JP3937556B2 - Double row tapered roller bearing device - Google Patents

Double row tapered roller bearing device Download PDF

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Publication number
JP3937556B2
JP3937556B2 JP04659298A JP4659298A JP3937556B2 JP 3937556 B2 JP3937556 B2 JP 3937556B2 JP 04659298 A JP04659298 A JP 04659298A JP 4659298 A JP4659298 A JP 4659298A JP 3937556 B2 JP3937556 B2 JP 3937556B2
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tapered roller
roller bearing
gear
contact angle
roller bearings
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JPH11247848A (en
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彰 石丸
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NSK Ltd
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NSK Ltd
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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
    • F16C19/54—Systems consisting of a plurality of bearings with rolling friction
    • F16C19/56—Systems consisting of a plurality of bearings with rolling friction in which the rolling bodies of one bearing differ in diameter from those of another
    • 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
    • F16C19/22—Bearings with rolling contact, for exclusively rotary movement with bearing rollers essentially of the same size in one or more circular rows, e.g. needle bearings
    • F16C19/34—Bearings with rolling contact, for exclusively rotary movement with bearing rollers essentially of the same size in one or more circular rows, e.g. needle bearings for both radial and axial load
    • F16C19/36—Bearings with rolling contact, for exclusively rotary movement with bearing rollers essentially of the same size in one or more circular rows, e.g. needle bearings for both radial and axial load with a single row of rollers
    • F16C19/364—Bearings with rolling contact, for exclusively rotary movement with bearing rollers essentially of the same size in one or more circular rows, e.g. needle bearings for both radial and axial load with a single row of rollers with tapered rollers, i.e. rollers having essentially the shape of a truncated cone
    • 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
    • F16C19/54—Systems consisting of a plurality of bearings with rolling friction
    • F16C19/546—Systems with spaced apart rolling bearings including at least one angular contact bearing
    • F16C19/547—Systems with spaced apart rolling bearings including at least one angular contact bearing with two angular contact rolling bearings
    • F16C19/548—Systems with spaced apart rolling bearings including at least one angular contact bearing with two angular contact rolling bearings in O-arrangement
    • 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
    • F16C2361/00—Apparatus or articles in engineering in general
    • F16C2361/61—Toothed gear systems, e.g. support of pinion shafts
    • 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
    • F16H—GEARING
    • F16H48/00—Differential gearings
    • F16H48/38—Constructional details
    • F16H48/42—Constructional details characterised by features of the input shafts, e.g. mounting of drive gears thereon
    • F16H2048/423—Constructional details characterised by features of the input shafts, e.g. mounting of drive gears thereon characterised by bearing arrangement

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  • Engineering & Computer Science (AREA)
  • General Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • General Details Of Gearings (AREA)
  • Rolling Contact Bearings (AREA)

Description

【０００１】
【発明の属する技術分野】
この発明は、自動車の駆動系に設けられるデファレンシャルギヤを構成するピニオン軸をケーシング内に回転自在に支持する、複列円すいころ軸受装置の改良に関する。
【０００２】
【従来の技術】
自動車のプロペラシャフトと車輪の駆動軸（アクセル軸）との間には、デファレンシャルギヤを設ける事により、プロペラシャフトと駆動軸との間での動力伝達を行なうと共に、進路変更に伴う左右の駆動輪の回転速度差を吸収し、更に動力の向きを変え、最終的な減速を行なう様にしている。この様なデファレンシャルギヤとして従来から、例えば特開平９−２８７６１７号公報、実公平４−５１２３９号公報に記載された様な構造が知られている。図３は、これら公報等に記載されて従来から一般的に知られているデファレンシャルギヤの１例を示している。
【０００３】
このデファレンシャルギヤでは、ギヤケース１の一端部内側にピニオン軸２を、本発明の対象となる複列円すいころ軸受装置を構成する１対の円すいころ軸受３ａ、３ｂにより、回転自在に支持している。そして、このピニオン軸２の内端部（上記ギヤケース１の中心寄り端部で、図３の左端部）でインナー側（次述するリングギヤ５側で、図３の左側）の円すいころ軸受３ａよりも突出した部分に固設したピニオンギヤ４と、上記ギヤケース１内に回転自在に支承したリングギヤ５とを、互いに噛合させている。この様な構成により、上記ピニオン軸２の外端部（図３の右端部）に結合固定した、図示しないドライブシャフトの回転力を、上記ピニオンギヤ４を介して、上記リングギヤ５に伝達自在としている。更に、このリングギヤ５の回転を、図示しない別の円すいころ軸受により上記ギヤケース１に支承した、図示しない左右１対の駆動軸に伝達自在としている。
【０００４】
上述の様なデファレンシャルギヤの運転時には、上記ピニオンギヤ４とリングギヤ５との噛合に基づいて、上記ピニオン軸２の内端部にラジアル荷重が加わる。この様なラジアル荷重は、上記１対の円すいころ軸受３ａ、３ｂに加わり、これら各円すいころ軸受３ａ、３ｂ部分でアキシアル荷重を発生させる。即ち、これら各円すいころ軸受３ａ、３ｂを構成する外輪６ａ、６ｂの内周面には円すい凹面状の外輪軌道７ａ、７ｂを設けており、同じく内輪８ａ、８ｂの外周面には円すい凸面状の内輪軌道９ａ、９ｂを設けている。そして、これら各外輪軌道７ａ、７ｂと各内輪軌道９ａ、９ｂとの間に、それぞれ複数個ずつの円すいころ１０、１０を転動自在に設けている。上記各円すいころ軸受３ａ、３ｂにラジアル荷重が加わると、上記各円すいころ１０、１０の転動面と、上記各外輪軌道７ａ、７ｂ及び各内輪軌道９ａ、９ｂとの係合に基づいて、上記各円すいころ軸受３ａ、３ｂを構成する各外輪軌道７ａ、７ｂと各内輪軌道９ａ、９ｂとを軸方向に離す方向のスラスト荷重が加わる。
【０００５】
この様にして発生するスラスト荷重に拘らず、上記各内輪８ａ、８ｂを外嵌固定した前記ピニオン軸２が軸方向にずれ動く事を防止する為、図３に鎖線ａ、ｂで示す様に、上記各円すいころ軸受３ａ、３ｂの接触角の方向を互いに逆にしている。従って、上記ラジアル荷重に基づいてこれら各円すいころ軸受３ａ、３ｂ内で発生するアキシアル荷重の作用方向は互いに逆になり、上記ピニオン軸２内で相殺されるか、相殺されない場合でも何れかの円すいころ軸受３ａ、３ｂがこのアキシアル荷重を支承する為、このピニオン軸２が軸方向に変位する事はない。
【０００６】
この様な、上記ピニオン軸２を支承する１対の円すいころ軸受３ａ、３ｂとして従来は、接触角がほぼ同じものを使用していた。この理由は、上記ラジアル荷重に基づいてこれら各円すいころ軸受３ａ、３ｂ内で発生するアキシアル荷重の大きさをほぼ同じにし、これら各円すいころ軸受３ａ、３ｂ内で発生したアキシアル荷重を上記ピニオン軸２内でできるだけ相殺する事を意図した為である。又、上記アキシアル荷重を相殺し切れない場合でも、何れかの円すいころ軸受３ａ、３ｂが十分にこのアキシアル荷重を支承できる様にすべく、これら各円すいころ軸受３ａ、３ｂの接触角を或る程度大きくしていた。又、上記ラジアル荷重が、上記ピニオンギヤ４から上記ピニオン軸２の内端部に加わった場合でも、このピニオン軸２の支持剛性を確保してこのピニオン軸２の中心軸がずれる事を防止すべく、上記１対の円すいころ軸受３ａ、３ｂ同士の間隔を広くあけて、支持スパンを大きくしていた。
【０００７】
【発明が解決しようとする課題】
図３に示した様な従来構造の場合、デファレンシャルギヤを通じて伝達するトルクの増大に拘らず、十分な耐久性を確保しようとした場合には、ピニオン軸２を支持する為の複列円すいころ軸受及びこの複列円すいころ軸受を組み込んだデファレンシャルギヤが大型化する。即ち、エンジンの高出力化等により、上記トルクが増大すると、ピニオンギヤ４からピニオン軸２の内端部に加わるラジアル荷重が大きくなる。図３に示した従来構造のまま、このラジアル荷重に対する支持剛性を確保しようとした場合には、少なくとも上記ピニオンギヤ４に近い、インナー側の円すいころ軸受３ａを大型化する必要がある。
【０００８】
この様な円すいころ軸受３ａの大型化は、上記複列円すいころ軸受及びこの複列円すいころ軸受を組み込んだデファレンシャルギヤの大型化に結び付き、この大型化は自動車の重量増大及び燃費性能の悪化に結び付く為、好ましくない。
この様な事情に鑑みて本発明者が、上記複列円すいころ軸受の構成各部の挙動に就いて考察したところ、１対の円すいころ軸受３ａ、３ｂ内で発生するアキシアル荷重を相殺する為には、これら両円すいころ軸受３ａ、３ｂの接触角を同じにする必要がない事に気付いた。即ち、上記ラジアル荷重の入力部である上記ピニオンギヤ４に近い、インナー側の円すいころ軸受３ａには大きなラジアル荷重が作用し、この円すいころ軸受３ａ内で発生するアキシアル荷重もその分大きくなる。これに対して、上記ピニオンギヤ４から遠い、アウター側の円すいころ軸受３ｂに作用するラジアル荷重は小さくなり、この円すいころ軸受３ｂ内で発生するアキシアル荷重もその分だけ小さくなる。
【０００９】
従って、上記インナー側の円すいころ軸受３ａの接触角をアウター側の円すいころ軸受３ｂの接触角より小さくしても、デファレンシャルギヤの運転時にこれら両円すいころ軸受３ａ、３ｂ内で発生するアキシアル荷重の大きさにあまり大きな差を生じさせる事はない。一方、円すいころ軸受のラジアル剛性を高くし、当該円すいころ軸受の近傍に支持したギヤと相手ギヤとの噛合状態を正規の状態に維持する為には、当該円すいころ軸受の接触角を小さくする事が好ましい。
本発明の複列円すいころ軸受装置は、この様な事情に鑑みて発明したものである。
【００１０】
【課題を解決するための手段】
本発明の複列円すいころ軸受装置は、前述したデファレンシャルギヤに組み込まれて従来から知られている複列円すいころ軸受装置と同様に、接触角の方向を互いに逆方向とすると共に、それぞれの外輪を固定部分に内嵌固定した第一、第二の円すいころ軸受と、これら両円すいころ軸受の内輪に内嵌固定する事により、上記固定部分に回転自在に支持した回転軸とを備える。そして、この回転軸の一端部で上記第一の円すいころ軸受から上記第二の円すいころ軸受と反対側に突出した部分に固定したギヤを他のギヤと噛合させた状態で使用する。
【００１１】
特に、本発明の複列円すいころ軸受装置に於いては、上記第二の円すいころ軸受の接触角αを上記第一の円すいころ軸受の接触角βよりも５度以上大きくする。これと共に、上記両円すいころ軸受の内輪の両端面のうち、互いに反対側の外端面同士の間隔を、上記第二の円すいころ軸受の内輪の内径の２．８倍以下にしている。
【００１２】
【作用】
上述の様に構成する本発明の複列円すいころ軸受装置によれば、第一の円すいころ軸受の接触角を小さくする事により、この第一の円すいころ軸受のラジアル剛性を十分に高くできる。この為、回転軸の端部に固定したギヤから受ける大きなラジアル荷重に拘らず、このギヤのラジアル方向に亙る変位を抑えて、このギヤと他のギヤとの噛合状態を正規状態に維持できる。
第一の円すいころ軸受の接触角を第二の円すいころ軸受の接触角よりも小さくした事に伴い、これら両円すいころ軸受に同じ大きさのラジアル荷重が作用したと仮定した場合には、第二の円すいころ軸受内で発生するアキシアル荷重が、第一の円すいころ軸受内で発生するアキシアル荷重よりも大きくなる。但し、実際には、ギヤに近い第一の円すいころ軸受に作用するラジアル荷重が、ギヤから遠い第二の円すいころ軸受に作用するラジアル荷重よりも大きくなる。従って、上記第一、第二の円すいころ軸受の内部でそれぞれ発生するアキシアル荷重に、大きな差が生じる事はない。
【００１３】
【発明の実施の形態】
図１〜２は、本発明の実施の形態の１例を示している。尚、本発明の特徴は、第二の円すいころ軸受であるアウター側（運転時にラジアル荷重を受けるピニオンギヤ４から遠い側で、図１〜２の右側）の円すいころ軸受３ｂの接触角αを、第一の円すいころ軸受であるインナー側（ピニオンギヤ４に近い側で、図１〜２の左側）の円すいころ軸受３ａの接触角βよりも大きくする事により、必要とする剛性を確保しつつ、上記両円すいころ軸受３ａ、３ｂの間隔を狭めて小型・軽量化を可能にした点にある。その他の部分の構造及び作用は、前述の図３に示した従来構造と同様であるから、同等部分には同一符号を付して重複する説明を省略若しくは簡略にし、以下、本発明の特徴部分を中心に説明する。
【００１４】
上記アウター側の円すいころ軸受３ｂの接触角αを、上記インナー側の円すいころ軸受３ａの接触角βよりも５度以上大きくしている（α≧β＋５°）。この場合に、好ましくは、上記アウター側の円すいころ軸受３ｂの接触角αを３４度以下とし、上記インナー側の円すいころ軸受３ａの接触角βを１２度以上とする（１２°≦β、β＋５°≦α≦３４°）。上記インナー側の円すいころ軸受３ａの接触角βを１２度以上とする理由は、この円すいころ軸受３ａのラジアル剛性を高くする反面、必要とするアキシアル剛性を確保する為である。又、アウター側、インナー側両円すいころ軸受３ｂ、３ａの接触角α、βの差を５度以上にする理由は、アウター側の円すいころ軸受３ｂのアキシアル剛性を高くして、リングギヤ５と噛合するピニオンギヤ４から加わる大きなラジアル荷重に基づいて、上記インナー側の円すいころ軸受３ａ内で発生するアキシアル荷重を、十分に支承可能にする為である。更に、上記アウター側の円すいころ軸受３ｂの接触角αを３４度以下にする理由は、この円すいころ軸受３ｂのアキシアル剛性を高くする反面、必要とするラジアル剛性を確保する為である。
【００１５】
又、ピニオン軸２の長さを小さくすると共に、上記両円すいころ軸受３ａ、３ｂを構成する内輪８ａ、８ｂの内端面（互いに対向する端面）同士の間に挟持する円筒状のスペーサ１１の軸方向寸法を小さくしている。そして、上記内輪８ａ、８ｂの両端面のうち、互いに反対側の外端面同士の間隔Ｄ8 を、上記アウター側の円すいころ軸受３ｂを構成する内輪８ｂの内径Ｒ8bの２．８倍以下（Ｄ8 ≦２．８Ｒ8b）にしている。この様に、上記外端面同士の間隔Ｄ8 を上記内輪８ｂの内径Ｒ8bの２．８倍以下にする理由は、上記両円すいころ軸受３ａ、３ｂの接触角α、βを上記範囲に規制した事に伴い、これら両円すいころ軸受３ａ、３ｂに加わるラジアル荷重及びアキシアル荷重をバランスさせつつ、複列円すいころ軸受装置の小型化を図る為である。尚、上記外端面同士の間隔Ｄ8 の最小値は、上記内輪８ｂの内径Ｒ8bの１．５倍程度必要である。
【００１６】
上述の様に構成する本発明の複列円すいころ軸受装置によれば、インナー側の円すいころ軸受３ａの接触角βを小さくする事により、このインナー側の円すいころ軸受３ａのラジアル剛性を十分に高くできる。従って、回転軸であるピニオン軸２の内端部に固定したピニオンギヤ４から受ける大きなラジアル荷重に拘らず、このピニオンギヤ４のラジアル方向に亙る変位を抑えて、このピニオンギヤ４と、他のギヤであるリングギヤ５との噛合状態を正規状態に維持できる。この為、これらピニオンギヤ４とリングギヤ５との噛合部の伝達効率を確保すると共に、この噛合部で異音が発生したり、この噛合部の摩耗が著しくなる事を防止できる。
【００１７】
尚、上記インナー側の円すいころ軸受３ａの接触角βをアウター側の円すいころ軸受３ｂの接触角αよりも５度以上小さくした事に伴い、これら両円すいころ軸受３ａ、３ｂに同じ大きさのラジアル荷重が作用したと仮定した場合には、アウター側の円すいころ軸受３ｂ内で発生するアキシアル荷重が、インナー側の円すいころ軸受３ａ内で発生するアキシアル荷重よりも大きくなる。但し、実際には、上記ピニオンギヤ４に近いインナー側の円すいころ軸受３ａに作用するラジアル荷重が、上記ピニオンギヤ４から遠いアウター側の円すいころ軸受３ｂに作用するラジアル荷重よりも大きくなる。従って、上記インナー側、アウター側、両円すいころ軸受３ａ、３ｂの内部でそれぞれ発生するアキシアル荷重に、大きな差が生じる事はない。この為、これら両円すいころ軸受３ａ、３ｂの内部でそれぞれ発生するアキシアル荷重にアンバランスが生じても、その大きさは小さく、何れの円すいころ軸受３ａ、３ｂによっても、このアンバランス量を十分に支承できる。
【００１８】
【発明の効果】
本発明の複列円すいころ軸受装置は、以上に述べた通り構成され作用するので、例えばデファレンシャルギヤの耐久性及び伝達効率を確保しつつ、小型・軽量化を図れる。この為、デファレンシャルギヤを組み込んだ自動車の軽量化により性能向上を図る等、各種機械装置の性能向上に寄与できる。
【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、デファレンシャルギヤの部分断面図。
【図２】１対の円すいころ軸受のみを取り出して、図１と同方向から見た断面図。
【図３】従来構造の１例を示す、デファレンシャルギヤの部分断面図。
【符号の説明】
１ ギヤケース
２ ピニオン軸
３ａ、３ｂ 円すいころ軸受
４ ピニオンギヤ
５ リングギヤ
６ａ、６ｂ 外輪
７ａ、７ｂ 外輪軌道
８ａ、８ｂ 内輪
９ａ、９ｂ 内輪軌道
１０ 円すいころ
１１ スぺーサ[0001]
BACKGROUND OF THE INVENTION
The present invention relates to an improvement in a double-row tapered roller bearing device that rotatably supports a pinion shaft constituting a differential gear provided in a drive system of an automobile in a casing.
[0002]
[Prior art]
By providing a differential gear between the propeller shaft of the automobile and the drive shaft (accelerator shaft) of the wheel, power is transmitted between the propeller shaft and the drive shaft, and the left and right drive wheels accompanying the course change The difference in rotational speed is absorbed, the direction of power is changed, and the final deceleration is performed. As such a differential gear, a structure as described in, for example, Japanese Patent Application Laid-Open No. 9-287617 and Japanese Utility Model Publication No. 4-512239 is conventionally known. FIG. 3 shows an example of a differential gear described in these publications and the like and generally known from the past.
[0003]
In this differential gear, the pinion shaft 2 is rotatably supported by a pair of tapered roller bearings 3a and 3b constituting a double-row tapered roller bearing device which is an object of the present invention inside one end portion of the gear case 1. . Then, from a tapered roller bearing 3a on the inner side (the left end in FIG. 3 at the center end of the gear case 1 and the left end in FIG. 3) of the pinion shaft 2 on the inner side (the ring gear 5 side to be described below, the left side in FIG. 3). Also, the pinion gear 4 fixed to the protruding portion and the ring gear 5 rotatably supported in the gear case 1 are engaged with each other. With such a configuration, the rotational force of a drive shaft (not shown) coupled and fixed to the outer end (right end in FIG. 3) of the pinion shaft 2 can be transmitted to the ring gear 5 via the pinion gear 4. . Further, the rotation of the ring gear 5 can be transmitted to a pair of drive shafts (not shown) supported on the gear case 1 by another tapered roller bearing (not shown).
[0004]
During the operation of the differential gear as described above, a radial load is applied to the inner end portion of the pinion shaft 2 based on the meshing between the pinion gear 4 and the ring gear 5. Such a radial load is applied to the pair of tapered roller bearings 3a and 3b, and an axial load is generated at each of the tapered roller bearings 3a and 3b. That is, conical concave outer ring raceways 7a and 7b are provided on the inner peripheral surfaces of the outer rings 6a and 6b constituting the tapered roller bearings 3a and 3b, and the outer ring surfaces of the inner rings 8a and 8b are also formed in a tapered convex shape. Inner ring raceways 9a and 9b are provided. A plurality of tapered rollers 10, 10 are provided between the outer ring raceways 7a, 7b and the inner ring raceways 9a, 9b, respectively, so that they can roll. When a radial load is applied to each of the tapered roller bearings 3a and 3b, based on the engagement between the rolling surfaces of the tapered rollers 10 and 10 and the outer ring raceways 7a and 7b and the inner ring raceways 9a and 9b, A thrust load is applied in a direction that separates the outer ring raceways 7a and 7b and the inner ring raceways 9a and 9b constituting the tapered roller bearings 3a and 3b in the axial direction.
[0005]
Regardless of the thrust load generated in this way, the pinion shaft 2 to which the inner rings 8a and 8b are externally fitted and fixed is prevented from moving in the axial direction, as shown by chain lines a and b in FIG. The contact angle directions of the tapered roller bearings 3a and 3b are opposite to each other. Accordingly, the acting directions of the axial loads generated in the tapered roller bearings 3a and 3b based on the radial load are opposite to each other, and even if they are canceled in the pinion shaft 2 or are not canceled, any one of the cones. Since the roller bearings 3a and 3b support this axial load, the pinion shaft 2 is not displaced in the axial direction.
[0006]
Conventionally, the pair of tapered roller bearings 3a and 3b for supporting the pinion shaft 2 have the same contact angle. This is because the axial loads generated in the tapered roller bearings 3a and 3b based on the radial load are made substantially the same, and the axial loads generated in the tapered roller bearings 3a and 3b are set to the pinion shaft. This is because it was intended to cancel as much as possible within 2. Further, even when the axial load cannot be completely offset, the contact angle of each of the tapered roller bearings 3a and 3b is set so that any one of the tapered roller bearings 3a and 3b can sufficiently support the axial load. It was about bigger. Further, even when the radial load is applied from the pinion gear 4 to the inner end portion of the pinion shaft 2, the support rigidity of the pinion shaft 2 is secured to prevent the center axis of the pinion shaft 2 from shifting. The pair of tapered roller bearings 3a and 3b is widely spaced to increase the support span.
[0007]
[Problems to be solved by the invention]
In the case of the conventional structure as shown in FIG. 3, a double row tapered roller bearing for supporting the pinion shaft 2 when sufficient durability is to be ensured regardless of an increase in torque transmitted through the differential gear. And the differential gear incorporating this double row tapered roller bearing becomes larger. That is, when the torque increases due to high engine output, the radial load applied from the pinion gear 4 to the inner end of the pinion shaft 2 increases. In the case where it is intended to secure the supporting rigidity against the radial load with the conventional structure shown in FIG. 3, it is necessary to enlarge the tapered roller bearing 3a on the inner side, which is at least close to the pinion gear 4.
[0008]
Such an increase in the size of the tapered roller bearing 3a leads to an increase in the size of the double row tapered roller bearing and a differential gear incorporating the double row tapered roller bearing. This increase in weight increases the weight of the automobile and deteriorates the fuel efficiency. It is not preferable because it is tied.
In view of such circumstances, the present inventor has considered the behavior of each part of the double-row tapered roller bearing, and in order to cancel the axial load generated in the pair of tapered roller bearings 3a and 3b. Noticed that the contact angles of these tapered roller bearings 3a and 3b do not have to be the same. That is, a large radial load acts on the tapered roller bearing 3a on the inner side close to the pinion gear 4 which is the input portion of the radial load, and the axial load generated in the tapered roller bearing 3a is also increased accordingly. On the other hand, the radial load acting on the outer tapered roller bearing 3b, which is far from the pinion gear 4, is reduced, and the axial load generated in the tapered roller bearing 3b is also reduced accordingly.
[0009]
Therefore, even if the contact angle of the inner tapered roller bearing 3a is smaller than the contact angle of the outer tapered roller bearing 3b, the axial load generated in the tapered roller bearings 3a and 3b during the operation of the differential gear is reduced. It doesn't make much difference in size. On the other hand, in order to increase the radial rigidity of the tapered roller bearing and maintain the meshing state between the gear supported in the vicinity of the tapered roller bearing and the counterpart gear in a normal state, the contact angle of the tapered roller bearing is reduced. Things are preferable.
The double row tapered roller bearing device of the present invention has been invented in view of such circumstances.
[0010]
[Means for Solving the Problems]
The double row tapered roller bearing device of the present invention is incorporated in the above-described differential gear, and in the same manner as the conventionally known double row tapered roller bearing device, the contact angles are opposite to each other, and each outer ring The first and second tapered roller bearings are internally fitted and fixed to the fixed portion, and the rotating shaft is rotatably supported on the fixed portion by being internally fitted and fixed to the inner rings of these tapered roller bearings. And the gear fixed to the part which protruded on the opposite side to said 2nd tapered roller bearing from said 1st tapered roller bearing in the one end part of this rotating shaft is used in the state which meshed | engaged with the other gear.
[0011]
In particular, in the double row tapered roller bearing device of the present invention, the contact angle α of the second tapered roller bearing is set to be 5 degrees or more larger than the contact angle β of the first tapered roller bearing. At the same time, the distance between the outer end faces on the opposite sides of the both end faces of the inner ring of the both tapered roller bearings is made 2.8 times or less the inner diameter of the inner ring of the second tapered roller bearing.
[0012]
[Action]
According to the double-row tapered roller bearing device of the present invention configured as described above, the radial rigidity of the first tapered roller bearing can be sufficiently increased by reducing the contact angle of the first tapered roller bearing. For this reason, regardless of the large radial load received from the gear fixed to the end portion of the rotating shaft, the gear can be kept in the normal state by suppressing the displacement of the gear in the radial direction.
Assuming that the radial load of the same magnitude is applied to both tapered roller bearings due to the contact angle of the first tapered roller bearing being smaller than the contact angle of the second tapered roller bearing, The axial load generated in the second tapered roller bearing is larger than the axial load generated in the first tapered roller bearing. However, in practice, the radial load acting on the first tapered roller bearing close to the gear is larger than the radial load acting on the second tapered roller bearing far from the gear. Accordingly, there is no great difference between the axial loads generated inside the first and second tapered roller bearings.
[0013]
DETAILED DESCRIPTION OF THE INVENTION
1 and 2 show an example of an embodiment of the present invention. The feature of the present invention is that the contact angle α of the tapered roller bearing 3b on the outer side (the side far from the pinion gear 4 that receives a radial load during operation and the right side in FIGS. While ensuring the required rigidity by making it larger than the contact angle β of the tapered roller bearing 3a on the inner side (the side close to the pinion gear 4 and the left side of FIGS. 1 and 2) which is the first tapered roller bearing, The point is that the distance between the tapered roller bearings 3a and 3b is reduced, thereby enabling a reduction in size and weight. Since the structure and operation of the other parts are the same as those of the conventional structure shown in FIG. 3 described above, the same parts are denoted by the same reference numerals, and redundant description is omitted or simplified. The explanation will be focused on.
[0014]
The contact angle α of the outer tapered roller bearing 3b is larger than the contact angle β of the inner tapered roller bearing 3a by 5 degrees or more (α ≧ β + 5 °). In this case, preferably, the contact angle α of the outer tapered roller bearing 3b is set to 34 ° or less, and the contact angle β of the inner tapered roller bearing 3a is set to 12 ° or more (12 ° ≦ β, β + 5 ° ≦ α ≦ 34 °). The reason why the contact angle β of the tapered roller bearing 3a on the inner side is set to 12 degrees or more is to increase the radial rigidity of the tapered roller bearing 3a, while ensuring the required axial rigidity. Also, the reason for the difference between the contact angles α and β of the outer side and inner side tapered roller bearings 3b and 3a to be 5 degrees or more is that the axial rigidity of the outer side tapered roller bearing 3b is increased and meshed with the ring gear 5. This is because the axial load generated in the tapered roller bearing 3a on the inner side can be sufficiently supported on the basis of the large radial load applied from the pinion gear 4 to be performed. Furthermore, the reason why the contact angle α of the tapered roller bearing 3b on the outer side is set to 34 degrees or less is to increase the axial rigidity of the tapered roller bearing 3b, but to ensure the required radial rigidity.
[0015]
In addition, the shaft of the cylindrical spacer 11 is made smaller while reducing the length of the pinion shaft 2 and sandwiching it between the inner end faces (end faces facing each other) of the inner rings 8a and 8b constituting the tapered roller bearings 3a and 3b. The direction dimension is reduced. Of the both end faces of the inner rings 8a and 8b, the distance D 8 between the opposite outer end faces is 2.8 times or less the inner diameter R 8b of the inner ring 8b constituting the outer tapered roller bearing 3b ( D 8 ≦ 2.8R 8b ). Thus, the reason for the distance D 8 between the outer end face below 2.8 times the inner diameter R 8b of the inner ring 8b is restricted the both tapered roller bearings 3a, the contact angle of 3b alpha, a β to the range This is to reduce the size of the double-row tapered roller bearing device while balancing the radial load and the axial load applied to both the tapered roller bearings 3a and 3b. The minimum value of the distance D 8 between the outer end surface, it is necessary 1.5 times the inner diameter R 8b of the inner ring 8b.
[0016]
According to the double-row tapered roller bearing device of the present invention configured as described above, the radial rigidity of the inner tapered roller bearing 3a is sufficiently increased by reducing the contact angle β of the inner tapered roller bearing 3a. Can be high. Therefore, regardless of the large radial load received from the pinion gear 4 fixed to the inner end of the pinion shaft 2 that is the rotating shaft, the pinion gear 4 and other gears are restrained from being displaced in the radial direction. The meshing state with the ring gear 5 can be maintained in a normal state. For this reason, it is possible to ensure the transmission efficiency of the meshing portion between the pinion gear 4 and the ring gear 5, and to prevent abnormal noise from being generated at the meshing portion and significant wear of the meshing portion.
[0017]
The contact angle β of the inner tapered roller bearing 3a is smaller than the contact angle α of the outer tapered roller bearing 3b by 5 degrees or more, so that both the tapered roller bearings 3a and 3b have the same size. Assuming that a radial load is applied, the axial load generated in the outer tapered roller bearing 3b is larger than the axial load generated in the inner tapered roller bearing 3a. However, in practice, the radial load acting on the inner tapered roller bearing 3 a close to the pinion gear 4 is larger than the radial load acting on the outer tapered roller bearing 3 b far from the pinion gear 4. Therefore, there is no great difference between the axial loads generated in the inner side, outer side, and both tapered roller bearings 3a and 3b. For this reason, even if the axial load generated inside each of these tapered roller bearings 3a and 3b is unbalanced, the magnitude of the axial load is small, and any of the tapered roller bearings 3a and 3b has a sufficient amount of unbalance. Can be supported.
[0018]
【The invention's effect】
Since the double row tapered roller bearing device of the present invention is configured and operates as described above, it is possible to reduce the size and weight while ensuring the durability and transmission efficiency of the differential gear, for example. For this reason, it can contribute to the performance improvement of various machine devices, such as aiming at a performance improvement by weight reduction of the motor vehicle incorporating a differential gear.
[Brief description of the drawings]
FIG. 1 is a partial sectional view of a differential gear, showing an example of an embodiment of the present invention.
FIG. 2 is a cross-sectional view taken from the same direction as FIG. 1 with only a pair of tapered roller bearings taken out.
FIG. 3 is a partial sectional view of a differential gear, showing an example of a conventional structure.
[Explanation of symbols]
DESCRIPTION OF SYMBOLS 1 Gear case 2 Pinion shaft 3a, 3b Tapered roller bearing 4 Pinion gear 5 Ring gear 6a, 6b Outer ring 7a, 7b Outer ring raceway 8a, 8b Inner ring 9a, 9b Inner ring raceway 10 Tapered roller 11 Spacer

Claims (2)

接触角の方向を互いに逆方向とすると共に、それぞれの外輪を固定部分に内嵌固定した第一、第二の円すいころ軸受と、これら両円すいころ軸受の内輪に内嵌固定する事により、上記固定部分に回転自在に支持した回転軸とを備え、この回転軸の一端部で上記第一の円すいころ軸受から上記第二の円すいころ軸受と反対側に突出した部分に固定したギヤを他のギヤと噛合させた状態で使用する複列円すいころ軸受装置に於いて、上記第二の円すいころ軸受の接触角αを上記第一の円すいころ軸受の接触角βよりも５度以上大きくすると共に、上記両円すいころ軸受の内輪の両端面のうち、互いに反対側の外端面同士の間隔を、上記第二の円すいころ軸受の内輪の内径の２．８倍以下にした事を特徴とする複列円すいころ軸受装置。The direction of the contact angle is opposite to each other, and the first and second tapered roller bearings in which the outer rings are fitted and fixed to the fixed portion, and the inner rings of these tapered roller bearings are fitted and fixed to the inner ring. A rotating shaft rotatably supported by a fixed portion, and a gear fixed to a portion protruding from the first tapered roller bearing to the opposite side of the second tapered roller bearing at one end of the rotating shaft. In a double row tapered roller bearing device used in a state of being engaged with a gear, the contact angle α of the second tapered roller bearing is made 5 degrees or more larger than the contact angle β of the first tapered roller bearing. Further, among the both end faces of the inner ring of the two tapered roller bearings, the distance between the opposite outer end faces is 2.8 times or less the inner diameter of the inner ring of the second tapered roller bearing. Row tapered roller bearing device. 第二の円すいころ軸受の接触角αを３４度以下とし、第一の円すいころ軸受の接触角βを１２度以上とした、請求項１に記載した複列円すいころ軸受装置。2. The double-row tapered roller bearing device according to claim 1, wherein the contact angle α of the second tapered roller bearing is 34 degrees or less and the contact angle β of the first tapered roller bearing is 12 degrees or more.
JP04659298A 1998-02-27 1998-02-27 Double row tapered roller bearing device Expired - Lifetime JP3937556B2 (en)

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JP04659298A JP3937556B2 (en) 1998-02-27 1998-02-27 Double row tapered roller bearing device

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JP3937556B2 true JP3937556B2 (en) 2007-06-27

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JP2003172345A (en) † 2001-12-07 2003-06-20 Koyo Seiko Co Ltd Axle pinion bearing device and vehicular final reduction gear
EP3001051A1 (en) * 2014-09-29 2016-03-30 Aktiebolaget SKF Bearing unit for pinions

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