cc / 4min / 1Brg
Axial screw constant at assembly 100N / μm
(The same spring constant is set so as not to be affected by preload.)
Drive method Belt
Cooling method Outer cylinder cooling Cooling oil temperature room temperature + 3 ° C (room temperature 22 ° C), 6NL / min
Evaluation method Temperature rise, burn-in limit
As shown in FIG. 14, the ball diameter is 7.938 mm, the number of balls is 26, the radius of curvature of the cross-section of the inner raceway groove and the outer raceway groove is Ri, Ro = 0.54 Da, the inner raceway groove and the outer raceway groove. In the case of a ball bearing (comparative example A) in which the contact angle of the ball with respect to the ball is 15 °, the material of the inner and outer rings is SUJ2, and the material of the ball is ceramic, the rotation speed of the inner ring is 24000 min ^-1 , The outer ring temperature rises sharply, and the rotation speed of the inner ring becomes 27000 min. ^-1 It can be seen that seizure occurs when the temperature reaches. On the other hand, the ball diameter is 8.731 mm, the number of balls is 25, the curvature radius of the inner raceway groove in the width direction cross section is Ri = 0.58 Da, and the curvature radius of the outer raceway groove in the width direction cross section is Ro = 0.54 Da. The contact angle of the ball with the inner raceway groove and the outer raceway groove is 18 °, and the material of the inner and outer rings is SHX (trade name of Nippon Seiko Co., Ltd .; C: 0.2 to 1.2% by weight; Si; 0.7 to 1. 5% by weight, Mo; 0.5 to 1.5% by weight, Cr; 0.5 to 2.0% by weight, steel material containing balance Fe and unavoidable impurity elements, surface carbon concentration: 0.8 to 1 0.3%, surface nitrogen concentration: 0.2 to 0.8%), and in the case of a ball bearing using ceramic as a ball material (Comparative Example B), the rotation speed of the inner ring is 31,000 min. ^-1 It can be seen that no seizure occurs even when the temperature reaches.
[0053]
Further, the ball diameter is 8.731 mm, the number of balls is 25, the radius of curvature of the cross section in the width direction of the inner raceway groove and the outer raceway groove is Ri, Ro = 0.54 Da, and the contact angle of the ball with the inner raceway groove and the outer raceway groove. 18 °, ball bearings made of SUJ2 inner and outer ring materials and ceramic ball material (Example A), or a ball diameter of 8.731 mm, 25 balls, width direction of inner raceway groove and outer raceway groove In the case of a ball bearing in which the radius of curvature of the cross section is Ri, Ro = 0.54 Da, the contact angle of the ball with the inner raceway groove and the outer raceway groove is 18 °, the material of the inner and outer rings is SHX, and the material of the ball is ceramic (Example B). In the same manner as in Comparative Example B, the rotation speed of the inner ring was 31,000 min. ^-1 It can be seen that no seizure occurs even when the temperature reaches.
[0054]
From the above, the curvature radius of the cross-section in the width direction of the inner raceway groove is Ri = 0.535 Da to 0.545 Da, the curvature radius of the cross-section in the width direction of the outer raceway groove is Ro = 0.515 Da to 0.545 Da, and the inner raceway groove is The contact angle of the ball with the outer raceway groove is 17 ° to 26 °, and the diameter of the ball is 0.4H or more with respect to the bearing cross-sectional height H (1/2 of the value obtained by subtracting the inner diameter of the inner ring from the outer diameter of the outer ring). By setting it to 0.56H, the amount of heat generated by high-speed rotation can be suppressed, seizure can be prevented, the rolling fatigue life can be prolonged, and the cost can be reduced.
[0055]
Further, the following experiment was conducted using the angular ball bearing of the first embodiment shown in FIG.
Inner diameter 65 mm, outer diameter 100 mm, width 18 mm, contact angle 18 °, ball diameter 7.144 mm, grease to be used is Isoflex NBU15 (manufactured by NOK Kluber Co., Ltd.), and the initial amount of grease is 15% (2 .3 cc), the rotating spindle of the testing machine was supported by a pair of angular ball bearings. The pair of angular contact ball bearings was a back-to-back combination, and the bearing interval was 100 mm.
[0056]
(Experiment 1)
The rotating spindle of the test machine was rotated at several set speeds, and the bearing durability time at each set speed was observed. Table 1 shows the results.
[Table 5]

Without grease replenishment, dmn 1.2 million could not achieve a durability time exceeding 10,000 hours. This indicates that the grease life was 10,000 hours. In the case of dmn 1.5 million, the grease life was 1000 hours. In the case of dmn 1.8 million, the grease life was 100 hours.
FIG. 15 shows a graph of the results in Table 5. FIG. 15 shows that the grease life time decreases exponentially with an increase in dmn.
[0057]
(Experiment 2)
Under the same conditions as in Experiment 1, the rotating main shaft of the testing machine was supported by a pair of angular ball bearings, and an appropriate running-in operation was performed. The outer ring temperature at 1.5mn dmn after the running-in operation was 55 ° C. The grease life at 1.5mn dmn is 1000 hours as described above.
In this experiment, after operating for 1,000 hours with dmn 1.5 million, grease replenishment of several set replenishment amounts was performed, and then starting from 0 to dmn 1.5 million in 5 seconds, operation was resumed, and the bearing was observed. Then, the durability time and the temperature rise of the bearing at each set supply amount were measured. Table 6 shows the results.
[Table 6]

When the replenishment amount was 4% or less, seizure occurred when exceeding 500 hours, which is half of the grease life of the initial filling.
On the other hand, when the replenishment amount exceeded 4%, the bearing abnormally heated (the outer ring rapidly rose to 75 ° C. or more) after the grease was replenished.
From the above results, it was found that the grease replenishment interval should be less than half the grease life time of the initial filling and the replenishment amount should be less than 4% of the bearing space volume.
[0058]
After the 2% grease was replenished, the grease scraped out beside the contact area inside the bearing was sampled and the amount was measured at the time of the break-in operation. It was found that 1.9% of the grease had been scraped off. This indicates that the minimum necessary grease was held as a lubricating film between the rolling elements (balls), the raceway grooves of the inner and outer rings, and the retainer, and excess grease was scraped out.
From this, it was found that the lower limit of the grease supply amount was 0.1% of the bearing space volume.
[0059]
FIG. 16 shows an angular contact ball bearing 20 according to a second embodiment of the present invention. The radius of curvature of the cross-section in the width direction of the inner raceway groove and the outer raceway groove, the contact angle of the ball, the diameter of the ball, and the amount of grease lubrication at one time are the same as in the first embodiment.
The angular ball bearing 20 includes an inner ring 21, an outer ring 22, a plurality of balls 23 disposed between an inner ring track 21 a of the inner and outer rings 21 and an outer ring track 22 a of the outer ring 22, and a retainer for holding the balls 23 at equal intervals in a circumferential direction. 24.
In the present embodiment, a supply hole 25 is provided on the counter bore side (the right side in the figure) of the outer ring 22 as a supply element that penetrates the outer ring 22 in the radial direction. A grease reservoir 25a is formed on the outer ring inner diameter surface side of the supply hole 25. The cross-sectional area of the grease reservoir 25a is larger than the cross-sectional areas of other portions of the supply hole 25. Since the grease reservoir 25a is provided, the supply hole 25 is a stepped cylindrical space. The grease reservoir 25a is located on the inner surface of the outer ring 22 at a position adjacent to the outer ring raceway 22a. Also in the embodiment described below, the supply hole may have a grease reservoir.
[0060]
FIG. 17 shows an angular contact ball bearing 30 according to a third embodiment of the present invention. The radius of curvature of the cross-section in the width direction of the inner raceway groove and the outer raceway groove, the contact angle of the ball, the diameter of the ball, and the amount of grease lubrication at one time are the same as in the first embodiment.
The angular ball bearing 30 includes an inner ring 31, an outer ring 32, a plurality of balls 33 disposed between an inner ring track 31 a of the inner ring 31 and an outer ring track 32 a of the outer ring 32, and a retainer 34 for holding the balls 33 at equal circumferential intervals. It has. This embodiment is an inner ring counterbore bearing.
In the present embodiment, a replenishing hole 35 as a replenishing element that penetrates the outer race 32 in the radial direction is opened on the outer raceway 32a of the outer race 32 on the side opposite to the side (the right side in the drawing) where the contact portion 32b is located. . Note that the supply hole 35 may be provided on the side where the contact portion 32b is located, or may be provided in a portion other than the contact portion 32b.
[0061]
FIG. 18 shows an angular contact ball bearing 40 according to a fourth embodiment of the present invention. The radius of curvature of the cross-section in the width direction of the inner raceway groove and the outer raceway groove, the contact angle of the ball, the diameter of the ball, and the amount of grease lubrication at one time are the same as in the first embodiment.
The angular ball bearing 40 includes an inner ring 41, an outer ring 42, a plurality of balls 43 arranged between an inner ring raceway 41 a of the inner ring 41 and an outer ring raceway 42 a of the outer ring 42, and an outer ring guide retainer 44. The present embodiment is an outer ring counterbore bearing.
In the present embodiment, a supply hole 45 is provided on the counter bore side (the right side in the drawing) of the outer ring 42 as a supply element that penetrates the outer ring 42 in the radial direction. The supply hole 45 is open toward a guide surface 44a on one side (the right side in the figure) of the retainer 44.
[0062]
FIG. 19 shows an angular contact ball bearing 50 according to a fifth embodiment of the present invention. The radius of curvature of the cross-section in the width direction of the inner raceway groove and the outer raceway groove, the contact angle of the ball, the diameter of the ball, and the amount of grease lubrication at one time are the same as in the first embodiment.
The angular ball bearing 50 includes an inner ring 51, an outer ring 52, a plurality of balls 53 disposed between an inner ring track 51 a of the inner ring 51 and an outer ring track 52 a of the outer ring 52, and an outer ring guide retainer 54. The present embodiment is an outer ring counterbore bearing.
In the present embodiment, a supply hole 55 as a supply element that penetrates the outer ring 52 in the radial direction is provided on a side of the outer ring 52 opposite to the counter bore (the left side in the figure). The supply hole 55 is open toward a guide surface 54 a on one side (the left side in the figure) of the retainer 54.
[0063]
FIG. 20 is a diagram showing a spindle device as a machine tool spindle device configured using the angular ball bearings described in the first to fifth embodiments of the present invention. Here, the angular ball bearing 10 of the first embodiment is used as an example.
[0064]
The bearing 10 fits over the main shaft 1 and fits inside the housing 7. The main shaft 1 is rotatable with respect to the housing 7 via a bearing 10. An inner ring spacer 5 and an outer ring spacer 6 are arranged between the inner ring and the outer ring of the bearing 10 along the main shaft 1 and the housing 7, respectively. At both ends in the axial direction of the inner ring spacer 5 and the outer ring spacer 6, an inner ring pressing member 8 and an outer ring pressing member 9 are arranged, respectively, and preload is applied to each bearing via each spacer. A gap (not shown) is formed between the inner ring pressing member 8 and the outer ring pressing member 9 to form a labyrinth between the pressing members.
[0065]
A nozzle (grease supply top) 4 that penetrates the housing 7 and supplies additional grease to a supply hole formed in an outer ring of each bearing 10 is fixed to the housing 7. Grease is supplied from the grease replenisher 2 to the nozzle 4 via the replenishment pipe 3, and is replenished radially into the bearing. The grease replenishing device 2 performs a grease shot at an appropriate timing (intermittently or periodically) so that a single replenishment amount is 0.1 to 4% of the bearing space volume.
[0066]
In FIG. 20, the bearing 10 of the first embodiment is described as an example, but of course, the other embodiments 2 to 5, or any combination thereof may be used instead.
Needless to say, the same effect can be expected even if the same supply hole is provided in the outer ring of the other bearing.
[0067]
FIG. 21 is a diagram showing a spindle device as a machine tool spindle device configured by using a rolling bearing 200 according to a sixth embodiment described below.
[0068]
The bearing 200 is fitted around the main shaft 1 and fitted inside the housing 7. The main shaft 1 is rotatable with respect to the housing 7 via a bearing 200. Between the inner ring and the outer ring of the bearing 200, inner ring spacers 500a, 500b, 500c, 500d, 500e and outer ring spacers 600a, 600b, 600c, 600d, 600e are arranged along the main shaft 1 and the housing 7, respectively. They are arranged in order from left to right.
[0069]
At both ends in the axial direction of the inner ring spacers 500a and 500e and the outer ring spacers 600a and 600e, inner ring pressing members 8a and 8b and outer ring pressing members 9a and 9b are arranged, respectively, and a preload is applied to each bearing via each spacer. ing. A gap (not shown) is formed between the inner ring pressing member 8a and the outer ring pressing member 9a and between the inner ring pressing member 8b and the outer ring pressing member 9b, and a labyrinth is formed between the two pressing members.
[0070]
FIG. 22 is an enlarged sectional view of the spindle device shown in FIG. Here, the angular ball bearing 200 and its peripheral structure according to the sixth embodiment of the present invention will be described.
[0071]
Each of the angular ball bearings 200 shown in FIG. 22 includes an inner ring 201, an outer ring 202, a plurality of balls 203 arranged between an inner ring raceway 201a of the inner ring 201 and an outer ring raceway 202a of the outer ring 202, and the like. A retainer 204 is provided for holding at intervals. The outer ring 202 has a tapered portion 202b on one side in the axial direction for holding the ball 203 with a contact angle. Hereinafter, one side in the axial direction where the tapered portion is formed is referred to as a front side, and the other side is referred to as a back side.
[0072]
In the present embodiment, a grease replenishing outer ring spacer 600b is arranged between the angular ball bearings 200. Two grease replenishing nozzles 4 penetrating the housing 7 are inserted into and fixed to the grease replenishing outer ring spacer 600b. Additional grease is supplied to the grease supply nozzle 4 from the external grease supply unit 2 via the supply pipe 3.
[0073]
The grease replenishing outer ring spacer 600b has a replenishing hole 205 as a replenishing element for replenishing additional grease into the inside of the angular ball bearing 200 from the tip of the nozzle 4. The supply hole 205 has a circular cross section with a diameter of 0.1 to 5 mm, and opens in the axial direction toward the inside of the bearing 200 (inner diameter side than the retainer 204). The supply hole 205 supplies additional grease between the inner ring 201 and the outer ring 202 in the axial direction from the rear side. The supplied grease is mainly supplied to the inner diameter side of the retainer 204.
[0074]
The supply holes 205 may be provided at a plurality of locations on the grease supply outer ring spacer 600b at intervals in the radial direction. It is preferable that the grease to be supplied is mainly supplied to the inner diameter side than the retainer 204, but may be supplied to the outer diameter side.
[0075]
The grease in an amount of 10 to 20% of the bearing space volume is initially sealed in the bearing space of each angular ball bearing 200. Then, after starting the use of the bearing, the grease supply unit 2 supplies the replenishment amount through the supply hole 205 at an appropriate timing (intermittently or periodically) to 0.1 to 4% of the bearing space volume. Grease shot so that it becomes. The grease additionally replenished inside the bearing accompanies the entire inside of the bearing as the ball 203 rolls, and makes up for the insufficient grease. This prevents the bearing from being damaged due to the lack of grease, and extends the life of the bearing.
The grease replenishing nozzle 4 penetrating the housing 7 is also inserted and fixed to the grease replenishing outer ring spacer 600d, and has the same configuration as described above.
[0076]
FIG. 23 is an enlarged cross-sectional view of a spindle device according to a first modification of the sixth embodiment.
The angular contact ball bearing 220 used in the present modified example rolls between the inner race 221 fitted on the shaft, the outer race 222 fitted on the housing 1000, and the inner raceway 221 a of the inner race 221 and the outer raceway 222 a of the outer race 222. It comprises a ball 223 that is freely arranged, and a holder 224 that holds the ball 223.
[0077]
The housing 1000 has a convex portion 1000a protruding radially inward. The outer ring 222 of the bearing 220 is in contact with the convex portion 1000a on the rear side in the axial direction. On the rear side in the axial direction of the inner ring 221, an inner ring spacer 510a axially facing the convex portion 1000a is disposed.
[0078]
On the other hand, a grease replenishing outer ring spacer 610 is provided on the axially front side of the outer ring 222. The grease replenishing outer ring spacer 610 is axially opposed to the inner ring spacer 510b. An opening 1000b for inserting the grease supply nozzle 400 into the grease supply outer ring spacer 610 is formed at a position corresponding to the outer diameter surface of the grease supply outer ring spacer 610 in the housing 1000. The base 400a of the grease supply nozzle 400 is fixed on the outer diameter surface of the housing 1000 by a fixing member 400b such as a screw, and the tip 400c extending from the base 400a is inserted into the grease supply outer ring spacer 610. Have been.
[0079]
The grease replenishing outer ring spacer 610 has a replenishing hole 225 as a replenishing element for replenishing additional grease into the inside of the bearing 220 from the tip part 400c of the grease replenishing nozzle 400. The supply hole 225 has a circular cross section having a diameter of 0.1 to 5 mm. The supply hole 225 supplies additional grease between the inner ring 221 and the outer ring 222 in the axial direction from the front side. This prevents the bearing from being damaged due to the lack of grease, and extends the life of the bearing.
Note that the supply holes 225 may be provided at a plurality of locations on the outer ring spacer 610 for grease supply at intervals in the radial direction.
[0080]
FIG. 24 is an enlarged sectional view of a spindle device according to a second modification of the sixth embodiment.
The angular contact ball bearing 230 used in the present modification rolling between the inner race 231 fitted on the shaft, the outer race 232 fitted on the housing 1100, and the inner race 231a of the inner race 231 and the outer race 232a of the outer race 232. The ball 233 is freely arranged, and the holder 234 holds the ball 233.
[0081]
The housing 1100 has a convex portion 1100a protruding radially inward. The outer ring 232 of the bearing 230 is in contact with the projection 1100a on the front side in the axial direction. On the front side of the inner ring 231, an inner ring spacer 520b axially facing the convex portion 1100a is arranged. On the other hand, an inner ring spacer 520a and an outer ring spacer 620 are disposed on the axially rear side of the outer ring 222 so as to face each other.
[0082]
An opening 1100b for inserting the grease supply nozzle 400 into the projection 1100a is formed on the outer diameter surface of the housing 1000 opposite to the projection 1100a. The base 400a of the grease replenishing nozzle 400 is fixed on the outer diameter surface of the housing 1100 by a fixing member 400b such as a screw, and the tip 400c extending from the base 400a is inserted into the projection 1100a.
[0083]
The protrusion 1100a has a supply hole 235 as a supply element for supplying additional grease into the bearing 230 from the tip portion 400c of the grease supply nozzle 400. The supply hole 235 has a circular cross section with a diameter of 0.1 to 5 mm. The supply hole 235 supplies additional grease between the inner ring 231 and the outer ring 232 in the axial direction from the front side. This prevents the bearing from being damaged due to the lack of grease, and extends the life of the bearing.
In addition, the supply holes 235 may be provided at a plurality of locations of the convex portion 1100 at intervals in the radial direction.
[0084]
FIG. 25 is an enlarged sectional view of a spindle device according to a third modification of the sixth embodiment.
In this modification, the front side and the back side of the angular ball bearing 230 of the second modification are interchanged, and the convex portion 1100a of the housing 1100 is provided on the axial back side of the angular ball bearing 230. Other configurations are the same as those shown in FIG.
[0085]
In this modification, additional grease is supplied axially from the back side between the inner ring 231 and the outer ring 232 from a supply hole 235 formed in the convex portion 1100a. This prevents the bearing from being damaged due to the lack of grease, and extends the life of the bearing.
[0086]
FIG. 26 is an enlarged sectional view of a spindle device according to a fourth modification of the sixth embodiment.
The angular contact ball bearing 240 used in this modification is rolling between the inner race 241 fitted on the shaft, the outer race 242 fitted on the housing 1200, and the inner raceway 241 a of the inner race 241 and the outer raceway 242 a of the outer race 242. The ball 243 is freely arranged, and the holder 244 holds the ball 243. At the front end of the outer race 1200, a convex portion 242b protruding radially inward from the tapered portion is formed.
[0087]
The outer ring 232 of the bearing 230 is in contact with the outer ring spacer 630a on the axial front side, that is, the protrusion 242b is in contact with the outer ring spacer 630b, and is in contact with the outer ring spacer 630a on the axial rear surface side. Inner ring spacers 530a and 530b radially opposed to outer ring spacers 630a and 630b are disposed on the rear and front sides of the inner ring 231 respectively.
[0088]
The housing 1200 has an opening 1200b on the outer diameter surface opposite to the convex portion 242b of the outer ring 242 for inserting the grease supply nozzle 400 into the convex portion 242b. The base 400a of the grease supply nozzle 400 is fixed on the outer diameter surface of the housing 1200 by a fixing member 400b such as a screw. It is inserted inside the part 242b.
[0089]
The convex portion 242b has a supply hole 245 as a supply element for supplying additional grease into the bearing 240 from the tip portion 400c of the grease supply nozzle 400. The supply hole 245 has a circular cross section with a diameter of 0.1 to 5 mm. The supply hole 245 supplies additional grease between the inner ring 241 and the outer ring 242 in the axial direction from the front side. This prevents the bearing from being damaged due to the lack of grease, and extends the life of the bearing.
Note that the supply holes 245 may be provided at a plurality of locations of the convex portion 242b at intervals in the radial direction.
[0090]
FIG. 27 is an enlarged sectional view of a spindle device according to a fifth modification of the sixth embodiment.
This modification is a modification of the outer ring 242 of the angular ball bearing 240 of the fourth modification, in which the convex portion 242b of the outer ring 242 is formed on the axial rear surface side of the angular ball bearing 240. Other configurations are the same as those shown in FIG.
[0091]
In this modification, additional grease is supplied axially from the back side between the inner ring 241 and the outer ring 242 from a supply hole 245 formed in the convex portion 242b. This prevents the bearing from being damaged due to the lack of grease, and extends the life of the bearing.
[0092]
By configuring as in the sixth embodiment and the first to fifth modifications of the sixth embodiment, additional grease can be replenished inside the bearing in the axial direction.
It is needless to say that the same effect can be expected even if the same supply holes are provided in other bearings.
[0093]
(Experiment 3)
Next, an endurance test was performed when additional grease was supplied from the axial direction using a rotation test device (intermediate S / P for 65 mm angular contact ball bearing) shown in FIG. In FIG. 28, the shaft 1 is configured to be rotatable with respect to the housing 1500 by angular ball bearings 250, 250 in a back-to-back (DB) combination.
[0094]
FIG. 29 is an enlarged sectional view showing the angular ball bearing 250 of FIG. 28 and its peripheral members. The angular ball bearing 250 includes an inner ring 251 fixed to the shaft 1, an outer ring 252 fixed to the housing, a rolling element 253 rotatably disposed between an inner race 251 a of the inner race 251 and an outer race 252 a of the outer race 252, and It comprises a retainer 254 for holding the rolling element 253.
[0095]
An inner ring spacer 550a and an outer ring spacer 650a are provided on the front side of each angular ball bearing 250, and an inner ring spacer 550 and an outer ring spacer 650b for supplying grease are provided on the rear side adjacent to the inner ring 251 and the outer ring 252, respectively. Are located. An outer ring spacer 660 is arranged between the grease supply outer ring spacers 650b.
[0096]
A grease supply nozzle 450 is inserted and fixed to the grease supply outer ring spacer 650b via a housing 1500. Grease is supplied to the grease supply nozzle 450 via a supply pipe 3 from a grease supply device (not shown). The grease supply outer ring spacer 650 has a supply hole 255 (φ2.0) opened on the front side of the angular ball bearing 250. The grease supplied to the grease supply nozzle 450 is supplied to the supply hole 255 from the tip end portion 450a, and is supplied axially into the bearing space of the bearing.
[0097]
In the experiment, an angular ball bearing having the inner diameter of 65 mm, the outer diameter of 100 mm, the width of 18 mm, the ball diameter of 7.144 mm, and the contact angle of 18 °, which is the configuration of the ball bearing according to the present invention, was used. The grease used for lubrication was Isoflex NBU15, and the initial grease filling amount was 15% of the bearing space volume. Both bearings 250, 250 are 20,000 min ^-1 (Dmn 1.8 million).
[0098]
In the experiment, for comparison, Test 1 in which additional grease was not supplied and Test 2 in which additional grease was supplied were performed. In Test 2, every 50 hours after the start of rotation, 0.3 cc (corresponding to 1.5% of the bearing space volume) of grease was additionally shot into the bearing space via the supply hole 255.
[0099]
As a result of the experiment, in Test 1, the bearing was seized 500 hours after the start of the experiment, so the experiment was stopped. On the other hand, in Test 2, no abnormality occurred even after 3000 hours had passed since the start of the experiment, and the experiment was completed successfully. From this experimental result, it was confirmed that when grease was supplied, no problem occurred in the bearing even after 3000 hours or more, and the life of the bearing was significantly extended.
[0100]
Next, a description will be given of a seventh embodiment of the present invention in which the amount of grease supplied at one time is changed in the angular ball bearings described in the first to sixth embodiments of the present invention.
In the first to sixth embodiments, in order to prevent the occurrence of seizure, the amount of grease supplied at one time is set to be 0.1 to 4% of the bearing space volume. Used in a main shaft of a device requiring strict accuracy, such as a machine tool for a vehicle, is desired not to cause temperature pulsation at the moment of replenishment. For this reason, in the seventh embodiment, grease is supplied so that the amount of grease supplied at one time is 0.004 cc to 0.1 cc.
[0101]
Grease in an amount of 10 to 20% of the bearing space volume is initially filled in the bearing space of the angular contact ball bearing. Here, the bearing space volume means a volume obtained by subtracting the volume of the rolling element and the volume of the cage from the space formed between the inner diameter of the outer ring and the outer diameter of the inner ring. When the bearing is used, the following grease replenishing method is applied. That is, the grease is shot at an appropriate timing (intermittently or periodically) through the supply hole 15 such that the amount of one supply is 0.004 cc to 0.1 cc. In consideration of the grease shot variation, the upper limit of the replenishment amount at one time is 0.12 cc. In addition, the amount of grease supplied at one time is preferably 0.01 cc to 0.03 cc in consideration of prevention of occurrence of temperature pulsation. By performing grease shots within the above range, abnormal temperature rise due to grease deterioration or insufficient oil film formation and damage to the bearings are prevented, and temperature pulsation during grease replenishment is suppressed. Deterioration of the axis accuracy of the machine tool to be mounted can be prevented. Note that the seventh embodiment is applicable to any of the configurations of the angular ball bearings shown in the first to sixth embodiments.
[0102]
Next, regarding the seventh embodiment of the present invention, the grease replenishing amount and the temperature were measured using angular ball bearings 740 and 750 having an inner ring inner diameter of 65 mm, an outer ring outer diameter of 100 mm, a width of 18 mm, a ball diameter of 7.144 mm, and a contact angle of 18 °. A temperature pulsation confirmation test was performed to confirm the pulsation relationship of the samples. The grease used in the temperature pulsation confirmation test was Isoflex NBU15, and the initial amount of the grease was 15% of the bearing space volume. The test condition was dmn = 1.8 million.
[0103]
This pulsation confirmation test was performed using the test spindle device 700 shown in FIG. The test spindle device 700 has a configuration in which a housing main body 703 is supported by a housing block 702 arranged on a support base 701. Angular ball bearings 740 and 750 are fitted inside the outer housing 703 in a rear arrangement. The angular ball bearing main shaft 740 is fitted around the main shaft 710, and rotatably supports the main shaft 710.
[0104]
Between the angular ball bearings 740 and 750, an inner ring spacer 711 is provided between inner rings of the angular ball bearings 740 and 750, and an outer ring spacer 712 is provided between outer rings of the angular ball bearings 740 and 750, respectively. A rear end outer ring retainer 713 is provided on the rear end side (the right side in the figure) of the angular ball bearing 750 in the axial direction.
An outer ring pressing member 714 and an inner ring pressing member 715 are provided on the axial front end side (left side in the figure) of the angular ball bearing 740. Each of the angular ball bearings 740 and 750 is pressed by the outer ring pressing member 714 and the inner ring pressing member 715 in the axial direction toward the rear end outer ring pressing 713 side.
[0105]
Nozzle members 730, 730 are attached to the housing 703 so as to correspond to the respective angular ball bearings 740, 750. The nozzle members 730, 730 supply grease into the bearing space from the hole side provided on the outer ring of each of the angular ball bearings 740, 750. In the present temperature pulsation confirmation test, grease was supplied every hour after the start of the test. The experiment was conducted five times with the amount of grease supplied to each bearing at one grease replenishment being 0.035 cc, 0.10 cc, 0.15 cc, 0.30 cc, and 0.60 cc. FIG. 31 is a graph showing the results of the temperature pulsation confirmation test. FIG. 31A shows the case where the grease supply amount is 0.035 cc, FIG. 31B shows the case where the grease supply amount is 0.10 cc, and FIG. (D) shows the case where the grease supply amount is 0.30 cc, and (e) shows the case where the grease supply amount is 0.60 cc.
[0106]
As shown in FIGS. 31A and 31B, when the grease replenishment amounts are 0.035 cc and 0.10 cc, the bearing temperature of the angular contact ball bearings 740 and 750 does not change at all even if the grease is replenished. Absent. However, as shown in FIG. 31 (c), when the grease replenishment amount is 0.15 cc, it can be seen that the temperature of the angular ball bearings 740 and 750 increases by about 1 ° C. every time grease is replenished. Similarly, as shown in FIGS. 31 (d) and (e), even when the grease replenishment amount is 0.30 cc and 0.60 cc, the temperature of the angular ball bearings 740 and 750 increases by 1 to 2 every time the grease is replenished. It can be seen that the temperature has increased by about ° C.
[0107]
Here, the angular ball bearing 740 and the bearing 750 have different temperatures in a steady state before grease replenishment. The temperature difference in the steady state is caused by a difference in the peripheral structure between the angular ball bearing 740 and the angular ball bearing 750, for example, a difference in distance from the housing block 702, a positional relationship with a cooling device (not shown), and the like. It is considered that the temperature in the steady state is different because the heat suppression rate is different.
[0108]
In any case, in both the angular ball bearing 740 and the angular ball bearing 750, as shown in FIGS. 31 (c) to 31 (e), when the grease supply amount is 0.15 cc or more, It is considered that the bearing generates heat due to the stirring resistance and the like, and the temperature of the bearing rises, that is, temperature pulsation occurs. Therefore, it has been found that by setting the amount of grease replenishment at one time to 0.1 cc or less, stable grease supply without temperature pulsation can be performed.
[0109]
【The invention's effect】
As described above, according to the present invention, the amount of heat generated due to high-speed rotation can be suppressed, seizure can be prevented with high rigidity, the rolling fatigue life can be extended, and when used as a bearing for a machine tool spindle. It is possible to provide a grease lubricated ball bearing that secures a static axial load capacity and is advantageous in terms of environment and cost. In addition, a spindle device for a machine tool having high reliability can be realized by using the rolling bearing.
[Brief description of the drawings]
FIG. 1 is a view showing an axial spring constant when a ball bearing rotates.
FIG. 2 is a diagram showing an inner ring PV value during rotation of a ball bearing.
FIG. 3 is a diagram showing an inner ring PV value and an outer ring PV value during rotation of a ball bearing.
FIG. 4 is a diagram showing a calorific value of a ball bearing.
FIG. 5 is a view showing a rolling fatigue life of a ball bearing.
FIG. 6 is a diagram showing a static allowable axial load of a ball bearing.
FIG. 7 is a diagram showing a relationship between an inner ring PV value of a ball bearing and a contact angle.
FIG. 8 is a diagram showing a relationship between a calorific value of a ball bearing and a contact angle.
FIG. 9 is a diagram showing a relationship between an outer ring PV value of a ball bearing and a contact angle.
FIG. 10 is a view showing a relationship between an inner ring PV value of a ball bearing and an inner and outer ring temperature difference.
FIG. 11 is a sectional view showing the first embodiment of the present invention.
FIG. 12 is a view showing a relationship between a ball and a raceway groove of the ball bearing according to the first embodiment of the present invention.
FIG. 13 is a view showing an example to which a ball bearing according to one embodiment of the present invention is applied.
FIG. 14 is a diagram showing the relationship between the outer ring temperature and the inner ring rotation speed of a ball bearing.
FIG. 15 is a graph showing an experimental result using the first embodiment of the present invention.
FIG. 16 is a sectional view showing a second embodiment of the present invention.
FIG. 17 is a sectional view showing a third embodiment of the present invention.
FIG. 18 is a sectional view showing a fourth embodiment of the present invention.
FIG. 19 is a sectional view showing a fifth embodiment of the present invention.
FIG. 20 is a cross-sectional view showing a spindle device configured using the rolling bearings described in the first to fifth embodiments of the present invention.
FIG. 21 is a cross-sectional view illustrating a spindle device configured using the rolling bearing according to the sixth embodiment of the present invention.
FIG. 22 is an enlarged sectional view of the spindle device shown in FIG. 21, showing a sixth embodiment of the present invention.
FIG. 23 is a sectional view showing a first modification of the sixth embodiment of the present invention.
FIG. 24 is a sectional view showing a second modification of the sixth embodiment of the present invention.
FIG. 25 is a sectional view showing a third modification of the sixth embodiment of the present invention.
FIG. 26 is a sectional view showing a fourth modification of the sixth embodiment of the present invention.
FIG. 27 is a sectional view showing a fifth modification of the sixth embodiment of the present invention.
FIG. 28 is a diagram showing a rotation test device.
FIG. 29 is an enlarged view of the rotation test device of FIG. 28.
FIG. 30 is a view showing a test spindle device used in a temperature pulsation test according to a seventh embodiment of the present invention.
FIGS. 31A and 31B are graphs showing the results of a pulsation confirmation test in the seventh embodiment of the present invention, where FIG. 31A shows a case where the grease supply amount is 0.035 cc, and FIG. 31B shows a case where the grease supply amount is 0.10 cc. , (C) shows the case where the grease supply amount is 0.15 cc, (d) shows the case where the grease supply amount is 0.30 cc, and (e) shows the case where the grease supply amount is 0.60 cc.
FIG. 32 is a view showing a relationship between a static allowable axial load of a ball bearing and a radius of curvature of an inner raceway groove.
[Explanation of symbols]
10, 20, 30, 40, 50 angular contact ball bearings (ball bearings)
11, 21, 31, 41, 51 Inner ring
12, 22, 32, 42, 52 Outer ring
13,23,33,43,53 ball
14,24,34,44,54 Cage
15, 25, 35, 45, 55 Supply hole

Claims (5)

An inner ring having an inner raceway groove on an outer peripheral surface, an outer race having an outer raceway groove facing the inner raceway groove on an inner peripheral surface, and a plurality of balls disposed between the inner raceway groove and the outer raceway groove; In the ball bearing having the inner raceway groove and the outer raceway groove formed in an arc-shaped cross section in the width direction,
The radius of curvature of the cross section of the inner raceway in the width direction is Ri, the radius of curvature of the cross section of the outer raceway in the width direction is Ro, the diameter of the ball is Da, and the value obtained by subtracting the inner diameter of the inner race from the outer diameter of the outer race. When 1/2 is the bearing section height H, the radius of curvature of the width section of the inner raceway groove is set to Ri = 0.535 Da to 0.545 Da, and the radius of curvature of the width section of the outer raceway groove is Ri. Is set to 0.515 Da to 0.545 Da, the contact angle of the ball with the inner raceway groove and the outer raceway groove is set to 17 ° to 26 °, and the diameter of the ball is set to Da = 0. A ball bearing set to 4H to 0.56H and lubricated by grease supplied from the outside. The ball bearing according to claim 1, wherein the ball is made of ceramic. At least one of the inner ring and the outer ring has a weight ratio C of 0.2 to 1.2%, Si of 0.4 to 1.5%, Mo of 0.5 to 1.5%, and Cr of 0.1 to 1.5%. It is made of a steel material containing 5 to 2.0%, the balance being Fe and unavoidable impurity elements, and is subjected to carbonitriding, so that the surface carbon concentration is set in the range of 0.8 to 1.3%. 3. The ball bearing according to claim 1, wherein the surface nitrogen concentration is set in a range of 0.2 to 0.8%. Additional grease is supplied through at least one of a supply hole penetrating the outer ring in the radial direction, a supply hole penetrating the outer ring in the axial direction, a supply hole provided in the outer ring spacer, or a supply hole provided in the housing. The ball bearing according to claim 1, wherein: A machine tool spindle device in which a ball bearing supporting the spindle is mounted in a housing,
A spindle device for a machine tool, wherein the ball bearing is the ball bearing according to any one of claims 1 to 4.

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