JP4225006B2 - Double row angular contact ball bearings for wheels - Google Patents

Double row angular contact ball bearings for wheels Download PDF

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Publication number
JP4225006B2
JP4225006B2 JP2002240319A JP2002240319A JP4225006B2 JP 4225006 B2 JP4225006 B2 JP 4225006B2 JP 2002240319 A JP2002240319 A JP 2002240319A JP 2002240319 A JP2002240319 A JP 2002240319A JP 4225006 B2 JP4225006 B2 JP 4225006B2
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Prior art keywords
ball
row
load
ring raceway
contact angle
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JP2004076892A (en
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淳太郎 佐原
宏敏 荒牧
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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
    • F16C19/02—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
    • F16C19/18—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 two or more rows of balls
    • F16C19/181—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 two or more rows of balls with angular contact
    • F16C19/183—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 two or more rows of balls with angular contact with two rows at opposite angles
    • F16C19/184—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 two or more rows of balls with angular contact with two rows at opposite angles in O-arrangement
    • F16C19/186—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 two or more rows of balls with angular contact with two rows at opposite angles in O-arrangement with three raceways provided integrally on parts other than race rings, e.g. third generation hubs
    • 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
    • F16C33/58—Raceways; Race rings
    • 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
    • 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
    • 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
    • F16C2326/00—Articles relating to transporting
    • F16C2326/01—Parts of vehicles in general
    • F16C2326/02—Wheel hubs or castors

Description

【０００１】
【発明の属する技術分野】
本発明の車輪用複列アンギュラ型玉軸受は、自動車の車輪を懸架装置に対して回転自在に支持する為に利用する。
【０００２】
【従来の技術】
自動車の車輪を懸架装置に対して回転自在に支持する為に従来から、例えば図５に示す様な車輪用複列アンギュラ型玉軸受が使用されている。この図５に示した車輪用複列アンギュラ型玉軸受は、駆動輪（ＦＦ車の前輪、ＦＲ車及びＲＲ車の後輪、４ＷＤ車の全輪）を支持する為のもので、外輪相当部材である外輪１と、内輪相当部材であるハブ２と、複数個の玉３、３とを備える。このうちの外輪１は、内周面に外向アンギュラ型の外輪軌道４ａ、４ｂを複列に形成すると共に、外周面に結合フランジ５を形成している。車両への組み付け時には、この結合フランジ５を懸架装置を構成するナックルに結合する。
【０００３】
又、上記ハブ２は、ハブ本体６と内輪７とを組み合わせて成る。このうちのハブ本体６の中心部にはスプライン孔８を、外周面の外端（軸方向に関して「外」とは、車両への組み付け状態で幅方向外側となる側を言い、図１、５の左側。反対に、車両への組み付け状態で幅方向中央側となる、図１、５の右側を「内」と言う。本明細書全体で同じ。）寄り部分には取付フランジ９を、それぞれ形成している。車両への組み付け時には、上記スプライン孔８に図示しない等速ジョイントに付属したスプライン軸を挿入すると共に、上記取付フランジ９に車輪を固定する。
【０００４】
又、上記ハブ本体６の外周面の中間部で、上記外輪１の内周面に形成した複列の外輪軌道４ａ、４ｂのうちのアウター側（車両への組み付け状態で幅方向外側となる側を言い、図１、５の左側。本明細書全体で同じ。）の外輪軌道４ｂと対向する部分には、アンギュラ型の内輪軌道１０ｂを形成している。更に、上記ハブ本体６の内端部に形成した小径段部１１に上記内輪７を外嵌している。そして、この内輪７の外周面に形成したアンギュラ型の内輪軌道１０ａを、上記複列の外輪軌道４ａ、４ｂのうちのインナー側（車両への組み付け状態で幅方向中央側となる側を言い、図１、５の右側。本明細書全体で同じ。）の外輪軌道４ａに対向させている。
【０００５】
そして、上記各外輪軌道４ａ、４ｂと上記各内輪軌道１０ａ、１０ｂとの間に前記各玉３、３を、それぞれ複数個ずつ、図示しない保持器により保持した状態で転動自在に設けている。この構成により、背面組み合わせである複列アンギュラ型の玉軸受を構成し、上記外輪１の内径側に前記ハブ２を、回転自在に、且つ、ラジアル荷重及びアキシアル荷重を支承自在に支持している。尚、図示の構造の場合、上記内輪７の外端面を上記小径段部１１の基端縁に設けた段差面１２に当接させた状態で、インナー側の玉列とアウター側の玉列とを構成する各玉３、３に、それぞれ所定の接触角α1 、α2 及び予圧が付与される様に、各部の形状及び寸法を規制している。又、上記外輪１の内周面と上記ハブ２の外周面との間で、上記複数個の玉３、３を設置した空間には、潤滑用のグリースを封入すると共に、この空間の両端開口を、それぞれシールリング１３ａ、１３ｂにより密閉している。
【０００６】
尚、上述の図５に示した構造を含む、従来の車輪用複列アンギュラ型玉軸受の場合、インナー側の玉列の接触角α1 とアウター側の玉列の接触角α2 とを、互いに等しく（α1 ＝α2 ）している。又、インナー側の内輪軌道１０ａの負荷側（負荷を受ける側で、図５の右側）の溝肩高さＨi1と、アウター側の内輪軌道１０ｂの負荷側（同じく、図５の左側）の溝肩高さＨi2とを、互いに等しく（Ｈi1＝Ｈi2）している。更に、インナー側の外輪軌道４ａの負荷側（同じく、図５の左側）の溝肩高さＨo1と、アウター側の外輪軌道４ｂの負荷側（同じく、図５の右側）の溝肩高さＨo2とを、互いに等しく（Ｈo1＝Ｈo2）している。
【０００７】
【発明が解決しようとする課題】
上述した様な車輪用複列アンギュラ型玉軸受には、車両の旋回走行時等にモーメント荷重が加わるが、この際、インナー側の玉列には、直進時に比べて大きなラジアル荷重が（アキシアル荷重よりも優勢に）負荷される。この様にインナー側の玉列にラジアル荷重が優勢に負荷されると、このインナー側の玉列の負荷圏が狭くなり、このインナー側の玉列を構成する個々の玉３に加わる荷重が大きくなる。この為、負荷分布に基づく計算寿命は、アウター側の玉列よりもインナー側の玉列の方が短くなる。
【０００８】
一方、車輪用複列アンギュラ型玉軸受の回転速度や負荷荷重が大きくなると、上記各玉３、３の転動面と前記外輪、内輪各軌道４ａ、４ｂ、１０ａ、１０ｂとの接触部に於ける、摩擦による発熱量が増大する。この様に当該接触部での発熱量が増大すると、潤滑用のグリースの劣化が促進されたり、当該接触部に油膜が形成されにくくなって、インナー側、アウター側の各玉列の疲れ寿命が低下する。特に、実際の運転条件に即した耐久試験に基づく実験寿命は、インナー側の玉列よりもアウター側の玉列の方が短くなる。
【０００９】
本発明の車輪用複列アンギュラ型玉軸受は、上述の様な事情に鑑み、インナー側の玉列とアウター側の玉列との寿命のバランスを取りつつ、車輪用複列アンギュラ型玉軸受全体としての寿命の延長を図れる構造を実現すべく発明したものである。
【００１０】
【課題を解決するための手段】
本発明の車輪用複列アンギュラ型玉軸受は、外輪相当部材と、内輪相当部材と、複数個の玉とを備える。
このうちの外輪相当部材は、内周面にアンギュラ型の外輪軌道を、複列に設けている。
又、上記内輪相当部材は、外周面にアンギュラ型の内輪軌道を、複列に設けている。
又、上記各玉は、上記各内輪軌道と上記各外輪軌道との間に、それぞれ複数個ずつ転動自在に設けられている。
【００１１】
特に、本発明の車輪用複列アンギュラ型玉軸受に於いては、上記懸架装置への組み付け時、車両の幅方向中央側となるインナー側の玉列の玉径と、同じく幅方向外側となるアウター側の玉列の玉径とが等しい。
又、インナー側の玉列の接触角を、アウター側の玉列の接触角よりも小さくしている。
【００１２】
又、アウター側の内輪軌道の負荷側の溝肩高さを、インナー側の内輪軌道の負荷側の溝肩高さよりも大きくしている。
【００１３】
更に、アウター側の外輪軌道の負荷側の溝肩高さを、インナー側の外輪軌道の負荷側の溝肩高さよりも大きくしている。
【００１４】
【作用】
上述の様に構成する本発明の車輪用複列アンギュラ型玉軸受は、インナー側の玉列の接触角をアウター側の玉列の接触角よりも小さくしている為、インナー側の玉列のラジアル荷重の支承能力を、アウター側の玉列に比べて大きくできる。従って、車両の旋回走行時等に車輪用複列アンギュラ型玉軸受にモーメント荷重が負荷され、これに伴ってインナー側の玉列に大きなラジアル荷重が（アキシアル荷重よりも優勢に）負荷される状態が生じる事に拘らず、このインナー側の玉列の寿命をアウター側の玉列の寿命と同程度にまで長くする事ができる。又、インナー側の玉列の接触角をアウター側の玉列の接触角よりも小さくする事により、これら両玉列が支承する荷重の分担をバランスさせる事ができる。この為、各玉の転動面と各軌道との接触部に於ける、摩擦による発熱量を少なくできる。この結果、潤滑用のグリースの劣化が促進される事や、上記接触部での油膜形成が不良になる事を有効に防止でき、インナー側、アウター側の各玉列の寿命を長くする事ができる。
【００１５】
又、インナー側の玉列の接触角を小さくしたので、このインナー側の玉列にラジアル荷重が負荷される際に発生する、アキシアル方向の分力を小さくできる。このアキシアル方向の分力は、アウター側の玉列に負荷されるアキシアル荷重の一部となる。この為、このアウター側の玉列に負荷されるアキシアル荷重を小さくできる。この結果、このアウター側の玉列を構成する各玉が軌道の肩部（幅方向端縁）に乗り上がりにくくなる為、このアウター側の玉列の寿命が低下する事を有効に防止できる。即ち、上記各玉が軌道の肩部に乗り上がると、これら各玉の転動面と軌道との接触面圧が過大になると共に、これら各玉の転動面が傷つき易くなる為、当該玉列の寿命が低下し易くなる。これに対し、本発明の場合には、上述した様にアウター側の玉列を構成する各玉が軌道の肩部に乗り上がりにくくなる為、このアウター側の玉列の寿命が低下する事を有効に防止できる。
【００１６】
更に、本発明によれば、アウター側の玉列を構成する各玉を、内輪、外輪各軌道の肩部に乗り上がりにくくして、このアウター側の玉列の寿命が低下する事を有効に防止できる。即ち、次述する実施例で示す様に、アウター側の玉列の接触角は、（各玉が軌道の肩部に乗り上がらない範囲内では）その値を大きくする程各玉と軌道との接触部での発熱量を少なく抑える事ができ、且つ、計算寿命を長くできる。この為、アウター側の玉列の接触角は、（各玉が軌道の肩部に乗り上がらない範囲内で）できるだけ大きくするのが好ましい。ところが、車輪用複列アンギュラ型玉軸受の構造上、アウター側の玉列を構成する各玉は、インナー側の玉列を構成する各玉に比べて、軌道の肩部に乗り上がり易い。従って、上述した様にアウター側の玉列の接触角を大きくすると、このアウター側の玉列を構成する各玉が、軌道の肩部に余計に乗り上がり易くなる。これに対し、本発明の場合には、アウター側の内輪、外輪各軌道の負荷側の溝肩高さを、インナー側の内輪、外輪各軌道の負荷側の溝肩高さよりも大きくしている為、アウター側の玉列を構成する各玉を、内輪、外輪各軌道の肩部に乗り上がりにくくする事ができる。従って、このアウター側の玉列の寿命が低下する事を有効に防止できる。
【００１７】
尚、インナー側の内輪、外輪各軌道の負荷側の溝肩高さも、アウター側の内輪、外輪各軌道の負荷側の溝肩高さと同程度に大きくすれば、インナー側の玉列を構成する各玉を、内輪、外輪各軌道の肩部に、より乗り上がりにくくする事ができる。但し、この様にすると、インナー側の内輪、外輪各軌道の表面の研磨面積が広がって加工コストが高くなると共に、車輪用複列アンギュラ型玉軸受の重量（車両のばね下荷重）が徒に大きくなる。従って、この様な不都合を防止する為に、インナー側の内輪、外輪各軌道の負荷側の溝肩高さは、必要最小限の値に抑えるのが好ましい。
【００１８】
【実施例】
本発明をなす過程で行なった実験（コンピュータシミュレーション）に就いて説明する。実験は、図１に略示する様な車輪用複列アンギュラ型玉軸受を対象に行なった。この図１に略示した車輪用複列アンギュラ型玉軸受の構造及び作用は、前述の図５に示した従来の車輪用複列アンギュラ型玉軸受とほぼ同様である為、同等部分には同一符号を付して重複する説明を省略する。実験では、この様な車輪用複列アンギュラ型玉軸受に就き、諸元の異なる複数種類のものを用意し、それぞれを実験の試料とした。これら各試料の諸元を、以下の表１に示す。
【表１】

Figure 0004225006
【００１９】
この表１に示す様に、試料の玉径、ピッチ円径、各列毎の玉３、３の数（玉数／列）、列間距離（列同士のピッチ）、予圧荷重、外輪軌道の負荷側の溝肩高さ、及び内輪軌道の負荷側の溝肩高さは、それぞれ上記各試料同士で互いに等しくした。尚、懸架装置への組み付け時、車両の幅方向中央側となるインナー側の玉列の玉径と、同じく幅方向外側となるアウター側の玉列の玉径とを等しくしている。但し、インナー側の外輪軌道４ａの負荷側の溝肩高さＨo1と、アウター側の外輪軌道４ｂの負荷側の溝肩高さＨo2とを、互いに等しく（Ｈo1＝Ｈo2＝４．７mm）する一方で、アウター側の内輪軌道１０ｂの負荷側の溝肩高さＨi2を、インナー側の内輪軌道１０ａの負荷側の溝肩高さＨi1よりも大きく｛Ｈi2（＝５．５mm）＞Ｈi1（＝５．０mm）｝した。これに対し、上記各試料毎に、インナー側の玉列の接触角α1 とアウター側の玉列の接触角α2 とを、それぞれ３０〜４５度の範囲で異ならせた。
【００２０】
そして、上述の様な各試料に就いて、軸受トルクから換算される発熱量と、Lundberg-Palmgren の寿命理論から計算される応力分布によるＬ10寿命と、アウター側の内輪軌道１０ｂでの玉の乗り上げ率とを、それぞれコンピュータシミュレーションによって求めた。試験条件は、以下の通りである。
Figure 0004225006
尚、本実験で、アウター側の内輪軌道１０ｂでの玉の乗り上げ率を調べた理由は、車輪用複列アンギュラ型玉軸受の構造上、インナー側の軌道よりもアウター側の軌道の方が玉の乗り上げが生じ易く、特に、外輪軌道よりも内輪軌道の方が、より玉の乗り上げが生じ易い為である。
【００２１】
上述の様にして行なったコンピュータシミュレーションの結果を、図２〜４に示す。先ず、図２は、インナー側、アウター側の各玉列の接触角と、軸受トルクから換算される発熱量との関係を示している。この図２に示した実験結果から明らかな通り、「インナー側の玉列の接触角＝アウター側の玉列の接触角」の場合を基準に見ると、「インナー側の玉列の接触角＜アウター側の玉列の接触角」の場合に発熱量が少なくなり、「インナー側の玉列の接触角＞アウター側の玉列の接触角」の場合に発熱量が多くなる。従って、「インナー側の玉列の接触角＜アウター側の玉列の接触角」とすれば、発熱量を少なくする事ができる。この結果、各玉３、３の設置部に封入した潤滑用のグリースの劣化が促進される事や、これら各玉３、３の転動面と外輪、内輪各軌道４ａ、４ｂ、１０ａ、１０ｂとの接触部での油膜形成が不良になる事を有効に防止でき、インナー側、アウター側の各玉列の寿命を延長させる事ができる。
【００２２】
次に、図３は、インナー側、アウター側の各玉列の接触角と、Lundberg-Palmgren の寿命理論から計算される応力分布によるＬ10寿命との関係を示している。この図３中、Ｌ10寿命は、「インナー側の玉列の接触角＝アウター側の玉列の接触角＝３５度」の場合のＬ10寿命を「１」としてこれに対する比率を示す、「寿命比」で表している。この図３に示した実験結果から明らかな通り、「インナー側の玉列の接触角＝アウター側の玉列の接触角」の場合を基準に見ると、「インナー側の玉列の接触角＜アウター側の玉列の接触角」の場合に計算寿命が長くなり、「インナー側の玉列の接触角＞アウター側の玉列の接触角」の場合に計算寿命が短くなる。従って、「インナー側の玉列の接触角＜アウター側の玉列の接触角」とすれば、計算寿命を延長させる事ができる。
【００２３】
次に、図４は、インナー側、アウター側の各玉列の接触角と、アウター側の内輪軌道１０ｂでの玉の乗り上げ率との関係を示している。尚、玉の乗り上げ率は、ヘルツの接触理論から計算される接触楕円が軌道（上記内輪軌道１０ｂ）の幅方向端縁からはみ出している長さの割合を示し、接触楕円の長半径をａ、この接触楕円が軌道の幅方向端縁からはみ出している長さをζとすると、（ζ／２ａ）×１００［％］で表される。又、玉の乗り上げ率は、符号が「正」である場合に、上記接触楕円が上記内輪軌道１０ｂの幅方向端縁からはみ出している事を示し、符号が「負」である場合に、上記接触楕円が上記内輪軌道１０ｂの幅方向端縁からはみ出していない（はみ出す事に対する余裕がある）事を示す。
【００２４】
図４に示した実験結果では、総ての試料に就いての玉の乗り上げ率の符号が負であるから、これら総ての試料が、玉の乗り上げに対する余裕を持っている事が分かる。この様な実験結果が得られた理由は、アウター側の内輪軌道１０ｂの負荷側の溝肩高さＨi2を、インナー側の内輪軌道１０ａの負荷側の溝肩高さＨi1よりも大きく（Ｈi2＞Ｈi1）した結果、アウター側の内輪軌道１０ｂの肩部に前記各玉３、３が乗り上がりにくくなった為である。尚、インナー側の内輪軌道１０ａの負荷側の溝肩高さＨi1も、アウター側の内輪軌道１０ｂの負荷側の溝肩高さＨi2と同程度に大きく（Ｈi2＞Ｈi1）すれば、インナー側の内輪軌道１０ａでの玉の乗り上げ率も、更に小さくする事ができる。但し、この様にすると、インナー側の内輪軌道１０ａの表面の研磨面積が広がって加工コストが高くなると共に、車輪用複列アンギュラ型玉軸受の重量（車両のばね下荷重）が徒に大きくなる。従って、本例の場合には、この様な不都合を防止する為、インナー側の内輪軌道１０ａの負荷側の溝肩高さＨi1を、必要最小限の値に抑えている。
【００２５】
又、図４に示した実験結果では、上述した様に、総ての試料に就いての玉の乗り上げ率の符号が負になる事が示された。ところが、実際の運転時には、タイヤが縁石に乗り上げる等により、車輪用複列アンギュラ型玉軸受に負荷されるアキシアル荷重が、前記試験条件で示したアキシアル荷重Ｆa よりも大きくなる可能性がある。そして、この様な場合に、車輪用複列アンギュラ型玉軸受を構成する各部材の弾性変形量が予想以上に大きくなって、上記玉の乗り上げ率の符号が負から正に変わる可能性がある。従って、この様な不都合が生じない様にすべく、車輪用複列アンギュラ型玉軸受の諸元は、図４に示した実験結果で、玉の乗り上げ率が−２５％以下（絶対値が２５％以上）になる様に設定するのが好ましい。即ち、図４に示す様に、アウター側の玉列の接触角を４５度まで大きくすると、玉の乗り上げ率が−２５％以上（絶対値が２５％以下）になって上述した条件を満足しなくなる為、アウター側の玉列の接触角は４０度以上４５度未満に設定するのが好ましい。又、前述の図２〜３に示した実験結果に基づく結論を踏まえて、インナー側の玉列の接触角は、アウター側の玉列の接触角（４０度以上４５度未満）よりも小さく設定するのが好ましい。
【００２６】
又、上述した様な計算機シミュレーションの結果に基づいて、車輪用複列アンギュラ型玉軸受のインナー側の玉列の接触角を３３±２度とし、アウター側の玉列の接触角を４０±２度として実際の運転を行なったところ、この車輪用複列アンギュラ型玉軸受の実験寿命が、従来構造の実験寿命の２〜５倍に延びた。以上の点から、本発明を実施する場合に、インナー側の玉列の接触角α1 を３０〜４０度、アウター側の玉列の接触角α2 を３５〜４５度の範囲内に収め、且つ、アウター側の玉列の接触角α2 をインナー側の玉列の接触角α1 よりも５度以上大きく（α2 −α1 ≧５度）する事が好ましい。
【００２７】
尚、上述した構造では、各試料のインナー側の外輪軌道４ａの負荷側の溝肩高さＨo1と、アウター側の外輪軌道４ｂの負荷側の溝肩高さＨo2とを、互いに等しくした。但し、本発明の様に、アウター側の外輪軌道４ｂの負荷側の溝肩高さＨo2をインナー側の外輪軌道４ａの負荷側の溝肩高さＨo1よりも大きく（Ｈo2＞Ｈo1）すれば、アウター側の外輪軌道４ｂでの玉の乗り上げ率を、インナー側の外輪軌道４ａでの玉の乗り上げ率と同程度に小さくする事ができる。
【００２８】
又、本発明の車輪用複列アンギュラ型玉軸受は、前述の図１、５に示した様な駆動輪用のものに限らず、従動輪用のものにも適用可能である。又、駆動輪用、従動輪用とも、各種構造の車輪用複列アンギュラ型玉軸受に適用できる。
【００２９】
【発明の効果】
本発明の車輪用複列アンギュラ型玉軸受は、以上に述べた様に構成され作用する為、インナー側の玉列とアウター側の玉列との寿命のバランスを取りつつ、車輪用複列アンギュラ型玉軸受全体としての寿命の延長を図れる。
【図面の簡単な説明】
【図１】 本発明をなす過程で行った実験の対象となる構造を、一部を切断して示す略図。
【図２】インナー側、アウター側の各玉列の接触角と発熱量との関係を示す三次元グラフ。
【図３】インナー側、アウター側の各玉列の接触角と寿命比との関係を示す三次元グラフ。
【図４】インナー側、アウター側の各玉列の接触角と、アウター側の内輪軌道での玉の肩部への乗り上げ率との関係を示す三次元グラフ。
【図５】車輪用複列アンギュラ型玉軸受の従来構造の１例を示す断面図。
【符号の説明】
１ 外輪
２ ハブ
３ 玉
４ａ、４ｂ 外輪軌道
５ 結合フランジ
６ ハブ本体
７ 内輪
８ スプライン孔
９ 取付フランジ
１０ａ、１０ｂ 内輪軌道
１１ 小径段部
１２ 段差面
１３ａ、１３ｂ シールリング[0001]
BACKGROUND OF THE INVENTION
The double-row angular ball bearing for a wheel of the present invention is used for rotatably supporting a wheel of an automobile with respect to a suspension device.
[0002]
[Prior art]
2. Description of the Related Art Conventionally, for example, a double-row angular ball bearing for a wheel as shown in FIG. 5 is used to rotatably support a vehicle wheel with respect to a suspension device. The double-row angular contact ball bearing shown in FIG. 5 is for supporting driving wheels (front wheels of FF vehicles, rear wheels of FR and RR vehicles, all wheels of 4WD vehicles) An outer ring 1, a hub 2 that is an inner ring equivalent member, and a plurality of balls 3 and 3. Out of these, the outer ring 1 has outwardly angular outer ring raceways 4a, 4b formed on the inner peripheral surface in a double row, and a coupling flange 5 formed on the outer peripheral surface. At the time of assembly to the vehicle, the coupling flange 5 is coupled to a knuckle constituting the suspension device.
[0003]
The hub 2 is formed by combining a hub body 6 and an inner ring 7. Of these, the spline hole 8 is formed at the center of the hub body 6, and the outer end of the outer peripheral surface (“outside” with respect to the axial direction refers to the side that is the outer side in the width direction when assembled to the vehicle. On the contrary, the right side in FIGS. 1 and 5 which is the center in the width direction when assembled to the vehicle is referred to as “inside.” The same applies throughout the present specification.) Forming. At the time of assembly to a vehicle, a spline shaft attached to a constant velocity joint (not shown) is inserted into the spline hole 8 and a wheel is fixed to the mounting flange 9.
[0004]
The outer side of the double-row outer ring raceways 4a and 4b formed on the inner peripheral surface of the outer ring 1 at the intermediate portion of the outer peripheral surface of the hub body 6 (the side that is the outer side in the width direction when assembled to the vehicle) The left side of FIGS. 1 and 5 (the same applies to the whole of the present specification) is formed with an angular inner ring raceway 10b at a portion facing the outer ring raceway 4b. Further, the inner ring 7 is externally fitted to a small diameter step portion 11 formed at the inner end of the hub body 6. Then, the angular inner ring raceway 10a formed on the outer peripheral surface of the inner ring 7 is referred to as the inner side of the double row outer ring raceways 4a, 4b (the side that becomes the center side in the width direction when assembled to the vehicle, The right side of FIGS. 1 and 5 (the same applies throughout the present specification).
[0005]
A plurality of balls 3, 3 are provided between the outer ring raceways 4a, 4b and the inner ring raceways 10a, 10b so that they can roll while being held by a cage (not shown). . With this configuration, a double-row angular ball bearing that is a combination of the rear surfaces is formed, and the hub 2 is supported on the inner diameter side of the outer ring 1 so as to be able to rotate and to support a radial load and an axial load. . In the case of the illustrated structure, with the outer end surface of the inner ring 7 in contact with the stepped surface 12 provided at the base end edge of the small-diameter stepped portion 11, The shape and size of each part are regulated so that predetermined contact angles α 1 , α 2 and a preload are applied to the balls 3, 3 constituting the lens. In addition, lubricating grease is sealed in the space where the plurality of balls 3, 3 are installed between the inner peripheral surface of the outer ring 1 and the outer peripheral surface of the hub 2, and both ends of the space are opened. Are sealed by seal rings 13a and 13b, respectively.
[0006]
Incidentally, including the structure shown in FIG. 5 described above, the conventional case of a double row angular type ball bearing for a wheel, ball row of contact angle alpha 1 and the outer side of the ball row on the inner side of the contact angle alpha 2, They are equal to each other (α 1 = α 2 ). Further, the groove shoulder height H i1 on the load side (the load receiving side on the right side in FIG. 5) of the inner side inner ring raceway 10a and the load side (same on the left side in FIG. 5) of the inner side raceway 10b on the outer side. The groove shoulder heights H i2 are equal to each other (H i1 = H i2 ). Furthermore, groove shoulder height of the load side of the inner side of the outer ring raceway 4a (also, the left side of FIG. 5) and the groove shoulder height H o1 of the load side of the outer side of the outer ring raceway 4b (again, the right side of FIG. 5) H o2 is equal to each other (H o1 = H o2 ).
[0007]
[Problems to be solved by the invention]
The double row angular contact ball bearings for wheels as described above are subject to moment load when the vehicle is turning, etc., but at this time, a larger radial load (axial load) is applied to the inner row of balls than when traveling straight. More preferentially). In this way, when a radial load is preferentially applied to the inner side ball row, the load area of the inner side ball row becomes narrower, and the load applied to the individual balls 3 constituting the inner side ball row becomes large. Become. For this reason, the calculated life based on the load distribution is shorter in the inner side ball row than in the outer side ball row.
[0008]
On the other hand, when the rotational speed and load load of the double row angular contact ball bearing for wheels increase, at the contact portion between the rolling surface of the balls 3 and 3 and the outer ring and inner ring raceways 4a, 4b, 10a and 10b. The amount of heat generated by friction increases. If the amount of heat generated at the contact portion increases in this way, deterioration of the grease for lubrication is promoted or an oil film is less likely to be formed at the contact portion, and the fatigue life of each of the inner and outer ball trains is reduced. descend. In particular, the experimental life based on the durability test according to the actual operating conditions is shorter for the outer side ball row than for the inner side ball row.
[0009]
The double-row angular contact ball bearing for wheels of the present invention is an overall double-row angular contact ball bearing for wheels while balancing the life of the inner and outer ball arrays in view of the circumstances as described above. The invention was invented to realize a structure capable of extending the service life as.
[0010]
[Means for Solving the Problems]
The double-row angular ball bearing for a wheel of the present invention includes an outer ring equivalent member, an inner ring equivalent member, and a plurality of balls.
Of these, the outer ring equivalent member has an angular outer ring raceway provided in double rows on the inner peripheral surface.
Further, the inner ring equivalent member is provided with angular inner ring raceways in double rows on the outer peripheral surface.
Further, a plurality of balls are provided between the inner ring raceways and the outer ring raceways so as to be capable of rolling.
[0011]
In particular, in the double-row angular contact ball bearing for a wheel of the present invention, when assembled to the suspension device, the ball diameter of the inner-side ball row , which is the center side in the width direction of the vehicle, is also the outer side in the width direction. The ball diameter of the outer row of balls is equal.
Also, the contact angle of the inner side ball row is made smaller than the contact angle of the outer side ball row.
[0012]
Also, the load-side groove shoulder height of the outer side inner ring raceway is made larger than the load side groove shoulder height of the inner side inner ring raceway.
[0013]
Furthermore, the load-side groove shoulder height of the outer-side outer ring raceway is made larger than the load-side groove shoulder height of the inner-side outer ring raceway.
[0014]
[Action]
The double row angular contact ball bearing for a wheel of the present invention configured as described above has a smaller contact angle of the inner side ball row than the contact angle of the outer side ball row. Radial load bearing capacity can be increased compared to the outer row of balls. Therefore, a moment load is applied to the double row angular contact ball bearings for wheels when the vehicle is turning, and a large radial load is applied to the inner side ball array (dominantly over the axial load). Regardless of the occurrence of this, the life of the inner side ball train can be extended to the same extent as that of the outer side ball train. Further, by making the contact angle of the inner side ball row smaller than the contact angle of the outer side ball row, it is possible to balance the sharing of the load supported by both the ball rows. For this reason, the amount of heat generated by friction at the contact portion between the rolling surface of each ball and each track can be reduced. As a result, it is possible to effectively prevent deterioration of the grease for lubrication and to prevent the formation of an oil film at the contact portion, and to prolong the service life of each inner row and outer row. it can.
[0015]
In addition, since the contact angle of the inner side ball row is reduced, the axial component force generated when a radial load is applied to the inner side ball row can be reduced. This axial component force is a part of the axial load applied to the outer ball array. For this reason, the axial load applied to this outer side ball train can be reduced. As a result, it is difficult for the balls constituting the outer side ball row to ride on the shoulder portion (width direction edge) of the track, so that the life of the outer side ball row can be effectively prevented from being reduced. That is, when the balls ride on the shoulders of the tracks, the contact surface pressure between the rolling surfaces of the balls and the tracks becomes excessive, and the rolling surfaces of the balls tend to be damaged. The lifetime of the row is likely to decrease. On the other hand, in the case of the present invention , as described above, it is difficult for the balls constituting the outer ball array to ride on the shoulders of the track, so that the life of the outer ball array is reduced. It can be effectively prevented.
[0016]
Furthermore, according to the present invention, it is possible to make it difficult for the balls constituting the outer row of balls to ride on the shoulders of the inner and outer races and to reduce the life of the outer row of balls. Can be prevented. That is, as shown in the embodiment described below, the contact angle of the outer ball array is such that the larger the value is (in the range where each ball does not ride on the shoulder of the track), The amount of heat generated at the contact portion can be reduced, and the calculation life can be extended. For this reason, it is preferable to make the contact angle of the outer side ball row as large as possible (within a range in which each ball does not ride on the shoulder of the track). However, due to the structure of the double row angular contact ball bearing for a wheel, each ball constituting the outer side ball row is more likely to ride on the shoulder of the track than each ball constituting the inner side ball row. Therefore, when the contact angle of the outer ball array is increased as described above, each ball constituting the outer ball array easily gets on the shoulder portion of the track. On the other hand, in the case of the present invention , the load-side groove shoulder height of the inner ring on the outer side and each track on the outer ring is made larger than the groove shoulder height on the load side of each track on the inner side and outer ring. Therefore, it is possible to make it difficult for the balls constituting the outer row of balls to ride on the shoulders of the inner and outer races. Therefore, it is possible to effectively prevent the life of the outer side ball train from being reduced.
[0017]
In addition, if the height of the shoulder on the load side of the inner ring on the inner side and the race on the outer ring is set to be approximately the same as the height of the shoulder on the load side of the inner ring on the outer side and the race on the outer ring, the inner row of balls is configured. Each ball can be made more difficult to get on the shoulder of each track of the inner ring and outer ring. However, if this is done, the polishing area of the inner and outer ring raceways on the inner side will increase and processing costs will increase, and the weight of the double-row angular contact ball bearing for wheels (the unsprung load of the vehicle) will increase. growing. Therefore, in order to prevent such inconveniences, it is preferable to keep the groove shoulder height on the load side of the inner ring on the inner side and the outer ring on the outer ring at a necessary minimum value.
[0018]
【Example】
An experiment (computer simulation) performed in the process of forming the present invention will be described. The experiment was conducted on a double row angular contact ball bearing for a wheel as schematically shown in FIG. The structure and operation of the double-row angular contact ball bearing for wheels schematically shown in FIG. 1 are substantially the same as those of the conventional double-row angular contact ball bearing for wheels shown in FIG. The description which attaches a code | symbol and overlaps is abbreviate | omitted. In the experiment, a double-row angular ball bearing for a wheel like this was used, and a plurality of types with different specifications were prepared, and each was used as a sample for the experiment. The specifications of these samples are shown in Table 1 below.
[Table 1]
Figure 0004225006
[0019]
As shown in Table 1, the ball diameter of the sample, the pitch circle diameter, the number of balls 3, 3 in each row (the number of balls / row), the distance between rows (the pitch between rows), the preload, the outer ring raceway The groove shoulder height on the load side and the groove shoulder height on the load side of the inner ring raceway were equal to each other for each of the samples. At the time of assembly to the suspension device, the ball diameter of the inner side ball row that is the center side in the width direction of the vehicle is made equal to the ball diameter of the outer side ball row that is also the outer side in the width direction. However, the load-side groove shoulder height H o1 of the inner-side outer ring raceway 4a and the load-side groove shoulder height H o2 of the outer-side outer ring raceway 4b are equal to each other (H o1 = H o2 = 4. 7 mm), the load-side groove shoulder height H i2 of the outer-side inner ring raceway 10b is larger than the load-side groove shoulder height H i1 of the inner-side inner ring raceway 10a {H i2 (= 5. 5 mm)> H i1 (= 5.0 mm)}. In contrast, in the each sample, and the contact angle alpha 2 of the ball row of contact angle alpha 1 and the outer side of the ball row on the inner side, it is made different in the ranges of 30 to 45 degrees.
[0020]
For each sample as described above, the calorific value converted from the bearing torque, the L 10 life due to the stress distribution calculated from the life theory of Lundberg-Palmgren, and the ball on the inner ring raceway 10b on the outer side The ride rate was determined by computer simulation. The test conditions are as follows.
Figure 0004225006
In this experiment, the reason for examining the ball climbing rate on the inner ring raceway 10b on the outer side is that the outer raceway is more ball than the inner race because of the structure of the double-row angular ball bearing for wheels. This is because the inner ring track is more likely to ride on the inner ring track than the outer ring track.
[0021]
The results of the computer simulation performed as described above are shown in FIGS. First, FIG. 2 shows the relationship between the contact angle of each ball array on the inner side and the outer side and the heat generation amount converted from the bearing torque. As is apparent from the experimental results shown in FIG. 2, when the case of “contact angle of the inner side ball row = contact angle of the outer side ball row” is taken as a reference, “contact angle of the inner side ball row < In the case of “the contact angle of the outer side ball row”, the heat generation amount is reduced, and in the case of “the contact angle of the inner side ball row> the contact angle of the outer side ball row”, the heat generation amount is increased. Therefore, if “the contact angle of the inner side ball row <the contact angle of the outer side ball row”, the amount of heat generation can be reduced. As a result, the deterioration of the grease for lubrication enclosed in the installation portions of the balls 3 and 3 is promoted, and the rolling surfaces and outer rings of the balls 3 and 3 and the races 4a, 4b, 10a and 10b of the inner rings are accelerated. It is possible to effectively prevent the formation of an oil film at the contact portion between the inner side and the outer side, and to extend the life of each ball array on the inner side and outer side.
[0022]
Next, FIG. 3 shows the relationship between the contact angle of each ball array on the inner side and the outer side and the L 10 life by the stress distribution calculated from the life theory of Lundberg-Palmgren. In FIG. 3, the L 10 life is expressed as a ratio to “1” as the L 10 life when “contact angle of the inner side ball row = contact angle of the outer side ball row = 35 degrees”. It is expressed as “life ratio”. As is apparent from the experimental results shown in FIG. 3, when the case of “contact angle of the inner side ball row = contact angle of the outer side ball row” is taken as a reference, “contact angle of the inner side ball row < In the case of “the contact angle of the outer side ball row”, the calculated life becomes longer, and in the case of “the contact angle of the inner side ball row> the contact angle of the outer side ball row”, the calculated life becomes shorter. Therefore, if “the contact angle of the inner side ball row <the contact angle of the outer side ball row” is satisfied, the calculation life can be extended.
[0023]
Next, FIG. 4 shows the relationship between the contact angle of each ball array on the inner side and the outer side, and the ball riding rate on the inner ring raceway 10b on the outer side. The ball climbing rate indicates the ratio of the length of the contact ellipse calculated from Hertz's contact theory protruding from the edge in the width direction of the track (the inner ring track 10b). When the length of the contact ellipse protruding from the edge in the width direction of the track is ζ, it is expressed as (ζ / 2a) × 100 [%]. The ball climbing rate indicates that when the sign is “positive”, the contact ellipse protrudes from the edge in the width direction of the inner ring raceway 10b, and when the sign is “negative”, This indicates that the contact ellipse does not protrude from the edge in the width direction of the inner ring raceway 10b (there is a margin for protruding).
[0024]
In the experimental results shown in FIG. 4, since the sign of the ball climbing rate for all the samples is negative, it can be seen that all these samples have a margin for the ball climbing. The reason why such an experimental result was obtained is that the load-side groove shoulder height H i2 of the inner ring raceway 10b on the outer side is larger than the groove shoulder height H i1 on the load side of the inner ring raceway 10a on the inner side ( As a result of H i2 > H i1 ), it is difficult for the balls 3 and 3 to get on the shoulder of the inner ring raceway 10b on the outer side. The load-side groove shoulder height H i1 of the inner-side inner ring raceway 10a is also made as large as the load-side groove shoulder height H i2 of the outer-side inner ring raceway 10b (H i2 > H i1 ). Further, the ball climbing rate on the inner ring raceway 10a can be further reduced. However, if this is done, the polishing area of the surface of the inner ring raceway 10a on the inner side is increased, the processing cost is increased, and the weight of the double-row angular contact ball bearing for wheels (the unsprung load of the vehicle) is increased. . Therefore, in the case of this example, in order to prevent such inconvenience, the load-side groove shoulder height H i1 of the inner-side inner raceway 10a is suppressed to a necessary minimum value.
[0025]
Further, the experimental results shown in FIG. 4 showed that the sign of the ball climbing rate for all the samples was negative as described above. However, in actual operation, by such tire rides on the curb, axial load applied to the double row angular type ball bearing wheels, can be greater than the axial load F a shown in the test conditions. In such a case, there is a possibility that the amount of elastic deformation of each member constituting the double-row angular contact ball bearing for wheels becomes larger than expected, and the sign of the above-mentioned ball riding rate may change from negative to positive. . Therefore, in order to prevent such inconvenience, the specifications of the double-row angular contact ball bearing for wheels are based on the experimental results shown in FIG. 4 and the ball climbing rate is -25% or less (the absolute value is 25 % Or more) is preferable. That is, as shown in FIG. 4, when the contact angle of the outer ball array is increased to 45 degrees, the ball climbing rate is -25% or more (absolute value is 25% or less), and the above-mentioned conditions are satisfied. In order to eliminate this, it is preferable to set the contact angle of the outer ball array to 40 degrees or more and less than 45 degrees. In addition, based on the conclusion based on the experimental results shown in FIGS. 2 to 3, the contact angle of the inner ball array is set to be smaller than the contact angle of the outer ball array (40 degrees or more and less than 45 degrees). It is preferable to do this.
[0026]
Further, based on the result of the computer simulation as described above, the contact angle of the inner side ball row of the double row angular contact ball bearing for wheels is set to 33 ± 2 degrees, and the contact angle of the outer side ball row is set to 40 ± 2. When actual operation was performed as a result, the experimental life of this double-row angular contact ball bearing for wheels was extended to 2 to 5 times the experimental life of the conventional structure. From the above points, when carrying out the present invention, the contact angle α 1 of the inner side ball row is set to 30 to 40 degrees, and the contact angle α 2 of the outer side ball row is set to a range of 35 to 45 °, In addition, it is preferable that the contact angle α 2 of the outer side ball row is larger by 5 degrees or more (α 2 −α 1 ≧ 5 °) than the contact angle α 1 of the inner side ball row.
[0027]
In the above structure, the groove shoulder height H o1 on the load side of the outer ring raceway 4a on the inner side of each sample, the load side of the outer side of the outer ring raceway 4b and the groove shoulder height H o2, and equal to each other . However, as in the present invention, greater than the groove shoulder height H o1 the load side of the outer side of the groove shoulder height H o2 on the load side of the outer ring raceway 4b on the inner side outer ring raceway 4a (H o2> H o1 ), The ball climbing rate on the outer-side outer ring raceway 4b can be made as small as the ball climbing rate on the inner-side outer ring raceway 4a.
[0028]
The double-row angular ball bearing for a wheel according to the present invention is not limited to that for driving wheels as shown in FIGS. Moreover, it can be applied to double-row angular ball bearings for wheels of various structures for both driving wheels and driven wheels.
[0029]
【The invention's effect】
The double-row angular contact ball bearing for a wheel of the present invention is configured and operates as described above, so that the double-row angular contact for the wheel is achieved while balancing the life of the inner and outer ball arrays. The life of the entire type ball bearing can be extended.
[Brief description of the drawings]
FIG. 1 is a schematic view showing a part of a structure which is an object of an experiment conducted in the process of forming the present invention.
FIG. 2 is a three-dimensional graph showing the relationship between the contact angle of each ball array on the inner side and outer side and the amount of heat generated.
FIG. 3 is a three-dimensional graph showing the relationship between the contact angle of each ball array on the inner side and the outer side and the life ratio.
FIG. 4 is a three-dimensional graph showing the relationship between the contact angle of each ball array on the inner side and the outer side, and the riding rate on the shoulder of the ball on the inner ring raceway on the outer side.
FIG. 5 is a sectional view showing an example of a conventional structure of a double-row angular contact ball bearing for a wheel.
[Explanation of symbols]
DESCRIPTION OF SYMBOLS 1 Outer ring 2 Hub 3 Ball 4a, 4b Outer ring raceway 5 Coupling flange 6 Hub main body 7 Inner ring 8 Spline hole 9 Mounting flange 10a, 10b Inner ring raceway 11 Small diameter step 12 Step surface 13a, 13b Seal ring

Claims (1)

内周面にアンギュラ型の外輪軌道を複列に設けた外輪相当部材と、外周面にアンギュラ型の内輪軌道を複列に設けた内輪相当部材と、これら各内輪軌道と上記各外輪軌道との間に、それぞれ複数個ずつ転動自在に設けられた玉とを備え、上記外輪相当部材と上記内輪相当部材とのうちの一方の部材を懸架装置に支持し、同じく他方の部材に車輪を支持する状態で使用する車輪用複列アンギュラ型玉軸受に於いて、上記懸架装置への組み付け時、車両の幅方向中央側となるインナー側の玉列の玉径と同じく幅方向外側となるアウター側の玉列の玉径とが等しく、インナー側の玉列の接触角をアウター側の玉列の接触角よりも小さくしており、アウター側の内輪軌道の負荷側の溝肩高さをインナー側の内輪軌道の負荷側の溝肩高さよりも、アウター側の外輪軌道の負荷側の溝肩高さをインナー側の外輪軌道の負荷側の溝肩高さよりも、それぞれ大きくした事を特徴とする車輪用複列アンギュラ型玉軸受。An outer ring equivalent member in which an angular outer ring raceway is provided in a double row on the inner peripheral surface, an inner ring equivalent member in which an angular type inner ring raceway is provided in a double row on the outer peripheral surface, and the inner ring raceway and the outer ring raceways. In between, a plurality of balls are provided so as to freely roll, and one member of the outer ring equivalent member and the inner ring equivalent member is supported by the suspension device, and the wheel is also supported by the other member. In a double-row angular contact ball bearing for a wheel to be used in the state of being used, the outer side which is the outer side in the width direction is the same as the ball diameter of the inner side ball row which is the center side in the width direction of the vehicle when assembled to the suspension device The contact diameter of the inner ball train is smaller than the contact angle of the outer ball train, and the groove shoulder height on the load side of the inner ring raceway on the outer side is the inner side. Than the groove shoulder height on the load side of the inner ring raceway Than groove shoulder height of the load side of the outer ring raceway of the inner side groove shoulder height of the load side of the outer ring raceway side, wheel double row angular type ball bearing, characterized in that the larger, respectively.
JP2002240319A 2002-08-21 2002-08-21 Double row angular contact ball bearings for wheels Expired - Fee Related JP4225006B2 (en)

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JP2011196888A (en) * 2010-03-23 2011-10-06 Nsk Ltd Load measuring apparatus for wheel of automobile
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