JP2004211862A - Pulley supporting device - Google Patents

Pulley supporting device Download PDF

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Publication number
JP2004211862A
JP2004211862A JP2003001993A JP2003001993A JP2004211862A JP 2004211862 A JP2004211862 A JP 2004211862A JP 2003001993 A JP2003001993 A JP 2003001993A JP 2003001993 A JP2003001993 A JP 2003001993A JP 2004211862 A JP2004211862 A JP 2004211862A
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JP
Japan
Prior art keywords
ball
rolling bearing
shaft
bearing
flange
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Pending
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JP2003001993A
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Japanese (ja)
Inventor
Motoji Kawamura
基司 河村
Hirobumi Momoji
博文 百々路
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Koyo Seiko Co Ltd
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Koyo Seiko Co Ltd
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Priority to JP2003001993A priority Critical patent/JP2004211862A/en
Publication of JP2004211862A publication Critical patent/JP2004211862A/en
Pending legal-status Critical Current

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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16C—SHAFTS; FLEXIBLE SHAFTS; ELEMENTS OR CRANKSHAFT MECHANISMS; ROTARY BODIES OTHER THAN GEARING ELEMENTS; BEARINGS
    • F16C19/00—Bearings with rolling contact, for exclusively rotary movement
    • F16C19/02—Bearings with rolling contact, for exclusively rotary movement with bearing balls essentially of the same size in one or more circular rows
    • F16C19/14—Bearings with rolling contact, for exclusively rotary movement with bearing balls essentially of the same size in one or more circular rows for both radial and axial load
    • F16C19/16—Bearings with rolling contact, for exclusively rotary movement with bearing balls essentially of the same size in one or more circular rows for both radial and axial load with a single row of balls
    • F16C19/163—Bearings with rolling contact, for exclusively rotary movement with bearing balls essentially of the same size in one or more circular rows for both radial and axial load with a single row of balls with angular contact
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16C—SHAFTS; FLEXIBLE SHAFTS; ELEMENTS OR CRANKSHAFT MECHANISMS; ROTARY BODIES OTHER THAN GEARING ELEMENTS; BEARINGS
    • F16C33/00—Parts of bearings; Special methods for making bearings or parts thereof
    • F16C33/30—Parts of ball or roller bearings
    • F16C33/38—Ball cages
    • F16C33/3887—Details of individual pockets, e.g. shape or ball retaining means
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16C—SHAFTS; FLEXIBLE SHAFTS; ELEMENTS OR CRANKSHAFT MECHANISMS; ROTARY BODIES OTHER THAN GEARING ELEMENTS; BEARINGS
    • F16C33/00—Parts of bearings; Special methods for making bearings or parts thereof
    • F16C33/72—Sealings
    • F16C33/76—Sealings of ball or roller bearings
    • F16C33/78—Sealings of ball or roller bearings with a diaphragm, disc, or ring, with or without resilient members
    • F16C33/784—Sealings of ball or roller bearings with a diaphragm, disc, or ring, with or without resilient members mounted to a groove in the inner surface of the outer race and extending toward the inner race
    • F16C33/7843—Sealings of ball or roller bearings with a diaphragm, disc, or ring, with or without resilient members mounted to a groove in the inner surface of the outer race and extending toward the inner race with a single annular sealing disc
    • F16C33/7846—Sealings of ball or roller bearings with a diaphragm, disc, or ring, with or without resilient members mounted to a groove in the inner surface of the outer race and extending toward the inner race with a single annular sealing disc with a gap between the annular disc and the inner race
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16C—SHAFTS; FLEXIBLE SHAFTS; ELEMENTS OR CRANKSHAFT MECHANISMS; ROTARY BODIES OTHER THAN GEARING ELEMENTS; BEARINGS
    • F16C2240/00—Specified values or numerical ranges of parameters; Relations between them
    • F16C2240/30—Angles, e.g. inclinations
    • F16C2240/34—Contact angles
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16C—SHAFTS; FLEXIBLE SHAFTS; ELEMENTS OR CRANKSHAFT MECHANISMS; ROTARY BODIES OTHER THAN GEARING ELEMENTS; BEARINGS
    • F16C2240/00—Specified values or numerical ranges of parameters; Relations between them
    • F16C2240/40—Linear dimensions, e.g. length, radius, thickness, gap
    • F16C2240/46—Gap sizes or clearances
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16C—SHAFTS; FLEXIBLE SHAFTS; ELEMENTS OR CRANKSHAFT MECHANISMS; ROTARY BODIES OTHER THAN GEARING ELEMENTS; BEARINGS
    • F16C2240/00—Specified values or numerical ranges of parameters; Relations between them
    • F16C2240/40—Linear dimensions, e.g. length, radius, thickness, gap
    • F16C2240/70—Diameters; Radii
    • F16C2240/76—Osculation, i.e. relation between radii of balls and raceway groove
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16C—SHAFTS; FLEXIBLE SHAFTS; ELEMENTS OR CRANKSHAFT MECHANISMS; ROTARY BODIES OTHER THAN GEARING ELEMENTS; BEARINGS
    • F16C2361/00—Apparatus or articles in engineering in general
    • F16C2361/63—Gears with belts and pulleys
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16C—SHAFTS; FLEXIBLE SHAFTS; ELEMENTS OR CRANKSHAFT MECHANISMS; ROTARY BODIES OTHER THAN GEARING ELEMENTS; BEARINGS
    • F16C2361/00—Apparatus or articles in engineering in general
    • F16C2361/65—Gear shifting, change speed gear, gear box
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16C—SHAFTS; FLEXIBLE SHAFTS; ELEMENTS OR CRANKSHAFT MECHANISMS; ROTARY BODIES OTHER THAN GEARING ELEMENTS; BEARINGS
    • F16C33/00—Parts of bearings; Special methods for making bearings or parts thereof
    • F16C33/30—Parts of ball or roller bearings
    • F16C33/38—Ball cages
    • F16C33/3837—Massive or moulded cages having cage pockets surrounding the balls, e.g. machined window cages
    • F16C33/3843—Massive or moulded cages having cage pockets surrounding the balls, e.g. machined window cages formed as one-piece cages, i.e. monoblock cages

Abstract

<P>PROBLEM TO BE SOLVED: To improve the load bearing characteristic and anti-seizing characteristic with respect to a supporting bearing 15 of a cylinder shaft 2 having a movable flange 7. <P>SOLUTION: A belt 8 is hung between opposite faces of a fixed flange 6 integrally formed on one end side of a rotating shaft 1, and the movable flange 7 integrally formed on one end side of the cylinder shaft 2 externally mounted on an outer diameter of the rotating shaft 1 in a state of being integrally rotatable and axially slidable, and a wrapping diameter of the belt 8 can be variably adjusted by axially displacing the movable flange 7 with the cylinder shaft 2 by a feed screw mechanism 10. The rolling bearing 15 for supporting the cylinder shaft 2 is a lubricant-filled angular ball bearing. Whereby the supporting rigidity of the cylinder shaft 2 is improved and the heat generation is inhibited in comparison with a conventional case using a deep groove ball bearing. <P>COPYRIGHT: (C)2004,JPO&NCIPI

Description

【０００１】
【発明の属する技術分野】
本発明は、プーリ支持装置に関する。
【０００２】
【従来の技術】
車両用の無段変速機では、二つ一対のプーリにベルトを巻き掛け、各プーリに対するベルトの巻き掛け径を変えることにより変速を行うようになっている（特許文献１参照）。
【０００３】
このような無段変速機に用いる二つのプーリは、それぞれ、軸方向に不動の固定フランジ（固定プーリとも言う）と軸方向に変位可能な可動フランジ（可動プーリとも言う）とで構成されており、これら固定フランジと可動フランジの対向面間にできるＶ溝にベルトが巻き掛けられている。上記固定フランジは、両端がケースに対して第１、第２の転がり軸受を介して支持される回転軸の一端側に対して、また、上記可動フランジは、前記回転軸に対してスプライン嵌合される筒軸の一端に対して、それぞれ一体に形成されている。なお、回転軸において固定フランジ側の転がり軸受を第１転がり軸受、反固定フランジ側の転がり軸受を第２の転がり軸受とする。上記プーリに対するベルト巻き掛け径を可変調節するために、送りねじ機構を用いて上記可動フランジを筒軸と共に軸方向にスライドさせるようになっている。上記送りねじ機構は、ねじ軸部材とそれの外径側に螺合されるナット部材とからなり、ねじ軸部材およびナット部材は筒軸の外径側に配設されている。ねじ軸部材の一端はケースと回転軸支持用の第２の転がり軸受との間に固定されており、また、ナット部材の一端は筒軸に対して第３の転がり軸受を介して支持されている。上記第１〜第３の転がり軸受は、すべて深溝型玉軸受とされている（特許文献２参照）。
【０００４】
【特許文献１】
特開２００１−２６０６７０号公報
【特許文献２】
特開平１０−２４６２９８号公報
【０００５】
【発明が解決しようとする課題】
上記従来例では、常時において、ベルトの張力により可動フランジを固定フランジから遠ざける向きのアキシアル荷重が可動フランジに対して作用しており、このアキシアル荷重は、第２、第３の転がり軸受に対してかかる。これらの転がり軸受を深溝型玉軸受としていたのではアキシアル荷重の負荷能力が不十分であり、また、発熱しやすくなる。したがって、上記従来例では、高温、高負荷、高速回転など条件が厳しい状況での使用に耐えられない。
【０００６】
【課題を解決するための手段】
本発明のプーリ支持装置は、両端が第１、第２の転がり軸受を介してケースに支持される回転軸と、この回転軸に対して一体回転可能かつ軸方向スライド可能に外装される筒軸とを有し、前記回転軸において前記第１転がり軸受よりも内側の領域に一体形成された径方向外向きのフランジと、前記筒軸において前記回転軸のフランジ側に一体形成された径方向外向きのフランジとでベルトが巻き掛けられるプーリが構成され、前記筒軸側のフランジを軸方向にスライドさせて前記プーリにおけるベルト巻き掛け径を可変調節する送り機構を有し、前記送り機構は、一端側がケースに対して非回転かつ軸方向不動に取り付けられる内側部材と、この内側部材に対して螺合されかつ前記筒軸に対して第３の転がり軸受を介して支持される外側部材とで構成され、前記第３の転がり軸受が、潤滑剤封入タイプのアンギュラ玉軸受とされている。
【０００７】
この場合、筒軸を支持する第３の転がり軸受には、ベルトの張力により常時においてアキシアル荷重が作用するが、この第３の転がり軸受を、アキシアル荷重の負荷容量が大きなアンギュラ玉軸受としているので、筒軸の支持剛性が従来例に比べて向上する。しかも、この第３の転がり軸受を潤滑剤封入タイプとしているので、当該軸受の潤滑条件が良好となり、発熱抑制効果が増すことになる。
【０００８】
なお、上記第３転がり軸受としてのアンギュラ玉軸受は、軸方向一側にカウンタボアを有する外輪と、この外輪の反カウンタボア側に対応する領域に肉ぬすみを有する内輪と、前記外輪の軌道面と前記内輪の軌道面との間に介装される複数の玉とを含み、前記玉の接触角は、２５度以上５０度以下に設定されているとともに、前記内輪の軌道面の曲率半径は、玉の直径の５３％以上５５％以下に、また、前記外輪の軌道面の曲率半径は、玉の直径の５２．５％以上５４．５％以下にそれぞれ設定されたものとすることができる。
【０００９】
この場合、第３の転がり軸受としてのアンギュラ玉軸受の接触角や軌道面の曲率半径を一般的な規格品と異なるように規定しており、それによって耐荷重性のさらなる向上と発熱のさらなる抑制とを図るうえで有利となる。
【００１０】
【発明の実施形態】
図１から図９に本発明の一実施形態を示している。図１に示すプーリ装置において、１は回転軸、２は筒軸である。
【００１１】
上記回転軸１は、その両端がケース３に対してそれぞれ第１、第２の転がり軸受４，５を介して回転自在かつ軸方向不動に支持されている。上記筒軸２は、回転軸１に対して一体回転可能かつ軸方向スライド可能に外装されている。
【００１２】
回転軸１において第１の転がり軸受４よりも内側の領域には、径方向外向きのフランジ６が一体に形成されており、また、筒軸２において回転軸１のフランジ６側には、径方向外向きのフランジ７が一体に形成されている。なお、回転軸１のフランジ６は、回転軸１と共に軸方向に不動なので、固定フランジと言い、筒軸２のフランジ７は、筒軸２と共に軸方向に変位するので、可動フランジと言う。
【００１３】
これら２つのフランジ６，７の対向面はテーパ形に形成されていて、この二つのフランジ６，７がＶ溝形のプーリを構成している。つまり、このフランジ６，７の対向面間にできるＶ溝にベルト８が巻き掛けられている。
【００１４】
なお、上記可動フランジ７を有する筒軸２は、送りねじ機構１０により軸方向にスライドされるようになっている。つまり、可動フランジ７を筒軸２と共に軸方向にスライドさせると、上記プーリに対するベルト８の巻き掛け径を可変調節できるようになっている。
【００１５】
上記送りねじ機構１０は、内側部材としてのねじ軸部材１１と、それの外径側に螺合される外側部材としてのナット部材１２とで構成されている。ねじ軸部材１１は、上記第１の転がり軸受４とケース３との間に嵌入されて非回転かつ軸方向不動に取り付けられており、ナット部材１２は、筒軸２の外径側に第３の転がり軸受１５を介して回転自在に外装されている。
【００１６】
次に、動作を説明する。送りねじ機構１０のナット部材１２に対して任意の方向の回転動力を入力すると、当該ナット部材１２が螺旋回転して軸方向一方へ移動し、このナット部材１２と一体に結合された筒軸２が軸方向一方にスライドさせられる。これにより、固定フランジ６と可動フランジ７との対向面の間隔が大小調節されるので、固定フランジ６と可動フランジ７との対向面間に巻き掛けられるベルト８の巻き掛け径が変更される。なお、図１においては、上半分に固定フランジ６と可動フランジ７との対向面の間隔を狭くした状態を、また、下半分に固定フランジ６と可動フランジ７との対向面の間隔を広くした状態を示している。
【００１７】
ところで、上記無段変速機では、ベルト８の張力により可動フランジ７を固定フランジ６から遠ざける向きのアキシアル荷重が可動フランジ７に対して常時作用している。つまり、前記アキシアル荷重は、可動フランジ７を有する筒軸２を支持する第３の転がり軸受１５と、回転軸１の反固定フランジ６側を支持する第２の転がり軸受５とで受けられる。
【００１８】
このような無段変速機の使用状況を考慮し、本発明では、上記第２、第３の転がり軸受５，１５を、潤滑剤封入タイプのアンギュラ玉軸受としている。但し、回転軸１の固定フランジ６側を支持する第１の転がり軸受４については、従来例と同様、深溝型玉軸受とされている。
【００１９】
具体的に、上記アンギュラ玉軸受からなる第２、第３の転がり軸受５，１５は、共に、内輪２１と、外輪２２と、複数の玉２３と、保持器２４と、二つのシール２５，２５とを備えている。なお、第３の転がり軸受１５のみを、図２に拡大して示しているが、第２の転がり軸受５と第３の転がり軸受１５の基本構成は同じであり、相違点は、第２の転がり軸受５の内輪２１が、内径を小さくして厚肉としていることである。
【００２０】
外輪２１は、軌道面２１ａを有するとともに軸方向一端にカウンタボア２１ｂを有している。内輪２２は、軌道面２２ａを有するとともに軸方向他端にカウンタボア２２ｂを有している。玉２３は、内輪２１の軌道面２１ａと外輪２２の軌道面２２ａとの間に介装されている。保持器２４は、複数の玉２３それぞれを円周ほぼ等間隔に配置するものである。シール２５は、第２、第３の転がり軸受５，１５の軸方向両端に配設されていて、軸受内部空間を密封するものである。
【００２１】
なお、上記第３の転がり軸受１５の内輪２１は、筒軸２に対して圧入により「しまりばめ」されているとともに、止め輪１６により抜け止めされている。また、第３の転がり軸受１５の外輪２２は、ナット部材１２の内周面に設けられている大径部分に対して「すきまばめ」されており、止め輪１７により軸方向に位置決めされている。
【００２２】
ところで、上記保持器２４については、本願出願人が出願している特願２００２−２０６７４７号に示されるものを用いれば、トルク上昇や発熱を軽減するうえで好ましい。この保持器２４は、図４から図９に示すように、複数の玉２３によって回転案内される過程において、玉２３の周速の小さい位置（Ａ１，Ａ２）に接触するように設計されており、それによって、トルク上昇や発熱を軽減できるようになっている。図中、３０は保持器２４のポケット、３１は大輪部、３２は小輪部、３３は架橋部、Ｘは荷重作用線、Ｙは仮想円筒面である。まず、図４に示すように、上記大輪部３１の第１凹部３４について、玉２３の曲率と同一の球状の凹面とし、この第１凹部３４の曲率半径を玉２３の曲率半径よりも大きく設定する。一方、上記小輪部３２の第２凹部３５については、その外径領域に上記第１凹部３４と同一の曲率および曲率半径に設定した球状の凹面３５ａを設け、径方向中間領域に半円筒面３５ｂを設け、内径側領域に上記第１凹部３４と同一の曲率半径でかつ曲率中心Ｐを内径側にβずらした球状の凹面３５ｃを設けている。この内径側の凹面３５ｃの曲率中心Ｐのずらし量βは、上記半円筒面３５ｂの径方向長さと同じに設定している。そして、上記第１凹部３４の外径側開口の軸方向最深部Ａ１から第２凹部３５の外径側開口の軸方向最深部Ａ２までの間隔Ｗ１と、第１凹部３４の内径側開口の軸方向最深部Ｂ１から第２凹部３５の内径側開口の軸方向最深部Ｂ２までの間隔Ｗ２とを、玉２３の直径ｒよりも小さく、かつ前記間隔Ｗ１を間隔Ｗ２よりも小さく設定している。このようなことにより、ポケット３０の中心と玉２３の中心とを一致させた状態において、第２凹部３５の内径縁と玉２３との間のラジアル隙間Δ２を、第１凹部３４の外径縁と玉２３との間のラジアル隙間Δ１よりも大きくしている。なお、上記ラジアル隙間Δ２は、保持器２４の熱膨張時の拡径量よりも大きく設定している。
【００２３】
また、上記シール２５は、環状芯金２５ａの内周部分にリップ２５ｂを加硫接着した構成であり、外周部分が外輪２２の軸方向両端に対して装着されており、リップ２５ｂが内輪２１に対して微小隙間を介して対向されて非接触密封部を作っている。このシール２５で密封した軸受内部空間には、所定量の潤滑剤が封入されている。特に、アンギュラ玉軸受の場合、深溝型玉軸受に比べて、玉の組み込みの関係より玉の使用数を多くできるので、潤滑剤の封入量を必要最小限に抑えることができて、攪拌抵抗ならびにトルクを軽減するうえで有利となる。
【００２４】
前記潤滑剤については、玉２３の周辺にとどまって基油のみを供給するチャネリングタイプと呼ばれるものが好ましい。このような潤滑剤としては、例えば日本グリース株式会社製の商品名ＫＮＧ２５０や、協同油脂株式会社製の商品名マルテンプＳＢＭと呼ばれるものが好適に用いられる。上記ＫＮＧ２５０は、基油をエーテル系合成油、増ちょう剤をジウレアとしたもので、使用温度範囲は−３０℃〜＋１７０℃である。上記マルテンプＳＢＭは、基油をエステル系合成油、増ちょう剤を（脂肪族＋脂環式ジウレア）としたもので、使用温度範囲は−４０℃〜＋１６０℃である。
【００２５】
そして、上記第２、第３の転がり軸受５，１５としてのアンギュラ玉軸受の仕様を一般的な規格品と異なるように設計しているので、以下で詳細に説明する。
【００２６】
具体的に、玉２３の呼び接触角αを、２５度以上５０度以下、好ましくは３５度以上４５度以下に設定している。また、内輪２１の軌道面２１ａの曲率半径Ｒ１を玉２３の直径ｒの５３％以上５５％以下に、好ましくは５５％に、さらに、外輪２２の軌道面２２ａの曲率半径Ｒ２を玉２３の直径ｒの５２．５％以上５４．５％以下に、好ましくは５３％にそれぞれ設定している。
【００２７】
このように、上記呼び接触角αを設定すれば、発熱を抑制することができる。発熱に関しては、内・外輪２１，２２に対して玉２３がスピン滑りする領域におけるＰＶ値が大きく影響する。前記ＰＶ値とは、軸受の発熱に対する使用限界を表すもので、内・外輪２１，２２と玉２３との接触面圧Ｐと、玉２３のスピン速度Ｖとの積で求められる。図３に、アンギュラ玉軸受の接触角αと前記ＰＶ値との関係を示している。条件としては、内輪２１の回転数を１２０００〔ｒｐｍ〕、荷重を高速条件時の値としている。結果は、図示するように、接触角αを２５度を越えるに従い、ＰＶ値が３５０〔ｋｇｆ／ｍｍ２・（ｍ／ｓ）〕から徐々に小さくなっている。参考までに、第２、第３の転がり軸受２０，３０を従来例のように深溝型玉軸受とした場合だと、ＰＶ値は７９０〔ｋｇｆ／ｍｍ２・（ｍ／ｓ）〕となる。したがって、この実施形態では、第２、第３の転がり軸受２０，３０の発熱が抑制されるので、耐焼付き性が向上し、高速回転の使用に好適なものとなる。
【００２８】
ところで、そもそも、高速回転時には玉２３が外輪２２側へ付勢されて、内輪２１側で滑りやすくなるが、上記のように、第２、第３の転がり軸受５，１５において、内輪２１の軌道面２１ａの曲率半径Ｒ１を外輪２２の軌道面２２ａの曲率半径Ｒ２よりも大きく設定していれば、内輪２１の軌道面２１ａに対する玉２３の接触楕円を小さくできて、玉２３のスピン滑りを減少できる。このことは、上述したように発熱を抑制できるようになった一因と言える。
【００２９】
以上説明したように、筒軸２を支持する第３の転がり軸受１５と、回転軸１の反固定フランジ６側を支持する第２の転がり軸受５とを、潤滑剤封入タイプのアンギュラ玉軸受にしているとともに、これら第２、第３の転がり軸受５，１５としてのアンギュラ玉軸受の諸元（呼び接触角αや、内輪２１の軌道面２１ａの曲率半径Ｒ１と外輪２２の軌道面２２ａの曲率半径Ｒ２）を一般的な規格品と異ならせているから、深溝型玉軸受を用いる従来例に比べて、耐荷重性、耐焼付き性を高めることができる。
【００３０】
【発明の効果】
本発明のプーリ支持装置では、可動フランジを有する筒軸の支持軸受について耐荷重性および耐焼付き性を高めることができるので、高温、高負荷、高速回転など条件が厳しい状況での使用に適したものとすることができる。
【図面の簡単な説明】
【図１】本発明の一実施形態に係るプーリ支持装置の使用状態を示す断面図
【図２】図１中の第３の転がり軸受の上半分を拡大して示す図
【図３】図１中のアンギュラ玉軸受の接触角とＰＶ値との関係を示す図表
【図４】図２中の保持器の各部を詳しく説明するための断面図
【図５】図２中の保持器を小輪部側から見た斜視図
【図６】図２中の保持器を大輪部側から見た斜視図
【図７】図２中の保持器を小輪部側から見た一部の側面図
【図８】図２中の保持器の外径側から見た一部の平面展開図
【図９】図２中の保持器の内径側から見た一部の平面展開図
【符号の説明】
１ 回転軸
２ 筒軸
３ ケース
４ 第１の転がり軸受
５ 第２の転がり軸受
６ 固定フランジ
７ 可動フランジ
８ ベルト
１０ 送りねじ機構
１１ ねじ軸部材
１２ ナット部材
１５ 第３の転がり軸受[0001]
TECHNICAL FIELD OF THE INVENTION
The present invention relates to a pulley support device.
[0002]
[Prior art]
In a continuously variable transmission for a vehicle, a belt is wound around a pair of pulleys, and the speed is changed by changing a winding diameter of the belt around each pulley (see Patent Document 1).
[0003]
Each of the two pulleys used in such a continuously variable transmission is constituted by a fixed flange that is immovable in the axial direction (also referred to as a fixed pulley) and a movable flange that is displaceable in the axial direction (also referred to as a movable pulley). The belt is wound around a V-groove formed between the opposed surfaces of the fixed flange and the movable flange. The fixed flange is spline-fitted to one end of a rotary shaft whose both ends are supported by a case via first and second rolling bearings, and the movable flange is spline-fitted to the rotary shaft. Are formed integrally with one end of the cylindrical shaft. In the rotary shaft, the rolling bearing on the fixed flange side is referred to as a first rolling bearing, and the rolling bearing on the side opposite to the fixed flange is referred to as a second rolling bearing. In order to variably adjust the belt winding diameter of the pulley, the movable flange is slid in the axial direction together with the cylinder shaft using a feed screw mechanism. The feed screw mechanism includes a screw shaft member and a nut member screwed to the outer diameter side of the screw shaft member, and the screw shaft member and the nut member are disposed on the outer diameter side of the cylindrical shaft. One end of the screw shaft member is fixed between the case and a second rolling bearing for supporting the rotating shaft, and one end of the nut member is supported on the cylindrical shaft via a third rolling bearing. I have. The first to third rolling bearings are all deep groove ball bearings (see Patent Document 2).
[0004]
[Patent Document 1]
JP 2001-260670 A [Patent Document 2]
JP 10-246298 A
[Problems to be solved by the invention]
In the above conventional example, an axial load in a direction of moving the movable flange away from the fixed flange due to the tension of the belt always acts on the movable flange, and this axial load is applied to the second and third rolling bearings. Take it. If these rolling bearings are formed as deep groove ball bearings, the ability to load an axial load is insufficient, and heat is easily generated. Therefore, the above conventional example cannot withstand use under severe conditions such as high temperature, high load, and high speed rotation.
[0006]
[Means for Solving the Problems]
A pulley supporting device according to the present invention includes a rotary shaft having both ends supported by a case via first and second rolling bearings, and a cylindrical shaft externally rotatable integrally with the rotary shaft and slidable in the axial direction. A radially outward flange integrally formed on a region inside the first rolling bearing on the rotary shaft, and a radially outward flange integrally formed on the flange side of the rotary shaft on the cylindrical shaft. A pulley around which the belt is wound with the orientation flange is configured, and has a feed mechanism that variably adjusts the belt winding diameter of the pulley by sliding the flange on the cylindrical shaft side in the axial direction, and the feed mechanism includes: An inner member one end of which is non-rotatably and axially fixed to the case; and an outer member screwed to the inner member and supported by the cylindrical shaft via a third rolling bearing. In the configuration, the third rolling bearing, there is a lubricant sealed type angular contact ball bearing.
[0007]
In this case, an axial load always acts on the third rolling bearing that supports the cylindrical shaft due to the tension of the belt. However, since the third rolling bearing is an angular ball bearing having a large load capacity for the axial load. The support rigidity of the cylindrical shaft is improved as compared with the conventional example. In addition, since the third rolling bearing is of a lubricant-filled type, the lubrication conditions for the bearing are improved, and the heat generation suppressing effect is increased.
[0008]
In addition, the angular contact ball bearing as the third rolling bearing includes an outer ring having a counterbore on one side in the axial direction, an inner ring having a thin wall in a region corresponding to the counterbore side of the outer ring, and a raceway surface of the outer ring. And a plurality of balls interposed between the inner ring and the raceway surface, wherein the contact angle of the ball is set to 25 degrees or more and 50 degrees or less, and the radius of curvature of the raceway surface of the inner ring is , And the radius of curvature of the raceway surface of the outer race may be set to 52.5% to 54.5% of the diameter of the ball. .
[0009]
In this case, the contact angle and the radius of curvature of the raceway surface of the angular ball bearing as the third rolling bearing are specified so as to be different from those of a general standard product, thereby further improving load resistance and further suppressing heat generation. This is advantageous in achieving the following.
[0010]
DETAILED DESCRIPTION OF THE INVENTION
1 to 9 show an embodiment of the present invention. In the pulley device shown in FIG. 1, 1 is a rotary shaft, and 2 is a cylindrical shaft.
[0011]
The rotating shaft 1 is rotatably and axially immovably supported at both ends thereof via a first and second rolling bearings 4 and 5 with respect to a case 3. The cylindrical shaft 2 is externally rotatable with respect to the rotating shaft 1 and slidable in the axial direction.
[0012]
A radially outwardly directed flange 6 is integrally formed in a region inside the first rolling bearing 4 on the rotating shaft 1, and a radially outward flange 6 is formed on the cylindrical shaft 2 on the flange 6 side of the rotating shaft 1. A flange 7 facing outward in the direction is integrally formed. Note that the flange 6 of the rotating shaft 1 is fixed in the axial direction together with the rotating shaft 1, so it is called a fixed flange, and the flange 7 of the cylindrical shaft 2 is called a movable flange because it is displaced in the axial direction together with the cylindrical shaft 2.
[0013]
Opposing surfaces of these two flanges 6 and 7 are formed in a tapered shape, and these two flanges 6 and 7 constitute a V-groove type pulley. That is, the belt 8 is wound around the V groove formed between the facing surfaces of the flanges 6 and 7.
[0014]
The cylindrical shaft 2 having the movable flange 7 is slid in the axial direction by a feed screw mechanism 10. That is, when the movable flange 7 is slid in the axial direction together with the cylindrical shaft 2, the diameter of the belt 8 wound around the pulley can be variably adjusted.
[0015]
The feed screw mechanism 10 includes a screw shaft member 11 as an inner member and a nut member 12 as an outer member screwed to the outer diameter side thereof. The screw shaft member 11 is fitted between the first rolling bearing 4 and the case 3 so as to be non-rotatably and axially immovable. Is rotatably mounted via a rolling bearing 15.
[0016]
Next, the operation will be described. When a rotational power in an arbitrary direction is input to the nut member 12 of the feed screw mechanism 10, the nut member 12 spirally rotates and moves in one axial direction, and the cylindrical shaft 2 integrated with the nut member 12 is integrated. Is slid in one axial direction. Thereby, the distance between the opposing surfaces of the fixed flange 6 and the movable flange 7 is adjusted, so that the winding diameter of the belt 8 wound between the opposing surfaces of the fixed flange 6 and the movable flange 7 is changed. In FIG. 1, the upper half has a narrower space between the opposing surfaces of the fixed flange 6 and the movable flange 7, and the lower half has a wider space between the opposing surfaces of the fixed flange 6 and the movable flange 7. The state is shown.
[0017]
Incidentally, in the above-described continuously variable transmission, an axial load in a direction of moving the movable flange 7 away from the fixed flange 6 by the tension of the belt 8 always acts on the movable flange 7. That is, the axial load is received by the third rolling bearing 15 that supports the cylindrical shaft 2 having the movable flange 7 and the second rolling bearing 5 that supports the rotating shaft 1 on the side opposite to the fixed flange 6.
[0018]
In consideration of the usage of such a continuously variable transmission, in the present invention, the second and third rolling bearings 5 and 15 are angular ball bearings of a lubricated type. However, the first rolling bearing 4 supporting the fixed flange 6 side of the rotating shaft 1 is a deep groove ball bearing as in the conventional example.
[0019]
More specifically, the second and third rolling bearings 5 and 15 made of the above-mentioned angular ball bearings each include an inner ring 21, an outer ring 22, a plurality of balls 23, a retainer 24, and two seals 25 and 25. And Although only the third rolling bearing 15 is shown in an enlarged manner in FIG. 2, the basic configuration of the second rolling bearing 5 and the third rolling bearing 15 is the same, and the difference is that The inner ring 21 of the rolling bearing 5 has a small inner diameter and a large wall thickness.
[0020]
The outer race 21 has a raceway surface 21a and a counter bore 21b at one end in the axial direction. The inner race 22 has a raceway surface 22a and a counter bore 22b at the other end in the axial direction. The ball 23 is interposed between the raceway surface 21a of the inner race 21 and the raceway surface 22a of the outer race 22. The retainer 24 arranges the plurality of balls 23 at substantially equal intervals around the circumference. The seals 25 are provided at both axial ends of the second and third rolling bearings 5 and 15 and seal the bearing internal space.
[0021]
The inner ring 21 of the third rolling bearing 15 is “tightly fitted” into the cylindrical shaft 2 by press fitting, and is prevented from falling off by the retaining ring 16. Further, the outer ring 22 of the third rolling bearing 15 is "clear-fit" with respect to the large diameter portion provided on the inner peripheral surface of the nut member 12, and is positioned in the axial direction by the retaining ring 17. I have.
[0022]
By the way, as for the cage 24, it is preferable to use a cage shown in Japanese Patent Application No. 2002-206747 filed by the present applicant in order to reduce torque increase and heat generation. As shown in FIGS. 4 to 9, the retainer 24 is designed to come into contact with the position (A1, A2) where the peripheral speed of the ball 23 is low in the process of being rotationally guided by the plurality of balls 23. Thereby, torque rise and heat generation can be reduced. In the drawing, 30 is a pocket of the cage 24, 31 is a large wheel portion, 32 is a small wheel portion, 33 is a bridge portion, X is a load acting line, and Y is a virtual cylindrical surface. First, as shown in FIG. 4, the first concave portion 34 of the large wheel portion 31 has the same spherical concave surface as the curvature of the ball 23, and the radius of curvature of the first concave portion 34 is set to be larger than the radius of curvature of the ball 23. I do. On the other hand, as for the second concave portion 35 of the small wheel portion 32, a spherical concave surface 35a having the same curvature and radius of curvature as the first concave portion 34 is provided in the outer diameter region, and the semi-cylindrical surface is formed in the radially intermediate region. 35b, and a spherical concave surface 35c having the same radius of curvature as that of the first concave portion 34 and having the center of curvature P shifted by β toward the inner diameter side is provided in the inner diameter side region. The shift amount β of the center of curvature P of the concave surface 35c on the inner diameter side is set to be the same as the radial length of the semi-cylindrical surface 35b. The distance W1 from the axially deepest portion A1 of the outer diameter side opening of the first concave portion 34 to the axially deepest portion A2 of the outer diameter side opening of the second concave portion 35, and the axis of the inner diameter side opening of the first concave portion 34 The interval W2 from the deepest portion B1 in the direction to the axially deepest portion B2 of the inner diameter side opening of the second concave portion 35 is set smaller than the diameter r of the ball 23, and the interval W1 is set smaller than the interval W2. Thus, in a state where the center of the pocket 30 and the center of the ball 23 coincide with each other, the radial gap Δ2 between the inner diameter edge of the second recess 35 and the ball 23 is changed to the outer diameter edge of the first recess 34. Is larger than the radial gap Δ1 between the ball 23 and the ball 23. Note that the radial gap Δ2 is set to be larger than the diameter expansion amount of the cage 24 during thermal expansion.
[0023]
The seal 25 has a configuration in which a lip 25b is vulcanized and bonded to an inner peripheral portion of an annular core 25a, and an outer peripheral portion is attached to both axial ends of the outer ring 22. On the other hand, they face each other via a minute gap to form a non-contact sealing portion. A predetermined amount of lubricant is sealed in the bearing internal space sealed by the seal 25. In particular, in the case of angular contact ball bearings, the number of balls used can be increased due to the incorporation of balls compared to deep groove ball bearings, so the amount of lubricant to be filled can be kept to the minimum necessary, and stirring resistance and This is advantageous in reducing the torque.
[0024]
The lubricant is preferably a so-called channeling type which stays around the ball 23 and supplies only the base oil. As such a lubricant, for example, a product called KNG250 (trade name, manufactured by Nippon Grease Co., Ltd.) or Multemp SBM (trade name, manufactured by Kyodo Yushi Co., Ltd.) is suitably used. The KNG250 uses an ether-based synthetic oil as a base oil and a diurea as a thickener, and has a working temperature range of -30 ° C to + 170 ° C. The Multemp SBM uses an ester-based synthetic oil as the base oil and (aliphatic and alicyclic diurea) as the thickener, and has a working temperature range of -40C to + 160C.
[0025]
Since the specifications of the angular ball bearings as the second and third rolling bearings 5 and 15 are designed so as to be different from general standard products, they will be described in detail below.
[0026]
Specifically, the nominal contact angle α of the ball 23 is set to 25 degrees or more and 50 degrees or less, preferably 35 degrees or more and 45 degrees or less. The radius of curvature R1 of the raceway surface 21a of the inner ring 21 is set to 53% or more and 55% or less, preferably 55% of the diameter r of the ball 23, and the radius of curvature R2 of the raceway surface 22a of the outer ring 22 is set to the diameter of the ball 23. r is set to 52.5% or more and 54.5% or less, preferably 53%.
[0027]
As described above, if the nominal contact angle α is set, heat generation can be suppressed. Regarding heat generation, the PV value in a region where the ball 23 spins and slides on the inner and outer rings 21 and 22 has a great influence. The PV value represents a use limit with respect to heat generation of the bearing, and is determined by a product of a contact surface pressure P between the inner and outer rings 21 and 22 and the ball 23 and a spin speed V of the ball 23. FIG. 3 shows the relationship between the contact angle α of the angular contact ball bearing and the PV value. As conditions, the number of rotations of the inner ring 21 is set to 12000 [rpm], and the load is set to a value in a high-speed condition. As shown in the drawing, the PV value gradually decreases from 350 [kgf / mm 2 · (m / s)] as the contact angle α exceeds 25 degrees. For reference, when the second and third rolling bearings 20, 30 are formed as deep groove ball bearings as in the conventional example, the PV value is 790 [kgf / mm 2 · (m / s)]. Therefore, in this embodiment, since the heat generation of the second and third rolling bearings 20 and 30 is suppressed, the seizure resistance is improved, which is suitable for use at high speed rotation.
[0028]
By the way, in the first place, the ball 23 is urged toward the outer ring 22 at the time of high-speed rotation, and becomes slippery on the inner ring 21 side. However, as described above, in the second and third rolling bearings 5 and 15, If the radius of curvature R1 of the surface 21a is set to be larger than the radius of curvature R2 of the raceway surface 22a of the outer ring 22, the contact ellipse of the ball 23 with the raceway surface 21a of the inner race 21 can be reduced, and the spin sliding of the ball 23 is reduced. it can. This can be said to be one of the reasons why heat generation can be suppressed as described above.
[0029]
As described above, the third rolling bearing 15 that supports the cylindrical shaft 2 and the second rolling bearing 5 that supports the non-fixed flange 6 side of the rotating shaft 1 are formed as lubrication-enclosed angular ball bearings. And the specifications of the angular ball bearings as the second and third rolling bearings 5 and 15 (the nominal contact angle α, the curvature radius R1 of the raceway surface 21a of the inner race 21 and the curvature of the raceway surface 22a of the outer race 22). Since the radius R2) is different from a general standard product, the load resistance and the seizure resistance can be improved as compared with a conventional example using a deep groove type ball bearing.
[0030]
【The invention's effect】
In the pulley support device of the present invention, the load bearing and seizure resistance of the cylindrical shaft support bearing having the movable flange can be improved, so that the pulley support device is suitable for use under severe conditions such as high temperature, high load, and high speed rotation. Things.
[Brief description of the drawings]
FIG. 1 is a sectional view showing a use state of a pulley support device according to an embodiment of the present invention; FIG. 2 is an enlarged view showing an upper half of a third rolling bearing in FIG. 1; FIG. FIG. 4 is a table showing the relationship between the contact angle of the angular contact ball bearing and the PV value in FIG. 4 FIG. 4 is a cross-sectional view for explaining in detail each part of the retainer in FIG. 2 FIG. FIG. 6 is a perspective view of the retainer in FIG. 2 viewed from the large wheel side. FIG. 7 is a partial side view of the retainer in FIG. 2 viewed from the small wheel side. 8 is a partially developed plan view of the retainer shown in FIG. 2 as viewed from the outer diameter side. FIG. 9 is a partially developed plan view of the retainer shown in FIG. 2 as viewed from the inner diameter side.
REFERENCE SIGNS LIST 1 rotating shaft 2 cylindrical shaft 3 case 4 first rolling bearing 5 second rolling bearing 6 fixed flange 7 movable flange 8 belt 10 feed screw mechanism 11 screw shaft member 12 nut member 15 third rolling bearing

Claims (2)

両端が第１、第２の転がり軸受を介してケースに支持される回転軸と、この回転軸に対して一体回転可能かつ軸方向スライド可能に外装される筒軸とを有し、
前記回転軸において前記第１転がり軸受よりも内側の領域に一体形成された径方向外向きのフランジと、前記筒軸において前記回転軸のフランジ側に一体形成された径方向外向きのフランジとでベルトが巻き掛けられるプーリが構成され、
前記筒軸側のフランジを軸方向にスライドさせて前記プーリにおけるベルト巻き掛け径を可変調節する送り機構を有し、
前記送り機構は、一端側がケースに対して非回転かつ軸方向不動に取り付けられる内側部材と、この内側部材に対して螺合されかつ前記筒軸に対して第３の転がり軸受を介して支持される外側部材とで構成され、
前記第３転がり軸受が、潤滑剤封入タイプのアンギュラ玉軸受とされている、プーリ支持装置。A rotating shaft supported at both ends by a case via first and second rolling bearings, and a cylindrical shaft externally rotatable and axially slidable with respect to the rotating shaft;
A radially outward flange integrally formed on a region inside the first rolling bearing on the rotary shaft, and a radially outward flange integrally formed on a flange side of the rotary shaft on the cylindrical shaft. A pulley around which the belt is wound is configured,
A feed mechanism for variably adjusting the belt winding diameter of the pulley by sliding the flange on the cylinder shaft side in the axial direction,
The feed mechanism includes an inner member whose one end is non-rotatably and axially immovable with respect to the case, and is screwed to the inner member and supported to the cylindrical shaft via a third rolling bearing. Outer member,
A pulley supporting device, wherein the third rolling bearing is a lubricant-enclosed angular contact ball bearing. 前記第３転がり軸受としてのアンギュラ玉軸受は、軸方向一側にカウンタボアを有する外輪と、この外輪の反カウンタボア側に対応する領域に肉ぬすみを有する内輪と、前記外輪の軌道面と前記内輪の軌道面との間に介装される複数の玉とを含み、
前記玉の接触角は、２５度以上５０度以下に設定されているとともに、前記内輪の軌道面の曲率半径は、玉の直径の５３％以上５５％以下に、また、前記外輪の軌道面の曲率半径は、玉の直径の５２．５％以上５４．５％以下にそれぞれ設定されている、請求項１のプーリ支持装置。The angular contact ball bearing as the third rolling bearing includes an outer ring having a counterbore on one side in the axial direction, an inner ring having a thin wall in a region corresponding to the counterbore side of the outer ring, and a raceway surface of the outer ring. Including a plurality of balls interposed between the raceway surface of the inner ring,
The contact angle of the ball is set at 25 degrees or more and 50 degrees or less, the radius of curvature of the raceway surface of the inner ring is 53% or more and 55% or less of the diameter of the ball, and the raceway surface of the outer ring is 2. The pulley supporting device according to claim 1, wherein the curvature radii are set to 52.5% or more and 54.5% or less of the diameter of the ball.
JP2003001993A 2003-01-08 2003-01-08 Pulley supporting device Pending JP2004211862A (en)

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JP2013148115A (en) * 2012-01-17 2013-08-01 Ntn Corp Deep groove ball bearing and bearing device
WO2014171162A1 (en) * 2013-04-16 2014-10-23 日本精工株式会社 Angular ball bearing cage
WO2016194980A1 (en) * 2015-06-03 2016-12-08 株式会社ジェイテクト Rolling bearing
EP3225863A1 (en) * 2015-12-30 2017-10-04 Hiwin Technologies Corp. Ball bearing having a window cage of which the pockets include grooves for the storage of lubricant
US10174790B2 (en) 2016-01-26 2019-01-08 Jtekt Corporation Rolling bearing
US10260561B2 (en) 2017-03-22 2019-04-16 Jtekt Corporation Rolling bearing
US10539183B2 (en) 2017-03-31 2020-01-21 Jtekt Corporation Rolling bearing
US10788075B2 (en) 2017-10-04 2020-09-29 Jtekt Corporation Ball bearing
US10977968B2 (en) * 2016-10-12 2021-04-13 Patrick V. Cleeves Apparatus and methods for displaying and storing a banner or advertisement on a horizontal wind turbine

Cited By (14)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JP2013148115A (en) * 2012-01-17 2013-08-01 Ntn Corp Deep groove ball bearing and bearing device
WO2014171162A1 (en) * 2013-04-16 2014-10-23 日本精工株式会社 Angular ball bearing cage
JP2014209006A (en) * 2013-04-16 2014-11-06 日本精工株式会社 Cage for angular ball bearing
CN104302935A (en) * 2013-04-16 2015-01-21 日本精工株式会社 Angular ball bearing cage
US10221893B2 (en) 2015-06-03 2019-03-05 Jtekt Corporation Rolling bearing
WO2016194980A1 (en) * 2015-06-03 2016-12-08 株式会社ジェイテクト Rolling bearing
JP2016223598A (en) * 2015-06-03 2016-12-28 株式会社ジェイテクト Rolling bearing
EP3225863A1 (en) * 2015-12-30 2017-10-04 Hiwin Technologies Corp. Ball bearing having a window cage of which the pockets include grooves for the storage of lubricant
US10174790B2 (en) 2016-01-26 2019-01-08 Jtekt Corporation Rolling bearing
US10253813B2 (en) * 2016-01-26 2019-04-09 Jtekt Corporation Rolling bearing
US10977968B2 (en) * 2016-10-12 2021-04-13 Patrick V. Cleeves Apparatus and methods for displaying and storing a banner or advertisement on a horizontal wind turbine
US10260561B2 (en) 2017-03-22 2019-04-16 Jtekt Corporation Rolling bearing
US10539183B2 (en) 2017-03-31 2020-01-21 Jtekt Corporation Rolling bearing
US10788075B2 (en) 2017-10-04 2020-09-29 Jtekt Corporation Ball bearing

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