JP3252587B2

JP3252587B2 - Ball bearing device

Info

Publication number: JP3252587B2
Application number: JP02940394A
Authority: JP
Inventors: 晋高野
Original assignee: NSK Ltd
Current assignee: NSK Ltd
Priority date: 1994-02-28
Filing date: 1994-02-28
Publication date: 2002-02-04
Anticipated expiration: 2017-02-04
Also published as: JPH07238926A

Description

DETAILED DESCRIPTION OF THE INVENTION

【０００１】[0001]

【産業上の利用分野】この発明に係る玉軸受装置は、ス
クリューコンプレッサの回転軸を回転自在に支持する為
に利用する。BACKGROUND OF THE INVENTION A ball bearing device according to the present invention is used for rotatably supporting a rotary shaft of a screw compressor.

【０００２】[0002]

【従来の技術】スクリューコンプレッサの回転軸等、高
速で回転する軸を支承する為、図１に示す様な玉軸受装
置が、従来から使用されている。この玉軸受装置は、ス
クリューコンプレッサを構成するロータ１を固定した回
転軸２の外周面と、ハウジング３の内周面との間に設け
られる。尚、図１で４は、ラジアル方向の荷重Ｆ_r を支
承する為のころ軸受である。2. Description of the Related Art A ball bearing device as shown in FIG. 1 is conventionally used to support a high-speed rotating shaft such as a rotary shaft of a screw compressor. This ball bearing device is provided between an outer peripheral surface of a rotating shaft 2 to which a rotor 1 constituting a screw compressor is fixed and an inner peripheral surface of a housing 3. Incidentally, 4 in FIG. 1 is a roller bearing for supporting a load F _r in the radial direction.

【０００３】本発明の対象である玉軸受装置は、上記回
転軸２の軸方向（図１の左右方向）に亙るアキシャル荷
重Ｆ_a を支承する為のもので、それぞれがアンギュラ型
である第一、第二の玉軸受５、６を組み合わせる事で構
成される。これら第一、第二の玉軸受５、６の接触角α
（後述する図２参照）の方向は、互いに逆方向で、且つ
互いの正面同士を対向させる正面組み合わせ（以下『Ｄ
Ｆ』とする。）としている。この為、上記回転軸２が図
１の左方に変位しようとする時は、同図で左方の第一の
玉軸受５がアキシャル荷重を支承し、同じく右方に変位
しようとする時は、右方の第二の玉軸受６がアキシャル
荷重を支承して、回転軸２及びロータ１が、ハウジング
３に対し変位する事を防止する。[0003] ball bearing apparatus is the subject of the present invention is intended for supporting an axial load F _a over the axial direction of the rotary shaft 2 (lateral direction in FIG. 1), the first respectively are angular type , And the second ball bearings 5 and 6 are combined. The contact angle α of these first and second ball bearings 5 and 6
The directions of (see FIG. 2 described later) are opposite to each other, and a front combination (hereinafter, referred to as “D
F ". ). For this reason, when the rotating shaft 2 is to be displaced to the left in FIG. 1, the first ball bearing 5 on the left in FIG. The right second ball bearing 6 supports the axial load to prevent the rotating shaft 2 and the rotor 1 from being displaced with respect to the housing 3.

【０００４】尚、アンギュラ型の玉軸受を１対組み合わ
せる事で何れの方向のアキシャル荷重も支承できる様に
する組み合わせの型としては、図１に示した構造とは逆
に、第一、第二の玉軸受の背面同士を対向させて組み合
わせる、所謂背面組み合わせがある。しかしながら、こ
の様な背面組み合わせの構造を採用した場合には、上記
第一、第二の玉軸受の内輪側の作用点の間隔が大きくな
って、回転軸が傾斜する事に対する曲げ剛性が大きくな
る。従って、スクリューコンプレッサの回転軸２の様
に、傾斜が比較的大きくなる部材の支持に上記背面組み
合わせを使用すると、上記作用点部分に過大な面圧が作
用し易くなる。過大な面圧は、異常発熱や玉の転動面並
びに軌道面の疲れ寿命が低下する原因となる為、好まし
くない。従って、本発明の対象となる玉軸受装置は、図
１に示す様なＤＦである。[0004] Incidentally, as a combination type in which an axial load in any direction can be supported by combining a pair of angular type ball bearings, the first and second types are opposite to the structure shown in FIG. There is a so-called back combination, in which the back surfaces of the ball bearings are combined so as to face each other. However, when such a back-to-back combination structure is employed, the interval between the action points on the inner ring side of the first and second ball bearings increases, and the bending rigidity against the inclination of the rotating shaft increases. . Therefore, when the back combination is used to support a member having a relatively large inclination, such as the rotating shaft 2 of the screw compressor, an excessive surface pressure is likely to act on the action point portion. Excessive surface pressure is not preferable because it causes abnormal heat generation and shortens the fatigue life of the ball rolling surface and the raceway surface. Therefore, the ball bearing device to which the present invention is applied is a DF as shown in FIG.

【０００５】又、本発明の対象となる玉軸受装置は、第
一、第二の玉軸受５、６を構成する玉１１、１２の転動
面と、内輪１０、１０外周面の内輪軌道１５、１５、外
輪１３、１３内周面の外輪軌道１４、１４との間にプラ
ス隙間（実際に存在する隙間）を残した状態で運転され
るものである。この理由は、やはり発熱や剥離寿命の低
下を防止する為である。本発明の対象となる玉軸受装置
が組み込まれるスクリューコンプレッサは、超高速（例
えばd_mn が７０万〜２００万mm・rpm）で回転する。この
為、上記各玉１１、１２に予圧付与を行ない、所謂マイ
ナスの隙間を持った第一第二の玉軸受５、６を使用し
た場合には、予圧に基づいてこれら第一、第二の玉軸受
５、６の内部荷重が増加し、やはり異常発熱や疲れ寿命
の低下の原因となる。The ball bearing device to which the present invention is applied includes a rolling surface of balls 11 and 12 constituting first and second ball bearings 5 and 6 and an inner ring raceway 15 of inner ring 10 and outer peripheral surface. , 15 and the outer races 13, 13 are operated with a plus clearance (a clearance that actually exists) between the inner races and the outer raceways 14, 14. The reason for this is also to prevent heat generation and a decrease in the peeling life. Screw compressors ball bearing apparatus to which the present invention is incorporated, to rotate at ultra high speed (e.g., d _m n is 700,000 to 2,000,000 mm · rpm). For this reason, when a preload is applied to each of the balls 11 and 12 and the first and second ball bearings 5 and 6 having a so-called minus gap are used, these first and second ball bearings are set based on the preload. The internal load of the ball bearings 5 and 6 increases, which also causes abnormal heat generation and shortens the fatigue life.

【０００６】従って、本発明の対象となる玉軸受装置
は、それぞれがプラス隙間を持ったアンギュラ型の第
一、第二の玉軸受５、６を、ＤＦで使用する。又、この
玉軸受装置により支持される回転軸２には、上記スクリ
ューコンプレッサの使用時に、前記ロータ１の回転に伴
なってほぼ一定方向（図１では右から左方向へ）のアキ
シャル荷重Ｆ_a が加わる。これは、ロータ１に加わる圧
力の方向が決まっている為である。尚、図１で７、８は
間座、９は抑え金である。Accordingly, the ball bearing device to which the present invention is applied uses the angular first and second ball bearings 5 and 6 each having a plus clearance in the DF. Further, the rotary shaft 2 which is supported by the ball bearing apparatus, in use of the screw compressor, axial load F _a substantially constant direction with the rotation of the rotor 1 (in Fig. 1 from right to left) Is added. This is because the direction of the pressure applied to the rotor 1 is determined. In FIG. 1, reference numerals 7 and 8 denote spacers and reference numeral 9 denotes a presser.

【０００７】ところで、上述の様に、それぞれがアンギ
ュラ型である第一、第二の玉軸受５、６をＤＦで組み合
わせて成る玉軸受装置は、回転軸２が高速で回転した場
合には、必ずしも十分な軸受寿命を得られない。この様
に高速回転に伴って軸受寿命が低下する原因に就いて、
図２により説明する。By the way, as described above, the ball bearing device in which the first and second ball bearings 5 and 6 each of which is of the angular type is combined with the DF, when the rotating shaft 2 rotates at a high speed, A sufficient bearing life cannot always be obtained. Regarding the cause of the shortening of the bearing life with high speed rotation,
This will be described with reference to FIG.

【０００８】図２は、高速回転で生じる遠心力に基づい
て、第一、第二の玉軸受５、６に加わる力を説明する為
の図である。上記回転軸２（図１）を高速で回転させた
場合、第一、第二の玉軸受５、６を構成する玉１１、１
２に遠心力Ｆ_c1が加わる。そして、これら第一、第二の
玉軸受５、６を構成する玉１１、１２は、この遠心力Ｆ
_c1に基づいて外輪軌道１４、１４に、Ｑ_in1 の力で、接
触角α方向から押し付けられる。そして、これら第一、
第二の玉軸受５、６を構成する玉１１、１２は、この力
Ｑ_in1 のアキシャル方向の分力Ｆ_ain1で各外輪軌道１
４、１４に、アキシャル方向に亙り押し付けられる。FIG. 2 is a view for explaining the force applied to the first and second ball bearings 5 and 6 based on the centrifugal force generated by high-speed rotation. When the rotating shaft 2 (FIG. 1) is rotated at a high speed, the balls 11, 1 constituting the first and second ball bearings 5, 6 are formed.
2, a centrifugal force F _c1 is applied. The balls 11 and 12 constituting the first and second ball bearings 5 and 6 have the centrifugal force F
Based on _c1 , they are pressed against the outer ring raceways 14, 14 with a force of Q _in1 from the direction of the contact angle α. And these first,
The balls 11, 12 constituting the second ball bearings 5, 6 are _subjected to the axial component component _Fain ₁ of the force Q _in1 to each outer raceway 1.
4, 14 are pressed in the axial direction.

【０００９】そして、この分力によりＦ_ain1＋Ｆ_ain1に
相当する内部アキシャル荷重が、互いに直列に組み合わ
された第一、第二の玉軸受５、６内で発生し、上記外輪
軌道１４、１４と玉１１、１２との接触面圧を増大させ
て、これら第一、第二の玉軸受５、６の疲れ寿命を低下
させてしまう。尚、高速回転時に於ける玉の遠心力を考
慮した玉軸受の疲れ寿命の計算は、１９５２年７月発行
のTrans. ASME 中に記載された論文である『The Life o
f High-Speed Ball Bearings』に示された理論に基づい
て行える。Then, an internal axial load corresponding to F _ain1 + F _ain1 is generated in the first and second ball bearings 5 and 6 which are combined in series with each other due to this component force. The contact surface pressure with the balls 11 and 12 is increased, and the fatigue life of the first and second ball bearings 5 and 6 is reduced. The calculation of the fatigue life of a ball bearing in consideration of the centrifugal force of the ball during high-speed rotation is described in “The Life o ASME” published in Trans. ASME issued in July 1952.
f High-Speed Ball Bearings ”.

【００１０】上述の様な原因による寿命低下を防止すべ
く、第一、第二の玉軸受５、６を構成する玉１１、１２
を、軽量なセラミック材により造ったり、或は玉１１、
１２として小径のものを使用し、遠心力Ｆ_c1を小さくす
る試みもなされている。更には、上記第一、第二の玉軸
受５、６の接触角αを何れも小さくする事で、上記遠心
力Ｆ_c1に基づくアキシャル方向の分力Ｆ_ain1を小さくす
る試みも行なわれている。In order to prevent the life from being shortened due to the above-mentioned causes, the balls 11, 12 constituting the first and second ball bearings 5, 6 are formed.
Is made of lightweight ceramic material, or ball 11,
Attempts have also been made to reduce the centrifugal force _Fc1 by using a small diameter as 12. Further, attempts have been made to reduce the axial component F _ain1 based on the centrifugal force F _c1 by reducing the contact angles α of the first and second ball bearings 5 and 6. .

【００１１】ところが、玉１１、１２を、縦弾性係数の
大きなセラミック材により造ると、特に外部荷重（例え
ば上記アキシャル荷重Ｆ_a ）を支承する側の第一の玉軸
受５に於いて、外輪軌道１４と玉１１との接触面圧が、
軸受鋼製の玉を使用した場合よりも増大する。この結
果、第一の玉軸受５の疲れ寿命が低下し、第一、第二の
玉軸受５、６を組み合わせて成る玉軸受装置全体の疲れ
寿命が低下してしまう。又、玉１１、１２として小径の
ものを使用したり接触角αを小さくすると、第一、第二
の玉軸受５、６のアキシャル方向の負荷容量（基本動定
格荷重）が小さくなり、やはり疲れ寿命が低下してしま
う。[0011] However, the balls 11 and 12, when made by a large ceramic material modulus, in the first ball bearing 5 side in particular supporting the external loads (e.g. the axial load F _a), the outer ring raceway The contact surface pressure between 14 and ball 11 is
It increases more than when a ball made of bearing steel is used. As a result, the fatigue life of the first ball bearing 5 decreases, and the fatigue life of the entire ball bearing device formed by combining the first and second ball bearings 5 and 6 decreases. Also, if small diameter balls 11 and 12 are used or the contact angle α is reduced, the axial load capacity (basic dynamic load rating) of the first and second ball bearings 5 and 6 is reduced, and the tires are also fatigued. The life will be shortened.

【００１２】この様な遠心力に基づく内部アキシャル荷
重を低減し、上記第一、第二の玉軸受５、６の疲れ寿命
を延長すべく、特開平５−２４８４３１号公報、同５−
２８０４８２号公報には、図３〜４に示す様な玉軸受装
置が記載されている。この玉軸受装置の場合には、使用
時に外部から回転軸２に加わるアキシャル荷重Ｆ_a を支
承する負荷側（図３〜４の左側）に設けられた、第一の
玉軸受５ａの接触角α₁ よりも、使用時に上記アキシャ
ル荷重Ｆ_a を支承しない反負荷側（図３〜４の右側）に
設けられた第二の玉軸受６ａの接触角α₂ を小さく（α
₂ ＜α₁ ）している。In order to reduce the internal axial load based on such centrifugal force and extend the fatigue life of the first and second ball bearings 5, 6, JP-A-5-248431 and JP-A-5-248431.
Japanese Patent Publication No. 280482 discloses a ball bearing device as shown in FIGS. In the case of the ball bearing apparatus is provided in axial load F _a the bears load applied from the outside to the rotary shaft 2 (left side in FIGS. 3-4) at the time of use, the contact angle of the first ball bearing 5a alpha than _1, the axial load F anti-load side without supporting the _a small contact angle alpha ₂ of the second ball bearing 6a provided (right side in FIGS. 3-4) at the time of use (alpha
₂ <α ₁ ).

【００１３】上述の様に、第一の玉軸受５ａの接触角α
₁ を第二の玉軸受６ａの接触角α₂よりも大きくする事
で、上記第一の玉軸受５ａの負荷容量を十分に確保でき
る。この場合に於いて、上記接触角α₁ を、前記従来か
ら一般的に知られた玉軸受装置に組み込まれた第一、第
二の玉軸受５、６の接触角αと同じ（α＝α₁ ）とすれ
ば、遠心力に基づいてこの第一の玉軸受５ａに作用する
アキシャル方向分力は、Ｆ_ain1となる。又、第二の玉軸
受６ａに加わるアキシャル方向分力はＦ_ain2となる。こ
の場合に於いて、この第二の玉軸受６ａの接触角α₂
は、上記第一の玉軸受５ａの接触角α₁ よりも小さい。
この為、上記第二の玉軸受６ａに加わるアキシャル方向
分力Ｆ_ain2は、前記従来装置の第二の玉軸受６に作用す
るアキシャル方向分力Ｆ_ain1よりも小さく（Ｆ_ain2＜Ｆ
_ain1）なる。この結果、第一、第二の玉軸受５ａ、６ａ
を組み合わせて成る玉軸受装置に発生する、遠心力に基
づく内部アキシャル荷重は、Ｆ_ain1＋Ｆ_ain2となり、前
記従来装置に生じる内部アキシャル荷重Ｆ_ain1＋Ｆ_ain1
よりも小さくなる。As described above, the contact angle α of the first ball bearing 5a
₁ By larger than the contact angle alpha ₂ of the second ball bearing 6a, can be sufficiently ensured load capacity of the first ball bearing 5a. In this case, the contact angle α ₁ is the same as the contact angle α of the first and second ball bearings 5 and 6 incorporated in the conventionally generally known ball bearing device (α = α). _{If 1} ), the axial component force acting on the first ball bearing 5a based on the centrifugal force is _Fain1 . Further, the axial component force applied to the second ball bearing 6a is _Fain2 . In this case, the contact angle α _{2 of} the second ball bearing 6a
Is smaller than the contact angle α1 of the _first ball bearing 5a.
For this reason, the axial component force _Fain2 applied to the second ball bearing 6a is smaller than the axial _component force _Fain1 acting on the second ball bearing 6 of the conventional device ( _Fain2 < _Fain2).
_ain1 ). As a result, the first and second ball bearings 5a, 6a
Occurs to the ball bearing apparatus comprising a combination of an internal axial load based on the centrifugal _{_force,} F ain1 + F ain2 next, the internal axial load _{_F} ain1 + F ain1 occurring conventional apparatus
Smaller than.

【００１４】この為、上記第一、第二の玉軸受５ａ、６
ａを構成する玉１１、１２と外輪軌道１４、１４との接
触面圧が低下する。そして、第一、第二の玉軸受５ａ、
６ａの転がり疲れ寿命の低下が抑えられ、これら第一、
第二の玉軸受５ａ、６ａを組み合わせて成る玉軸受装置
の寿命低下を抑える事ができる。接触角α₂ を小さくし
た第二の玉軸受６ａのアキシャル方向の負荷容量は小さ
くなるが、この第二の玉軸受６ａは稀に加わる軽荷重を
支承する、一種のバックアップ軸受としての機能しか持
たないので、負荷容量の減少は特に問題とはならない。For this reason, the first and second ball bearings 5a, 6
The contact surface pressure between the balls 11, 12 constituting the a and the outer raceways 14, 14 decreases. And the first and second ball bearings 5a,
6a, the reduction of the rolling fatigue life is suppressed.
It is possible to suppress a reduction in the life of the ball bearing device formed by combining the second ball bearings 5a and 6a. Load capacity of the axial direction of the second ball bearing 6a having a small contact angle alpha ₂ is reduced, the second ball bearing 6a is for supporting the light load applied to rare, only have the function as a kind of back-up bearing Since there is no load capacity reduction, there is no particular problem.

【００１５】[0015]

【発明が解決しようとする課題】ところが、上述の様な
各公報に記載された玉軸受装置の場合、反負荷側の第二
の玉軸受６ａの接触角α₂ を小さくした事に伴い、この
第二の玉軸受６ａの耐久性が不足しがちになる。この様
に、小さな接触角α₂ を有する第二の玉軸受６ａの耐久
性が不足する理由は次の通りである。However, in the case of the ball bearing devices described in the above publications, the contact angle α ₂ of the second ball bearing 6a on the non-load side is reduced, The durability of the second ball bearing 6a tends to be insufficient. Thus, the reason why the durability of the second ball bearing 6a is insufficient with a small contact angle alpha ₂ is as follows.

【００１６】前述した通り本発明の対象となる玉軸受装
置を構成する第一、第二の玉軸受５ａ、６ａは、プラス
隙間を持っている。そして、スクリューコンプレッサの
運転時に反負荷側の第二の玉軸受６ａには、アキシャル
荷重が殆ど加わらない（ラジアル荷重もころ軸受４によ
り支持されるので殆ど加わらない）。従って、この第二
の玉軸受６ａを構成する玉１２は、遠心力Ｆ_c2（図４）
に基づいて直径方向外方に変位する傾向となり、図５に
示す様に、各玉１２が外輪１３内周面の外輪軌道１４の
大径側端部（図５の左端部）に移動する。又、この状態
では、回転軸２（図３参照）から上記第第一の玉軸受５
ａの内輪１０に加わるアキシャル荷重Ｆ_a （図４）によ
り、内輪１０外周面の内輪軌道１５が、上記玉１２から
遠ざかる傾向となる。As described above, the first and second ball bearings 5a and 6a constituting the ball bearing device to which the present invention is applied have a plus clearance. During the operation of the screw compressor, the axial load is hardly applied to the second ball bearing 6a on the non-load side (the radial load is hardly applied because the roller load is also supported by the roller bearing 4). Therefore, the ball 12 constituting the second ball bearing 6a has a centrifugal force F _c2 (FIG. 4).
5, the balls 12 move to the large-diameter end (the left end in FIG. 5) of the outer raceway 14 on the inner peripheral surface of the outer race 13 as shown in FIG. In this state, the first ball bearing 5 is moved from the rotating shaft 2 (see FIG. 3).
Due to the axial load F _a (FIG. 4) applied to the inner ring 10 of a, the inner ring raceway 15 on the outer peripheral surface of the inner ring 10 tends to move away from the ball 12.

【００１７】この結果、図５に示す様に、上記各玉１２
と外輪軌道１４との接触角が０度になり、しかも各玉１
２の転動面と上記内輪軌道１５とが離れる傾向となる。
この状態では、上記各玉１２が転動しにくい為、各玉１
２の転動面と上記内輪軌道１５との接触部に滑りが生じ
る。そして、この滑りにより接触部が摩耗し、上記第二
の玉軸受６ａが早期に破損する。この為、図３〜４に示
す様に反負荷側の第二の玉軸受６ａの接触角α₂ を小さ
くした構造を採用した場合には、上記滑りに基づく早期
破損を防止する為、この第二の玉軸受６ａ、並びにこの
第二の玉軸受６ａと組み合わされた負荷側の第一の玉軸
受５ａにも予圧を付与した状態（マイナス隙間を持たせ
た状態）で使用せざるを得なかった。As a result, as shown in FIG.
The contact angle between the ball and the outer ring raceway 14 becomes 0 degree,
2 tends to separate from the inner raceway 15.
In this state, since each ball 12 is hard to roll, each ball 1
Slip occurs at the contact portion between the rolling surface of No. 2 and the inner raceway 15. Then, the contact portion is worn by the slip, and the second ball bearing 6a is damaged at an early stage. Therefore, in the case of employing a small structure, the contact angle alpha ₂ of the second ball bearing 6a of the anti-load side as shown in FIG. 3-4, to prevent premature failure based on the slip, the first The second ball bearing 6a and the first ball bearing 5a on the load side combined with the second ball bearing 6a must be used in a state where a preload is applied (a state in which a minus gap is provided). Was.

【００１８】ところが、この様に第一、第二の玉軸受５
ａ、６ａに予圧を付与した状態で、スクリューコンプレ
ッサの回転軸２を支持すると、予圧に基づいてこれら第
一、第二の玉軸受５ａ、６ａの内部荷重が増加し、異常
発熱や疲れ寿命の低下の原因となる事は、前述の通りで
ある。本発明の玉軸受装置は、上述の様な事情に鑑みて
発明したものである。However, as described above, the first and second ball bearings 5
When the rotary shaft 2 of the screw compressor is supported in a state in which the preload is applied to the a and 6a, the internal loads of the first and second ball bearings 5a and 6a increase based on the preload, resulting in abnormal heat generation and fatigue life. The cause of the decrease is as described above. The ball bearing device of the present invention has been invented in view of the above circumstances.

【００１９】尚、前述の論文『The Life of High-Speed
Ball Bearings』に示された理論は、本明細書中に記載
した計算寿命を求める場合の基礎となるものである。し
かしながら、この論文中には、玉軸受に予圧を付与せ
ず、プラス隙間を設けた状態で運転する場合の理論は開
示されていない。又、組み合わされた複数の玉軸受の接
触角等の条件を変えて遠心力の影響を減ずる方法も書か
れておらず、勿論、この様な方法による利害とその解決
方法も記載されていない。The above-mentioned paper "The Life of High-Speed"
The theory presented in Ball Bearings is the basis for determining the calculated lifetime described herein. However, this paper does not disclose a theory in which the ball bearing is operated without a preload and with a plus clearance provided. Also, there is no description of a method for reducing the effect of centrifugal force by changing conditions such as contact angles of a plurality of ball bearings combined, and, of course, does not describe the interests of such a method and the solution thereof.

【００２０】更に、接触角、玉径等が互いに異なる複数
の玉軸受を組み合わせた構造としては、前述した各公報
の他、MRC Bearing Services発行の『PUMPAC The MRC B
earing System 』、１９８５年のFAG のカタログ02105/
2EA の第９３頁、実開昭５３−９７７０１号公報、John
Wiley＆Sons,Inc.,発行の『Rolling Bearing Analysi
s』、綿林英一編著の『転がり軸受の選び方・使い方』
等に記載されたものも、従来から知られている。しかし
ながら、これら各刊行物に記載された構造は、本発明の
玉軸受装置と用途を異にするものである。従って、これ
ら何れの刊行物にも、上述の様な滑りに基づく摩耗、並
びにその解決方法に関する記載はない。Further, as a structure combining a plurality of ball bearings having different contact angles, ball diameters, etc., besides each of the above-mentioned publications, “PUMPAC The MRC B” issued by MRC Bearing Services
earing System ", FAG catalog 1985/02105 /
2EA, page 93, Japanese Utility Model Application Laid-Open No. 53-97701, John
Rolling Bearing Analysi, published by Wiley & Sons, Inc.
s ", Eiichi Watabayashi," How to choose and use rolling bearings "
And the like are conventionally known. However, the structure described in each of these publications has a different application from the ball bearing device of the present invention. Accordingly, none of these publications describes the above-mentioned slip-based wear and its solution.

【００２１】[0021]

【課題を解決するための手段】本発明の玉軸受装置は何
れも、前述の各公報等に記載された従来の玉軸受装置と
同様に、軸の外周面とハウジングの内周面との間に設け
られた、接触角の方向が互いに異なるアンギュラ型の第
一、第二の玉軸受を備え、使用時に上記軸とハウジング
との間に外部からほぼ一定方向に加わるアキシャル荷重
を、第二の玉軸受の接触角に比べて大きな接触角を有す
る第一の玉軸受により支承する。In each of the ball bearing devices of the present invention, like the conventional ball bearing devices described in the above-mentioned publications, the distance between the outer peripheral surface of the shaft and the inner peripheral surface of the housing is increased. Provided with angular first and second ball bearings having different contact angle directions from each other, and apply an axial load applied in a substantially constant direction from the outside between the shaft and the housing during use to the second. It is supported by a first ball bearing having a larger contact angle than the contact angle of the ball bearing.

【００２２】特に、本発明の玉軸受装置に於いては、上
記第一、第二の玉軸受はＤＦで配列されて、それぞれプ
ラスの隙間を持っている。又、上記第二の玉軸受の外輪
軌道の断面形状の曲率半径をｒ_e2とし、この第二の玉軸
受を構成する玉の外径をｄとした場合に、０．５３ｄ≦ｒ_e2≦０．５６ｄを満たし、上記第二の玉
軸受の内輪軌道の断面形状の曲率半径をｒ_i2とし、この
第二の玉軸受を構成する玉の外径をｄとした場合に、０．５０５ｄ≦ｒ_i2≦０．５２ｄを満たす事を特徴とし
ている。In particular, in the ball bearing device of the present invention, the first and second ball bearings are arranged in DF, each having a positive gap. When the radius of curvature of the cross-sectional shape of the outer ring raceway of the second ball bearing is r _e2, and the outer diameter of the ball constituting the second ball bearing is d, 0.53 d ≦ re ₂ ≦ 0 .56d, the radius of curvature of the cross-sectional shape of the inner ring raceway of the second ball bearing is r _i2, and the outer diameter of the ball constituting the second ball bearing is d, 0.505d ≦ r It is characterized by satisfying _i2 ≦ 0.52d.

【００２３】[0023]

【作用】上述の様に構成される本発明の玉軸受装置は、
前述の各公報に記載された玉軸受装置と同様に、使用時
に外部から加わるアキシャル荷重を支承する第一の玉軸
受の接触角が大きく、この第一の玉軸受の負荷容量が大
きい為、十分に大きなアキシャル荷重を支承できる。
又、使用時にこのアキシャル荷重を支承しない第二の玉
軸受は、遠心力に基づく内部アキシャル荷重の増大を抑
える。この為、１対のアンギュラ型の玉軸受を組み合わ
せて成る玉軸受装置全体としての、疲れ寿命の低下を抑
える事ができる。The ball bearing device of the present invention configured as described above has the following features.
Similarly to the ball bearing devices described in the above-mentioned publications, the contact angle of the first ball bearing that supports an axial load applied from the outside during use is large, and the load capacity of the first ball bearing is large. A large axial load can be supported.
The second ball bearing which does not support the axial load during use suppresses an increase in the internal axial load due to the centrifugal force. Therefore, it is possible to suppress a decrease in fatigue life of the entire ball bearing device formed by combining a pair of angular type ball bearings.

【００２４】特に、本発明の玉軸受装置の場合には、第
一、第二の玉軸受がＤＦで配列されており、しかもプラ
スの隙間を持っている為、この玉軸受装置の運転時にも
予圧が加わる事がなく、予圧に基づく発熱や疲れ寿命の
低下を来す事がない。しかも、第二の玉軸受の外輪軌道
及び内輪軌道の曲率半径を規制した為、使用時に殆ど荷
重が加わらない第二の玉軸受の玉と軌道との間で滑りが
生じる事がない。In particular, in the case of the ball bearing device of the present invention, since the first and second ball bearings are arranged in DF and have a positive clearance, even when the ball bearing device is operated. No preload is applied, and no heat or fatigue life is reduced due to preload. In addition, since the radius of curvature of the outer raceway and the inner raceway of the second ball bearing is regulated, there is no slippage between the ball of the second ball bearing to which a load is hardly applied during use and the raceway.

【００２５】即ち、上記第二の玉軸受の外輪軌道の断面
形状の曲率半径を大きくした事に伴い、遠心力に基づい
て玉が直径方向外方に変位した場合でも、この玉と外輪
軌道との接触角が０度とはならない。言い換えれば、玉
と外輪軌道とは常に接触角を持って接触する。従って、
玉の転動面と外輪軌道及び内輪軌道は常に接触したまま
となり、第二の玉軸受の回転時には、玉が確実に転動す
る。この結果、上記転動面と外輪軌道及び内輪軌道との
間に異常な滑りが発生する事がなくなり、この滑りに基
づく異常摩耗を防止できる。That is, even if the ball is displaced radially outward due to the centrifugal force due to the increase in the radius of curvature of the cross-sectional shape of the outer raceway of the second ball bearing, the ball and the outer raceway are not moved. Does not become 0 degree. In other words, the ball and the outer ring raceway always contact with a contact angle. Therefore,
The rolling surface of the ball, the outer ring raceway and the inner ring raceway always remain in contact, and the ball surely rolls when the second ball bearing rotates. As a result, abnormal slippage does not occur between the rolling surface and the outer raceway and inner raceway, and abnormal wear due to the slippage can be prevented.

【００２６】[0026]

【実施例】本発明の玉軸受装置は、前述の各公報に記載
された玉軸受装置と同様、図３〜４に示す様に、使用時
に外部から回転軸２に加わるアキシャル荷重Ｆ_a を支承
する、負荷側（図３〜４の左側）に設けられた第一の玉
軸受５ａの接触角α₁ よりも、使用時に上記アキシャル
荷重Ｆ_a を支承しない、反負荷側（図３〜４の右側）に
設けられた第二の玉軸受６ａの接触角α₂ を小さく（α
₂ ＜α₁ ）している。EXAMPLES ball bearing apparatus of the present invention, like the ball bearing apparatus described in each of the aforementioned publication, as shown in FIGS. 3-4, support the axial load F _a externally applied to the rotating shaft 2 at the time of use to, than the contact angle alpha ₁ of the first ball bearing 5a provided on the load side (left side in FIGS. 3-4), no support the axial load F _a at the time of use, the anti-load side (in FIGS. 3-4 The contact angle α ₂ of the second ball bearing 6a provided on the right side) is reduced (α
₂ <α ₁ ).

【００２７】この様に、第一の玉軸受５ａの接触角α₁
を第二の玉軸受６ａの接触角α₂ よりも大きくする事
で、上記第一の玉軸受５ａの負荷容量を十分に確保でき
る。又、第二の玉軸受６ａの接触角α₂ は、上記第一の
玉軸受５ａの接触角α₁ よりも小さい為、上記第二の玉
軸受６ａに加わる軸方向分力Ｆ_ain2は、前記従来装置の
第二の玉軸受６（図１〜２）に作用する軸方向分力Ｆ
_ain1よりも小さく（Ｆ_ain2＜Ｆ_ain1）なる。この結果、
第一、第二の玉軸受５ａ、６ａを組み合わせて成る玉軸
受装置に発生する、遠心力に基づく内部アキシャル荷重
が、前記図１〜２に示した従来装置に比べて小さくな
る。As described above, the contact angle α _{1 of} the first ball bearing 5a is
Is larger than the contact angle α ₂ of the second ball bearing 6a, so that the load capacity of the first ball bearing 5a can be sufficiently ensured. Also, since the contact angle α ₂ of the second ball bearing 6a is smaller than the contact angle α _{1 of} the first ball bearing 5a, the axial _component force _Fain2 applied to the second ball bearing 6a is Axial component force F acting on the second ball bearing 6 (FIGS. 1-2) of the conventional device
_smaller than _ain1 ( _Fain2 < _Fain1 ). As a result,
The internal axial load based on the centrifugal force generated in the ball bearing device formed by combining the first and second ball bearings 5a and 6a is smaller than that of the conventional device shown in FIGS.

【００２８】この為、上記各第一の玉軸受５ａ、６ａを
構成する玉１１、１２と外輪軌道１４、１４及び内輪軌
道１５、１５との接触面圧が低下する。そして、第一、
第二の玉軸受５ａ、６ａの転がり疲れ寿命が低下する事
が防止され、それぞれがアンギュラ型である第一、第二
の玉軸受５ａ、６ａを組み合わせて成る玉軸受装置の寿
命低下を抑える事ができる。For this reason, the contact surface pressure between the balls 11, 12 constituting each of the first ball bearings 5a, 6a, the outer raceways 14, 14, and the inner raceways 15, 15 is reduced. And first,
A reduction in the rolling fatigue life of the second ball bearings 5a and 6a is prevented, and a reduction in the life of the ball bearing device formed by combining the first and second ball bearings 5a and 6a, each of which is an angular type, is suppressed. Can be.

【００２９】例えば、本発明者が内径２５mm、外径６２
mm、幅１７mmのアンギュラ型玉軸受（７３０５型）を２
個、製作時の測定アキシャル隙間が０．０３０mmとなる
様にして、ＤＦで組み合わせ、６７kgf のアキシャル荷
重を受けつつ、回転数２３０００r.p.m.で回転する回転
軸に装着した場合に於ける、第一、第二の玉軸受の転が
り疲れ寿命を計算した結果を、下記の第１表に示す。
尚、アキシャル荷重を受ける第一の玉軸受（負荷側軸
受）の接触角は総て３０度とし、第二の玉軸受（反負荷
側軸受）の接触角は、１５度、３０度、４０度の３種類
に就いて計算した。For example, the present inventor has set an inner diameter of 25 mm and an outer diameter of 62 mm.
mm, 17 mm wide angular contact ball bearings (Type 7305)
When the DF is combined with a DF so that the measured axial gap at the time of manufacture is 0.030 mm, and it is mounted on a rotating shaft that rotates at 23000 rpm while receiving an axial load of 67 kgf, The results of calculating the rolling fatigue life of the second ball bearing are shown in Table 1 below.
The contact angles of the first ball bearings (load-side bearings) that receive an axial load are all 30 degrees, and the contact angles of the second ball bearings (anti-load-side bearings) are 15, 30, and 40 degrees. It calculated about three kinds of.

【００３０】尚、この計算中、回転軸２の外周面と内輪
１０、１０の内周面との間の締め代を０．０１２mm、ハ
ウジング３の内周面と外輪１３、１３の外周面との間の
締め代を０mm、外輪１３、１３の温度を８０℃、内輪１
０、１０の温度を８５℃、回転軸２の材質を鋼、ハウジ
ング３の材質を鋳鉄とした。尚、d_mn （玉１１、１２の
ピッチ円の直径d_m（mm）と回転数n （r.p.m.）との積）
は約１００万となる。又、ラジアル方向の荷重Ｆ_r は０
とした。これらの条件は、スクリューコンプレッサ用玉
軸受装置として一般的な条件である。そして、この様な
条件によれば、第一、第二の玉軸受５ａ、６ａに適正な
プラス隙間を付与する事で、d_mn が７０〜２００万、更
にはd_mn が８０〜３００万と言った様な、超高速回転を
連続して行なわせる事が可能となる。In this calculation, the interference between the outer peripheral surface of the rotating shaft 2 and the inner peripheral surfaces of the inner rings 10 and 10 is 0.012 mm, and the inner peripheral surface of the housing 3 and the outer peripheral surfaces of the outer rings 13 and 13 are different from each other. The interference between the outer ring is 0 mm, the temperature of the outer rings 13 and 13 is 80 ° C, and the inner ring 1
The temperature of 0 and 10 was 85 ° C., the material of the rotating shaft 2 was steel, and the material of the housing 3 was cast iron. Incidentally, d _m n (diameter of the pitch circle of the balls 11, 12 d _m (mm) and the rotational speed n (rpm) product of)
Is about one million. Also, the radial load _Fr is zero.
And These conditions are general conditions for a ball bearing device for a screw compressor. Then, according to these conditions, the first, second ball bearing 5a, By imparting an appropriate positive gap 6a, d _m n is from 70 to 2,000,000, more is d _m n 80 to 300 It is possible to continuously perform ultra-high-speed rotation as if it were a lot.

【００３１】第１表Table 1

【表１】 [Table 1]

【００３２】又、上記した各条件のうち、回転速度（d_m
n ）及びラジアル荷重Ｆ_r とアキシャル荷重Ｆ_a との比
（Ｆ_r ／Ｆ_a ）を変えて、これら各要素（d_mn 、Ｆ_r ／
Ｆ_a）が玉軸受装置の疲れ寿命に及ぼす影響に就いて計
算したところ、図６に示す様な結果が得られた。尚、こ
の計算の前提条件として、負荷側の第一の玉軸受５ａの
接触角は３０度、反負荷側の第二の玉軸受６ａの接触角
は１５度とした。この図６で、Ｘ軸は上記ラジアル荷重
Ｆ_r とアキシャル荷重Ｆ_a との比（Ｆ_r ／Ｆ_a）を、Ｙ
軸は回転速度（d_mn ）を、Ｚ軸は前記図１〜２に示した
従来構造の寿命Ｌ₀ に対する寿命の比（Ｌ／Ｌ₀ ）を、
それぞれ表している。この図６の記載から明らかな通
り、本発明の玉軸受装置によれば、回転速度（d_mn ）が
７０〜２００万、荷重比（Ｆ_r ／Ｆ_a ）が２以下（Ｆ_r
≦２Ｆ_a ）の範囲で顕著な効果を発揮する。In each of the above conditions, the rotational speed (d _m
by changing n) and the ratio of the radial load F _r and axial load F _a the (F _r / F _a), each of these elements (d _m n, F _r /
When the effect of F _a ) on the fatigue life of the ball bearing device was calculated, the results shown in FIG. 6 were obtained. As a precondition for this calculation, the contact angle of the first ball bearing 5a on the load side was 30 degrees, and the contact angle of the second ball bearing 6a on the non-load side was 15 degrees. In FIG. 6, the ratio of the X-axis is the radial load F _r and axial load F _a the _{_{(F r / F a),}} Y
Shaft rotational speed (d _m n), the ratio of the lifetime Z-axis with respect to the lifetime L ₀ of the conventional construction shown in FIG. 1-2 the (L / L _0),
Each is represented. As is evident from FIG. 6, according to the ball bearing apparatus of the present invention, the rotational speed (d _m n) is from 70 to 2,000,000, a load ratio (F _r / F _a) is 2 or less (F _r
A remarkable effect is exhibited in the range of ≦ 2F _a ).

【００３３】上述の計算結果から明らかな通り、反負荷
側の第二の玉軸受６ａの接触角α₂を小さくした玉軸受
装置は、玉１２と外輪軌道１４及び内輪軌道１５との間
に滑りが生じない限り、従来の玉軸受装置に比べて大幅
に寿命が長くなる。但し、上記表並びに図６に記載した
寿命は、反負荷側の第二の玉軸受６ａを構成する玉１２
の転動面と外輪軌道１４及び内輪軌道１５との間に滑り
が発生しない場合の数値を示している。この滑りが発生
しない限り、上記反負荷側の第二の玉軸受６ａの接触角
α₂ を小さくし、且つ第一、第二の玉軸受５ａ、６ａに
プラス隙間を付与する事で、図１〜２に示した従来装置
に比べて大幅な寿命延長を図れる。ところが、この様な
小さな接触角α₂ とプラス隙間とを組み合わせた場合に
は、しばしば玉１２の転動面と外輪軌道１４及び内輪軌
道１５との間に滑りが発生し、かえって玉軸受装置の寿
命を著しく短くする場合がある事は、前述の通りであ
る。As is clear from the above calculation results, the ball bearing device in which the contact angle α ₂ of the second ball bearing 6a on the non-load side is reduced has a sliding between the ball 12 and the outer raceway 14 and the inner raceway 15. As long as the occurrence does not occur, the life is greatly extended as compared with the conventional ball bearing device. However, the life described in the above table and FIG. 6 is the same as the life of the ball 12 constituting the second ball bearing 6a on the non-load side.
Are shown in the case where no slippage occurs between the rolling surface of No. 1 and the outer raceway 14 and the inner raceway 15. As long as the slip does not occur, the contact angle α ₂ of the second ball bearing 6a on the non-load side is reduced, and the first and second ball bearings 5a, 6a are provided with a plus clearance, as shown in FIG. The life can be greatly extended as compared with the conventional apparatuses shown in FIGS. However, when such a small contact angle α ₂ and the plus clearance are combined, slip often occurs between the rolling surface of the ball 12 and the outer ring raceway 14 and the inner raceway 15, and the ball bearing device has As described above, the life may be significantly shortened.

【００３４】この為、本発明の玉軸受装置の場合には、
上記第二の玉軸受６ａの外輪軌道１４の曲率半径ｒ_e2及
び内輪軌道１５の曲率半径ｒ_i2を、玉１２の外径ｄとの
関係で規制する事により、上記寿命の短縮に結び付く様
な滑りの発生を防止している。即ち、本発明の玉軸受装
置の場合には、図７に示す様に、第二の玉軸受６ａを構
成する外輪１３の外輪軌道１４の断面形状の曲率半径ｒ
_e2と、上記第二の玉軸受６ａを構成する内輪１０の内輪
軌道１５の断面形状の曲率半径ｒ_i2と、この第二の玉軸
受６ａを構成する玉１２の外径ｄとの関係を、０．５３ｄ≦ｒ_e2≦０．５６ｄであり、且つ０．５０５ｄ≦ｒ_i2≦０．５２ｄとしている。For this reason, in the case of the ball bearing device of the present invention,
The radius of curvature r _i2 radius of curvature r _e2 and inner ring raceway 15 of the outer ring raceway 14 of the second ball bearing 6a, By regulating the relationship between the outer diameter d of the ball 12, such as lead to shortening of the life span It prevents slippage. That is, in the case of the ball bearing device of the present invention, as shown in FIG. 7, the curvature radius r of the cross-sectional shape of the outer raceway 14 of the outer race 13 constituting the second ball bearing 6a.
The relationship between _e2 , the radius of curvature r _i2 of the cross-sectional shape of the inner raceway 15 of the inner race 10 of the second ball bearing 6a, and the outer diameter d of the ball 12 of the second ball bearing 6a is as follows: a 0.53d ≦ r _e2 ≦ 0.56d, is set to and 0.505d ≦ r _i2 ≦ 0.52d.

【００３５】上記各軌道１４、１５の断面形状の曲率半
径ｒ_e2、ｒ_i2を上述の範囲に規制する事により、これら
各軌道１４、１５と玉１２の転動面との滑りを防止でき
る理由に就いて、図７により説明する。この図７の鎖線
は、従来から一般的に知られたアンギュラ型玉軸受の外
輪軌道１４ａの断面形状を表している。この一般的な玉
軸受の場合、外輪軌道１４ａの断面形状の曲率半径ｒ_e2
´は、同図に実線で示した、本発明に於ける外輪軌道１
４の断面形状の曲率半径ｒ_e2よりも小さい（ｒ_e2´＜ｒ
_e2）。The reason that by restricting the radii of curvature r _e2 , r _i2 of the cross-sectional shape of each of the tracks 14, 15 to the above-mentioned range, slippage between these tracks 14, 15 and the rolling surface of the ball 12 can be prevented. Will be described with reference to FIG. The chain line in FIG. 7 represents the cross-sectional shape of the outer raceway 14a of the conventionally generally known angular ball bearing. In the case of this general ball bearing, the radius of curvature r _e2 of the cross-sectional shape of the outer raceway 14a is shown.
'Indicates the outer raceway 1 according to the present invention, which is indicated by a solid line in FIG.
Curvature of the fourth cross-sectional shape smaller than the radius r _{_e2} (r _e2 '<r
_e2 ).

【００３６】一方、玉１２の転動面と上記各軌道１４
ａ、１５との間にプラスの隙間を設けた場合には、遠心
力に基づいて玉１２が直径方向外方（図７の上方）に変
位する。そしてこの変位に基づいて、上記玉１２の転動
面と外輪軌道１４ａとの接触点Ｐが、この外輪軌道１４
ａの外方（内径が大きくなった部分）に移動する。そし
て、上記鎖線で示した従来形状の場合には、この接触点
Ｐが正規の位置からｌだけ変位した時点で、上記玉２と
外輪軌道１４ａとの接触角が０度になる。この様に接触
角が０度になった状態では、この外輪軌道１４ａが上記
玉１２を内輪軌道１５に向けて押圧する力が殆どなくな
る。従って、玉１２の転動面と内輪軌道１５との接触圧
が殆どなくなり、この玉１２は遠心力に基づいて上記外
輪軌道１４ａに押し付けられたまま、殆ど転動しなくな
る。この結果、この玉１２の転動面と内輪軌道１５との
間で異常な滑りが発生し、これら転動面及び内輪軌道１
５に著しい摩耗を発生する。On the other hand, the rolling surface of the ball 12 and each of the orbits 14
When a positive gap is provided between the balls 12 and 15, the ball 12 is displaced radially outward (upward in FIG. 7) based on the centrifugal force. Based on this displacement, the contact point P between the rolling surface of the ball 12 and the outer raceway 14a is determined by the outer raceway 14a.
It moves to the outside of (a) (the part where the inside diameter becomes large). In the case of the conventional shape indicated by the chain line, the contact angle between the ball 2 and the outer raceway 14a becomes 0 degree when the contact point P is displaced by 1 from the normal position. In the state where the contact angle becomes 0 degree, the force of the outer raceway 14a pressing the ball 12 toward the inner raceway 15 is almost eliminated. Therefore, the contact pressure between the rolling surface of the ball 12 and the inner raceway 15 is almost eliminated, and the ball 12 hardly rolls while being pressed against the outer raceway 14a based on the centrifugal force. As a result, abnormal sliding occurs between the rolling surface of the ball 12 and the inner raceway 15, and the rolling surface and the inner raceway 1
5 produces significant wear.

【００３７】これに対して本発明の場合には、第二の玉
軸受６ａを構成する外輪１３内周面の外輪軌道１４の断
面形状の曲率半径ｒ_e2を０．５３ｄ〜０．５６ｄと、内
輪軌道１５の断面形状の曲率半径ｒ_i2（＝０．５０５ｄ
〜０．５２ｄ）よりも大きくしている（ｒ_e2＞ｒ_i2）
為、上述の様な滑りに基づく著しい摩耗が発生しない。
即ち、上記曲率半径ｒ_e2を大きくした事に伴い、玉１２
の接触角が０度になる為には、上記接触点Ｐが正規の位
置からＬ（＞ｌ）だけ変位しなければならなくなる。こ
れに対して内輪軌道１５の断面形状の曲率半径ｒ_i2は比
較的小さい為、上記玉１２がＬ分変位する以前に、この
玉１２の転動面と上記内輪軌道１５とが当接する。この
結果、上記玉１２が遠心力に基づいて直径方向外方に変
位した場合でも、この玉１２の転動面と外輪軌道１４及
び内輪軌道１５との当接状態が確保され、且つ、当接部
に十分な面圧が作用する。従って、上記玉１２は玉軸受
装置の運転に伴って転動する。この結果、この玉１２の
転動面と内輪軌道１５との間に滑りが発生する事を防止
できて、前述した様な著しい摩耗の発生を確実に防止で
きる。[0037] In the case of the present invention contrary, the 0.53d~0.56d curvature radius r _e2 of the cross-sectional shape of the second outer ring raceway 14 of the outer ring 13 inner peripheral surfaces constituting the ball bearing 6a, The radius of curvature r _i2 (= 0.505d) of the cross-sectional shape of the inner raceway 15
Ｒ0.52d) (r _e2 > r _i2 )
Therefore, there is no remarkable abrasion due to slippage as described above.
That is, as the curvature radius _re2 is increased, the ball 12
In order that the contact angle becomes zero, the contact point P must be displaced from the normal position by L (> l). On the other hand, since the radius of curvature r _i2 of the cross-sectional shape of the inner raceway 15 is relatively small, the rolling surface of the ball 12 and the inner raceway 15 abut before the ball 12 is displaced by L. As a result, even when the ball 12 is displaced radially outward due to the centrifugal force, the contact state between the rolling surface of the ball 12 and the outer ring raceway 14 and the inner ring raceway 15 is ensured, and the contact state is maintained. Sufficient surface pressure acts on the part. Therefore, the ball 12 rolls with the operation of the ball bearing device. As a result, it is possible to prevent the occurrence of slippage between the rolling surface of the ball 12 and the inner raceway 15, and it is possible to reliably prevent the occurrence of the above-mentioned remarkable wear.

【００３８】上述の説明から明らかな通り、プラス隙間
を付与して無負荷状態で運転される第二の玉軸受６ａ
に、著しい摩耗に結び付く様な滑りが発生するのを防止
する為には、内輪軌道１５の断面の曲率半径ｒ_i2を小さ
く、外輪軌道１４の断面の曲率半径ｒ_e2を大きくすれば
良い。但し、軌道面の曲率半径は、玉１２の転動を円滑
に行なわせる必要上、玉１２の外径ｄとの関係で、或る
程度以上（０．５０５ｄ以上）確保する必要がある。従
って、内輪軌道１５の断面の曲率半径ｒ_i2も、０．５０
５ｄ以上としなければならない。一方、軌道面の曲率半
径を大きくし過ぎると、当該軌道面と玉１２の転動面と
の接触面積が狭くなり、上記第二の玉軸受６ａに負荷が
加わった場合、或は遠心力に基づいて玉１２が外輪軌道
１４に押し付けられた場合に、当該接触部に加わる面圧
が大きくなり過ぎて、この第二の玉軸受６ａの寿命を短
くする原因となる。従って、上記外輪軌道１４の断面の
曲率半径ｒ_e2を０．５６ｄを越えて大きくする事は好ま
しくない。As is clear from the above description, the second ball bearing 6a which is operated with no load by providing a plus clearance.
In order to prevent the occurrence of slippage which may lead to significant wear, the radius of curvature r _i2 of the cross section of the inner raceway 15 may be reduced and the radius of curvature r _e2 of the cross section of the outer raceway 14 may be increased. However, the radius of curvature of the raceway surface needs to be secured to a certain degree or more (0.505d or more) in relation to the outer diameter d of the ball 12 in order to smoothly roll the ball 12. Accordingly, the radius of curvature r _i2 of the cross section of the inner raceway 15 is also 0.50
Must be at least 5d. On the other hand, if the radius of curvature of the raceway surface is too large, the contact area between the raceway surface and the rolling surface of the ball 12 is reduced, and when a load is applied to the second ball bearing 6a, or due to centrifugal force. When the ball 12 is pressed against the outer ring raceway 14 based on this, the surface pressure applied to the contact portion becomes too large, which causes the life of the second ball bearing 6a to be shortened. Therefore, by increasing the curvature radius r _e2 of the cross section of the outer ring raceway 14 beyond 0.56d it is not preferable.

【００３９】以上の説明から、内輪軌道１５の断面形状
の曲率半径ｒ_i2は０．５０５ｄ以上（ｒ_i2≧０．５０５
ｄ）とし、外輪軌道１４の断面の曲率半径ｒ_e2は０．５
６ｄ以下（ｒ_e2≦０．５６ｄ）としなければならない事
が解る。更に、上記滑りの発生を防止する為には、内輪
軌道１５の断面の曲率半径ｒ_i2を外輪軌道１４の断面の
曲率半径ｒ_e2よりも十分に小さくする（ｒ_i2＜ｒ_e2）必
要がある事は、やはり前述の説明から明らかである。こ
れらの理由から上記各曲率半径ｒ_i2、ｒ_e2は、次の範囲
に規制しなければならない。From the above description, the radius of curvature r _i2 of the cross-sectional shape of the inner raceway 15 is 0.505d or more (r _i2 ≧ 0.505).
d), the radius of curvature _re2 of the cross section of the outer raceway 14 is 0.5
6d below (r _e2 ≦ 0.56d) and it can be seen that must be. Further, in order to prevent the occurrence of the slip, it is necessary to make the radius of curvature r _i2 of the cross section of the inner raceway 15 sufficiently smaller than the radius of curvature r _e2 of the cross section of the outer raceway 14 (r _i2 <r _e2 ). The thing is clear from the above description. For these reasons, the respective radii of curvature r _i2 and re ₂ must be restricted to the following ranges.

【００４０】０．５０５ｄ≦ｒ_i2≦０．５２ｄ０．５３ｄ≦ｒ_e2≦０．５６ｄ更に好ましくは、０．５０５ｄ≦ｒ_i2＜０．５１ｄ０．５３ｄ＜ｒ_e2≦０．５６ｄとする。そして、最も好ましい値は、ｒ_i2＝０．５０５ｄｒ_e2＝０．５４ｄである。上記各曲率半径ｒ_i2、ｒ_e2をこの様な範囲に規
制する事で、玉１２の転動面と内輪軌道１５及び外輪軌
道１４との当接部の接触面圧を低く抑えつつ、前述した
様な滑りに基づく異常摩耗の発生を防止できる。[0040] More preferably _{0.505d ≦ r i2 ≦ 0.52d 0.53d ≦} r e2 ≦ 0.56d, and _{0.505d ≦ r i2 <0.51d 0.53d <} r e2 ≦ 0.56d. The most preferred values are _{_{r i2 = 0.505d r e2 = 0.54d}} . By regulating the respective curvature radii r _i2 , re _e2 to such ranges, the contact surface pressure at the contact portion between the rolling surface of the ball 12 and the inner ring raceway 15 and the outer ring raceway 14 can be reduced, and the above-mentioned condition can be obtained. The occurrence of abnormal wear due to such slippage can be prevented.

【００４１】更に、運転時に負荷を受けつつ回転する第
一の玉軸受５ａの内輪軌道１５及び外輪軌道１４の断面
の曲率半径を規制する事で、プラス隙間を付与して無負
荷状態で運転される第二の玉軸受６ａに滑りが発生する
可能性を、より少なくできる。この理由に就いて図５に
より説明する。上記滑りは、第二の玉軸受６ａを構成す
る玉１２が遠心力に基づいて外輪軌道１４の大径側に変
位し、しかもこの玉１２の転動面と内輪軌道１５との当
接圧が零若しくは零に近くなった状態で発生する。従っ
て、上記滑りの発生を防止する為には、上記第二の玉軸
受６ａを構成する内輪１０外周面の内輪軌道１５が、玉
１２から離れる方向（図５の左方向）に変位するのを防
止する事が効果がある。一方、この内輪１０は、上記第
一の玉軸受５ａに加わるアキシャル荷重が大きい場合に
は、この第一の玉軸受５ａを構成する玉１１の転動面と
外輪軌道１４及び内輪軌道１５との当接部分の弾性変形
に基づいて、図５の左方向に変位する。この事から明ら
かな通り、負荷状態で運転される第一の玉軸受５ａの弾
性変形を少なく抑える事が、無負荷状態で運転される第
二の玉軸受６ａの滑り防止に効果がある。Further, by regulating the radii of curvature of the cross sections of the inner raceway 15 and the outer raceway 14 of the first ball bearing 5a, which rotates while receiving a load during operation, the operation is performed with no load by providing a plus clearance. The possibility that slippage will occur in the second ball bearing 6a can be further reduced. This reason will be described with reference to FIG. Due to the above-mentioned slippage, the ball 12 constituting the second ball bearing 6a is displaced toward the large-diameter side of the outer raceway 14 due to the centrifugal force, and the contact pressure between the rolling surface of the ball 12 and the inner raceway 15 is reduced. Occurs at zero or near zero. Therefore, in order to prevent the occurrence of the slip, the inner raceway 15 on the outer peripheral surface of the inner race 10 constituting the second ball bearing 6a must be displaced in a direction away from the ball 12 (leftward in FIG. 5). Prevention is effective. On the other hand, when the axial load applied to the first ball bearing 5a is large, the inner ring 10 is positioned between the rolling surface of the ball 11 constituting the first ball bearing 5a and the outer ring raceway 14 and the inner ring raceway 15. It is displaced leftward in FIG. 5 based on the elastic deformation of the contact portion. As is clear from this, suppressing the elastic deformation of the first ball bearing 5a operated under a load condition is effective in preventing the second ball bearing 6a operated under no load condition from slipping.

【００４２】上記弾性変形を少なく抑える為には、上記
第一の玉軸受５ａの外輪軌道１４の断面形状の曲率半径
ｒ_e1と、内輪軌道１５の断面形状の曲率半径ｒ_i1とを、
この第一の玉軸受５ａを構成する玉１１の外径ｄ´（本
実施例の場合にはｄ´＝ｄ）の１／２に近づける（但
し、前述した理由で０．５０５ｄ´以上）事が効果があ
る。但し、上記両曲率半径ｒ_e1、ｒ_i1を何れも上記外径
ｄ´の１／２に近づけると、外輪１３の中心軸と内輪１
０の中心軸とが傾斜した場合に、玉１１の転動面と外輪
軌道１４及び内輪軌道１５との接触部分に無理な力が加
わり、上記第一の玉軸受５ａが破損し易くなる。前述の
様に、本発明の玉軸受装置が組み込まれるスクリューコ
ンプレッサの場合、上記内輪１０を外嵌した回転軸２
（図３）が傾斜し易い為、この様な事態を避けるべく、
上記各曲率半径ｒ_e1、ｒ_i1を或る程度大きくする必要が
ある。In order to reduce the elastic deformation, the radius of curvature r _e1 of the cross-sectional shape of the outer raceway 14 of the first ball bearing 5a and the radius of curvature r _i1 of the cross-sectional shape of the inner raceway 15 are calculated as _follows .
The outer diameter d 'of the ball 11 constituting the first ball bearing 5a (d' = d in the case of this embodiment) should be close to 1/2 (however, 0.505d 'or more for the above-mentioned reason). Is effective. However, if both the radii of curvature r _e1 and r _i1 are close to の of the outer diameter d ′, the center axis of the outer ring 13 and the inner ring 1
When the center axis of 0 is inclined, an excessive force is applied to the contact portion between the rolling surface of the ball 11 and the outer raceway 14 and the inner raceway 15, and the first ball bearing 5a is easily damaged. As described above, in the case of the screw compressor in which the ball bearing device of the present invention is incorporated, the rotary shaft 2 on which the inner ring 10 is externally fitted is used.
(Fig. 3) is easy to tilt, so to avoid such a situation,
It is necessary to increase the curvature radii _re1 and _ri1 to some extent.

【００４３】そこで、本実施例の場合には、回転軸２の
傾斜により第一の玉軸受５ａが破損する事を防止しつ
つ、上記第一の玉軸受５ａを構成する内輪１０が図５の
左方向に変位する事を防止する為、上記曲率半径ｒ_e1、
ｒ_i1を、次の（１）〜（３）の条件を満たす様に規制す
る。ｒ_i1≦ｒ_e1 −−− （１）ｒ_i1≦０．５２ｄ´ −−− （２）ｒ_i1＋ｒ_e1≧１．０３ｄ´ −−− （３）Therefore, in the case of the present embodiment, the inner ring 10 constituting the first ball bearing 5a is prevented from being damaged by the inclination of the rotating shaft 2 while the inner race 10 of FIG. In order to prevent displacement to the left, the radius of curvature r _e1 ,
r _i1 is regulated so as to satisfy the following conditions (1) to (3). _{_{r i1 ≦ r e1 --- (1}} ) r i1 ≦ 0.52d' --- (2) r i1 + r e1 ≧ 1.03d' --- (3)

【００４４】上記（１）〜（３）の条件の内、（１）
（２）の条件は、玉１２の転動面と内輪軌道１５との当
接部の弾性変形を小さくして（当接部の接触面積を広く
して）、内輪１０が図５の左方向に変位する事を防止す
る為に必要である。又、（３）の条件は、回転軸２が傾
斜した場合に、第一の玉軸受５ａに無理な応力が加わる
事を防止する為に必要である。接触角α₂ が小さく、無
負荷状態で運転される第二の玉軸受６ａの外輪軌道１４
及び内輪軌道１５の断面形状の曲率半径ｒ_e2、ｒ_i2を前
述の様に規制するだけでなく、接触角α₁ が大きく、負
荷状態で運転される第一の玉軸受５ａの断面形状の曲率
半径ｒ_e1、ｒ_i1を上述の様に規制する事で、これら第
一、第二の玉軸受５ａ、６ａを組み合わせて成る玉軸受
装置の寿命を、前記従来の玉軸受装置に比べて大幅に延
長できる。又、前記各公報に記載された玉軸受装置に比
べても、寿命延長効果が確実になる。Of the above conditions (1) to (3), (1)
The condition (2) is such that the elastic deformation of the contact portion between the rolling surface of the ball 12 and the inner raceway 15 is reduced (the contact area of the contact portion is increased), and the inner ring 10 is moved leftward in FIG. It is necessary to prevent displacement. The condition (3) is necessary to prevent the excessive stress from being applied to the first ball bearing 5a when the rotating shaft 2 is inclined. Contact angle alpha ₂ is small, the second ball bearing 6a which is operated at no-load condition the outer ring raceway 14
In addition to restricting the radii of curvature r _e2 and _ri ₂ of the cross-sectional shape of the inner raceway 15 as described above, the curvature of the cross-sectional shape of the first ball bearing 5a operated under a load with a large contact angle α ₁ is also provided. By regulating the radii _re1 and _ri1 as described above, the life of the ball bearing device formed by combining the first and second ball bearings 5a and 6a is greatly increased as compared with the conventional ball bearing device. Can be extended. Further, the effect of extending the life is ensured as compared with the ball bearing devices described in the above publications.

【００４５】図８は、本発明の効果を確認する為に行な
った計算の結果を示している。計算の前提条件は、アキ
シャル荷重を１００kgf とした以外は、前記第１表を作
成した場合の条件と同じである。又、負荷側である第一
の玉軸受５ａの接触角は３０度、反負荷側である第二の
玉軸受６ａの接触角は１５度とした。この図８で、横軸
は第一、第二の玉軸受５ａ、６ａを組み合わせて成る玉
軸受装置の運転中のアキシャル方向の隙間（有効アキシ
ャル隙間）を、縦軸は計算寿命を、それぞれ表してい
る。又、実線Ａは、本発明の実施例の計算寿命を、破線
Ｂは本発明の限界部分から外れた曲率を持つ例であり、
実施する場合に注意を要する比較例の計算寿命を、それ
ぞれ表している。FIG. 8 shows the results of calculations performed to confirm the effects of the present invention. The prerequisites for the calculation are the same as those in the case of preparing Table 1 except that the axial load was set to 100 kgf. The contact angle of the first ball bearing 5a on the load side was 30 degrees, and the contact angle of the second ball bearing 6a on the non-load side was 15 degrees. In FIG. 8, the horizontal axis represents the axial clearance (effective axial clearance) during operation of the ball bearing device formed by combining the first and second ball bearings 5a and 6a, and the vertical axis represents the calculated life. ing. Further, a solid line A is an example having a calculated life of the embodiment of the present invention, and a broken line B is an example having a curvature deviating from the limit of the present invention,
The calculated lifetimes of the comparative examples that require attention when performing are shown.

【００４６】尚、上記実線Ａで表された本発明の実施
例、及び破線Ｂで表された比較例の、第一の玉軸受５ａ
及び第二の玉軸受６ａの外輪軌道１４及び内輪軌道１５
の曲率半径は、次の様に規制した。尚、玉１１、１２の
外径ｄ、ｄ´は同じ（ｄ＝ｄ´）である。実施例第一の玉軸受５ａｒ_i1＝０．５０５ｄ´ ｒ_e1＝０．５２５ｄ´ 第二の玉軸受６ａｒ_i2＝０．５０５ｄｒ_e2＝０．５４０ｄ比較例第一の玉軸受５ａｒ_i1＝０．５１５ｄ´ ｒ_e1＝０．５１５ｄ´ 第二の玉軸受６ａｒ_i2＝０．５１５ｄｒ_e2＝０．５１５ｄThe first ball bearing 5a of the embodiment of the present invention represented by the solid line A and the comparative example represented by the broken line B
Outer raceway 14 and inner raceway 15 of second ball bearing 6a
Was controlled as follows. The outer diameters d and d 'of the balls 11 and 12 are the same (d = d'). EXAMPLE first ball bearing _{_{5a r i1 = 0.505d' r e1 =}} 0.525d' second ball bearing _{_{6a r i2 = 0.505d r e2 =}} 0.540d Comparative Example first ball bearing 5a r _i1 = 0.515d' r _e1 = 0.515d' second ball bearing _{_{6a r i2 = 0.515d r e2 =}} 0.515d

【００４７】図８の記載から明らかな通り、本発明の玉
軸受装置によれば、従来構造に比べて大幅な寿命延長を
図れる。しかも、十分な有効アキシャル隙間を設けた状
態で長寿命を実現できる為、スクリューコンプレッサ用
回転軸の様に、しばしば傾斜角度を変化させつつ超高速
で回転する回転軸の支持を確実に行なえる。即ち、本発
明の実施例によれば、例えば上記有効アキシャル隙間を
５〜２５μｍ程度と比較的大きくしても、十分な耐久性
（計算寿命）を得る事ができる。As is clear from the description of FIG. 8, the life of the ball bearing device according to the present invention can be greatly extended as compared with the conventional structure. In addition, since a long service life can be realized with a sufficient effective axial clearance provided, a rotating shaft that rotates at an extremely high speed, such as a rotating shaft for a screw compressor, often changing the inclination angle can be surely supported. That is, according to the embodiment of the present invention, sufficient durability (calculated life) can be obtained even if the effective axial gap is relatively large, for example, about 5 to 25 μm.

【００４８】これに対して比較例の場合には、元々得ら
れる耐久性が実施例に比べて劣るだけでなく、有効アキ
シャル隙間が３〜６μｍを越えると、前述した様な滑り
の発生に基づいて耐久性が急激に低下してしまう。従っ
て、使用時には、有効アキシャル隙間が０〜６μｍの範
囲に収まる様に注意する必要がある。この様に、本発明
から外れた曲率半径を持つ場合には、耐久性が急激に低
下する有効アキシャル隙間の値の範囲がより小さいもの
となり、殆ど実用に供し得なくなる。On the other hand, in the case of the comparative example, the durability originally obtained is not only inferior to the example, but also when the effective axial gap exceeds 3 to 6 μm, the above-mentioned slippage is caused. And the durability drops sharply. Therefore, during use, care must be taken so that the effective axial gap falls within the range of 0 to 6 μm. As described above, when the radius of curvature deviates from the present invention, the range of the value of the effective axial gap at which the durability is sharply reduced becomes smaller, and it is hardly practical.

【００４９】尚、玉軸受装置を構成する第一の玉軸受５
ａの計算寿命Ｌ_5aと第二の玉軸受６ａの計算寿命Ｌ_6aと
の関係は、３Ｌ_5a≦Ｌ_6aとする事が、これら両玉軸受５
ａ、６ａを組み合わせて成る玉軸受装置の計算寿命Ｌ_T
を確保する面から好ましい。この理由は次の通りであ
る。複数の玉軸受を組み合わせて成る玉軸受装置の計算
寿命は、最も計算寿命が短い玉軸受（短寿命軸受）の計
算寿命と一致するのではなく、これよりも短くなる事、
計算寿命の長い玉軸受（長寿命軸受）の計算寿命に影響
される事、更には長寿命軸受の計算寿命が長くなる程、
玉軸受装置の計算寿命Ｌ_T が短寿命軸受の計算寿命に近
づく事は、従来から知られている。The first ball bearing 5 constituting the ball bearing device
The relationship between the calculated life L _5a of the second ball bearing 6a and the calculated life L _{6a of} the second ball bearing 6a is set to 3L _5a ≦ L _6a.
Calculated life L _{T of a} ball bearing device composed of a and 6a
It is preferable from the viewpoint of securing The reason is as follows. The calculated life of a ball bearing device composed of a combination of multiple ball bearings is not the same as the calculated life of the shortest ball bearing (short life bearing), but is shorter than this.
Being affected by the calculated life of a ball bearing with a long calculated life (long life bearing), and the longer the calculated life of a long life bearing,
Be calculated life L _T of the ball bearing device approaches the calculated life of the short life the bearings are known in the art.

【００５０】一方、使用時にアキシャル荷重を受ける第
一の玉軸受５ａの計算寿命Ｌ_5aを長くする事は難しい反
面、この様なアキシャル荷重を受けない第二の玉軸受６
ａの計算寿命Ｌ_6aを長くする事は容易である。従って、
本発明の玉軸受装置の場合には、第一の玉軸受５ａの計
算寿命Ｌ_5aを短寿命軸受の計算寿命と考え、第二の玉軸
受６ａの計算寿命Ｌ_6aを長寿命軸受の計算寿命と考える
事ができる。そこで、これら第一、第二の玉軸受５ａ、
６ａの計算寿命Ｌ_5a、Ｌ_6aの比（Ｌ_6a／Ｌ_5a）と、これ
ら両玉軸受５ａ、６ａを組み合わせて成る玉軸受装置の
計算寿命Ｌ_T と上記第一の玉軸受５ａの計算寿命Ｌ_5aと
の比（Ｌ_T ／Ｌ_5a）の関係を示すと、図９の様になる。
この図９は、従来から知られた組み合わせ軸受の寿命の
計算式に基づいて描いたものである。この図９から明ら
かな通り、第二の玉軸受６ａの計算寿命Ｌ_6aを第一の玉
軸受５ａの計算寿命Ｌ_5aの３倍以上にすれば、玉軸受装
置の計算寿命Ｌ_T を第一の玉軸受５ａの計算寿命Ｌ_5aの
８０％以上にできる。On the other hand, the other hand it is difficult to increase the calculated life L _5a of the first ball bearing 5a for receiving the axial load, in use, the second ball bearing 6 is not subjected to such axial load
It is easy to extend the calculated life _L6a of a. Therefore,
If the ball bearing apparatus of the present invention, the calculated life L _5a of the first ball bearing 5a considered calculated life of short life bearing, calculated life L _6a a long life bearing of calculated life of the second ball bearing 6a Can be considered. Therefore, these first and second ball bearings 5a,
6a the calculated life L _5a, the ratio of L _6a and (L _6a / L _5a), the calculated life of the calculated life L _T and the first ball bearing 5a of both ball bearings 5a, ball bearing apparatus comprising a combination of 6a When showing the relationship between the specific (L _T / L _5a) of the L _5a, it becomes as in FIG.
FIG. 9 is based on a conventionally known formula for calculating the life of a combination bearing. As is clear from FIG. 9, when the calculated life L _6a of the second ball bearing 6a is set to be three times or more the calculated life L _5a of the first ball bearing 5a, the calculated life L _T of the ball bearing device is made first. It can be of more than 80% of the calculated life L _5a of the ball bearing 5a.

【００５１】勿論、第二の玉軸受６ａの計算寿命Ｌ_6aを
第一の玉軸受５ａの計算寿命Ｌ_5aに比べて大幅に長くす
れば、上記玉軸受装置の計算寿命Ｌ_T を第一の玉軸受５
ａの計算寿命Ｌ_5aに、より近づける事ができる。但し、
第二の玉軸受６ａの計算寿命Ｌ_6aを第一の玉軸受５ａの
計算寿命Ｌ_5aの３倍を大幅に越えて延長しても、計算寿
命Ｌ_6aを伸ばす事に要するコストに比べて、それにより
得られる玉軸受装置の計算寿命Ｌ_T の延長効果は少な
い。従って、全体のコストを考えた場合、第二の玉軸受
６ａの計算寿命Ｌ_6aは、上記第一の玉軸受５ａの計算寿
命Ｌ_5aの３倍を越える程度にする事が適当である。Of course, if the calculated life L _6a of the second ball bearing 6a is made much longer than the calculated life L _5a of the first ball bearing 5a, the calculated life L _T of the above-mentioned ball bearing device can be made first. Ball bearing 5
It can be closer to the calculated life _L5a of a. However,
Even if the calculated life L _6a of the second ball bearing 6a is extended significantly more than three times the calculated life L _5a of the first ball bearing 5a, compared to the cost required to extend the calculated life L _6a , prolonging effect of the calculated life L _T as obtained thereby ball bearing apparatus is small. Therefore, considering the overall cost, it is appropriate that the calculated life L _6a of the second ball bearing 6a exceeds three times the calculated life L _5a of the first ball bearing 5a.

【００５２】次に、図１０は本発明の第二実施例を示し
ている。本実施例の場合、第一、第二の玉軸受５ａ、６
ａの玉１１、１２を、保持器１６、１６により転動自在
に保持している。各保持器１６、１６は、所謂もみ抜き
保持器と呼ばれるもので、円筒状の主部１７、１７に玉
１１、１２を転動自在に保持する為のポケット１８、１
８を形成して成る。この様な保持器１６、１６は、それ
ぞれ第一、第二の玉軸受５ａ、６ａを構成する外輪１
３、１３の内側に、所謂外輪案内で装着している。特に
図示の実施例では、これら各保持器１６、１６の軸方向
（図１０の左右方向）両端部外周面と、外輪１３、１３
の内周面両肩部、即ち外輪軌道１４、１４から外れた部
分とを、微小隙間を介して互いに対向させて、所謂外輪
両肩案内としている。Next, FIG. 10 shows a second embodiment of the present invention. In the case of the present embodiment, the first and second ball bearings 5a, 6
The balls 11 and 12 of a are held by the retainers 16 and 16 so as to freely roll. Each of the retainers 16, 16 is a so-called milled retainer, and has pockets 18, 1 for rollingly holding the balls 11, 12 in cylindrical main parts 17, 17.
8 is formed. Such retainers 16, 16 are used for the outer race 1 constituting the first and second ball bearings 5a, 6a, respectively.
The outer ring guides 3 and 13 are mounted inside. In particular, in the illustrated embodiment, the outer peripheral surfaces of both ends of the retainers 16, 16 in the axial direction (left-right direction in FIG. 10) and the outer races 13, 13
Of the inner peripheral surface, i.e., the portions deviated from the outer ring raceways 14, 14, are opposed to each other with a small gap therebetween, thereby forming a so-called outer ring double shoulder guide.

【００５３】この様に、上記各保持器１６、１６の案内
状態を外輪案内とする理由は、次の通りである。即ち、
玉軸受装置としての一般的な使用状態、即ち、d_mn が７
０万以下で、且つ第一、第二の玉軸受５ａ、６ａに予圧
が付与されて、運転時に振動の発生が少ない状態であれ
ば、保持器１６、１６の案内条件を特に規制する必要は
ない。一方、本発明の対象となる、スクリューコンプレ
ッサに組み込まれる玉軸受装置の場合には、d_mn が７０
万を上回る超高速領域で、プラスのアキシャル隙間を付
与した状態で運転される。しかも、スクリューコンプレ
ッサ特有のアキシャル荷重の変動により振動が多くな
る。この様な使用状態で、上記各保持器１６、１６の案
内面の摩耗を抑える為には、この保持器１６、１６の傾
斜を抑える事ができる外輪案内とする事が必要である。The reason why the guide state of each of the cages 16 and 16 is set to the outer ring guide is as follows. That is,
General state of use of the ball bearing apparatus, i.e., d _m n is 7
If the pressure is not more than 100,000 and the first and second ball bearings 5a, 6a are preloaded and vibration is less generated during operation, it is not necessary to particularly restrict the guiding conditions of the retainers 16, 16. Absent. On the other hand, in the case of the subject to, the ball bearing device incorporated in the screw compressor of the present invention, d _m n is 70
It is operated in a super-high speed range exceeding 10,000 with a positive axial clearance. Moreover, the vibration increases due to the fluctuation of the axial load peculiar to the screw compressor. In such a state of use, in order to suppress the wear of the guide surfaces of the retainers 16, 16, it is necessary to provide an outer ring guide capable of suppressing the inclination of the retainers 16, 16.

【００５４】これに対して、これら保持器１６、１６を
転動体案内、或は内輪案内とした場合には、運転時の遠
心力等に基づいて（内輪１０及び外輪１３に比べて剛性
が低い）保持器１６、１６の径が弾性的に広がる事に伴
い、案内面の間隔が広がって、十分な案内を行なえなく
なる。この状態でこれら保持器１６、１６に運転時の振
動が加わると、保持器１６、１６の表面と相手面とが狭
い面積で接触し、異常摩耗等の不都合の原因となる。外
輪案内とした場合には、遠心力に伴って保持器１６、１
６の径が広がる傾向となっても、案内面（保持器１６、
１６の外周面及び外輪１３、１３の内周面）の間隔が広
がる事はない。又、この案内面の間には潤滑油が存在す
る為、この案内面同士が直接擦れ合う事はない。従っ
て、保持器１６、１６を外輪案内とする事で、上記案内
面の異常摩耗を有効に防止できる。尚、保持器１６、１
６として銅合金製の保持器を使用すれば、この保持器１
６、１６の強度を高くして、保持器１６、１６の損傷防
止を有効に図れる。On the other hand, when the cages 16 and 16 are used as rolling element guides or inner ring guides, the rigidity is lower than the inner ring 10 and the outer ring 13 due to centrifugal force and the like during operation. ) As the diameters of the retainers 16 and 16 are elastically increased, the interval between the guide surfaces is increased, and sufficient guidance cannot be performed. When vibrations during operation are applied to the cages 16 in this state, the surfaces of the cages 16 and 16 come into contact with each other with a small area, which causes inconvenience such as abnormal wear. When the outer ring guide is used, the retainers 16, 1
Even if the diameter of 6 becomes wider, the guide surface (cage 16,
The distance between the outer peripheral surface of the outer ring 16 and the inner peripheral surface of the outer rings 13, 13 does not increase. Since lubricating oil exists between the guide surfaces, the guide surfaces do not directly rub against each other. Therefore, by using the cages 16 and 16 as outer ring guides, abnormal wear of the guide surfaces can be effectively prevented. In addition, the retainers 16, 1
If a copper alloy cage is used as 6, this cage 1
By increasing the strength of the retainers 6, 16, the retainers 16, 16 can be effectively prevented from being damaged.

【００５５】その他、プラス隙間を設ける点、第二の玉
軸受６ａの接触角を第一の玉軸受５ａの接触角よりも小
さくする点、各玉軸受５ａ、６ａの外輪軌道１４、１４
及び内輪軌道１５、１５の断面形状の曲率半径を規制す
る点は、上述した第一実施例と同様である。尚、使用時
のd_mn がそれ程高くなく、従って保持器１６、１６に加
わる遠心力が限られたものであれば、図１１に示す第三
実施例の様に、各保持器１６、１６を外輪片肩案内とし
ても良い。In addition, a point where a plus clearance is provided, a point where the contact angle of the second ball bearing 6a is smaller than a contact angle of the first ball bearing 5a, and the outer ring raceways 14 and 14 of each of the ball bearings 5a and 6a are provided.
The point of restricting the radius of curvature of the cross-sectional shape of the inner raceways 15, 15 is the same as in the first embodiment described above. Incidentally, as long as d _m n in use is not so high, therefore the centrifugal force applied to the cage 16, 16 is limited, as in the third embodiment shown in FIG. 11, the retainer 16, 16 May be used as the outer ring one shoulder guide.

【００５６】次に、図１２は本発明の第四実施例とし
て、スクリューコンプレッサの回転軸の支持部分のより
具体化した構造を示している。回転軸２の外周面とハウ
ジング３の内周面との間には、このハウジング３内に収
納されたロータ１の側から順に、ラビリンスシール等の
非接触型のシール、或はメカニカルシール等の接触型の
シール等のシール装置１９と、ラジアル荷重を支承する
為のころ軸受４と、本発明の玉軸受装置を構成する第一
の玉軸受５ａ及び第二の玉軸受６ａとを、互いに直列に
配設している。又、上記ころ軸受４の内輪２０と上記第
一の玉軸受５ａとの間には間座７を、上記ころ軸受４の
外輪２１と上記第一の玉軸受５ａの外輪１３との間には
ノズルリング２２を、それぞれ挟持している。Next, FIG. 12 shows, as a fourth embodiment of the present invention, a more specific structure of a supporting portion of a rotary shaft of a screw compressor. Between the outer peripheral surface of the rotating shaft 2 and the inner peripheral surface of the housing 3, in order from the side of the rotor 1 housed in the housing 3, a non-contact type seal such as a labyrinth seal or a mechanical seal A sealing device 19 such as a contact type seal, a roller bearing 4 for supporting a radial load, and a first ball bearing 5a and a second ball bearing 6a constituting the ball bearing device of the present invention are connected in series with each other. It is arranged in. Further, the spacer 7 is provided between the inner ring 20 of the roller bearing 4 and the first ball bearing 5a, and between the outer ring 21 of the roller bearing 4 and the outer ring 13 of the first ball bearing 5a. The nozzle rings 22 are respectively sandwiched.

【００５７】上記ノズルリング２２の両側面内周寄り部
分は、内周縁に向かう程互いに近づく方向に傾斜した、
円錐凹面状の傾斜面としている。そして両傾斜面の円周
方向１個所乃至は複数個所に、ノズル孔２３、２３を開
口させている。各ノズル孔２３、２３は、上記傾斜面に
対して垂直方向に形成している。上記傾斜面を設けるの
は、傾斜方向のノズル孔２３、２３を形成する作業の容
易化を図る為である。従って、これら各ノズル孔２３、
２３は、直径方向斜め内方に向けて開口している。そし
て、これら各ノズル孔２３、２３には、上記ハウジング
３及び上記ノズルリング２２内に形成した給油通路２４
を介して、潤滑油を送り込み自在としている。The inner peripheral portions of both sides of the nozzle ring 22 are inclined so as to approach each other toward the inner peripheral edge.
It is a conical concave inclined surface. The nozzle holes 23 are opened at one or a plurality of positions in the circumferential direction of both inclined surfaces. Each of the nozzle holes 23 is formed in a direction perpendicular to the inclined surface. The provision of the inclined surface is for facilitating the operation of forming the nozzle holes 23 in the inclined direction. Therefore, each of these nozzle holes 23,
Reference numeral 23 is open diagonally inward in the diameter direction. Each of the nozzle holes 23 has an oil supply passage 24 formed in the housing 3 and the nozzle ring 22.
, Lubricating oil can be sent freely.

【００５８】一方、上記第一、第二の玉軸受５ａ、６ａ
に組み込まれた保持器１６ａ、１６ａの軸方向（図１２
の左右方向）両端部内周面は、端縁部に向かう程内径が
大きくなる、円錐凹面状の傾斜面としている。従って、
上記各玉軸受５ａ、６ａを構成する内輪１０、１０の外
周面と上記各保持器１６ａ、１６ａの内周面との間に存
在する空間の、直径方向（図１２の上下方向）に亙る幅
寸法は、開口端部に向かう程大きくなる。この結果、上
記ノズル孔２３から噴出する潤滑油を、上記第一、第二
の玉軸受５ａ、６ａ内に送り込む効率が向上する。従っ
て、これら各玉軸受５ａ、６ａにより構成される玉軸受
装置が超高速で運転される場合にも、これら各玉軸受５
ａ、６ａ内に十分量の潤滑油を取り込む事が可能とな
る。On the other hand, the first and second ball bearings 5a, 6a
The axial direction of the retainers 16a, 16a incorporated in
The inner peripheral surfaces at both ends are inclined surfaces having a conical concave shape, the inner diameter of which increases toward the edge. Therefore,
The width of the space existing between the outer peripheral surfaces of the inner rings 10 and 10 constituting the ball bearings 5a and 6a and the inner peripheral surfaces of the retainers 16a and 16a in the diametric direction (vertical direction in FIG. 12). The size increases toward the opening end. As a result, the efficiency with which the lubricating oil ejected from the nozzle hole 23 is fed into the first and second ball bearings 5a and 6a is improved. Therefore, even when the ball bearing device constituted by the respective ball bearings 5a and 6a is operated at an extremely high speed, the respective ball bearings 5a and 6a are operated.
It becomes possible to take in a sufficient amount of lubricating oil into a and 6a.

【００５９】ノズル孔２３が開口する方向、並びに上記
各保持器１６ａ、１６ａの内周面の形状を上述の様にし
た理由は、次の通りである。即ち、本発明の玉軸受装置
は、前述した従来装置に比べて寿命を大幅に延長する事
ができる。従って、従来と同じ大きさで造った場合に要
求寿命を大幅に上回る様な場合には、第一、第二の玉軸
受５ａ、６ａの寸法系列を小さくして、内径を変える事
なく外径を小さくする事ができる。これにより、玉軸受
装置の小型化、低廉化が可能となるが、寸法系列の小さ
な玉軸受は内部空間が狭い。この為、スクリューコンプ
レッサで一般的に行なわれているジェット給油（高速回
転に基づいて軸受の内部に発生する空気の壁を破って、
軸受の内部に潤滑油を送り込む潤滑方法）を行ないにく
くなる。そこで、上述の様なノズル孔２３、２３の開口
方向と保持器１６ａ、１６ａの内周面形状とを採用する
事により、上記ジェット給油を可能とする。The reason why the direction in which the nozzle holes 23 are opened and the shapes of the inner peripheral surfaces of the retainers 16a and 16a are as described above is as follows. That is, the life of the ball bearing device of the present invention can be greatly extended as compared with the above-described conventional device. Therefore, in the case where the required life is significantly exceeded when the bearing is manufactured in the same size as the conventional one, the dimensional series of the first and second ball bearings 5a and 6a is reduced, and the outer diameter is changed without changing the inner diameter. Can be reduced. As a result, the size and cost of the ball bearing device can be reduced, but the internal space of a ball bearing having a small size series is narrow. For this reason, jet lubrication generally performed in screw compressors (breaks the wall of air generated inside the bearing based on high speed rotation,
Lubrication method for feeding lubricating oil into the bearing). Therefore, the above-described jet refueling is enabled by adopting the opening directions of the nozzle holes 23, 23 and the inner peripheral surface shapes of the retainers 16a, 16a as described above.

【００６０】尚、上記ノズル孔２３の回転軸２に対する
傾斜角度は、１５度程度が最適であるが、１０〜２０度
の範囲で設定可能である。又、このノズル孔２３を有す
る、外輪間座としての機能を兼ね備えたノズルリング２
２は、硬度が HRc５６〜６６（更に好ましくは HRc６０
〜６２）の範囲の鋼材により造るのが好ましい。これ
は、スクリューコンプレッサの運転時に生じる特有の振
動により、互いに当接するこのノズルリング２２の両側
面と、ころ軸受４及び第一の玉軸受５ａの外輪２１、１
３の端面とにフレッチング摩耗が発生する事を防止する
為である。The angle of inclination of the nozzle hole 23 with respect to the rotation axis 2 is optimally about 15 degrees, but can be set in the range of 10 to 20 degrees. Further, the nozzle ring 2 having the nozzle hole 23 and having a function as an outer ring spacer is also provided.
2 has a hardness of HRc 56 to 66 (more preferably HRc 60
-62). This is because the two sides of the nozzle ring 22 which come into contact with each other and the outer races 21 and 1 of the roller bearing 4 and the first ball bearing 5a due to the specific vibration generated during the operation of the screw compressor.
This is for preventing the occurrence of fretting wear on the end face 3.

【００６１】その他、プラス隙間を設ける点、第二の玉
軸受６ａの接触角を第一の玉軸受５ａの接触角よりも小
さくする点、各玉軸受５ａ、６ａの外輪軌道１４、１４
及び内輪軌道１５、１５の断面形状の曲率半径を規制す
る点は、前述した第一実施例及び上述した第二〜第三実
施例と同様である。In addition, a point where a plus clearance is provided, a point where the contact angle of the second ball bearing 6a is smaller than a contact angle of the first ball bearing 5a, and the outer ring raceways 14 and 14 of each of the ball bearings 5a and 6a are provided.
The point of restricting the radius of curvature of the cross-sectional shape of the inner ring raceways 15, 15 is the same as in the above-described first embodiment and the above-described second to third embodiments.

【００６２】又、図１３は本発明の第五実施例を示して
いる。本実施例の場合には、反負荷側の第二の玉軸受６
ａを構成する玉１２ａを、負荷側の第一の玉軸受５ａを
構成する玉１１よりも小径にしている。玉１２ａを小径
にした分だけ、使用時にこの玉１２ａに加わる遠心力に
基づいて玉軸受装置の内部で発生する内部アキシャル荷
重が小さくなり、玉軸受装置の耐久性をより向上させる
事ができる。小径の玉１２ａを使用する事で負荷容量は
小さくなるが、第二の玉軸受６ａに加わる負荷は小さい
ので、小径の玉１２ａを使用した場合でも、実用上十分
な耐久性を確保できる。その他、プラス隙間を設ける
点、第二の玉軸受６ａの接触角を第一の玉軸受５ａの接
触角よりも小さくする点、第一、第二の玉軸受５ａ、６
ａの外輪軌道１４、１４及び内輪軌道１５、１５の断面
形状の曲率半径を規制する点は、上述した各実施例と同
様である。FIG. 13 shows a fifth embodiment of the present invention. In the case of this embodiment, the second ball bearing 6 on the non-load side
The diameter of the ball 12a constituting the first ball bearing a is smaller than the diameter of the ball 11 constituting the first ball bearing 5a on the load side. Due to the smaller diameter of the ball 12a, the internal axial load generated inside the ball bearing device based on the centrifugal force applied to the ball 12a during use becomes smaller, and the durability of the ball bearing device can be further improved. Although the load capacity is reduced by using the small-diameter ball 12a, the load applied to the second ball bearing 6a is small, so that even when the small-diameter ball 12a is used, practically sufficient durability can be secured. In addition, a point where a plus gap is provided, a point where the contact angle of the second ball bearing 6a is smaller than a contact angle of the first ball bearing 5a, and the first and second ball bearings 5a and 6
The point of restricting the radius of curvature of the cross-sectional shape of the outer raceways 14, 14 and the inner raceways 15, 15 of a is the same as in the above-described embodiments.

【００６３】更に、図示は省略したが、反負荷側の第二
の玉軸受を構成する玉の数を、負荷側の第一の玉軸受を
構成する玉の数よりも少なくする事で、上記内部アキシ
ャル荷重を小さくする事もできる。玉の数を減らす事で
負荷容量は小さくなるが、第二の玉軸受に加わる負荷は
小さいので、玉の数を減らした場合でも、実用上十分な
耐久性を確保できる。この様に第二の玉軸受を構成する
玉の数ｍを第一の玉軸受を構成する玉の数ｎよりも少な
く（ｍ＜ｎ）する場合、上記少ない数ｍを多い数ｎの７
０〜８０％（ｍ＝（０．７〜０．８）ｎ）に規制する事
が好ましい。８０％を越える数の玉を組み込んだ場合に
は、内部アキシャル荷重を低減する効果が不十分にな
る。反対に、７０％に満たない場合には、隣り合う玉の
間隔が広くなり過ぎて、円滑な回転を妨げてしまう。
尚、この様に、反負荷側の第二の玉軸受を構成する玉の
数を、負荷側の第一の玉軸受を構成する玉の数よりも少
なくする技術は、本発明と組み合わせて実施できる他、
本発明とは独立した形でも実施できる事は明らかであ
る。Although not shown, the number of balls constituting the second ball bearing on the non-load side is made smaller than the number of balls constituting the first ball bearing on the load side. The internal axial load can be reduced. Although the load capacity is reduced by reducing the number of balls, the load applied to the second ball bearing is small, so that practically sufficient durability can be ensured even when the number of balls is reduced. As described above, when the number m of the balls constituting the second ball bearing is smaller than the number n of the balls constituting the first ball bearing (m <n), the small number m is increased to the number n of 7
It is preferable to regulate to 0 to 80% (m = (0.7 to 0.8) n). When the number of balls exceeding 80% is incorporated, the effect of reducing the internal axial load becomes insufficient. On the other hand, if it is less than 70%, the distance between adjacent balls becomes too wide, which hinders smooth rotation.
Incidentally, as described above, the technology of reducing the number of balls constituting the second ball bearing on the non-load side to be smaller than the number of balls constituting the first ball bearing on the load side is implemented in combination with the present invention. Other than
Obviously, the present invention can be implemented in a form independent of the present invention.

【００６４】同様に、反負荷側の第二の玉軸受を構成す
る玉のみを、軸受鋼に比べて軽量なセラミック製とする
事で、上記内部アキシャル荷重の低減を図る事もでき
る。セラミック製の玉を使用する事により、前述の様に
玉の転動面と当接する軌道面に加わる面圧が高くなる
が、第二の玉軸受に加わる負荷は小さいので、過大な面
圧が作用する事はなく、実用上十分な耐久性を確保でき
る。勿論、この様に反負荷側の玉軸受を構成する玉の径
を小さくしたり、玉の数を減らしたり、セラミック製の
玉を使用する技術は、単独で使用する他、組み合わせて
使用する事もできる。この様に、内部アキシャル荷重を
より小さくできる結果、d_mn が８０〜３００と言った様
なより超高速回転で、ラジアル荷重Ｆ_r がアキシャル荷
重Ｆ_a の１／５以下（５Ｆ_r ≦Ｆ_a ）と言った、より厳
しい条件下での使用も可能になる。Similarly, the internal axial load can be reduced by forming only the balls constituting the second ball bearing on the non-load side from ceramic which is lighter than the bearing steel. By using a ceramic ball, the surface pressure applied to the raceway surface that comes into contact with the rolling surface of the ball increases as described above, but the load applied to the second ball bearing is small. It does not act, and sufficient durability for practical use can be secured. Of course, the technology of reducing the diameter of the balls constituting the ball bearing on the non-load side, reducing the number of balls, and using ceramic balls can be used alone or in combination. Can also. Thus, the results can be further reduced internal axial load, d _m n is an ultra high speed than like said 80 to 300, radial load F _r the following 1/5 of the axial load F _a (5F _r ≦ F Use under more severe conditions such as _a ) is also possible.

【００６５】尚、図示の実施例は何れも、１対のアンギ
ュラ型玉軸受をＤＦで組み合わせた例に就いて示した
が、運転時にアキシャル荷重Ｆ_a を支承する第一の玉軸
受を複数個、互いに並列組み合わせで設け、これら複数
個の第一の玉軸受と、アキシャル荷重Ｆ_a を支承しない
１個の第二の玉軸受とをＤＦで組み合わせる事もでき
る。In each of the illustrated embodiments, an example is shown in which a pair of angular ball bearings are combined with a DF. However, _a plurality of first ball bearings that support an axial load Fa during operation are provided. , provided in parallel combination with each other, it is also possible to combine the plurality of first ball bearings thereof, and one second ball bearing that does not support the axial load F _a at DF.

【００６６】[0066]

【発明の効果】本発明は、以上に述べた通り構成され作
用する為、外部アキシャル荷重に対する負荷容量を十分
に確保しつつ、内部アキシャル荷重の低減を図って、玉
軸受装置の耐久性向上を図れる。しかも、回転軸の傾斜
に対する許容度も大きく、且つ滑りに伴う異常摩耗も確
実に防止できる為、より優れた耐久性を確実に得る事が
できる。Since the present invention is constructed and operated as described above, the internal axial load is reduced while the load capacity against the external axial load is sufficiently secured, and the durability of the ball bearing device is improved. I can do it. In addition, the tolerance for the inclination of the rotating shaft is large, and abnormal wear caused by slippage can be surely prevented, so that more excellent durability can be surely obtained.

[Brief description of the drawings]

【図１】従来装置の第１例を示す断面図。FIG. 1 is a sectional view showing a first example of a conventional device.

【図２】同じく要部断面図。FIG. 2 is a sectional view of an essential part of the same.

【図３】同第２例を示す断面図。FIG. 3 is a cross-sectional view showing the second example.

【図４】同じく要部断面図。FIG. 4 is a sectional view of an essential part of the same.

【図５】反負荷側の玉軸受に滑りが発生する状態を示
す、図４と同様の断面図。FIG. 5 is a sectional view similar to FIG. 4, showing a state in which slippage occurs in the ball bearing on the non-load side.

【図６】回転速度並びにアキシャル荷重とラジアル荷重
との比が玉軸受装置の寿命に及ぼす影響を示す線図。FIG. 6 is a diagram showing the influence of the rotational speed and the ratio between the axial load and the radial load on the life of the ball bearing device.

【図７】本発明の玉軸受装置を構成する第二の玉軸受の
断面を、軌道の曲率半径を誇張して示す断面図。FIG. 7 is a cross-sectional view showing a cross section of a second ball bearing constituting the ball bearing device of the present invention, exaggerating a radius of curvature of a track.

【図８】軌道の曲率半径とアキシャル隙間とが玉軸受装
置の計算寿命に及ぼす影響を示す線図。FIG. 8 is a diagram showing the influence of the radius of curvature of the track and the axial gap on the calculated life of the ball bearing device.

【図９】第一の玉軸受の計算寿命と第二の玉軸受の計算
寿命との比と、これらが組み合わされて成る玉軸受装置
の計算寿命と第一の玉軸受の計算寿命との比との関係を
示す線図。FIG. 9 shows the ratio between the calculated life of the first ball bearing and the calculated life of the second ball bearing, and the ratio between the calculated life of the ball bearing device formed by combining them and the calculated life of the first ball bearing. FIG.

【図１０】本発明の第二実施例を示す要部断面図。FIG. 10 is a sectional view of a main part showing a second embodiment of the present invention.

【図１１】同第三実施例を示す要部断面図。FIG. 11 is an essential part cross sectional view showing the third embodiment.

【図１２】同第四実施例を示す要部断面図。FIG. 12 is a sectional view showing a main part of the fourth embodiment.

【図１３】同第五実施例を示す要部断面図。FIG. 13 is a sectional view showing a main part of the fifth embodiment.

[Explanation of symbols]

１ロータ２回転軸３ハウジング４ころ軸受５、５ａ第一の玉軸受６、６ａ第二の玉軸受７、８間座９抑え金１０内輪１１玉１２、１２ａ玉１３外輪１４、１４ａ外輪軌道１５内輪軌道１６、１６ａ保持器１７主部１８ポケット１９シール装置２０内輪２１外輪２２ノズルリング２３ノズル孔２４給油通路 DESCRIPTION OF SYMBOLS 1 Rotor 2 Rotation shaft 3 Housing 4 Roller bearing 5, 5a First ball bearing 6, 6a Second ball bearing 7, 8 Spacer 9 Retainer 10 Inner ring 11 Ball 12, 12a Ball 13 Outer ring 14, 14a Outer ring track 15 Inner ring raceway 16, 16a Cage 17 Main part 18 Pocket 19 Sealing device 20 Inner ring 21 Outer ring 22 Nozzle ring 23 Nozzle hole 24 Oil supply passage

フロントページの続き (58)調査した分野(Int.Cl.⁷，ＤＢ名) F16C 19/00 - 19/56 F16C 33/00 - 33/66 Continuation of front page (58) Field surveyed (Int. Cl. ⁷ , DB name) F16C 19/00-19/56 F16C 33/00-33/66

Claims

(57) [Claims]

A first and a second ball bearings each provided between an outer peripheral surface of a shaft and an inner peripheral surface of a housing, each of which is of an angular type and whose contact angle directions are different from each other. In a ball bearing device in which an axial load applied in a substantially constant direction from the outside between the shaft and the housing is supported by a first ball bearing having a larger contact angle than a contact angle of a second ball bearing, The first and second ball bearings are arranged in a front combination, each having a positive gap, and the radius of curvature of the cross-sectional shape of the outer raceway of the second ball bearing is r.
and _e2, the outer diameter of the ball configuring the second ball bearing in the case of the d, meet 0.53d ≦ r _e2 ≦ 0.56d, the curvature of the sectional shape of the inner raceway of the second ball bearing Radius r
_i2 and then, the outer diameter of the ball configuring the second ball bearing in the case of the d, ball bearing apparatus characterized in that satisfy 0.505d ≦ r _i2 ≦ 0.52d.

2. When the radius of curvature of the cross-sectional shape of the outer raceway of the first ball bearing is r _e1, and the outer diameter of the ball constituting the first ball bearing is d ′, 0.505d ′ ≦ r _e1 ≦ 0.53d ′, and the radius of curvature of the cross-sectional shape of the inner raceway of the first ball bearing is r
and _i1, when the outer diameter of the ball configuring the first ball bearing and d', meet 0.505d' ≦ r _i1 ≦ 0.52d', further, r _e1 + r _i1 ≧ 1.03d' The ball bearing device according to claim 1, which satisfies the following.