JP2004019898A - Rotating support device for pulley - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To increase a durability while meeting the specified requirements for assuring an axial rigidity with a small and inexpensive structure. <P>SOLUTION: A single-row, deep-groove radial ball bearing 14a is used as a rolling bearing for supporting a driven pulley. Where the curvature radius of the cross section of an outer ring raceway 18a is Re, the curvature radius of the cross section of an inner ring raceway 19a is R<SB>i</SB>, the diameter of balls 17 and 17 is D<SB>a</SB>, and the pitch circle diameter of the balls 17 and 17 is D<SB>p</SB>, the dimensions of the parts are controlled so as to meet both the requirements of 1 < (R<SB>e</SB>+ R<SB>i</SB>)/D<SB>a</SB>≤ 0.9746 (D<SB>a</SB>/D<SB>p</SB>)<SP>-0.0304</SP>and 0.05 ≤ D<SB>a</SB>/D<SB>p</SB>≤ 0.16. <P>COPYRIGHT: (C)2004,JPO

Description

【００２５】
又、この表１から明らかな様に、各ケース１〜４のラジアル玉軸受１４ａで、各玉１７、１７のピッチ円直径Ｄ_Ｐ　に対する各玉の直径Ｄ_ａ　の比Ｄ_ａ　／Ｄ_Ｐ　は、互いに異なっている。そして、これら各ケース１〜４のラジアル玉軸受１４ａで、内輪、外輪各軌道１９ａ、１８ａの断面形状の曲率半径Ｒ_ｉ　、Ｒ_ｅ　を変化させた場合での、内輪１６ａに対する外輪１５ａの最大アキシアル変位量を計算した。又、この計算では、９８０Ｎ（＝１００ｋｇｆ　）のラジアル荷重をラジアル玉軸受１４ａの中心（各玉１７、１７の中心）から軸方向に１０ｍｍずれた（オフセット）位置に負荷した。又、上記計算では、簡略化の為に、上記内輪、外輪両軌道１９ａ、１８ａの断面形状の曲率半径Ｒ_ｉ　、Ｒ_ｅ　を同じにした（Ｒ_ｉ　＝Ｒ_ｅ　）。
【００２６】
この様な条件で行なった計算の結果を、図３に示している。この図３の横軸は、各玉１７、１７の直径Ｄ_ａ　に対する内輪、外輪各軌道１９ａ、１８ａの曲率半径Ｒ_ｉ　、Ｒ_ｅ　の割合を％［｛Ｒ_ｉ　（＝Ｒ_ｅ　）／Ｄ_ａ　｝×１００］で表し、縦軸は、内輪１６ａに対する外輪１５ａの最大アキシアル変位量をｍｍで表している。又、この図３には、一般的な複列アンギュラ玉軸受及び４点接触型の単列ラジアル玉軸受の場合を、参考として示している。そして、上記図３に示した計算結果から、上記各ケース毎に、上記最大アキシアル変位量が０．１５ｍｍと十分に小さくなる、内輪、外輪各軌道１９ａ、１８ａの曲率半径Ｒ_ｉ　、Ｒ_ｅ　を求め、求めたＲ_ｉ　、Ｒ_ｅ　と、上記各ケースでの各玉１７、１７の直径Ｄ_ａ　と、これら各玉１７、１７のピッチ円直径Ｄ_Ｐ　とから、最小自乗法によって、前記（１）式が得られた。
言い換えれば、前記（１）式が満たされる場合には、最大アキシアル変位量を、０．１５ｍｍと十分に小さくでき、ラジアル玉軸受１４ａの軸方向の剛性を十分に確保できる事を確認できた。
【００２７】
一方、上記図３で示した計算結果から明らかな様に、前記（１）式で、左辺を右辺よりも小さくする程、上記最大アキシアル変位量をより小さくでき、上記軸方向の剛性の向上を図れる。但し、この様に、前記（１）式の左辺である、比（Ｒ_ｅ　＋Ｒ_ｉ　）／Ｄ_ａ　を小さくした場合には、各玉１７、１７の転動面と内輪、外輪各軌道１９ａ、１８ａとの接触部に形成される接触楕円が大きくなり、この接触楕円部分での相対滑りに基づく発熱量が多くなる。逆に、前記（１）式が満たされる場合には、この発熱量を十分に少なく抑えられる可能性がある。この結果、前記（１）式を満たす様に各部の寸法を規制した場合には、上記軸方向の剛性を確保しつつ、ラジアル玉軸受１４ａでの発熱量を十分に少なく抑えられる可能性がある事を確認できた。従って、前述した様に、各玉１７、１７の外径Ｄ_ａ　を、ピッチ円直径Ｄ_Ｐ　の５〜１６％、好ましくは５〜１４％、より好ましくは５〜１０％の範囲に納める事により、上記ラジアル玉軸受１４ａの内部の温度上昇を抑えられる事を確認する為、本発明者が行なった計算の前提として、前記（１）式が満たされるとした。又、この計算では、簡略化の為に、内輪軌道１９ａの曲率半径Ｒ_ｉ　と外輪軌道１８ａの曲率半径Ｒ_ｅ　とを同じにした（Ｒ_ｉ　＝Ｒ_ｅ　）。
【００２８】
上述の様な仕様を有するラジアル玉軸受１４ａを、１０００Ｎのラジアル荷重をこのラジアル玉軸受１４ａの中心（各玉１７、１７の中心）から軸方向に４．３５ｍｍずれた（オフセット）位置に負荷した状態で、外輪１５ａを１００００ｍｉｎ^−１　の速度で回転させるとした。そして、この条件下で当該各転がり接触部分の発熱量を算出した。算出方法（解析手法）は、「谷口、荒牧、正田：４点接触玉軸受の性能解析」（社団法人、日本トライボロジー学会、トライボロジー会議９６春　東京　講演予稿集）に記載された方法を採用した。尚、この文献に記載された解析方法によって計算される軸受の摩擦トルクは、実験によるトルク測定結果に一致する事が報告されており、十分に信頼性のある計算結果を得られるものである。
【００２９】
本発明者は、上記解析方法を利用して、上記各玉１７、１７の直径Ｄ_ａ　を変えた場合に於ける、上記ラジアル玉軸受１４ａ内部での局所的な発熱、即ち、上記各玉１７、１７毎に２点ずつ存在する転がり接触部分の発熱量を計算した。又、この計算は、これら各玉１７、１７の直径Ｄ_ａ　を、同じくピッチ円直径Ｄ_Ｐ　の３．７〜２５．５％の間で変化させて、それぞれに就いて行なった。この場合に於いて、ラジアル玉軸受１４ａとして実用的な構造を考慮し、上記各玉１７、１７の直径Ｄ_ａ　が変化した場合でも、この直径Ｄ_ａ　とこれら各玉１７、１７の数との積がほぼ一定となる様に（直径Ｄ_ａ　が小さくなる程数を多くして）玉１７、１７の数を決定し、計算に用いた。
【００３０】
この様な条件で行なった計算の結果を、図４に示している。この図４の横軸は、上記各玉１７、１７のピッチ円直径Ｄ_Ｐ　に対するこれら各玉１７、１７の直径Ｄ_ａ　の割合を％｛（Ｄ_ａ　／Ｄ_Ｐ　）×１００｝で表し、縦軸は玉１個毎の発熱量をＷで表している。尚、この発熱量は、上記ラジアル玉軸受１４ａに組み込まれた上記各玉１７、１７のうちで、前記１０００Ｎなるラジアル荷重を負荷する位置に存在する玉１７、１７の転がり接触部の摩擦損失に基づいて求めた。これは、一方向のラジアル荷重を受けつつ運転されるラジアル玉軸受１４ａでは、このラジアル荷重の作用側に存在する玉１７、１７の転動面が、外輪軌道１８ａ及び内輪軌道１９ａと、大きな面圧で接触しているからである。
【００３１】
この様にラジアル荷重の作用側に存在する玉１７、１７は、他の位置、即ちラジアル荷重が作用しない部分に存在する玉１７、１７に比べて転がり接触部で発生する摩擦が極端に大きく、この転がり接触部での発熱量も多くなる。この様な転がり接触部での発熱量には、上記各玉１７、１７の転動面と外輪軌道１８ａ及び内輪軌道１９ａとの転がり摩擦による損失に基づくものと、上記各玉１７、１７の転動面とこれら外輪軌道１８ａ及び内輪軌道１９ａとの転がり接触部に形成される接触楕円内部での相対的な滑り摩擦による損失に基づくものとが含まれる。何れにしても、上記ラジアル玉軸受１４ａの耐久性低下に結び付く、高温下でのグリース劣化は、上記玉１７、１７毎、更には接触楕円毎と言った、局所的な発熱、温度上昇が影響していると考えられる。即ち、摩擦により接触楕円部分で著しい温度上昇が発生すると、当該接触楕円近傍に存在するグリースが熱劣化する結果、グリース全体が次第に熱劣化して、このグリースを封入したラジアル玉軸受１４ａの耐久性低下の原因になると考えられている。従って、グリースの熱劣化に起因するラジアル玉軸受１４ａの耐久性低下を防止する為には、ラジアル荷重の作用側に存在する玉１７、１７に関する接触楕円部分と言った、局所的な発熱をできるだけ小さく抑える事が必要であると考えられる。
【００３２】
この様な観点で図４を見れば明らかな通り、上記各玉１７、１７の外径Ｄ_ａ　を、ピッチ円直径Ｄ_Ｐ　の５〜１６％、好ましくは５〜１４％、より好ましくは５〜１０％の範囲に納めれば、上記発熱を低く抑えて、ラジアル玉軸受１４ａの耐久性向上を図れる。即ち、上記ラジアル荷重の作用側に存在する玉１７、１７に関する接触楕円部分に関しては、玉１７、１７の直径Ｄ_ａ　がピッチ円直径Ｄ_Ｐ　の１０％以上（Ｄ_ａ　≧０．１Ｄ_Ｐ　）であれば、この直径Ｄ_ａ　が小さい程、摩擦損失も小さい。特に、この直径Ｄ_ａ　がピッチ円直径Ｄ_Ｐ　の５〜１０％の範囲では、摩擦損失はほぼ一定の小さい値をとる。これに対して、上記直径Ｄ_ａ　がピッチ円直径Ｄ_Ｐ　の５％未満（Ｄ_ａ　＜０．０５Ｄ_Ｐ　）になると、逆にこの直径Ｄ_ａ　が小さくなる程、摩擦損失が急激に大きくなる。
【００３３】
一方、図４に示した解析結果を得たラジアル玉軸受１４ａのサイズ（ピッチ円直径＝４３．５ｍｍ）に近いサイズを有する、標準的な単列深溝型玉軸受である、呼び番号が６００５（内径＝２５ｍｍ、外径＝４７ｍｍ、幅＝１２ｍｍ）、６００６（内径＝３０ｍｍ、外径＝５５ｍｍ、幅＝１３ｍｍ）、６００７（内径＝３５ｍｍ、外径＝６２ｍｍ、幅＝１４ｍｍ）では、玉１７、１７の直径Ｄ_ａ　はピッチ円直径Ｄ_Ｐ　の１６〜１７％程度である。上述した様に、玉１７、１７の直径Ｄ_ａ　がピッチ円直径Ｄ_Ｐ　の１０％以上であれば、この直径Ｄ_ａ　が小さい程、摩擦損失も小さくなる事から、玉１７の直径Ｄ_ａ　を上記標準的な単列深溝型玉軸受の値（ピッチ直径Ｄ_Ｐ　の１６％）よりも小さくする事によって、この標準的な単列深溝型玉軸受と同じ直径の玉１７、１７を採用するよりも局所的な発熱が小さくなり、軸受の耐久時間延長を図れる事が分かる。
【００３４】
又、計算結果を示す図４では、玉１７、１７の直径Ｄ_ａ　がピッチ円直径Ｄ_Ｐ　の１４％である部分を境に、これ以下では玉１７、１７の発熱量が５％以上低下している。この事から、接触楕円部分での局所発熱を更に低減する為には、玉１７、１７の直径Ｄ_ａ　をピッチ円直径Ｄ_Ｐ　の１４％以下とする事が好ましい事が分かる。尚、玉１７、１７の直径Ｄ_ａ　を小さくすると、ラジアル玉軸受１４ａの負荷容量が低下するので、転がり疲れ寿命を確保する面からは不利になるが、高負荷容量を要求されない場合には、接触楕円部分での局所発熱を抑え、グリース寿命を確保する面からは、更に玉１７、１７の直径Ｄ_ａ　をピッチ円直径Ｄ_Ｐ　の１０％以下とする事が好ましい事も分かる。但し、上記図４から明らかな通り、玉１７、１７の直径Ｄ_ａ　をピッチ円直径Ｄ_Ｐ　の５％よりも小さくすると、接触楕円部分での局所発熱が急激に増大する。この面から、グリース寿命を確保する事を考慮した場合でも、玉１７、１７の直径Ｄ_ａ　はピッチ円直径Ｄ_Ｐ　の５％以上、好ましくは５．５％以上確保する必要がある事が分かる。
【００３５】
尚、上述した計算結果は、（Ｒ_ｅ　＋Ｒ_ｉ　）／Ｄ_ａ　＝０．９７４６（Ｄ_ａ　／Ｄ_Ｐ　）^{−０．０３０４}　を前提として計算した場合に就いて説明した。但し、本発明は、この様な条件を満たす場合に限定するものではない。例えば、（Ｒ_ｅ　＋Ｒ_ｉ　）／Ｄ_ａ　＜０．９７４６（Ｄ_ａ　／Ｄ_Ｐ　）^{−０．０３０４}　を満たす場合には、各玉１７、１７のピッチ円直径Ｄ_Ｐ　に対する各玉１７、１７の直径Ｄ_ａ　の割合と、各玉１７、１７毎の発熱量との関係を表す曲線が、上述の図４に示した曲線よりも、全体的に発熱量が多くなる方向にずれる。但し、この場合に得られる曲線の場合も、上述の図４に示した曲線の場合と同様に、各玉１７、１７のピッチ円直径Ｄ_Ｐ　に対する各玉１７、１７の直径Ｄ_ａ　の比Ｄ_ａ　／Ｄ_Ｐ　に関して、０．０５≦Ｄ_ａ　／Ｄ_Ｐ　≦０．１６を満たす場合に、各玉１７、１７毎の発熱量が少なくなる傾向になる。この為、本発明によれば、この様な場合でも、ラジアル玉軸受１４ａの内部での発熱量を抑えて、耐久性の向上を図れる。
【００３６】
又、以上の説明は、プーリと回転軸とを係脱する為の電磁クラッチを設けた、プーリ用回転支持装置に本発明を適用した場合に就いて示したが、本発明は、プーリから回転軸に回転力の伝達を自在とした構造であれば、電磁クラッチを設けない、単なるプーリ用回転支持装置にも適用できる。即ち、例えば特開平１１−２１０６１９号公報、或は実開昭６４−２７４８２号公報に記載された様な、斜板式可変容量型コンプレッサの場合には、斜板の傾斜角度を極く小さく（更には傾斜角度をゼロに）する事により、コンプレッサの回転軸の回転トルクを極く小さくできる。
【００３７】
この様な構造の場合には、図７に示す様に、ケーシング２の端部に形成した支持筒部３の周囲に転がり軸受２１を介して回転自在に支持した従動プーリ４ｃと回転軸１とを、トルクチューブとして機能する緩衝材２２を介して、過大なトルクが加わらない限り回転力の伝達自在に結合し、電磁クラッチを設けない場合もある。この様な構造で、上記転がり軸受２１として、前述の図１〜２で示した様に、内部諸元を規制したものを使用すれば、本発明の作用・効果を得られる。尚、上記緩衝材２２を省略する代わりに、回転軸１の端部に固定した円環状部材の側面と、従動プーリ４ｃの端面との間に、過大なトルクが加わらない限り、これら両部材４ｃ同士の回転力の伝達を自在とするトルクリミッタ機構を設ける場合もある。
【００３８】
又、本発明は、ベルト式無段変速機で、ハウジング等の固定の支持部分にプーリを支持する、プーリ用回転支持装置に適用する事もできる。例えば、この場合には、それぞれの外周面に傾斜面部を設けた１対のプーリ素子をそれぞれの外周面に固設した、第一の回転軸と第二の回転軸とを組み合わせて、回転軸を構成する。そして、上記第一、第二の各回転軸の端部を、ハウジング等の固定の支持部分に対し、転がり軸受により回転自在に支持する。この様な構造で、この転がり軸受として、前述の図１〜２で示した様に内部諸元を規制したものを使用する場合にも、本発明の作用・効果を得られる。
【００３９】
【発明の効果】
本発明のプーリ用回転支持装置は、以上に述べた通り、小型且つ安価な構造で、所定の条件を満たしつつ、玉軸受の転がり接触部分での発熱量を抑えて、この玉軸受内部の温度上昇を抑え、グリースの劣化を防止する事による玉軸受の耐久性向上を図れる。
【図面の簡単な説明】
【図１】本発明の実施の形態の１例を示す部分断面図。
【図２】ラジアル玉軸受のみを取り出して示す拡大断面図。
【図３】各玉の直径に対する内輪、外輪各軌道の曲率半径の割合が、内輪に対する外輪の最大アキシアル変位量に及ぼす影響を示す線図。
【図４】各玉のピッチ円直径に対するこれら各玉の直径の割合が発熱量に及ぼす影響を示す線図。
【図５】従来構造の第１例を示す部分断面図。
【図６】同第２例を、ラジアル玉軸受のみを取り出して示す部分断面図。
【図７】本発明の対象となる構造の別例を示す断面図。
【符号の説明】
１　　回転軸
２　　ケーシング
３　　支持筒部
４、４ａ、４ｂ、４ｃ　従動プーリ
５　　複列アンギュラ玉軸受
６　　ソレノイド
７　　取付ブラケット
８　　環状板
９　　板ばね
１０　　電磁クラッチ
１１　　無端ベルト
１２　　外輪
１３　　内輪
１４、１４ａ　ラジアル玉軸受
１５、１５ａ　外輪
１６、１６ａ　内輪
１７　　玉
１８、１８ａ　外輪軌道
１９、１９ａ　内輪軌道
２１　　転がり軸受
２２　　緩衝材[0001]
TECHNICAL FIELD OF THE INVENTION
The rotation supporting device for a pulley according to the present invention, for example, rotatably supports a driven pulley for rotationally driving the compressor on a fixed portion such as a housing constituting a compressor for an air conditioner for an automobile, Used to rotatably support the pulley of the belt-type continuously variable transmission on a fixed part such as a housing.
[0002]
[Prior art]
A compressor that is incorporated in an air conditioner for a vehicle and compresses a refrigerant is rotationally driven by a traveling engine. For this reason, an endless belt is bridged between a driven pulley provided at an end of a rotary shaft of the compressor and a drive pulley fixed to an end of a crankshaft of the traveling engine, and based on the circulation of the endless belt. Thus, the rotation shaft is driven to rotate.
[0003]
FIG. 5 shows the structure of the rotary drive portion of the rotary shaft 1 of the compressor. The rotating shaft 1 is rotatably supported in the casing 2 by a rolling bearing (not shown). A driven pulley 4 is rotatably supported by a double-row angular ball bearing 5 around a support cylindrical portion 3 provided on an outer surface of an end portion of the casing 2 and corresponding to a support portion described in claims. The driven pulley 4 has a U-shaped cross section and is formed in an annular shape as a whole. A solenoid 6 fixed to an end face of the casing 2 is disposed in an internal space of the driven pulley 4. On the other hand, a mounting bracket 7 is fixed to a portion protruding from the casing 2 at the end of the rotating shaft 1, and an annular plate 8 made of a magnetic material is provided around the mounting bracket 7 via a leaf spring 9. I support it. When the solenoid 6 is not energized, the annular plate 8 is separated from the driven pulley 4 as shown in FIG. 5 by the resilience of the leaf spring 9. However, when the solenoid 6 is energized, the driven pulley 4 to allow the transmission of rotational force from the driven pulley 4 to the rotary shaft 1. That is, the solenoid 6, the annular plate 8, and the leaf spring 9 constitute an electromagnetic clutch 10 for engaging and disengaging the driven pulley 4 and the rotary shaft 1.
[0004]
In the case of the structure in which the driven pulley 4 is rotatably supported by the double-row angular ball bearing 5 as described above, a slight eccentric load is applied to the driven pulley 4 from the endless belt 11 wrapped around the driven pulley 4. In this case, the center axis of the outer ring 12 and the center axis of the inner rings 13, 13 constituting the double-row angular contact ball bearing 5 hardly become mismatched (inclined). Accordingly, the durability of the double row angular contact ball bearing 5 is sufficiently ensured, and the rotation center of the driven pulley 4 is prevented from being inclined, so that the uneven wear of the endless belt 11 can be prevented. When the double-row angular contact ball bearing 5 is used, the rigidity in the axial direction can be increased, so that one side (the left side surface in FIG. 5) of the driven pulley 4 and the annular shape when the electromagnetic clutch 10 is disconnected. The dimension in the axial direction of the gap between one side of the plate 8 (the right side in FIG. 5) can be easily regulated to a desired value. Therefore, the connection / disconnection state of the electromagnetic clutch 10 can be reliably switched, and when the electromagnetic clutch 10 is disconnected, the driven pulley 4 and the annular plate 8 are prevented from rubbing each other. Wear and abnormal noise can be suppressed.
[0005]
However, the use of the double-row angular contact ball bearing 5 inevitably increases the axial dimension. In recent years, there has been an increasing demand for miniaturization of automotive parts, and in order to respond to this demand, the rotation supporting portion of the driven pulley 4 must be installed in a limited space in many cases. For this reason, it is not preferable that the axial dimension increases. In addition, the cost of each component increases as the axial dimension increases, which causes an increase in the cost of the compressor.
[0006]
If a single-row deep-groove type radial ball bearing is used as the rolling bearing for supporting the driven pulley 4 in place of the double-row angular ball bearing 5 as described above, the axial space is reduced and the limited space is reduced. Installation inside is easy. However, in the case of a simple single-row deep-groove type radial ball bearing, unlike in the case of the double-row angular ball bearing 5, the rigidity in the axial direction is reduced, so that the clearance between the driven pulley 4 and the annular plate 8 is reduced. It becomes difficult to regulate the dimension in the axial direction to a desired value. For this reason, the connection and disconnection of the electromagnetic clutch 10 becomes defective, some components are easily worn, and abnormal noise is easily generated.
[0007]
In view of such circumstances, in order to support the driven pulley, using a single-row four-point contact type radial ball bearing, or using a single-row three-point contact type radial ball bearing, For example, as described in Japanese Patent Application Laid-Open Nos. 9-119510 and 11-33695, it has been conventionally considered. FIG. 6 shows a radial ball bearing 14 used in the second example of the conventional structure described in Japanese Patent Application Laid-Open No. H11-333695.
[0008]
The radial ball bearing 14 is a single-row, four-point contact type. That is, the radial ball bearing 14 includes an outer ring 15 and an inner ring 16 and a plurality of balls 17 arranged concentrically with each other. An outer raceway 18 is formed on the inner peripheral surface of the outer race 15 and an inner raceway 19 is formed on the outer peripheral surface of the inner race 16 over the entire circumference. The cross-sectional shape of each of the tracks 18 and 19 is a so-called Gothic arch shape in which arcs having a radius of curvature larger than １ ／ of the diameter of each ball 17 intersect at an intermediate portion. Therefore, each of the orbits 18 and 19 and the rolling surface of each of the balls 17 are in rolling contact with each other at two points, for each of the balls 17 at a total of four points.
[0009]
Although such a four-point contact type radial ball bearing 14 has a small and inexpensive structure, it has higher axial rigidity than a general single-row deep groove type radial ball bearing. Therefore, the connection and disconnection of the electromagnetic clutch 10 (see FIG. 5) can be switched satisfactorily, and the driven pulley 4 and the annular plate 8 (see FIG. 5) are prevented from rubbing each other. Wear and abnormal noise can be suppressed.
Also, the so-called three-point contact type single row deep groove type has two contact points between the rolling surface of each ball and the outer raceway and the inner raceway, but one contact point on the other raceway. Also in the case of a radial ball bearing, the rigidity in the axial direction is larger than that of a general single row deep groove type radial ball bearing. Therefore, by using a three-point contact type single row deep groove type radial ball bearing, it is possible to switch the connection and disconnection of the electromagnetic clutch satisfactorily and to suppress the wear of components and the generation of abnormal noise. it can.
[0010]
[Problems to be solved by the invention]
As described above, when the above-described three-point contact type or four-point contact type radial ball bearing is assembled to the rotation supporting portion of the driven pulley for driving the compressor, miniaturization and ensuring durability are ensured. There is a possibility that they can be balanced at a high level. However, in the case of a structure in which such a radial ball bearing is assembled, since the rolling surface of each ball and at least one of the outer raceway and the inner raceway are in rolling contact at two points, the rolling contact portion The sliding motion based on the spin at the point becomes large, and the calorific value at the rolling contact portion increases. When the amount of heat generated at the rolling contact portion increases, the temperature of the rolling contact portion rises remarkably, and damage such as early abrasion and seizure easily occurs.
[0011]
On the other hand, as a rolling bearing for rotatably supporting a driven pulley with respect to a fixed portion, a single-row deep groove type in which the total radius of curvature of the cross-sectional shape of the inner raceway and the outer raceway with respect to the diameter of each ball is reduced. The use of such radial ball bearings has been conventionally considered, as described in JP-A-2000-24063. According to the conventional structure described in this publication, the rigidity in the axial direction can be increased, so that the connection / disconnection state of the electromagnetic clutch can be reliably switched, and the occurrence of wear of component parts and generation of abnormal noise can be suppressed. . Moreover, since a single row deep groove type radial ball bearing is used, miniaturization and cost reduction can be achieved as in the case of using the above-described three-point contact type or four-point contact type radial ball bearing. .
[0012]
In addition, in the case of the conventional structure described in Japanese Patent Application Laid-Open No. 2000-240663, since the rolling surface of each ball and the inner raceway and the outer raceway make rolling contact at one point each, the rolling contact portion The sliding motion based on the spins at the point becomes smaller. Therefore, there is a possibility that the amount of heat generated at the rolling contact portion can be reduced.
[0013]
However, in the case of the conventional structure described in the above-mentioned Japanese Patent Application Laid-Open No. 2000-240663, since the specifications of each part are not always sufficiently examined, a sufficient effect is not necessarily obtained from the viewpoint of suppressing the amount of generated heat. there is a possibility. That is, in the case of this conventional structure, the radius of curvature of the cross-sectional shape of each of the inner ring and outer ring raceways tends to be small, so that the contact ellipse formed in the rolling contact portion tends to be large. For this reason, the relative sliding motion at the contact elliptical portion increases, and the amount of heat generated at the contact elliptical portion tends to increase. In particular, since the rotation supporting portion of the driven pulley for driving the compressor is installed in the engine room where the temperature is high, the temperature rise inside the radial ball bearing is increased in combination with the increase in the amount of heat generated at the contact elliptical portion. Tends to be remarkable. When the temperature rise becomes remarkable, the grease sealed in the radial ball bearing is thermally degraded and lubricity is impaired, and damage such as early wear and seizure is liable to occur.
In addition to the above-mentioned rotary support portion of the driven pulley for driving the compressor, a radial ball bearing in which the total of both inner and outer raceways is reduced is used for the rotary support portion of the pulley of the belt type continuously variable transmission. In such a case, the rotation supporting portion is exposed to the automatic fluid whose temperature rises during operation, so that the same disadvantage as the above-described rotation supporting portion of the driven pulley occurs.
In view of such circumstances, the rotation support device for a pulley of the present invention has been invented in order to improve durability while satisfying predetermined conditions for securing axial rigidity and the like with a small and inexpensive structure. It is.
[0014]
[Means for Solving the Problems]
The rotation supporting device for a pulley of the present invention includes a fixed supporting portion, a rolling bearing supported by the fixed supporting portion, and a rolling bearing similar to the above-described conventionally known rotation supporting device for a pulley. A pulley rotatably supported by the bearing with respect to the support portion, for passing the endless belt. The rolling bearing is provided with an inner ring having a deep groove type inner raceway on the outer peripheral surface, an outer ring having a deep groove type outer raceway on the inner peripheral surface, and rollingly provided between the inner raceway and the outer raceway. And a plurality of balls, and is a single row deep groove type radial ball bearing.
[0015]
In particular, in the rotation supporting device for a pulley of the present invention, the diameter of each ball is D _a And the pitch circle diameter of each of these balls is D _P , 0.05 ≦ D _a / D _P Satisfies ≦ 0.16.
[0016]
Further, in the case of the pulley rotation support device described in claim 2, the radius of curvature of the cross-sectional shape of the outer raceway is R. _e And the radius of curvature of the cross-sectional shape of the inner raceway is R _i 1.0 <(R _e + R _i ) / D _a ≤0.9746 (D _a / D _P ) ^-0.0304 Meet.
[0017]
[Action]
According to the pulley rotation support device of the present invention configured as described above, a small-sized and inexpensive structure can be obtained because a single-row deep-groove type radial ball bearing is used. Moreover, the radius of curvature R of the cross-sectional shape of each of the inner ring and outer ring raceways _i , R _e Temperature in the ball bearings by suppressing heat generation at the rolling contact areas between the rolling surface of each ball and the inner and outer raceways while enabling proper regulation of axial rigidity. Can be suppressed. Therefore, it is possible to prevent the grease sealed in the ball bearing from deteriorating and improve the durability of the ball bearing. In particular, as in the pulley rotation support device according to claim 2, each of the curvature radii R _i , R _e , Rigidity in the axial direction can be sufficiently ensured. For this reason, when the electromagnetic clutch is used as a rotation supporting portion of a pulley provided with the electromagnetic clutch, the connection / disconnection state of the electromagnetic clutch can be reliably switched, and wear of components and generation of abnormal noise can be suppressed.
[0018]
BEST MODE FOR CARRYING OUT THE INVENTION
1 and 2 show an example of an embodiment of the present invention corresponding to all claims. A feature of the present invention is that a single-row deep-groove type radial ball bearing 14a is used as a rolling bearing for rotatingly supporting the driven pulley 4b with respect to a fixed supporting portion such as the supporting cylindrical portion 3 of the casing 2. By properly regulating the specifications of the radial ball bearing 14a, durability is improved while satisfying predetermined conditions. Since the structure and operation of the other parts are the same as those of the conventional structure shown in FIG. 5 described above, the same parts are denoted by the same reference numerals, and redundant description is omitted or simplified. This will be mainly described.
[0019]
The radial ball bearing 14a includes an outer ring 15a and an inner ring 16a supported concentrically with each other, and a plurality of balls 17,17. A single-row, deep-groove type outer raceway 18a having a single circular cross section is provided on the inner peripheral surface of the outer ring 15a, and a single circular arc-shaped single ring-shaped outer raceway is also provided on the outer peripheral surface of the inner race 16a. An inner ring raceway 19a of a row deep groove type is formed respectively. Each of the balls 17, 17 is rotatably provided between the inner ring and outer ring tracks 19a, 18a.
[0020]
In particular, in the case of the present invention, the diameter of each of the balls 17, 17 is set to D _a And the pitch circle diameter of each of these balls 17, 17 is D _P , 0.05 ≦ D _a / D _P ≦ 0.16, preferably 0.05 ≦ D _a / D _P ≦ 0.14, more preferably 0.05 ≦ D _a / D _P The dimensions of each part are regulated so as to satisfy ≦ 0.10. In the case of this example, the radius of curvature of the cross-sectional shape of the outer raceway 18a is R _e And the radius of curvature of the cross-sectional shape of the inner raceway 19a is R _i Where 1 <(R _e + R _i ) / D _a ≤0.9746 (D _a / D _P ) ^-0.0304 The size of each part is regulated to satisfy
[0021]
According to the pulley rotation support device of the present invention configured as described above, a single-row deep groove type radial ball bearing 14b is used to support the driven pulley 4b with respect to the support cylinder 3, so that the size is small. In addition, an inexpensive structure can be obtained. Moreover, in the case of the present invention, the diameter D of each of the balls 17, 17 is set. _a Is the pitch circle diameter D _P It is regulated in relation to. Therefore, as in this example, the radius of curvature R of the cross-sectional shape of each of the inner ring and outer ring raceways 19a and 18a is as follows. _i , R _e Is the diameter D of each of the balls 17, 17 _a And the pitch circle diameter D _P By appropriately regulating the relationship with the above, it is possible to suppress heat generation at the rolling contact portions between the rolling surfaces of the balls 17, 17 and the raceways 18a, 19a while securing rigidity in the axial direction. . For this reason, even when the ambient temperature is high, the temperature inside the radial ball bearing 14a can be prevented from rising, and the grease sealed inside can be prevented from deteriorating. Performance can be improved. In particular, in the case of this example, 1 <(R _e + R _i ) / D _a ≤0.9746 (D _a / D _P ) ^-0.0304 Is satisfied, sufficient rigidity in the axial direction can be secured. For this reason, when the electromagnetic clutch 10 is provided between the driven pulley 4b and the rotary shaft 1 as in the present embodiment, the connection / disconnection state of the electromagnetic clutch 10 can be reliably switched, and the wear of the components is reduced. And generation of abnormal noise can be suppressed. In the case of this example, 1 <(R _e + R _i ) / D _a , R _e / D _a > 0.5, R _i / D _a > 0.5. For this reason, the rolling surfaces of the balls 17, 17 and the raceways 18a, 19a are brought into rolling contact near the bottoms of the raceways 18a, 19a, so that the balls 17, 17 in the radial ball bearing 14a are formed. Rolling can be performed smoothly.
[0022]
Next, the diameter D of each of the balls 17, 17 _a With the pitch circle diameter D _P 0.05 to 0.16 times (5 to 16%), preferably 0.05 to 0.14 times (5 to 14%), more preferably 0.05 to 0.10 times (5 to 10%) The following describes calculations performed by the inventor of the present invention in order to confirm that the heat generation at the rolling contact portion of the radial ball bearing 14a can be suppressed by setting it within the range described above.
This calculation is based on the calorific value of each of the balls 17, 17 when the radial ball bearing 14a having the specifications shown in FIG. 2 is operated while applying a radial load (for one ball 17, 17). The calorific value based on the frictional loss at the two rolling contact portions) was performed by a computer (simulated). As a premise of this calculation, the specification of the radial ball bearing 14a is set based on the pitch circle diameter D of each of the balls 17, 17. _P Was set to 43.5 mm, and the effective radial gap was set to 0.0 mm.
[0023]
It is assumed that the following equation (1) is satisfied as a premise of the above calculation.
(R _e + R _i ) / D _a = 0.9746 (D _a / D _P ) ^-0.0304 −−− (1)
The reason is that, when the relationship of the expression (1) is satisfied, there is a possibility that heat generation at the rolling contact portion can be sufficiently suppressed while sufficiently securing the axial rigidity of the radial ball bearing 14a. Because. Next, calculation performed to confirm that axial rigidity can be secured when the above equation (1) is satisfied will be described. For this calculation, four types of single row deep groove type radial ball bearings 14a having the specifications shown in Table 1 below were used.
[0024]
[Table 1]

[0025]
Further, as is apparent from Table 1, in the radial ball bearings 14a of the respective cases 1 to 4, the pitch circle diameter D of each of the balls 17, 17 is determined. _P Diameter D of each ball _a Ratio D _a / D _P Are different from each other. Then, in the radial ball bearings 14a of the cases 1 to 4, the radius of curvature R of the cross-sectional shape of each of the inner ring and outer ring raceways 19a and 18a is calculated. _i , R _e Was calculated, the maximum axial displacement of the outer ring 15a with respect to the inner ring 16a was calculated. In this calculation, a radial load of 980 N (= 100 kgf) was applied to a position (offset) shifted by 10 mm in the axial direction from the center of the radial ball bearing 14a (the center of each ball 17, 17). In the above calculation, for the sake of simplicity, the radius of curvature R of the cross-sectional shape of the inner and outer raceways 19a, 18a is shown. _i , R _e (R _i = R _e ).
[0026]
FIG. 3 shows the results of calculations performed under such conditions. The horizontal axis in FIG. 3 is the diameter D of each ball 17, 17. _a Radius of curvature R of the inner and outer raceways 19a, 18a with respect to _i , R _e % [｛R _i (= R _e ) / D _a ｝ × 100], and the vertical axis represents the maximum axial displacement of the outer ring 15a with respect to the inner ring 16a in mm. FIG. 3 shows, for reference, the case of a general double-row angular contact ball bearing and a single-point radial ball bearing of a four-point contact type. From the calculation results shown in FIG. 3, the radius of curvature R of each of the inner and outer raceways 19a and 18a is such that the maximum axial displacement becomes sufficiently small as 0.15 mm for each case. _i , R _e R _i , R _e And the diameter D of each ball 17, 17 in each case described above. _a And the pitch circle diameter D of each of these balls 17, 17 _P From the above, the above equation (1) was obtained by the least squares method.
In other words, when the expression (1) is satisfied, it has been confirmed that the maximum axial displacement can be sufficiently reduced to 0.15 mm, and the axial rigidity of the radial ball bearing 14a can be sufficiently secured.
[0027]
On the other hand, as is clear from the calculation results shown in FIG. 3, as the left side is smaller than the right side in the equation (1), the maximum axial displacement can be reduced, and the rigidity in the axial direction can be improved. I can do it. However, as described above, the ratio (R _e + R _i ) / D _a Is smaller, the contact ellipse formed at the contact portion between the rolling surface of each ball 17, 17 and each of the races 19a, 18a of the inner ring and the outer ring becomes large, and the heat generated by the relative slip at the contact ellipse portion is generated. The amount increases. Conversely, when the above expression (1) is satisfied, there is a possibility that this heat generation can be sufficiently reduced. As a result, when the size of each part is regulated so as to satisfy the expression (1), there is a possibility that the amount of heat generated in the radial ball bearing 14a can be sufficiently reduced while securing the rigidity in the axial direction. I was able to confirm that. Therefore, as described above, the outer diameter D of each ball 17, 17 _a Is the pitch circle diameter D _P 5 to 16%, preferably 5 to 14%, more preferably 5 to 10%, to confirm that the temperature rise inside the radial ball bearing 14a can be suppressed. As a premise of the calculations performed, it was assumed that the above-mentioned equation (1) was satisfied. In this calculation, for the sake of simplicity, the radius of curvature R _i And the radius of curvature R of the outer raceway 18a _e And the same (R _i = R _e ).
[0028]
The radial ball bearing 14a having the above-described specification is loaded with a radial load of 1000N at a position (offset) shifted by 4.35 mm in the axial direction from the center of the radial ball bearing 14a (the center of each ball 17, 17). In the state, the outer ring 15a is ^-1 At a speed of. Then, under these conditions, the calorific value of each rolling contact portion was calculated. As a calculation method (analysis method), a method described in “Taniguchi, Aramaki, Masada: Performance analysis of four-point contact ball bearing” (Japan Tribological Society, Tribology Conference 96 Spring Tokyo Lecture Transcript) was adopted. It is reported that the friction torque of the bearing calculated by the analysis method described in this document coincides with the torque measurement result obtained by an experiment, and a sufficiently reliable calculation result can be obtained.
[0029]
The present inventor uses the above-described analysis method to calculate the diameter D of each of the balls 17, 17. _a Was calculated, the amount of heat generated locally in the radial ball bearing 14a, that is, the amount of heat generated at two rolling contact portions for each of the balls 17, 17 was calculated. Also, this calculation is based on the diameter D of each of these balls 17, 17 _a Is also the pitch circle diameter D _P Was carried out for each of the samples, varying from 3.7 to 25.5%. In this case, considering the practical structure of the radial ball bearing 14a, the diameter D of each of the balls 17, 17 is considered. _a This diameter D _a And the product of the number of these balls 17 and 17 is substantially constant (diameter D _a The number of the balls 17, 17 was determined and used for calculation.
[0030]
FIG. 4 shows the results of calculations performed under such conditions. The horizontal axis of FIG. 4 is the pitch circle diameter D of each of the balls 17, 17. _P The diameter D of each of these balls 17, 17 _a % ｛(D _a / D _P ) × 100 °, and the ordinate represents the heat value of each ball in W. The calorific value is determined by the friction loss of the rolling contact portions of the balls 17, 17 located at the position where the radial load of 1000N is applied, among the balls 17, 17 incorporated in the radial ball bearing 14a. Determined based on This is because, in the radial ball bearing 14a which is operated while receiving a radial load in one direction, the rolling surfaces of the balls 17, 17 existing on the side on which the radial load acts are larger surfaces than the outer raceway 18a and the inner raceway 19a. This is because they are in contact with each other by pressure.
[0031]
As described above, the balls 17, 17 existing on the side on which the radial load is applied have extremely large friction generated at the rolling contact portion as compared with the balls 17, 17 existing at other positions, that is, the portions where the radial load is not applied, The amount of heat generated at the rolling contact portion also increases. The amount of heat generated at the rolling contact portion is based on the loss due to the rolling friction between the rolling surface of each of the balls 17, 17 and the outer raceway 18a and the inner raceway 19a. Included are those based on the loss due to relative sliding friction within the contact ellipse formed at the rolling contact between the moving surface and the outer raceway 18a and inner raceway 19a. In any case, the deterioration of grease at high temperatures, which leads to a decrease in the durability of the radial ball bearing 14a, is affected by local heat generation and temperature rise, such as the balls 17, 17 and contact ellipses. it seems to do. That is, when a remarkable temperature rise occurs in the contact ellipse due to friction, the grease existing in the vicinity of the contact ellipse is thermally degraded. As a result, the entire grease is gradually thermally degraded, and the durability of the radial ball bearing 14a in which the grease is sealed is increased. It is thought to cause the decline. Therefore, in order to prevent a decrease in the durability of the radial ball bearing 14a due to the thermal deterioration of the grease, local heat generation, such as a contact elliptical portion relating to the balls 17, 17 existing on the side where the radial load acts, is minimized. It is necessary to keep it small.
[0032]
As is apparent from FIG. 4 from such a viewpoint, the outer diameter D of each of the balls 17, 17 is apparent. _a Is the pitch circle diameter D _P Within the range of 5% to 16%, preferably 5% to 14%, and more preferably 5% to 10%, the heat generation can be suppressed low and the durability of the radial ball bearing 14a can be improved. That is, with respect to the contact elliptical portion relating to the balls 17, 17 existing on the side on which the radial load acts, the diameter D of the balls 17, 17 _a Is the pitch circle diameter D _P 10% or more (D _a ≧ 0.1D _P ), This diameter D _a Is smaller, the friction loss is smaller. In particular, this diameter D _a Is the pitch circle diameter D _P In the range of 5 to 10%, the friction loss takes a substantially constant small value. In contrast, the diameter D _a Is the pitch circle diameter D _P Less than 5% (D _a <0.05D _P ), On the contrary, this diameter D _a Becomes smaller, the friction loss increases sharply.
[0033]
On the other hand, a standard single-row deep-groove type ball bearing having a size close to the size (pitch circle diameter = 43.5 mm) of the radial ball bearing 14a from which the analysis result shown in FIG. In the inner diameter = 25 mm, outer diameter = 47 mm, width = 12 mm), 6006 (inner diameter = 30 mm, outer diameter = 55 mm, width = 13 mm) and 6007 (inner diameter = 35 mm, outer diameter = 62 mm, width = 14 mm), the ball 17 17 diameter D _a Is the pitch circle diameter D _P About 16 to 17% of the total. As described above, the diameter D of the balls 17, 17 _a Is the pitch circle diameter D _P 10% or more of the diameter D _a Is smaller, the friction loss is smaller, so the diameter D of the ball 17 is smaller. _a Of the above standard single row deep groove ball bearing (pitch diameter D _P 16%), the local heat generation is smaller than in the case of using balls 17 having the same diameter as this standard single row deep groove ball bearing, and the durability time of the bearing can be extended. I understand.
[0034]
In FIG. 4 showing the calculation results, the diameters D of the balls 17, 17 are shown. _a Is the pitch circle diameter D _P Below this, the calorific value of the balls 17, 17 is reduced by 5% or more. From this, in order to further reduce the local heat generation at the contact elliptical portion, the diameter D of the balls 17 _a Is the pitch circle diameter D _P It is understood that it is preferable to set it to 14% or less. The diameter D of the balls 17, 17 _a Is smaller, the load capacity of the radial ball bearing 14a decreases, which is disadvantageous from the viewpoint of securing the rolling fatigue life.However, when a high load capacity is not required, local heat generation at the contact elliptical portion is suppressed. From the viewpoint of ensuring the grease life, the diameter D of the balls 17, 17 is further increased. _a Is the pitch circle diameter D _P It is also understood that it is preferable to set it to 10% or less. However, as is apparent from FIG. _a Is the pitch circle diameter D _P If it is smaller than 5%, the local heat generation at the contact elliptical portion sharply increases. From this aspect, even when the grease life is taken into consideration, the diameter D of the balls 17 _a Is the pitch circle diameter D _P It is understood that it is necessary to secure 5% or more, preferably 5.5% or more of the above.
[0035]
Note that the above calculation result is (R _e + R _i ) / D _a = 0.9746 (D _a / D _P ) ^-0.0304 Has been described on the assumption that However, the present invention is not limited to the case where such a condition is satisfied. For example, (R _e + R _i ) / D _a <0.9746 (D _a / D _P ) ^-0.0304 Is satisfied, the pitch circle diameter D of each ball 17, 17 _P Of each ball 17, 17 with respect to _a And the curve representing the relationship between the ratio of the heat and the heat value of each of the balls 17 and 17 are shifted in the direction in which the heat value becomes larger as a whole than the curve shown in FIG. However, in the case of the curve obtained in this case, similarly to the case of the curve shown in FIG. _P Of each ball 17, 17 with respect to _a Ratio D _a / D _P For 0.05 ≦ D _a / D _P When ≦ 0.16 is satisfied, the calorific value of each ball 17 tends to decrease. For this reason, according to the present invention, even in such a case, the amount of heat generated inside the radial ball bearing 14a can be suppressed, and the durability can be improved.
[0036]
The above description has been given of the case where the present invention is applied to a pulley rotation support device provided with an electromagnetic clutch for engaging and disengaging a pulley and a rotating shaft. As long as the shaft is capable of transmitting the rotational force freely, the present invention can be applied to a simple pulley rotation support device without an electromagnetic clutch. That is, in the case of a swash plate type variable displacement compressor as described in, for example, JP-A-11-210619 or Japanese Utility Model Laid-Open Publication No. 64-27482, the inclination angle of the swash plate is extremely small (further, By setting the inclination angle to zero), the rotational torque of the rotary shaft of the compressor can be extremely reduced.
[0037]
In the case of such a structure, as shown in FIG. 7, a driven pulley 4c rotatably supported via a rolling bearing 21 around a support cylindrical portion 3 formed at an end of a casing 2, and a rotating shaft 1, May be connected via a cushioning material 22 functioning as a torque tube so that torque can be transmitted unless an excessive torque is applied, and an electromagnetic clutch may not be provided. With such a structure, the operation and effect of the present invention can be obtained by using the rolling bearing 21 whose internal specifications are regulated as shown in FIGS. Instead of omitting the cushioning member 22, unless an excessive torque is applied between the side surface of the annular member fixed to the end of the rotating shaft 1 and the end surface of the driven pulley 4c, both members 4c In some cases, a torque limiter mechanism that allows the transmission of the rotational force between them may be provided.
[0038]
The present invention can also be applied to a rotation supporting device for pulleys that supports a pulley on a fixed supporting portion such as a housing in a belt-type continuously variable transmission. For example, in this case, a pair of pulley elements each having an inclined surface portion provided on each outer peripheral surface are fixed to each outer peripheral surface. Is composed. The ends of the first and second rotating shafts are rotatably supported by rolling bearings on a fixed supporting portion such as a housing. With such a structure, the operation and effect of the present invention can be obtained even when a rolling bearing whose internal specifications are regulated as shown in FIGS.
[0039]
【The invention's effect】
As described above, the rotation supporting device for a pulley according to the present invention has a small and inexpensive structure, suppresses the heat generation at the rolling contact portion of the ball bearing while satisfying predetermined conditions, and reduces the temperature inside the ball bearing. It is possible to improve the durability of the ball bearing by suppressing the rise and preventing the grease from deteriorating.
[Brief description of the drawings]
FIG. 1 is a partial sectional view showing an example of an embodiment of the present invention.
FIG. 2 is an enlarged sectional view showing only a radial ball bearing.
FIG. 3 is a diagram showing the effect of the ratio of the radius of curvature of each of the inner ring and outer ring raceways to the diameter of each ball on the maximum axial displacement of the outer ring with respect to the inner ring.
FIG. 4 is a diagram showing the influence of the ratio of the diameter of each ball to the pitch circle diameter of each ball on the amount of heat generated.
FIG. 5 is a partial cross-sectional view showing a first example of a conventional structure.
FIG. 6 is a partial cross-sectional view showing the second example by extracting only a radial ball bearing.
FIG. 7 is a cross-sectional view showing another example of a structure to which the present invention is applied.
[Explanation of symbols]
1 Rotary axis
2 casing
3 Support tube
4, 4a, 4b, 4c driven pulley
5 Double row angular contact ball bearing
6 solenoid
7 Mounting bracket
8 annular plate
9 leaf spring
10 Electromagnetic clutch
11 Endless belt
12 Outer ring
13 Inner ring
14, 14a Radial ball bearing
15, 15a Outer ring
16, 16a Inner ring
17 balls
18, 18a Outer ring track
19, 19a Inner ring track
21 Rolling bearing
22 cushioning material

Claims (4)

The rolling bearing, comprising: a fixed supporting portion; a rolling bearing supported by the fixed supporting portion; and a pulley for supporting an endless belt rotatably supported on the supporting portion by the rolling bearing. Are an inner ring having a deep groove type inner raceway on the outer peripheral surface, an outer ring having a deep groove type outer raceway on the inner peripheral surface, and a plurality of balls rotatably provided between the inner raceway and the outer raceway. in bets on a single row deep groove pulley rotary supporting device is a radial ball bearing with a diameter of the balls and D _a, the pitch circle diameter of the balls in the case of the D _P, 0. A rotation support device for a pulley characterized by satisfying 05 ≦ D _a / D _P ≦ 0.16. The radius of curvature of the cross-sectional shape of the outer ring raceway and _{R e,} the radius of curvature of the cross-sectional shape of the inner raceway when the _{R i, 1.0 <(R e} + R i) / D a ≦ 0.9746 (D a / D _P) ^-0.0304 meet, rotary supporting device for a pulley according to claim 1. Satisfy _{0.05 ≦ D a / D P ≦} 0.14, the rotating support device pulley according to claim 2. The rotation supporting device for a pulley according to claim 3, wherein 0.05 ≦ D _a / D _P ≦ 0.10 is satisfied.

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