JP2003314542A - Tapered roller bearing - Google Patents

Tapered roller bearing

Info

Publication number
JP2003314542A
JP2003314542A JP2002113784A JP2002113784A JP2003314542A JP 2003314542 A JP2003314542 A JP 2003314542A JP 2002113784 A JP2002113784 A JP 2002113784A JP 2002113784 A JP2002113784 A JP 2002113784A JP 2003314542 A JP2003314542 A JP 2003314542A
Authority
JP
Japan
Prior art keywords
tapered roller
tapered
roller bearing
diameter side
large diameter
Prior art date
Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
Pending
Application number
JP2002113784A
Other languages
Japanese (ja)
Other versions
JP2003314542A5 (en
Inventor
Hiromichi Takemura
浩道 武村
Yoshitaka Hayashi
善貴 林
Current Assignee (The listed assignees may be inaccurate. Google has not performed a legal analysis and makes no representation or warranty as to the accuracy of the list.)
NSK Ltd
Original Assignee
NSK Ltd
Priority date (The priority date is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the date listed.)
Filing date
Publication date
Application filed by NSK Ltd filed Critical NSK Ltd
Priority to JP2002113784A priority Critical patent/JP2003314542A/en
Publication of JP2003314542A publication Critical patent/JP2003314542A/en
Publication of JP2003314542A5 publication Critical patent/JP2003314542A5/ja
Pending legal-status Critical Current

Links

Classifications

    • 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
    • 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
    • F16C2240/50—Crowning, e.g. crowning height or crowning radius
    • 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

Landscapes

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

Abstract

<P>PROBLEM TO BE SOLVED: To cope with both the assurance of a bearing life and a lower torque at a high grade. <P>SOLUTION: When a contact angle of a tapered roller bearing 6a is set to be α, a half of a tapered angle of a tapered roller 9a is set to be β, and the number of the tapered roller 9a is set to be z, dimensions of respective portions are set to satisfy 15°<α<22° and 2.0≤(β×z)/α≤3.0. <P>COPYRIGHT: (C)2004,JPO

Description

【発明の詳細な説明】Detailed Description of the Invention

【０００１】[0001]

【発明の属する技術分野】この発明は、例えば自動車の
デファレンシャルギヤ（最終減速機）を構成するピニオ
ン軸や、トランスミッションを構成するギヤ等を、ケー
シング（デフケース又はミッションケース）の内側に回
転自在に支持する為の円すいころ軸受の改良に関する。BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention rotatably supports, for example, a pinion shaft constituting a differential gear (final reduction gear) of an automobile, a gear constituting a transmission, etc. inside a casing (a differential case or a transmission case). To improve the tapered roller bearings.

【０００２】[0002]

【従来の技術】自動車の動力伝達系の途中に設けてプロ
ペラシャフトの回転を減速すると同時に回転方向を直角
に変換するデファレンシャルギヤは、図１に示す様に構
成している。ケーシング１の内側前寄り（図１の右寄
り）部分に、ピニオン軸２を配設している。又、このピ
ニオン軸２の前端部（図１の右端部）で上記ケーシング
１の前端開口部から突出した部分に固設した結合フラン
ジ３に、図示しないプロペラシャフトの後端部を連結自
在としている。又、上記ピニオン軸２の後端部（図１の
左端部）に減速小歯車４を固定し、この減速小歯車４と
減速大歯車５とを互いに噛合させている。この減速大歯
車５は、上記ケーシング１の後部（図１の左部）内側
に、回転のみ自在に支持している。又、上記ピニオン軸
２の中間部前後２個所位置を、前後１対の円すいころ軸
受６ａ、６ｂにより、上記ケーシング１に対し回転自在
に支持している。2. Description of the Related Art A differential gear, which is provided in the middle of a power transmission system of an automobile, reduces the rotation of a propeller shaft and at the same time changes the rotation direction to a right angle, is constructed as shown in FIG. A pinion shaft 2 is arranged in a portion of the casing 1 on the inner front side (on the right side in FIG. 1). Further, a rear end of a propeller shaft (not shown) can be connected to a coupling flange 3 fixedly provided at a front end of the pinion shaft 2 (right end in FIG. 1) protruding from a front end opening of the casing 1. . A reduction gear 4 is fixed to the rear end of the pinion shaft 2 (the left end in FIG. 1), and the reduction gear 4 and the reduction gear 5 are meshed with each other. The reduction gear wheel 5 is rotatably supported inside the rear portion (left portion in FIG. 1) of the casing 1. Further, two positions in the front and rear of the intermediate portion of the pinion shaft 2 are rotatably supported with respect to the casing 1 by a pair of front and rear tapered roller bearings 6a and 6b.

【０００３】上記各円すいころ軸受６ａ、６ｂは、図２
に詳示する様に、それぞれ１個ずつの外輪７ａ、７ｂ及
び内輪８ａ、８ｂと、それぞれ複数個ずつの円すいころ
９ａ、９ｂとから構成している。外輪７ａ、７ｂの内周
面には円すい凹面状の外輪軌道１０ａ、１０ｂを、内輪
８ａ、８ｂの中間部外周面には円すい凸面状の内輪軌道
１１ａ、１１ｂを、それぞれ形成している。又、これら
各内輪８ａ、８ｂの外周面で大径側端部に外向フランジ
状の大径側鍔部２１を、同じく小径側端部にやはり外向
フランジ状の小径側鍔部２２を、それぞれ形成してい
る。そして、上記各円すいころ９ａ、９ｂは、上記各外
輪軌道１０ａ、１０ｂと各内輪軌道１１ａ、１１ｂとの
間に、図２にのみ示した保持器２０により複数ずつ保持
した状態で、転動自在に設けている。そして、上記外輪
７ａ、７ｂを前記ケーシング１の一部に内嵌固定し、上
記内輪８ａ、８ｂを前記ピニオン軸２の中間部前後２個
所位置に外嵌固定している。尚、一般的に、上記内、外
輪各軌道１１ａ、１１ｂ、１０ａ、１０ｂに接する接線
ｌ1 、ｌ2 と、各円すいころ９ａ、９ｂの中心軸の延長
線ｌ3 とは、内、外両輪８ａ、８ｂ、９ａ、９ｂの中心
軸の延長線上の１点で交わらせている。The tapered roller bearings 6a and 6b are shown in FIG.
As will be described in detail, each of them is composed of one outer ring 7a, 7b and one inner ring 8a, 8b, and a plurality of tapered rollers 9a, 9b. Conical concave outer ring raceways 10a, 10b are formed on the inner peripheral surfaces of the outer rings 7a, 7b, and conical convex inner ring raceways 11a, 11b are formed on the intermediate outer peripheral surfaces of the inner rings 8a, 8b. Further, on the outer peripheral surface of each of the inner rings 8a and 8b, an outward flange-shaped large-diameter side flange portion 21 is formed at the large-diameter side end portion, and an outward flange-shaped small-diameter side collar portion 22 is also formed at the small-diameter side end portion. is doing. Each of the tapered rollers 9a and 9b is freely rollable while being held between the outer ring raceways 10a and 10b and the inner ring raceways 11a and 11b by a retainer 20 shown only in FIG. It is provided in. Then, the outer rings 7a and 7b are internally fitted and fixed to a part of the casing 1, and the inner rings 8a and 8b are externally fitted and fixed at two positions before and after an intermediate portion of the pinion shaft 2. Incidentally, in general, the tangent lines l 1 and l 2 contacting the inner and outer races 11a, 11b, 10a and 10b and the extension line l 3 of the central axis of each of the tapered rollers 9a and 9b are the inner and outer rings. 8a, 8b, 9a, 9b intersect at one point on the extension line of the central axis.

【０００４】又、上記各円すいころ軸受６ａ、６ｂは、
一般的に、上記外輪軌道１０ａ、１０ｂの接線と上記
内、外両輪８ａ、８ｂ、９ａ、９ｂの中心軸とのなす角
度である接触角αの大きさに応じて、３つの型式に分け
られる。即ち、この接触角αが１５度未満である（α＜
１５°）場合には並勾配型に、同じく１５度以上で２
２．５度以下である（１５°≦α≦２２．５°）場合に
は中勾配型に、同じく２２．５度よりも大きい（α＞２
２．５°）場合には急勾配型に、それぞれ分けられる。
そして、上述の図１に示した様に、ケーシング１の内側
にピニオン軸２を、１対の円すいころ軸受６ａ、６ｂに
より片持ちで支持する場合には、前側（図１の右側）の
円すいころ軸受６ａに並勾配型を、後側（図１の左側）
の円すいころ軸受６ｂに急勾配型を、それぞれ使用する
事が軸受寿命及び軸受剛性を確保する為に有効である事
が、従来から知られている。The tapered roller bearings 6a and 6b are
Generally, there are three types according to the magnitude of the contact angle α, which is the angle formed by the tangent line of the outer ring raceways 10a, 10b and the center axis of the inner and outer wheels 8a, 8b, 9a, 9b. . That is, the contact angle α is less than 15 degrees (α <
15 °), in the case of parallel gradient type, if it is 15 ° or more, 2
When it is less than 2.5 degrees (15 ° ≦ α ≦ 22.5 °), it is a medium-gradient type and also larger than 22.5 degrees (α> 2).
2.5 °), it is divided into steep type.
As shown in FIG. 1 described above, when the pinion shaft 2 is cantilevered by the pair of tapered roller bearings 6a and 6b inside the casing 1, the front side (the right side in FIG. 1) cone is used. The roller bearing 6a is of a parallel gradient type, and the rear side (left side in FIG. 1)
It has been conventionally known that it is effective to use a steep grade for each of the tapered roller bearings 6b in order to secure the bearing life and the bearing rigidity.

【０００５】又、円すいころの数を（２〜３本）多くす
ると共に、円すいころの寸法を大きくして、高荷重が加
わる状態で使用するのに適した構造としたＨＲシリーズ
の円すいころ軸受は、円すいころ軸受の基本動定格荷重
を、一般的な構造に対し１．２０〜１．２５倍程度大き
くできて、軸受寿命の向上を図れる事も、従来から知ら
れている。Further, the number of tapered rollers is increased (2 to 3) and the dimensions of the tapered rollers are increased so that the tapered roller bearings of the HR series have a structure suitable for use under a high load. It has been known from the past that the basic dynamic load rating of a tapered roller bearing can be increased by about 1.20 to 1.25 times that of a general structure and the life of the bearing can be improved.

【０００６】[0006]

【発明が解決しようとする課題】近年、自動車の省燃費
化に対する要求が強くなっており、この為に、自動車の
回転支持部に組み込む円すいころ軸受に関して、小型化
及び軽量化を図ると共に、起動トルク及び動トルク（回
転抵抗）の低減を図る事が望まれている。特に、転がり
軸受のうちで円すいころ軸受の起動トルク及び動トルク
は、玉軸受の場合よりも一般的に大きい。この為、自動
車の省燃費化を図る為に、円すいころ軸受のトルクの低
減を図る事が重要であると考えられている。又、このト
ルクの低減を図る場合に、円すいころ軸受の軸受寿命を
確保する事が重要である事は勿論である。In recent years, there has been a strong demand for reducing fuel consumption of automobiles. For this reason, it is possible to reduce the size and weight of a tapered roller bearing to be incorporated in a rotation supporting portion of an automobile and to start it. It is desired to reduce torque and dynamic torque (rotational resistance). In particular, among the rolling bearings, the starting torque and the dynamic torque of the tapered roller bearing are generally larger than those of the ball bearing. Therefore, it is considered important to reduce the torque of the tapered roller bearing in order to save fuel consumption of the automobile. In addition, it is of course important to ensure the bearing life of the tapered roller bearing when reducing the torque.

【０００７】この様な円すいころ軸受のトルクを大きく
する原因として、次のに示す摩擦が存在する。 各円すいころ９ａ、９ｂの大径側端面と内輪８ａ、
８ｂの大径側端部外周面に設けた大径側鍔部２１（図
２）の片面との接触部に作用する滑り摩擦ＭS 各円すいころ９ａ、９ｂの転動面と内、外輪各軌道
１１ａ、１１ｂ、１０ａ、１０ｂとの接触部に作用する
転がり摩擦Ｍr 従って、上記円すい転がり軸受のトルクを低減する為に
は、これら滑り摩擦Ｍ S と転がり摩擦Ｍr との低減を図
る必要がある。尚、これら滑り摩擦ＭS 及び転がり摩擦
Ｍr の大きさは、それぞれ次の２式に従う。 ＭS ＝ｅ×μ×cos β×Ｆa −−−（１） Ｍr ∝（ｄm ×Ｍe ×ｚ）／Ｄa −−−（２） これら２式中、ｅは各円すいころ９ａ、９ｂの大径側端
面と大径側鍔部２１の片面との接触部の高さ（内輪軌道
１１ａ、１１ｂの母線の延長線から接触部中心までの距
離）を、μはこの接触部での摩擦係数を、βは各円すい
ころ９ａ、９ｂの円すい角度の１／２（円すいころ９
ａ、９ｂの母線と中心軸とのなす角度）を、Ｆa はアキ
シアル荷重を、それぞれ表している。又、ｚは円すいこ
ろ９ａ、９ｂの数を、Ｄa は各円すいころ９ａ、９ｂの
平均直径を、ｄm は複数の円すいころ９ａ、９ｂの大径
側端面でのピッチ円の直径を、Ｍe は各円すいころ９
ａ、９ｂの転動面と内、外輪各軌道１１ａ、１１ｂ、１
０ａ、１０ｂとの間に作用する転がり抵抗を、それぞれ
表している。Increase the torque of such tapered roller bearings
The following friction exists as a cause.   The large diameter side end surface of each tapered roller 9a, 9b and the inner ring 8a,
The large-diameter side flange portion 21 (Fig.
2) Sliding friction M acting on the contact surface with one sideS   Each tapered roller 9a, 9b rolling surface and inner and outer ring raceways
Acts on contact parts with 11a, 11b, 10a, 10b
Rolling friction Mr Therefore, in order to reduce the torque of the above-mentioned tapered roller bearing,
Is the sliding friction M S And rolling friction Mr To reduce
Need to Incidentally, these sliding friction MS And rolling friction
Mr The size of each follows the following two equations. MS = E × μ × cos β × Fa           --- (1) Mr ∝ (dm × Me Xz) / Da       --- (2) In these two equations, e is the large diameter side end of each tapered roller 9a, 9b.
Height of the contact surface between one surface of the large-diameter side flange 21 (inner ring raceway)
Distance from the extension of the busbars of 11a and 11b to the center of the contact part
), Μ is the coefficient of friction at this contact point, β is each cone
1/2 of the cone angle of rollers 9a and 9b (cone roller 9
The angle between the generatrix of a and 9b and the central axis is Fa Is Aki
Each sial load is shown. Also, z is a cone
The number of filters 9a and 9b is set to Da Is for each tapered roller 9a, 9b
The average diameter is dm Is the large diameter of multiple tapered rollers 9a, 9b
Set the diameter of the pitch circle at the side end face to Me Is each tapered roller 9
a, 9b rolling surfaces and inner and outer races 11a, 11b, 1
The rolling resistance acting between 0a and 10b is
It represents.

【０００８】一方、前述した様に、高荷重が加わる状態
で使用するのに適した、円すいころの数を多くすると共
に、各円すいころの寸法を大きくしたＨＲシリーズの円
すいころ軸受は、軸受寿命の向上を図れる事が知られて
いる。但し、この様なＨＲシリーズの構造を採用した場
合には、これら各円すいころの数が増える事により転が
り摩擦Ｍr が大きくなり、円すいころ軸受のトルクが増
大する。尚、各円すいころの平均直径の増大によるトル
ク低減は、この数増大によるトルク増大に比べて、限ら
れたものである。この為、軸受寿命の向上を図れても、
トルクの低減を図る事は難しい。On the other hand, as described above, the HR series tapered roller bearing, which has a large number of tapered rollers and is large in size, is suitable for use under a high load condition. It is known that the improvement of However, in the case of adopting the structure of such a HR series, friction M r rolling by these number of each tapered roller is increased becomes larger, the torque of the tapered roller bearing is increased. The torque reduction due to the increase in the average diameter of each tapered roller is limited as compared with the torque increase due to the increase in the number. Therefore, even if the bearing life is improved,
It is difficult to reduce the torque.

【０００９】又、前述の図１〜２に示した様な、自動車
のデファレンシャルギヤのピニオン軸２の支持部や、ト
ランスミッションの回転支持部に使用する円すいころ軸
受では、ギヤ同士の噛合により生じた鉄粉等の異物の侵
入に基づく寿命低下も考慮する必要がある。そして、円
すいころ軸受の各部の寸法（軸方向長さ）が限定されて
いる場合には、基本動定格荷重を大きくして寿命確保を
図るべく、接触角αが小さい、並勾配型や中間勾配型を
使用する必要が生じる。但し、この様に接触角αを小さ
くする事により寿命確保を図る従来技術では、トルクの
低減を図る事に関しては、未だ改良の余地がある。Further, in the tapered roller bearing used for the support portion of the pinion shaft 2 of the differential gear of an automobile and the rotation support portion of the transmission as shown in FIGS. It is also necessary to consider the shortening of service life due to the intrusion of foreign matter such as iron powder. When the dimensions (axial length) of each part of the tapered roller bearing are limited, the contact angle α is small, the parallel gradient type or the intermediate gradient type is used in order to increase the basic dynamic load rating and secure the service life. The need to use types arises. However, there is still room for improvement in the reduction of torque in the conventional technology that secures the life by reducing the contact angle α in this way.

【００１０】これに対して、特開平９−９６３５２号公
報には、デファレンシャルギヤのピニオン軸を支持する
為の円すいころ軸受で、接触角αと、各円すいころの端
面の直径及び長さの比と、ころ数係数ｋとを規制した円
すいころ軸受が記載されている。この公報に記載された
円すいころ軸受によれば、軸受寿命の確保と、トルクの
低減とを、高次元で両立できる。尚、上記ころ数係数ｋ
とは、各円すいころの大径側端部の直径をＤa ´とし、
複数の円すいころの大径側端部でのピッチ円の直径をｄ
m とし、円すいころの数をｚとした場合に、ｋ＝（ｄm
／Ｄa ´）・sin （１８０°／ｚ）で表されるものであ
る。但し、上記公報に記載された円すいころ軸受の場
合、接触角αが２２〜２８度と大きくなっている。この
様に接触角αが大きい場合には、基本動定格荷重が小さ
くなり易く、円すいころ軸受の寿命確保の面からは不利
である。そこで、本発明者は、上記公報に記載された円
すいころ軸受に更に改良を加え、上記接触角αを小さく
すると共に、他の諸元を適切に規制する事により、トル
クの低減と軸受寿命の向上とを、更に高次元で両立でき
る円すいころ軸受を発明した。On the other hand, Japanese Unexamined Patent Publication No. 9-96352 discloses a tapered roller bearing for supporting a pinion shaft of a differential gear, in which the contact angle α is the ratio of the diameter and length of the end face of each tapered roller. And a tapered roller bearing in which the roller number coefficient k is regulated. According to the tapered roller bearing described in this publication, ensuring the bearing life and reducing the torque can be achieved at a high level. In addition, the roller number coefficient k
And the diameter of the large diameter side end of each tapered roller is D a ′,
D is the diameter of the pitch circle at the large diameter end of multiple tapered rollers.
When m and the number of tapered rollers is z, k = (d m
/ D a ′) · sin (180 ° / z). However, in the case of the tapered roller bearing described in the above publication, the contact angle α is as large as 22 to 28 degrees. When the contact angle α is large in this way, the basic dynamic load rating tends to be small, which is disadvantageous in terms of ensuring the life of the tapered roller bearing. Therefore, the present inventor has further improved the tapered roller bearing described in the above publication to reduce the contact angle α and appropriately regulate other specifications to reduce the torque and the bearing life. We have invented a tapered roller bearing that can achieve both improvement and higher dimension.

【００１１】[0011]

【課題を解決するための手段】本発明の円すいころ軸受
は、従来から知られている円すいころ軸受と同様に、内
周面に円すい凹面状の外輪軌道を有する外輪と、外周面
のうちで中間部に円すい凸面状の内輪軌道を、大径側端
部に大径側鍔部を、小径側端部に小径鍔部を、それぞれ
有する内輪と、上記外輪軌道と内輪軌道との間に転動自
在に設けられた複数個の円すいころとを備える。A tapered roller bearing according to the present invention has an outer ring having an outer ring raceway having a conical concave surface on its inner peripheral surface and an outer peripheral surface, similarly to a conventionally known tapered roller bearing. The inner ring raceway has a conical convex inner ring raceway in the middle section, the large diameter side collar section at the large diameter side end, and the small diameter side collar section at the small diameter side end, and the inner ring raceway between the outer ring raceway and the inner ring raceway. And a plurality of tapered rollers movably provided.

【００１２】特に、本発明の円すいころ軸受に於いて
は、接触角（外輪軌道の接線と内、外両輪の中心軸との
なす角度）をαとし、これら各円すいころの円すい角度
の１／２（円すいころの母線と中心軸とのなす角度）を
βとし、円すいころの数をｚとし、軸方向長さをＬと
し、平均直径をＤa とし、クラウニング量をδとした場
合に、１５°＜α＜２２°であり、且つ、２．０≦（β
×ｚ）／α≦３．０を満たし、且つ、これら各円すいこ
ろの軸方向中央から軸方向両端側に０．４Ｌ外れた位置
で、０．１％≦δ／Ｄa ≦０．５％を満たす。In particular, in the tapered roller bearing of the present invention, the contact angle (the angle between the tangent line of the outer ring raceway and the central axes of the inner and outer rings) is α, and 1 / the taper angle of each of these tapered rollers. 2 (angle between the generatrix of the tapered roller and the central axis) is β, the number of tapered rollers is z, the axial length is L, the average diameter is D a , and the crowning amount is δ, 15 ° <α <22 ° and 2.0 ≦ (β
Xz) /α≦3.0, and 0.1% ≦ δ / D a ≦ 0.5% at a position deviated from the axial center of each of these tapered rollers by 0.4 L to both axial ends. Meet

【００１３】又、請求項２に記載した円すいころ軸受の
場合、上記各円すいころに関して、平均直径をＤa と
し、軸方向長さをＬとし、複数の円すいころの大径側端
面でのピッチ円の直径をｄm とし、これら複数の円すい
ころと外輪と内輪とを組み立てた状態での幅をＷとした
場合に、３．０≦（ｄm ×Ｌ）／（Ｄa ×Ｗ）≦６．０
を満たす。Further, in the tapered roller bearing according to the second aspect of the present invention, with respect to each of the tapered rollers, the average diameter is D a , the axial length is L, and the pitches of the plurality of tapered rollers on the large diameter side end surface are large. When the diameter of the circle is d m and the width of the plurality of tapered rollers, the outer ring and the inner ring assembled together is W, 3.0 ≦ (d m × L) / (D a × W) ≦ 6.0
Meet

【００１４】又、請求項３に記載した円すいころ軸受の
場合、各円すいころの転動面の粗さをＲ1 とし、これら
各円すいころの大径側端面の粗さをＲ2 とし、内、外輪
各軌道の粗さをｒ1 とし、内輪の大径側端部外周面に設
けた大径側鍔部の片面で、上記各円すいころの大径側端
面と対向する面の粗さをｒ2 とした場合に、Ｒ2 及びｒ
2 が０．１５μｍＲａ以下であり、Ｒ2 ／Ｒ1 ＜１で、
且つ、ｒ2 ／ｒ1 ＜１であり、２．０≦（Ｒ1 ×ｒ1 ）
／（Ｒ2 ×ｒ2 ）≦５．０を満たす。Further, in the case of the tapered roller bearing according to claim 3, the roughness of the rolling surface of each tapered roller is R 1, and the roughness of the large diameter side end surface of each of these tapered rollers is R 2 , , R 1 is the roughness of each outer ring raceway, and the roughness of the surface facing the large diameter end face of each of the above tapered rollers is one side of the large diameter side flange portion provided on the large diameter side end outer peripheral surface of the inner ring. in the case of the r 2, R 2 and r
2 is 0.15 μmRa or less, R 2 / R 1 <1, and
In addition, r 2 / r 1 <1, and 2.0 ≦ (R 1 × r 1 ).
/ (R 2 × r 2 ) ≦ 5.0 is satisfied.

【００１５】[0015]

【作用】上述の様に構成する本発明の円すいころ軸受の
場合には、接触角αを２２度未満と小さくしている為、
基本動定格荷重を大きくできる。又、この様に接触角α
を小さくした場合には、円すいころ軸受の諸元を決定付
ける他の要素との関係で、トルク（動トルク）を小さく
した円すいころ軸受を実現する事が難しくなる。これに
対して、本発明の場合には、２．０≦（β×ｚ）／α≦
３．０を満たす様にしている為、上述の様に接触角αを
小さくしても、各円すいころの大径側端面と内輪の大径
側端部外周面に設けた大径側鍔部の片面との接触部に作
用する滑り摩擦と、これら各円すいころの転動面と内、
外輪各軌道との間に作用する転がり摩擦とを小さく抑え
る事ができる。この為、本発明によれば、トルクの低減
と軸受寿命の向上とを、高次元で両立できる。従って、
本発明の円すいころ軸受を組み込んだ自動車の省燃費化
と耐久性の向上とに寄与できる。尚、上記接触角αを１
５度以下とすれば、基本動定格荷重を更に大きくできる
が、アキシアル方向に関する許容荷重が不足する（アキ
シアル荷重に関する定格荷重が過小になる）。この為、
本発明では、上記接触角αを１５度よりも大きくした。In the tapered roller bearing of the present invention constructed as described above, the contact angle α is made smaller than 22 degrees,
The basic dynamic load rating can be increased. In addition, the contact angle α
When is smaller, it is difficult to realize a tapered roller bearing having a small torque (dynamic torque) because of the relationship with other factors that determine the specifications of the tapered roller bearing. On the other hand, in the case of the present invention, 2.0 ≦ (β × z) / α ≦
Since 3.0 is satisfied, even if the contact angle α is reduced as described above, the large-diameter side flange portion provided on the large-diameter side end surface of each tapered roller and the large-diameter side end outer peripheral surface of the inner ring Of sliding friction acting on the contact part with one side of, and the rolling surface of each of these tapered rollers,
Rolling friction acting between each outer ring raceway can be suppressed to be small. Therefore, according to the present invention, both reduction of torque and improvement of bearing life can be achieved at a high level. Therefore,
It can contribute to the fuel saving and the improvement of the durability of the automobile incorporating the tapered roller bearing of the present invention. The contact angle α is 1
If it is 5 degrees or less, the basic dynamic load rating can be further increased, but the allowable load in the axial direction will be insufficient (the rated load related to the axial load will be too small). Therefore,
In the present invention, the contact angle α is larger than 15 degrees.

【００１６】更に、本発明の円すいころ軸受によれば、
各円すいころのクラウニング量を最適に規制している
為、トルクの低減と、軸受寿命の確保とを、より高次元
で両立できる。これに対して、上記各円すいころの軸方
向中央から軸方向両端側に０．４Ｌ外れた位置での比δ
／Ｄa を、本発明の下限値よりも小さく（δ／Ｄ＜０．
１％）した場合には、クラウニング量が過小になる為、
クラウニングを設ける事によるトルクの低減を図れなく
なる。又、この場合には、各円すいころの転動面から
内、外輪各軌道にエッヂロードが加わり易くなり、加わ
った場合には、表面起点型の剥離を生じ易くなる。逆
に、上記比δ／Ｄa が、本発明の上限値よりも大きくな
った（δ／Ｄ＞０．５％）場合には、各円すいころの転
動面の母線の曲率が過大になる事で、この転動面と内、
外輪各軌道との接触面積が狭くなり、この接触部の面圧
が大きくなる為、内部起点型の剥離を生じ易くなる。本
発明の場合には、０．１≦δ／Ｄ≦０．５％を満たして
いる為、この様な不都合をなくして、トルクの低減と、
軸受寿命の確保とを、更に高次元で両立できる。Further, according to the tapered roller bearing of the present invention,
Since the amount of crowning of each tapered roller is optimally regulated, reduction of torque and securing of bearing life can be achieved at a higher level. On the other hand, the ratio δ at the position deviated from the axial center of each of the above tapered rollers by 0.4 L to both axial ends.
/ D a smaller than the lower limit of the present invention (δ / D <0.
1%), the crowning amount will be too small.
It becomes impossible to reduce the torque by providing the crowning. Further, in this case, the edge roads are likely to be added to the inner and outer races from the rolling surface of each tapered roller, and when they are added, the surface-origin type separation is likely to occur. On the contrary, when the ratio δ / D a becomes larger than the upper limit of the present invention (δ / D> 0.5%), the curvature of the generatrix of the rolling surface of each tapered roller becomes excessive. By the way, this rolling surface and inside,
Since the contact area with each race of the outer ring is narrowed and the surface pressure of this contact portion is increased, internal-origin-type delamination is likely to occur. In the case of the present invention, since 0.1 ≦ δ / D ≦ 0.5% is satisfied, such inconvenience is eliminated and the torque is reduced.
It is possible to achieve both higher bearing life and longer bearing life.

【００１７】又、請求項２に記載した円すいころ軸受に
よれば、限られた寸法範囲内で基本動定格荷重を大きく
できて、軸受寿命を確保しつつトルクを更に低減でき
る。即ち、請求項３に記載した円すいころ軸受で規制す
る、比（ｄm ×Ｌ）／（Ｄa ×Ｗ）は、ｄm ／Ｄa とＬ
／Ｗとの積になる。そして、このうちの１未満である、
比Ｌ／Ｗを大きくした場合には、基本動定格荷重を大き
くできる。但し、この場合には、各円すいころの転動面
と内、外輪各軌道との接触面積が大きくなる為、トルク
が高くなる傾向になる。これに対して、請求項３に記載
した円すいころ軸受の場合、上記比（ｄm ×Ｌ）／（Ｄ
a ×Ｗ）を所定の範囲に規制している為、上記比Ｌ／Ｗ
が大きくなるのに伴って、１よりも大きい、上記比ｄm
／Ｄa を小さくできる。又、この比ｄm ／Ｄa に関す
る、複数の円すいころのピッチ円の直径ｄm は、円すい
ころの数ｚと各円すいころの平均直径Ｄa との積に比例
する（ｄm ∝ｚ×Ｄa ）。従って、Ｄa の大きさに拘ら
ずｚを小さくして、トルクを低減できる。この様に、上
記比（ｄm ×Ｌ）／（Ｄa ×Ｗ）を所定の範囲に規制す
る、請求項３に記載した円すいころ軸受の場合には、基
本動定格荷重を大きくできて、軸受寿命を確保しつつト
ルクを更に低減できる。この為、実用上要求される性能
を確保しつつ、円すいころ軸受の小型化を図り易くな
る。According to the tapered roller bearing of the second aspect, the basic dynamic load rating can be increased within a limited size range, and the bearing life can be secured while the torque can be further reduced. That is regulated by a tapered roller bearing according to claim 3, the ratio (d m × L) / ( D a × W) is, d m / D a and L
It becomes the product of / W. And less than 1 of these,
When the ratio L / W is increased, the basic dynamic load rating can be increased. However, in this case, the contact area between the rolling surface of each tapered roller and each of the inner and outer ring raceways becomes large, so that the torque tends to increase. In contrast, when the tapered roller bearing according to claim 3, said ratio (d m × L) / ( D
a × W) is regulated within a predetermined range, the above ratio L / W
Is larger than 1, the ratio d m is larger than 1.
/ D a can be reduced. Further, the pitch circle diameter d m of the plurality of tapered rollers with respect to the ratio d m / D a is proportional to the product of the number z of the tapered rollers and the average diameter D a of each tapered roller (d m ∝z × D a ). Therefore, the torque can be reduced by reducing z regardless of the magnitude of D a . Thus, in the case of the tapered roller bearing according to claim 3, wherein the ratio (d m × L) / (D a × W) is restricted within a predetermined range, the basic dynamic load rating can be increased, The torque can be further reduced while ensuring the bearing life. Therefore, it is easy to reduce the size of the tapered roller bearing while ensuring the performance required for practical use.

【００１８】又、請求項３に記載した円すいころ軸受に
よれば、コスト上昇を抑えつつ、更なるトルクの低減を
図れる。即ち、各円すいころの転動面の粗さＲ1 と、こ
れら各円すいころの大径側端面の粗さＲ2 と、内、外輪
各軌道の粗さｒ1 と、内輪の大径側鍔部の片面の粗さｒ
2 との総てを、十分に小さく（例えば０．１μｍＲa以
下に）すれば、トルクを更に低減できる。但し、この場
合には、加工コストが徒に嵩む原因となる。又、上記各
円すいころの転動面と上記内、外輪各軌道との転がり接
触部に関する粗さＲ1 、ｒ1 と、これら各円すいころの
大径側端面と上記大径側鍔部の片面との滑り接触部に関
する粗さＲ2 、ｒ2 とのうち、一方を他方に対し小さく
し過ぎた場合には、一方の粗さを小さくする事によりト
ルクの低減を図れる効果が小さくなる。本発明の場合に
は、これらの粗さに関する比（Ｒ 1 ×ｒ1 ）／（Ｒ2 ×
ｒ2 ）を所定の範囲に規制すると共に、上記各円すいこ
ろの大径側端面と上記大径側鍔部の片面との滑り接触部
に関する粗さＲ2 、ｒ2 を小さく抑えている為、コスト
上昇を小さく抑えつつ、更なるトルクの低減を図れる。Further, in the tapered roller bearing according to claim 3,
According to this, further reduction of torque while suppressing cost increase
Can be achieved. That is, the roughness R of the rolling surface of each tapered roller1 And this
Roughness R of the large diameter end surface of each tapered roller2 And inner and outer rings
Roughness of each track r1 And the roughness r on one side of the large-diameter side flange of the inner ring
2 And all are sufficiently small (for example, 0.1 μmRaSince
(Below), the torque can be further reduced. However, this place
If this happens, the processing cost will be excessive. In addition, each of the above
Rolling contact between the tapered roller rolling surface and each of the above inner and outer ring raceways
Roughness R about touch1 , R1 And each of these tapered rollers
Regarding the sliding contact part between the large diameter side end surface and one side of the large diameter side collar part
Roughness R2 , R2 And one is smaller than the other
If too much, reduce the roughness of one side
The effect of reducing luck is reduced. In the case of the present invention
Is the ratio (R 1 × r1 ) / (R2 ×
r2 ) Is regulated within a predetermined range, and
Sliding contact part between the large diameter side end face of the filter and one side of the large diameter side flange part
Roughness R2 , R2 Is kept small, so the cost
It is possible to further reduce the torque while suppressing the rise to be small.

【００１９】[0019]

【実施例】本発明の効果を確認する為に、本発明者が行
なった実験に就いて説明する。次の表１は、この実験に
使用した１４種類の実施例と５種類の比較例との諸元を
示している。尚、実験に使用した円すいころ軸受は、外
輪、内輪、円すいころの何れの部材も、高炭素クロム軸
受鋼であるＳＵＪ ２（ＪＩＳ Ｇ ４８０５）製の素
材に、従来から一般的に知られている熱処理を施したも
のにより造っている。但し、これら各部材は、クロム鋼
であるＳＣｒ ４２０Ｈ、ＳＣｒ ４３０Ｈ、ＳＣｒ
４４０Ｈ（ＪＩＳ Ｇ ４０５２）、中炭素モリブデン
鋼、中炭素クロムマンガン鋼により構成する事もでき
る。又、構成各部材の材料は、必ずしも同一である必要
はなく、各部材ごとに異種材料を組み合わせて使用する
事もできる。又、各円すいころ軸受の内径を２０〜５５
mmとし、複数の円すいころの大径側端面でのピッチ円の
直径ｄm （図２参照）を３８〜７０mmとした。又、基本
動定格荷重Ｃr を３５〜７０ｋＮとした。EXAMPLES In order to confirm the effects of the present invention, experiments conducted by the present inventors will be described. The following Table 1 shows the specifications of 14 kinds of Examples and 5 kinds of Comparative Examples used in this experiment. In the tapered roller bearing used in the experiment, the outer ring, the inner ring, and the tapered roller are all commonly known as a material made of SUJ 2 (JIS G 4805), which is high carbon chrome bearing steel. It is made of heat-treated material. However, these members are made of chromium steel such as SCr 420H, SCr 430H, and SCr.
440H (JIS G4052), medium carbon molybdenum steel, medium carbon chromium manganese steel can also be used. The materials of the constituent members are not necessarily the same, and different materials may be used in combination for each member. Also, set the inner diameter of each tapered roller bearing to 20 to 55.
mm, and the diameter d m (see FIG. 2) of the pitch circles on the large diameter side end faces of the plurality of tapered rollers was 38 to 70 mm. Further, the basic dynamic load rating C r is set to 35 to 70 kN.

【００２０】[0020]

【表１】 [Table 1]

【００２１】この表１に示した１９種類の円すいころ軸
受に就いて、図３に示す様な実験装置を使用して、耐久
試験とトルク測定とを行なった。先ず、この実験装置に
就いて説明する。円すいころ軸受６ａ（又は６ｂ）の内
輪８ａ（又は８ｂ）を外嵌固定したホルダ１２は、駆動
軸１３の上端部にテーパ嵌合して、この駆動軸１３によ
り回転駆動される。又、外輪７ａ（又は７ｂ）は外側ホ
ルダ１４を介してハウジング１５の内側に内嵌固定して
いる。このハウジング１５内には、給油孔１６を通じて
所定の潤滑油を供給自在としている。又、上記ハウジン
グ１５の上面には、静圧パッド１７を介して、所定のア
キシアル荷重を付与自在としている。更に、上記ハウジ
ング１５の外周面に固定した腕片１８の先端部と図示し
ない固定の部分との間にロードセル１９を設けて、上記
駆動軸１３の回転時に上記ハウジング１５に加わる動ト
ルク｛＝円すいころ軸受６ａ（又は６ｂ）の回転トル
ク｝と起動トルクとを測定自在としている。又、このハ
ウジング１５に、図示しない振動センサを設けている。
尚、動トルク及び起動トルクを低減させる目的は、前述
した様に省燃費化を図る為である。従って、省燃費化の
面からは影響の少ないラジアル荷重は、動トルク及び起
動トルクの測定時に付与しなかった。即ち、デファレン
シャルギヤやトランスミッションの運転時には、アキシ
アル荷重は常に加わったままとなるが、大きなラジアル
荷重が加わるのは、急加減速時等、限られた場合であ
り、運転時間全体に占める割合は少ない。従って、ラジ
アル荷重による動トルク及び起動トルクの変化が燃費性
能に及ぼす影響はアキシアル荷重に比べて小さい。そこ
で、動トルク及び起動トルクの測定の実験時に付与する
荷重は、アキシアル荷重のみとした。The 19 types of tapered roller bearings shown in Table 1 were subjected to a durability test and a torque measurement by using an experimental apparatus as shown in FIG. First, this experimental apparatus will be described. The holder 12 to which the inner ring 8a (or 8b) of the tapered roller bearing 6a (or 6b) is externally fitted and fixed is taper-fitted to the upper end portion of the drive shaft 13 and is rotationally driven by the drive shaft 13. The outer ring 7 a (or 7 b) is fitted and fixed inside the housing 15 via the outer holder 14. A predetermined lubricating oil can be supplied into the housing 15 through the oil supply hole 16. Further, a predetermined axial load can be applied to the upper surface of the housing 15 via a static pressure pad 17. Further, a load cell 19 is provided between the end of the arm piece 18 fixed to the outer peripheral surface of the housing 15 and a fixed portion (not shown) so that the dynamic torque applied to the housing 15 when the drive shaft 13 rotates {= cone The rotational torque of the roller bearing 6a (or 6b)} and the starting torque can be measured freely. Further, the housing 15 is provided with a vibration sensor (not shown).
The purpose of reducing the dynamic torque and the starting torque is to save fuel as described above. Therefore, the radial load, which has little influence from the viewpoint of fuel saving, was not applied at the time of measuring the dynamic torque and the starting torque. That is, while the differential gear and the transmission are operating, the axial load is always applied, but the large radial load is applied only in a limited time such as during sudden acceleration / deceleration, and the ratio to the total operating time is small. . Therefore, the influence of the change in the dynamic torque and the starting torque due to the radial load on the fuel consumption performance is smaller than that in the axial load. Therefore, the load applied during the experiment for measuring the dynamic torque and the starting torque was only the axial load.

【００２２】次の表２は、上述の様な実験装置を使用し
て行なった、円すいころ軸受の耐久試験の結果を示して
いる。この耐久試験は、潤滑油としてＶＧ１０のギヤオ
イルを供給すると共に、アキシアル荷重Ｆa として基本
動定格荷重Ｃr の０．２倍（Ｆa ＝０．２Ｃr ）を付与
しつつ行なった。又、上記駆動軸１３を、２０００min
-1 （r.p.m.）で回転させた。更に、潤滑油中には、硬
さがＨv ７５０で５０μｍ以下の異物を１００ｐｐｍ混
入させて、油膜パラメータΛを１程度に設定した。そし
て、表１にそれぞれ諸元を示した１９種類の円すいころ
軸受毎に、それぞれ１０個（Ｎ＝１０）ずつ耐久試験を
行なった。又、振動センサにより検出された振動値が初
期振動の５倍になった時点で試験を中止し、内、外輪各
軌道１１ａ、１０ａ（又は１１ｂ、１０ｂ）と各円すい
ころ９ａ（又は９ｂ）の転動面とでの剥離の発生の有無
を観察した。又、アキシアル荷重Ｆr と基本動定格荷重
Ｃrとの比Ｆr ／Ｃr ＝０．２で、計算により求められ
る軸受寿命Ｌc は２１４×１０6 回転である為、耐久試
験は、この軸受寿命ＬC よりも長い２５０×１０6 回転
で打ち切った。尚、剥離は、内、外輪各軌道１１ａ、１
０ａ（又は１１ｂ、１０ｂ）と各円すいころ９ａ（又は
９ｂ）の転動面とで、ほぼ同じ頻度で発生した。尚、表
２中、「剥離の有無」の欄で示した数字は、１０個の試
料のうち、剥離が発生した試料の割合を表している。Table 2 below shows the results of the durability test of the tapered roller bearing, which was carried out by using the above-described experimental apparatus. This durability test was performed while supplying VG10 gear oil as the lubricating oil and applying 0.2 times the basic dynamic load rating C r (F a = 0.2 C r ) as the axial load F a . In addition, the drive shaft 13 is set to 2000 min.
It was rotated at -1 (rpm). Further, 100 ppm of foreign matter having a hardness of Hv 750 and 50 μm or less was mixed in the lubricating oil, and the oil film parameter Λ was set to about 1. Then, 10 (N = 10) durability tests were conducted for each of the 19 types of tapered roller bearings whose specifications are shown in Table 1. The test was stopped when the vibration value detected by the vibration sensor became 5 times the initial vibration, and the inner and outer races 11a, 10a (or 11b, 10b) and the tapered rollers 9a (or 9b) It was observed whether or not peeling occurred on the rolling surface. Further, the ratio F r / C r = 0.2 the axial load F r and the basic dynamic load rating C r, for bearing life L c obtained by calculation is 214 × 10 6 rotates, the durability test, the The bearing was cut off at 250 × 10 6 rotations longer than the bearing life L C. In addition, peeling is performed for each of the inner and outer raceways 11a, 1
0a (or 11b, 10b) and the rolling surface of each tapered roller 9a (or 9b) occurred at almost the same frequency. In Table 2, the number shown in the column of "presence / absence of peeling" represents the ratio of the samples in which peeling occurred among the 10 samples.

【００２３】[0023]

【表２】 [Table 2]

【００２４】上記表２に示した各実施例１〜１４は、
２．０≦（β×ｚ）／α≦３．０で、且つ、０．１％≦
δ／Ｄa ≦０．５％で、且つ、３．０≦（ｄm ×Ｌ）／
（Ｄa×Ｗ）≦６．０を満たす様に、各部の寸法を規制
している。この為、各円すいころ９ａ（又は９ｂ）の転
動面と内、外輪各軌道１１ａ、１０ａ（又は１１ｂ、１
０ｂ）との接触部での転がり摩擦と、これら各円すいこ
ろ９ａ（又は９ｂ）の大径側端面と大径側鍔部２１の片
面との接触部での滑り摩擦とを小さく抑える事ができ
る。従って、各滑り接触部と各転がり接触部とに潤滑油
膜が形成されにくい状態でもこれら各接触部での温度上
昇を抑えて、異物により、早期剥離に結び付き易い、著
しい圧痕が形成されるのを防止できる。この結果、上記
表２の実験結果から明らかな様に、各実施例１〜１４の
寿命は、軸受寿命Ｌ10を計算で求められる寿命ＬC （＝
２１４×１０6 回転）よりも長くできた。Examples 1 to 14 shown in Table 2 above are
2.0 ≦ (β × z) /α≦3.0 and 0.1% ≦
δ / D a ≦ 0.5% and 3.0 ≦ (d m × L) /
The dimensions of each part are regulated so as to satisfy (D a × W) ≦ 6.0. Therefore, the rolling surface of each tapered roller 9a (or 9b) and each of the inner and outer races 11a, 10a (or 11b, 1
0b), the rolling friction at the contact portion and the sliding friction at the contact portion between the large diameter side end surface of each of the tapered rollers 9a (or 9b) and one surface of the large diameter side flange portion 21 can be suppressed to be small. . Therefore, even if a lubricating oil film is difficult to be formed on each sliding contact portion and each rolling contact portion, the temperature rise at each contact portion is suppressed, and a significant indentation that easily leads to early peeling due to foreign matter is formed. It can be prevented. As a result, as apparent from the experimental results of Table 2, the life of the Examples 1-14, the life is determined by calculating the bearing life L 10 L C (=
It was longer than 214 × 10 6 revolutions).

【００２５】特に、実施例４〜１１の場合には、接触角
度を小さく（２１°以下に）抑える事により上記大径側
端面と片面との接触部での滑り摩擦を十分に小さくし
て、この接触部での温度上昇を抑えている。しかも、
３．５≦（ｄm ×Ｌ）／（Ｄa ×Ｗ）≦４．９を満たす
様に規制している為、内、外輪各軌道１１ａ、１０ａ
（又は１１ｂ、１０ｂ）と各円すいころ９ａ（又は９
ｂ）との間に異物が噛み込まれにくくなる。この為、耐
久試験の打ち切り時間経過後でも、剥離は発生せず、軸
受寿命を他の実施例の場合よりも長くできた。従って、
本発明の円すいころ軸受では、好ましくは、上記接触角
αを２１°以下に抑えると共に、３．５≦（ｄm×Ｌ）
／（Ｄa ×Ｗ）≦４．９を満たす様に、各部の寸法を規
制する。In particular, in the case of Examples 4 to 11, by suppressing the contact angle to be small (21 ° or less), the sliding friction at the contact portion between the large-diameter side end surface and one surface was sufficiently reduced, The temperature rise at this contact part is suppressed. Moreover,
Since it is regulated to satisfy 3.5 ≦ (d m × L) / (D a × W) ≦ 4.9, the inner and outer ring raceways 11a, 10a
(Or 11b, 10b) and each tapered roller 9a (or 9
It becomes difficult for foreign matter to be caught between b). Therefore, peeling did not occur even after the end of the endurance test, and the bearing life could be made longer than in the other examples. Therefore,
The tapered roller bearing of the present invention, preferably, the contact angle α with suppressed to 21 ° or less, 3.5 ≦ (d m × L )
The dimensions of each part are regulated so that / (D a × W) ≦ 4.9 is satisfied.

【００２６】これに対して、比較例１の場合には、３．
０≦（ｄm ×Ｌ）／（Ｄa ×Ｗ）≦６．０を満たしてい
るが、２．０≦（β×ｚ）／α≦３．０を満たしていな
い。この為、各円すいころ９ａ（又は９ｂ）の大径側端
面と、内輪８ａ（又は８ｂ）に形成した大径側鍔部２１
の片面との接触部に作用する滑り摩擦と、上記各円すい
ころ９ａ（又は９ｂ）の転動面と内、外輪各軌道１１
ａ、１０ａ（又は１１ｂ、１０ｂ）との間に作用する転
がり摩擦とが大きくなる。又、上記比較例１の場合に
は、比δ／Ｄa が０．１％未満である。この為にこの比
較例１の場合には、表面起点型の剥離を生じ易くなる。
これらにより、上記比較例１の場合には、１０回の試験
の総てで、内、外輪各軌道１１ａ、１０ａ（又は１１
ｂ、１０ｂ）及び各円すいころ９ａ（又は９ｂ）の転動
面に剥離が発生し、軸受寿命Ｌ10が１１１×１０6 回転
と、計算で求められる寿命ＬC の約１／２になった。On the other hand, in the case of Comparative Example 1, 3.
0 ≦ (d m × L) / (D a × W) ≦ 6.0 is satisfied, but 2.0 ≦ (β × z) /α≦3.0 is not satisfied. Therefore, the large-diameter side end face of each tapered roller 9a (or 9b) and the large-diameter side flange portion 21 formed on the inner ring 8a (or 8b).
Sliding friction acting on the contact portion with one surface of each of the tapered rollers 9a (or 9b) and the inner and outer ring raceways 11
The rolling friction acting between a and 10a (or 11b and 10b) becomes large. Further, in the case of Comparative Example 1, the ratio δ / D a is less than 0.1%. For this reason, in the case of Comparative Example 1, the surface-originating type peeling is likely to occur.
As a result, in the case of Comparative Example 1 described above, the inner and outer ring raceways 11a, 10a (or 11) in all of the 10 tests.
b, 10b) and peeling occurred on the rolling surfaces of each tapered roller 9a (or 9b), the bearing life L 10 was 111 × 10 6 rotations, which was about 1/2 of the calculated life L C. .

【００２７】又、比較例２、３、５の場合には、２．０
≦（β×ｚ）／α≦３．０を満たしているが、δ／Ｄa
が０．１％未満である（δ／Ｄa ＜０．１％）為、表面
起点型の剥離を生じ易くなる。この結果、上記比較例
２、３、５の場合には、１０回の試験の総てで、内、外
輪各軌道１１ａ、１０ａ（又は１１ｂ、１０ｂ）及び各
円すいころ９ａ（又は９ｂ）の転動面に剥離が発生し、
軸受寿命Ｌ10が計算で求められる寿命ＬC の約１／３に
なった。In the case of Comparative Examples 2, 3 and 5, 2.0
≦ (β × z) /α≦3.0 is satisfied, but δ / D a
Is less than 0.1% (δ / D a <0.1%), the surface-originating type peeling is likely to occur. As a result, in the case of Comparative Examples 2, 3, and 5, the inner and outer ring raceways 11a, 10a (or 11b, 10b) and the tapered rollers 9a (or 9b) were rolled in all the 10 tests. Peeling occurs on the moving surface,
The bearing life L 10 is about 1/3 of the calculated life L C.

【００２８】更に、比較例４の場合には、３．０≦（ｄ
m ×Ｌ）／（Ｄa ×Ｗ）≦６．０を満たしているが、
２．０≦（β×ｚ）／α≦３．０を満たしていない。こ
の為、各円すいころ９ａ（又は９ｂ）の大径側端面と、
内輪８ａ（又は８ｂ）に形成した大径側鍔部２１の片面
との接触部に作用する滑り摩擦と、上記各円すいころ９
ａ（又は９ｂ）の転動面と内、外輪各軌道１１ａ、１０
ａ（又は１１ｂ、１０ｂ）との間に作用する転がり摩擦
とが大きくなる。又、上記比δ／Ｄa が０．６％と、本
発明の場合の上限値を大きく越えている。この為、上記
比較例４の場合には、上記各円すいころ９ａ（又は９
ｂ）の転動面と内、外輪各軌道１１ａ、１０ａ（又は１
１ｂ、１０ｂ）との接触部に作用する接触面圧が極めて
大きくなり、内部起点型の剥離を生じ易くなる。これら
により、上記比較例４の場合には、１０回の試験の総て
で剥離が発生し、軸受寿命Ｌ10が計算で求められる寿命
ＬC の約１／４になった。Further, in the case of Comparative Example 4, 3.0 ≦ (d
m × L) / (D a × W) ≦ 6.0,
2.0 ≦ (β × z) /α≦3.0 is not satisfied. Therefore, the large diameter side end surface of each tapered roller 9a (or 9b),
Sliding friction that acts on the contact portion of the large-diameter side flange portion 21 formed on the inner ring 8a (or 8b) with one surface, and the above tapered rollers 9
a (or 9b) rolling surface and inner and outer ring raceways 11a, 10
The rolling friction acting between a (or 11b, 10b) becomes large. Further, the ratio δ / D a is 0.6%, which greatly exceeds the upper limit of the present invention. Therefore, in the case of Comparative Example 4, each of the tapered rollers 9a (or 9
b) rolling surface and inner and outer ring raceways 11a, 10a (or 1)
The contact surface pressure acting on the contact portion with 1b, 10b) becomes extremely large, and the internal origin type peeling easily occurs. As a result, in the case of Comparative Example 4, peeling occurred in all of the 10 tests, and the bearing life L 10 became about 1/4 of the calculated life L C.

【００２９】次に、前述の図３に示した実験装置を使用
して行なった、トルク測定試験の結果に就いて説明す
る。この実験では、前述の表１、２に示した実施例６、
８、１１、１４及び比較例１、５の円すいころ軸受で、
それぞれ粗さＲ1 、Ｒ2 、ｒ1、ｒ2 の組み合わせを３
種類又は４種類に異ならせたものを使用した。又、この
実験では、潤滑油としてＶＧ６８のギヤオイルを供給す
ると共に、基本動定格荷重Ｃr の０．０５倍（Ｆa ＝
０．０５Ｃr ）のアキシアル荷重Ｆa を付与しつつ行な
った。又、駆動軸１３を、２０００min-1 で回転させ
た。更に、潤滑油としてＶＧ６８であるギヤオイルを使
用し、油温を４０℃一定とした。次の表３及び図４に、
この様にして行なったトルク測定試験の結果を示してい
る。尚、表３に示した、円すいころ軸受の種類を表す記
号で、「−」の前に付した数字は前述の表１に示した実
施例又は比較例の種類を、同じく後に付した数字は、当
該例中での試料番号を、それぞれ表している。Next, the results of the torque measurement test conducted using the experimental apparatus shown in FIG. 3 will be described. In this experiment, Example 6 shown in Tables 1 and 2 above,
In the tapered roller bearings of 8, 11, 14 and Comparative Examples 1, 5,
Each combination of roughness R 1 , R 2 , r 1 , r 2 is 3
Different types or four types were used. In addition, in this experiment, VG68 gear oil was supplied as lubricating oil, and 0.05 times the basic dynamic load rating C r (F a =
It was performed while applying an axial load F a of 0.05 C r ). The drive shaft 13 was rotated at 2000 min -1 . Further, a gear oil of VG68 was used as the lubricating oil, and the oil temperature was kept constant at 40 ° C. The following Table 3 and FIG.
The results of the torque measurement test thus performed are shown. In addition, in Table 3, a symbol indicating the type of the tapered roller bearing, the number before "-" is the type of the embodiment or comparative example shown in Table 1 above, and the number after the same is the same. , The sample numbers in the examples are shown.

【００３０】[0030]

【表３】 [Table 3]

【００３１】表３及び図４に示す実験結果から明らかな
様に、各実施例の場合には、２．０≦（Ｒ1 ×ｒ1 ）／
（Ｒ2 ×ｒ2 ）≦５．０を満たしている為、回転速度が
２０００min-1 での動トルクを小さく（３．５Ｎｍ以下
に）できた。又、起動トルクと動トルクとの比が１．４
以下になっており、起動トルクも小さくできる事が分か
った。特に、２．０≦（Ｒ1 ×Ｒ3 ）／（Ｒ2 ×Ｒ4 ）
≦４．０を満たしている、実施例６−１〜３、８−１、
８−２、１１−１〜３、１４−２、１４−５の場合に
は、上記動トルクが３．２Ｎｍ以下になり、より好まし
い事が分かった。As is clear from the experimental results shown in Table 3 and FIG. 4, 2.0 ≦ (R 1 × r 1 ) /
Since (R 2 × r 2 ) ≦ 5.0 is satisfied, the dynamic torque at the rotation speed of 2000 min −1 can be made small (3.5 Nm or less). Also, the ratio of the starting torque to the dynamic torque is 1.4.
It became the following, and it turned out that the starting torque can also be made small. In particular, 2.0 ≦ (R 1 × R 3 ) / (R 2 × R 4 ).
Examples 6-1 to 3 and 8-1, which satisfy ≦ 4.0,
In the cases of 8-2, 11-1 to 3, 14-2, and 14-5, the dynamic torque was 3.2 Nm or less, and it was found to be more preferable.

【００３２】これに対して、（Ｒ1 ×ｒ1 ）／（Ｒ2 ×
ｒ2 ）＞５．０である、比較例１−３、５−３の場合に
は、内、外輪各軌道１１ａ、１０ａ（又は１１ｂ、１０
ｂ）と各円すいころ９ａ（又は９ｂ）の転動面との接触
部での転がり摩擦が大きくなる為、上記動トルクが５．
１Ｎｍ以上と、大きくなった。逆に、（Ｒ1 ×ｒ1 ）／
（Ｒ2 ×ｒ2 ）＜２．０である、比較例１−１、１−
２、５−１、５−２の場合には、各円すいころ９ａ（又
は９ｂ）の大径側端面と内輪８ａ（又は８ｂ）の大径側
鍔部２１の片面との滑り接触部での摩擦が大きくなる
為、上記動トルクが４．２Ｎｍ以上と、やはり大きくな
った。しかも、この場合には、上記各円すいころ９ａ
（又は９ｂ）の大径側端面の粗さＲ2 と、内輪８ａ（又
は８ｂ）の大径側鍔部２１の片面の粗さｒ2 とが０．１
５μｍよりも大きい為、低速でのトルクが高くなり、上
記起動トルクと動トルクとの比が２．１以上と高くなっ
た。On the other hand, (R 1 × r 1 ) / (R 2 ×
In the case of Comparative Examples 1-3 and 5-3 in which r 2 )> 5.0, inner and outer ring raceways 11a, 10a (or 11b, 10).
Since the rolling friction at the contact portion between b) and each rolling surface of each tapered roller 9a (or 9b) becomes large, the dynamic torque becomes 5.
It became larger than 1 Nm. Conversely, (R 1 × r 1 ) /
(R 2 × r 2 ) <2.0, Comparative Examples 1-1, 1-
In the case of 2, 5-1, 5-2, at the sliding contact portion between the large diameter side end surface of each tapered roller 9a (or 9b) and one surface of the large diameter side flange portion 21 of the inner ring 8a (or 8b). Since the friction was large, the dynamic torque was 4.2 Nm or more, which was also large. Moreover, in this case, each of the above tapered rollers 9a
(Or 9b) the roughness R 2 of the large diameter side end surface and the roughness r 2 of one surface of the large diameter side collar portion 21 of the inner ring 8a (or 8b) are 0.1.
Since it was larger than 5 μm, the torque at low speed was high, and the ratio of the starting torque to the dynamic torque was 2.1 or higher.

【００３３】尚、上述のトルク測定に使用した各実施例
では、各円すいころ９ａ（又は９ｂ）の転動面の粗さＲ
1 と内、外輪各軌道１１ａ、１０ａ（又は１１ｂ、１０
ｂ）の粗さｒ1 とを、０．１０〜０．３０μｍＲａとす
ると共に、各円すいころ９ａ（又は９ｂ）の大径側端面
の粗さＲ2 と内輪８ａ（又は８ｂ）の大径側鍔部２１の
片面の粗さｒ2 とを、０．０７〜０．１５μｍＲａとし
た。但し、これら各円すいころ９ａ（又は９ｂ）の大径
側端面の粗さＲ2 と上記大径側鍔部２１の片面の粗さｒ
2 とを、０．０６μｍＲａ以下に更に小さくすれば、ト
ルクを更に低減できる。In each of the embodiments used for the above torque measurement, the roughness R of the rolling surface of each tapered roller 9a (or 9b) is used.
1 and inner and outer ring raceways 11a, 10a (or 11b, 10)
The roughness r 1 of b) is set to 0.10 to 0.30 μmRa, and the roughness R 2 of the large diameter side end surface of each tapered roller 9a (or 9b) and the large diameter side of the inner ring 8a (or 8b) are set. The roughness r 2 on one surface of the collar portion 21 was set to 0.07 to 0.15 μmRa. However, the roughness R 2 of the large-diameter side end surface of each of the tapered rollers 9a (or 9b) and the roughness r of one surface of the large-diameter side flange portion 21.
By further reducing 2 and 0.06 μmRa or less, the torque can be further reduced.

【００３４】尚、上述した各実施例は、軸受鋼の一種で
あるＳＵＪ ２を使用したが、肌焼き鋼に浸炭処理又は
浸炭窒化処理を施して、表面の残留オーステナイト量を
２０〜４５％で表面硬さ（ビッカース硬さ）をＨv ７０
０〜８５０としたもの等の、異物による表面損傷を生じ
にくい材料を使用すれば、軸受寿命を更に向上させる事
ができる。In each of the above-mentioned embodiments, SUJ 2 which is a kind of bearing steel was used. However, case hardening steel is carburized or carbonitrided so that the amount of retained austenite on the surface is 20 to 45%. Surface hardness (Vickers hardness) Hv 70
The bearing life can be further improved by using a material such as 0 to 850, which is unlikely to cause surface damage due to foreign matter.

【００３５】[0035]

【発明の効果】本発明の円すいころ軸受は、以上に述べ
た通り構成され作用するので、トルク低減と軸受寿命の
確保とを、高次元に両立できる。この結果、円すいころ
軸受を組み込んだ自動車の耐久性を確保しつつ省燃費化
を図れる。Since the tapered roller bearing of the present invention is constructed and operates as described above, it is possible to achieve both high torque reduction and long bearing life at the same time. As a result, it is possible to save fuel consumption while ensuring the durability of the automobile incorporating the tapered roller bearing.

【図面の簡単な説明】[Brief description of drawings]

【図１】円すいころ軸受を組み込んだデファレンシャル
ギヤの１例を示す縦断側面図。FIG. 1 is a vertical sectional side view showing an example of a differential gear incorporating a tapered roller bearing.

【図２】円すいころ軸受の１例を示す半部断面図。FIG. 2 is a half sectional view showing an example of a tapered roller bearing.

【図３】実験装置の縦断面図。FIG. 3 is a vertical cross-sectional view of the experimental device.

【図４】円すいころ軸受の滑り接触部及び転がり接触部
に関する粗さと、動トルクとの関係を示す図。FIG. 4 is a diagram showing a relationship between roughness and dynamic torque of a sliding contact portion and a rolling contact portion of a tapered roller bearing.

【符号の説明】[Explanation of symbols]

１ ケーシング ２ ピニオン軸 ３ 結合フランジ ４ 減速小歯車 ５ 減速大歯車 ６ａ、６ｂ 円すいころ軸受 ７ａ、７ｂ 外輪 ８ａ、８ｂ 内輪 ９ａ、９ｂ 円すいころ １０ａ、１０ｂ 外輪軌道 １１ａ、１１ｂ 内輪軌道 １２ ホルダ １３ 駆動軸 １４ 外側ホルダ １５ ハウジング １６ 給油孔 １７ 静圧パッド １８ 腕片 １９ ロードセル ２０ 保持器 ２１ 大径側鍔部 ２２ 小径側鍔部 1 casing 2 pinion shaft 3 coupling flange 4 reduction gears 5 reduction gears 6a, 6b tapered roller bearing 7a, 7b outer ring 8a, 8b Inner ring 9a, 9b tapered roller 10a, 10b Outer ring raceway 11a, 11b Inner ring raceway 12 holder 13 Drive axis 14 Outer holder 15 housing 16 lubrication hole 17 Static pressure pad 18 arm pieces 19 load cell 20 cage 21 Large diameter side flange 22 Small diameter side flange

Claims (3)

【特許請求の範囲】[Claims] 【請求項１】 内周面に円すい凹面状の外輪軌道を有す
る外輪と、外周面のうちで中間部に円すい凸面状の内輪
軌道を、大径側端部に大径側鍔部を、小径側端部に小径
鍔部を、それぞれ有する内輪と、上記外輪軌道と内輪軌
道との間に転動自在に設けられた複数個の円すいころと
を備えた円すいころ軸受に於いて、接触角をαとし、こ
れら各円すいころの円すい角度の１／２をβとし、円す
いころの数をｚとし、軸方向長さをＬとし、平均直径を
Ｄa とし、クラウニング量をδとした場合に、１５°＜
α＜２２°であり、且つ、２．０≦（β×ｚ）／α≦
３．０を満たし、且つ、これら各円すいころの軸方向中
央から軸方向両端側に０．４Ｌ外れた位置で、０．１％
≦δ／Ｄa ≦０．５％を満たす事を特徴とする円すいこ
ろ軸受。1. An outer ring having a conical concave outer ring raceway on its inner peripheral surface, a conical convex inner raceway at an intermediate portion of its outer peripheral surface, and a large diameter side flange portion at a large diameter side end portion with a small diameter. In a tapered roller bearing provided with an inner ring each having a small-diameter collar portion at the side end and a plurality of tapered rollers rotatably provided between the outer ring raceway and the inner ring raceway, the contact angle is α, 1/2 of the taper angle of each of these tapered rollers is β, the number of tapered rollers is z, the axial length is L, the average diameter is D a , and the crowning amount is δ, 15 ° <
α <22 ° and 2.0 ≦ (β × z) / α ≦
3.0% and 0.1% at a position deviated from the axial center of each of these tapered rollers by 0.4L on both axial ends.
A tapered roller bearing characterized by satisfying ≦ δ / D a ≦ 0.5%. 【請求項２】 各円すいころに関して、軸方向長さをＬ
とし、平均直径をＤ a とし、複数の円すいころの大径側
端面でのピッチ円の直径をｄm とし、これら複数の円す
いころと外輪と内輪とを組み立てた状態での幅をＷとし
た場合に、３．０≦（ｄm ×Ｌ）／（Ｄa ×Ｗ）≦６．
０を満たす、請求項１に記載した円すいころ軸受。2. The axial length of each tapered roller is L
And the average diameter is D a And the large diameter side of multiple tapered rollers
The diameter of the pitch circle at the end face is dm And these multiple circles
The width when the die, outer ring and inner ring are assembled is W
3.0 ≦ (dm × L) / (Da × W) ≦ 6.
The tapered roller bearing according to claim 1, which satisfies 0. 【請求項３】 各円すいころの転動面の粗さをＲ1 と
し、これら各円すいころの大径側端面の粗さをＲ2 と
し、内、外輪各軌道の粗さをｒ1 とし、内輪の大径側端
部外周面に設けた大径側鍔部の片面で、上記各円すいこ
ろの大径側端面と対向する面の粗さをｒ2 とした場合
に、Ｒ2 及びｒ2 が０．１５μｍＲａ以下であり、Ｒ2
／Ｒ1 ＜１で、且つ、ｒ2 ／ｒ1 ＜１であり、２．０≦
（Ｒ1 ×ｒ1）／（Ｒ2 ×ｒ2 ）≦５．０を満たす、請
求項１〜２の何れかに記載した円すいころ軸受。3. The roughness of the rolling surface of each tapered roller is R 1 , the roughness of the end surface on the large diameter side of each of these tapered rollers is R 2, and the roughness of each of the inner and outer races is r 1 . in one side of the large diameter side collar portion provided at the large diameter end portion outer peripheral surface of the inner ring, the roughness of the large diameter side end face and the opposing surfaces of the respective tapered rollers when the r 2, R 2 and r 2 Is 0.15 μmRa or less, R 2
/ R 1 <1, and r 2 / r 1 <1, 2.0 ≦
The tapered roller bearing according to claim 1, which satisfies (R 1 × r 1 ) / (R 2 × r 2 ) ≦ 5.0.
JP2002113784A 2002-04-16 2002-04-16 Tapered roller bearing Pending JP2003314542A (en)

Priority Applications (1)

Application Number Priority Date Filing Date Title
JP2002113784A JP2003314542A (en) 2002-04-16 2002-04-16 Tapered roller bearing

Applications Claiming Priority (1)

Application Number Priority Date Filing Date Title
JP2002113784A JP2003314542A (en) 2002-04-16 2002-04-16 Tapered roller bearing

Publications (2)

Publication Number Publication Date
JP2003314542A true JP2003314542A (en) 2003-11-06
JP2003314542A5 JP2003314542A5 (en) 2005-09-15

Family

ID=29533490

Family Applications (1)

Application Number Title Priority Date Filing Date
JP2002113784A Pending JP2003314542A (en) 2002-04-16 2002-04-16 Tapered roller bearing

Country Status (1)

Country Link
JP (1) JP2003314542A (en)

Cited By (10)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JP2005188738A (en) * 2003-12-02 2005-07-14 Ntn Corp Tapered roller bearing
JP2005351472A (en) * 2004-05-13 2005-12-22 Ntn Corp Tapered roller bearing
JP2006022824A (en) * 2004-07-05 2006-01-26 Ntn Corp Tapered roller bearing for differential
JP2006022823A (en) * 2004-07-05 2006-01-26 Ntn Corp Tapered roller bearing for transmission
EP1754900A1 (en) * 2005-08-18 2007-02-21 Jtekt Corporation Tapered roller bearing with crowned rolling contact surfaces for the support of an automotive pinion shaft
WO2008152921A1 (en) 2007-06-15 2008-12-18 Ntn Corporation Tapered roller bearing
WO2009037959A1 (en) * 2007-09-18 2009-03-26 Ntn Corporation Tapered roller bearing
JP2013040634A (en) * 2011-08-11 2013-02-28 Nsk Ltd Assembling method of conical rolling bearing and conical rolling bearing
JP2016053422A (en) * 2016-01-19 2016-04-14 日本精工株式会社 Assembly method of conical roller bearing
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Cited By (20)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US7789570B2 (en) 2003-12-02 2010-09-07 Ntn Corporation Tapered roller bearing
JP2005188738A (en) * 2003-12-02 2005-07-14 Ntn Corp Tapered roller bearing
JP2005351472A (en) * 2004-05-13 2005-12-22 Ntn Corp Tapered roller bearing
JP2006022824A (en) * 2004-07-05 2006-01-26 Ntn Corp Tapered roller bearing for differential
JP2006022823A (en) * 2004-07-05 2006-01-26 Ntn Corp Tapered roller bearing for transmission
US7484895B2 (en) 2005-08-18 2009-02-03 Jtekt Corporation Tapered roller bearing and automotive pinion shaft supporting apparatus utilizing same tapered roller bearing
EP1754900A1 (en) * 2005-08-18 2007-02-21 Jtekt Corporation Tapered roller bearing with crowned rolling contact surfaces for the support of an automotive pinion shaft
EP2157326A1 (en) * 2007-06-15 2010-02-24 NTN Corporation Tapered roller bearing
JP2008309270A (en) * 2007-06-15 2008-12-25 Ntn Corp Tapered roller bearing
WO2008152921A1 (en) 2007-06-15 2008-12-18 Ntn Corporation Tapered roller bearing
EP2157326A4 (en) * 2007-06-15 2011-11-16 Ntn Toyo Bearing Co Ltd Tapered roller bearing
US8226300B2 (en) 2007-06-15 2012-07-24 Ntn Corporation Tapered roller bearing
WO2009037959A1 (en) * 2007-09-18 2009-03-26 Ntn Corporation Tapered roller bearing
JP2009068676A (en) * 2007-09-18 2009-04-02 Ntn Corp Tapered roller bearing
EP2189670A1 (en) * 2007-09-18 2010-05-26 NTN Corporation Tapered roller bearing
EP2189670A4 (en) * 2007-09-18 2011-11-23 Ntn Toyo Bearing Co Ltd Tapered roller bearing
US8439574B2 (en) 2007-09-18 2013-05-14 Ntn Corporation Tapered roller bearing
JP2013040634A (en) * 2011-08-11 2013-02-28 Nsk Ltd Assembling method of conical rolling bearing and conical rolling bearing
JP2016053422A (en) * 2016-01-19 2016-04-14 日本精工株式会社 Assembly method of conical roller bearing
CN112145557A (en) * 2019-06-27 2020-12-29 纳博特斯克有限公司 Bearing and speed reducer

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