JP2024069478A - Double-row tapered roller bearing - Google Patents

Double-row tapered roller bearing Download PDF

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JP2024069478A
JP2024069478A JP2024040170A JP2024040170A JP2024069478A JP 2024069478 A JP2024069478 A JP 2024069478A JP 2024040170 A JP2024040170 A JP 2024040170A JP 2024040170 A JP2024040170 A JP 2024040170A JP 2024069478 A JP2024069478 A JP 2024069478A
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row
load
double
bearing
tapered roller
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偉達 顔
Weida Yan
智仁 伊藤
Tomohito Ito
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NTN Corp
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NTN Corp
NTN Toyo Bearing Co Ltd
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Priority claimed from JP2020092067A external-priority patent/JP7456851B2/en
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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
    • F16C33/00—Parts of bearings; Special methods for making bearings or parts thereof
    • F16C33/30—Parts of ball or roller bearings
    • F16C33/34—Rollers; Needles
    • F16C33/36—Rollers; Needles with bearing-surfaces other than cylindrical, e.g. tapered; with grooves in the bearing surfaces
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F03—MACHINES OR ENGINES FOR LIQUIDS; WIND, SPRING, OR WEIGHT MOTORS; PRODUCING MECHANICAL POWER OR A REACTIVE PROPULSIVE THRUST, NOT OTHERWISE PROVIDED FOR
    • F03D—WIND MOTORS
    • F03D80/00—Details, components or accessories not provided for in groups F03D1/00 - F03D17/00
    • F03D80/70—Bearing or lubricating arrangements
    • 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
    • F16C19/38—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 two or more rows of rollers
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16C—SHAFTS; FLEXIBLE SHAFTS; ELEMENTS OR CRANKSHAFT MECHANISMS; ROTARY BODIES OTHER THAN GEARING ELEMENTS; BEARINGS
    • F16C19/00—Bearings with rolling contact, for exclusively rotary movement
    • F16C19/54—Systems consisting of a plurality of bearings with rolling friction
    • F16C19/56—Systems consisting of a plurality of bearings with rolling friction in which the rolling bodies of one bearing differ in diameter from those of another
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16C—SHAFTS; FLEXIBLE SHAFTS; ELEMENTS OR CRANKSHAFT MECHANISMS; ROTARY BODIES OTHER THAN GEARING ELEMENTS; BEARINGS
    • 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
    • F16C35/00—Rigid support of bearing units; Housings, e.g. caps, covers
    • F16C35/08—Rigid support of bearing units; Housings, e.g. caps, covers for spindles
    • F16C35/12—Rigid support of bearing units; Housings, e.g. caps, covers for spindles with ball or roller bearings
    • Y—GENERAL TAGGING OF NEW TECHNOLOGICAL DEVELOPMENTS; GENERAL TAGGING OF CROSS-SECTIONAL TECHNOLOGIES SPANNING OVER SEVERAL SECTIONS OF THE IPC; TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
    • Y02—TECHNOLOGIES OR APPLICATIONS FOR MITIGATION OR ADAPTATION AGAINST CLIMATE CHANGE
    • Y02E—REDUCTION OF GREENHOUSE GAS [GHG] EMISSIONS, RELATED TO ENERGY GENERATION, TRANSMISSION OR DISTRIBUTION
    • Y02E10/00—Energy generation through renewable energy sources
    • Y02E10/70—Wind energy
    • Y02E10/72—Wind turbines with rotation axis in wind direction

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  • Engineering & Computer Science (AREA)
  • General Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • Life Sciences & Earth Sciences (AREA)
  • Sustainable Development (AREA)
  • Sustainable Energy (AREA)
  • Chemical & Material Sciences (AREA)
  • Combustion & Propulsion (AREA)
  • Rolling Contact Bearings (AREA)

Abstract

To provide a double-row tapered roller bearing which equalizes applied loads in double rows, and obtains a long life as a whole bearing under a load condition in which an axial load mainly acts from one direction due to the optimization of a contact angle, and an axial load may act in a reverse direction.SOLUTION: This bearing is a front-face combination double-row tapered roller bearing. A contact angle θA of an applied load row A of an axial load is larger than a contact angle θB of a non-load side row B. A difference between the contact angles θA, θB of both the rows A, B is equal to 15° or larger. Either of, or both a roller length and a roller diameter of a roller at the non-load side row B are longer and larger than a roller length and a roller diameter of a roller at the applied load row A. The contact angle θA of the 25°≤applied load row A may be ≤35°, and the contact angle θB of the 5°≤non-load side row B may be ≤15°.SELECTED DRAWING: Figure 1

Description

この発明は、左右列非対称設計の複列円すいころ軸受、例えば、風力発電装置主軸用軸受等の複列円すいころ軸受に関する。 This invention relates to a double-row tapered roller bearing with an asymmetric left-right design, such as a double-row tapered roller bearing for the main shaft of a wind power generator.

調心輪つき円すいころ軸受として、複列円すいころ軸受の左右列のころ長さ、ころ径等を互いに異ならせることにより、アキシアルを負荷する列の負荷容量を高めることが提案されている（例えば、特許文献１）。 It has been proposed to increase the load capacity of the row carrying the axial load by making the roller lengths and roller diameters of the left and right rows of a double-row tapered roller bearing with aligning rings different from each other (for example, Patent Document 1).

特開２００６－１７７４４６号公報JP 2006-177446 A

正面組合せの複列円すいころ軸受と円筒ころ軸受とを組み合わせる配置は、風車主軸用の軸受装置としてよく採用されている。
複列円すいころ軸受の強度不足の要因の一つは、アキシアル荷重により両列のころに不均衡負荷が発生し、両列の仕様が同じである場合、アキシアル荷重が主に負荷される側の列である荷重負荷列が先に疲労限度に至ると考えられる。そのため、通常の設計思想では荷重負荷列の負荷容量を大きくする。 An arrangement in which a face-to-face combination of a double row tapered roller bearing and a cylindrical roller bearing is often used as a bearing device for a wind turbine main shaft.
One of the reasons for the lack of strength of double row tapered roller bearings is that the axial load causes an unbalanced load on both rows of rollers, and if the specifications of both rows are the same, it is thought that the row on which the axial load is mainly applied, the load-bearing row, will reach its fatigue limit first. Therefore, the usual design concept is to increase the load capacity of the load-bearing row.

しかし、風車の風荷重は一定ではないため、非負荷側列に大きな荷重が負荷されることがあり、非負荷側列の高負荷容量も必要となる。そのため、従来の設計思想では、非負荷側列の負荷容量が不足し、非負荷側列の安全率が基準に対して厳しくなる。 However, because the wind load on a wind turbine is not constant, a large load can be applied to the non-load side row, requiring a high load capacity for the non-load side row. As a result, with conventional design concepts, the load capacity of the non-load side row is insufficient, and the safety factor of the non-load side row becomes stricter than the standards.

この発明の目的は、接触角の適正化により、アキシアル荷重が主に一方向から作用し、また逆方向にもアキシアル荷重が作用する場合が生じる荷重条件下で、両列の負荷荷重を均衡化し、軸受全体として長寿命化が達成できる複列円すいころ軸受を提供することである。 The object of this invention is to provide a double row tapered roller bearing that can balance the load on both rows and achieve a long life for the entire bearing under load conditions where the axial load acts mainly from one direction and also acts in the opposite direction by optimizing the contact angle.

この発明の複列円すいころ軸受は、正面組合せの複列円すいころ軸受であって、
アキシアル荷重が主に負荷される側の列である荷重負荷列の接触角が、反対側の列である非負荷側列の接触角よりも大きく、かつ両列の接触角の差が１５°以上である。
なお、前記「主に」とは、アキシアル荷重の方向が定まっていればその方向を、アキシアル荷重の方向が変動することがある場合は、多くの時間アキシアル荷重が作用する方向を示す。 The double row tapered roller bearing of the present invention is a face-to-face combination double row tapered roller bearing,
The contact angle of the loaded row, which is the row on which the axial load is mainly applied, is larger than the contact angle of the non-loaded row, which is the row on the opposite side, and the difference in contact angle between the two rows is 15° or more.
In addition, the above-mentioned "mainly" refers to the direction of the axial load if the direction of the axial load is fixed, and refers to the direction in which the axial load acts most of the time if the direction of the axial load varies.

この構成によると、荷重負荷列の接触角が非負荷側列の接触角よりも大きいため、荷重負荷列のアキシアル荷重の負荷能力が高まる。その反面、荷重負荷列のラジアル荷重の負荷能力が低減する。複列円すいころ軸受では、作用する荷重は一般的に、アキシアル荷重よりもラジアル荷重の方が大きく、ラジアル荷重は両列で負荷することになる。そのため複列円すいころ軸受は、軸受全体としてアキシアル荷重とラジアル荷重との負荷能力の均衡化を図る必要がある。
この場合に、両列の接触角の差を１５°以上と大きくしたため、荷重負荷列の接触角をある程度大きくしながら、非負荷側列の接触角を十分に小さくできる。非負荷側列の接触角が小さくなることで、非負荷側列のラジアル荷重の負荷能力が高まり、荷重負荷列のラジアル荷重の負荷能力の低減が補える。これにより、アキシアル荷重とラジアル荷重の両方を考慮すると、両列の負荷荷重が均衡化し、両列の寿命が均衡化して軸受全体の長寿命化が達成できる。 With this configuration, the contact angle of the loaded row is larger than the contact angle of the non-loaded row, so the load capacity of the loaded row for axial loads is increased. Conversely, the load capacity of the loaded row for radial loads is reduced. In double-row tapered roller bearings, the load acting is generally greater for radial loads than for axial loads, and the radial loads are borne by both rows. For this reason, it is necessary for double-row tapered roller bearings to balance the load capacity of the bearing as a whole between axial loads and radial loads.
In this case, the difference in contact angle between the two rows is set to 15° or more, so the contact angle of the non-loaded row can be made sufficiently small while the contact angle of the loaded row is made relatively large. The smaller contact angle of the non-loaded row increases the radial load capacity of the non-loaded row, compensating for the reduced radial load capacity of the loaded row. As a result, when both axial and radial loads are taken into consideration, the loads of the two rows are balanced, and the lives of the two rows are balanced, thereby achieving a long life for the entire bearing.

この発明において、非負荷側列のころのころ長さところ径のいずれか一方または両方が、荷重負荷列のころのころ長さところ径よりも大きくてもよい。
複列円すいころ軸受は、荷重の方向や大きさが大きく変動する条件で使用される場合がある。例えば、風力発電装置主軸用軸受として使用する場合、風車の風荷重は大きく変化するため、非負荷側列に大きな荷重が負荷される場合がある。このような場合に、非負荷側列のころのころ長さところ径のいずれか一方または両方が、荷重負荷列のころのころ長さところ径よりも大きく形成されていると、非負荷側列に大きなアキシアル荷重とラジアル荷重が負荷されても、安全率の基準を満足させることが容易となる。 In the present invention, either or both of the roller length and roller diameter of the rollers in the non-loaded row may be greater than the roller length and roller diameter of the rollers in the loaded row.
Double-row tapered roller bearings are sometimes used under conditions where the direction and magnitude of the load vary greatly. For example, when used as a bearing for the main shaft of a wind power generator, the wind load of the wind turbine varies greatly, so a large load may be applied to the non-load side row. In such cases, if either or both of the roller length and roller diameter of the rollers in the non-load side row are formed larger than the roller length and roller diameter of the rollers in the loaded row, it becomes easy to satisfy the safety factor standards even if a large axial load and radial load are applied to the non-load side row.

この発明において、両列の接触角の差を１５°以上するだけでなく、各列の接触角につき、次の条件を充足することがより好ましい。
２５°≦荷重負荷列の接触角≦３５°
５°≦非負荷側列の接触角≦１５°
荷重負荷列の接触角が２５°以上であり、非負荷側列の接触角が１５°以下であると、両列の軸受中心軸上における両列の作用点間距離Ｍが、左右対称設計の複列円すいころ軸受の５０％～７５％程度になる。そのため、非負荷側列のラジアル荷重の負荷割合が、左右対称設計の複列円すいころ軸受よりも高くなる。このように非負荷側列のラジアル荷重の負荷割合が高くなることで、アキシアル荷重が一方向に偏って作用する使用条件下で、対称設計よりも両列の荷重の均等化が実現できる。
荷重負荷列の接触角が３５°以上であると、荷重負荷列のラジアル荷重の負荷能力が低減するため、好ましくない。また、非負荷側列の接触角が５°以下であると、アキシアル荷重が反対方向に作用した場合に、アキシアル荷重の負荷能力が不足する。 In the present invention, it is more preferable that not only is the difference between the contact angles of both rows set to 15° or more, but also that the contact angles of each row satisfy the following condition.
25°≦Contact angle of the load application row≦35°
5°≦non-load side row contact angle≦15°
When the contact angle of the loaded row is 25° or more and the contact angle of the non-loaded row is 15° or less, the distance M between the application points of both rows on the bearing center axis is about 50% to 75% of that of a symmetrically designed double-row tapered roller bearing. Therefore, the load ratio of the radial load on the non-loaded row is higher than that of a symmetrically designed double-row tapered roller bearing. By increasing the load ratio of the radial load on the non-loaded row in this way, it is possible to achieve more equal load distribution on both rows than with a symmetrical design under operating conditions where the axial load acts biased in one direction.
If the contact angle of the loaded row is 35° or more, the radial load capacity of the loaded row is reduced, which is not preferable. Also, if the contact angle of the non-loaded row is 5° or less, the axial load capacity is insufficient when an axial load acts in the opposite direction.

この発明において、両例のころピッチ円直径の比（ＰＣＤＡ／ＰＣＤＢ）が、
０．９≦（ＰＣＤＡ／ＰＣＤＢ）≦１．１
であってもよい。
両列で接触角を異ならせると、両列の間で必要な中つばの高さに差が生じる。この中つばの高さの差を抑えるため、両列のピッチ円直径の比（ＰＣＤＡ／ＰＣＤＢ）を定量的に定めておくことが好ましく、ピッチ円直径の比（ＰＣＤＡ／ＰＣＤＢ）を、上記の、
０．９≦（ＰＣＤＡ／ＰＣＤＢ）≦１．１
となる範囲に設定することで、両列間で中つばの高さの差が大きくなり過ぎることを抑制できる。 In the present invention, the ratio of the roller pitch circle diameters ( PCDA / PCDB ) of the two examples is:
0.9≦( PCDA / PCDB )≦1.1
may be also possible.
If the contact angles of the two rows are made different, a difference in the height of the central flange required between the two rows will occur. In order to suppress this difference in the height of the central flange, it is preferable to quantitatively determine the ratio of the pitch circle diameters of the two rows ( PCDA / PCDB ) .
0.9≦( PCDA / PCDB )≦1.1
By setting the range so as to satisfy the above condition, it is possible to prevent the difference in height of the center brims between the two rows from becoming too large.

この発明において、内輪が両列の軌道面間に中つばを有する場合、この中つばの厚みＴＨが、
ＴＨ≦０．１５×内輪幅
であることが好ましい。
中つばは、アキシアル荷重が作用する場合にころの軸方向の移動を抑えるためにある程度の厚みＴＨ必要であるが、中つばの厚みＴＨが０．１５×内輪幅の範囲を超えて大きくなると、中つばによって軌道面幅、ころ長さが無駄に小さくなる。 In the present invention, when the inner ring has a center rib between the raceway surfaces of both rows, the thickness TH of the center rib is:
It is preferable that TH≦0.15×the inner ring width.
The inner rib needs to have a certain thickness TH to suppress axial movement of the rollers when an axial load is applied, but if the thickness TH of the inner rib exceeds the range of 0.15 x the inner ring width, the raceway width and roller length will be unnecessarily reduced by the inner rib.

この発明の風力発電装置主軸用軸受は、この発明の前記のうちのいずれかの構成の複列円すいころ軸受とされる。
風力発電装置主軸用軸受の場合、風車の風荷重が時によって大きく変動するため、この発明の複列円すいころ軸受の構成であることによる作用，効果が、効果的に発揮される。 The bearing for a main shaft of a wind turbine generator according to the present invention is a double row tapered roller bearing having any one of the above-mentioned configurations according to the present invention.
In the case of a bearing for a main shaft of a wind power generator, the wind load of the wind turbine varies greatly from time to time, so the action and effect of the double row tapered roller bearing configuration of the present invention are effectively exerted.

この発明の複列円すいころ軸受は、正面組合せの複列円すいころ軸受であって、アキシアル荷重が主に負荷される側の列である荷重負荷列の接触角が、反対側の列である非負荷側列の接触角よりも大きく、かつ両列の接触角の差が１５°以上であるため、接触角の適正化により、アキシアル荷重が主に一方向から作用し、かつ逆方向にもアキシアル荷重が作用する場合が生じる荷重条件下で、両列の負荷荷重を均衡化し、軸受全体として長寿命化が達成できる。 The double row tapered roller bearing of this invention is a face-to-face combination double row tapered roller bearing in which the contact angle of the loaded row, which is the row that is primarily subjected to axial load, is greater than the contact angle of the non-loaded row, which is the row on the opposite side, and the difference in contact angle between the two rows is 15° or more. By optimizing the contact angle, the loads of both rows can be balanced under loading conditions where the axial load acts primarily from one direction and there are also cases where the axial load acts in the opposite direction, thereby achieving a long life for the bearing as a whole.

また、この発明の風力発電装置主軸用軸受は、この発明の複列円すいころ軸受からなるため、風車の風荷重が大きく変動しても、両列の負荷荷重を均衡化し、軸受全体として長寿命化が達成できる。 In addition, because the wind power generator main shaft bearing of this invention is made of the double row tapered roller bearing of this invention, even if the wind load of the wind turbine fluctuates greatly, the load on both rows can be balanced, achieving a long life for the bearing as a whole.

この発明の第１の実施形態に係る複列円すいころ軸受の部分断面図である。1 is a partial cross-sectional view of a double-row tapered roller bearing according to a first embodiment of the present invention. 同複列円すいころ軸受の各部の寸法を示す断面図である。FIG. 2 is a cross-sectional view showing dimensions of each part of the double row tapered roller bearing. 同複列円すいころ軸受における両列のころの部分破断正面図である。FIG. 2 is a partially cutaway front view of both rows of rollers in the double row tapered roller bearing. 従来の対称品および第１の実施形態に係る設計(1)、設計(2)の複列円すいころ軸受の荷重負荷列における疲労荷重での転動体荷重分布のシミュレーション例を示す説明図である。FIG. 11 is an explanatory diagram showing an example of a simulation of rolling element load distribution under fatigue load in a loaded row of a double-row tapered roller bearing of a conventional symmetric product and of designs (1) and (2) according to the first embodiment. 従来の対称品および第１の実施形態に係る設計(1)、設計(2)の複列円すいころ軸受の非負荷側列における疲労荷重での転動体荷重分布のシミュレーション例を示す説明図である。FIG. 11 is an explanatory diagram showing a simulation example of rolling element load distribution under fatigue load in the non-loaded row of double-row tapered roller bearings of a conventional symmetric product and designs (1) and (2) according to the first embodiment. この発明の他の実施形態に係る複列円すいころ軸受の部分断面図である。FIG. 4 is a partial cross-sectional view of a double row tapered roller bearing according to another embodiment of the present invention. 同複列円すいころ軸受と円筒ころ軸受とを組み合わせた軸受装置の断面図である。4 is a cross-sectional view of a bearing device combining the double row tapered roller bearing and a cylindrical roller bearing. FIG. 第１の実施形態に係る複列円すいころ軸受を風力発電装置主軸用軸受として用いた風力発電装置の断面図である。1 is a cross-sectional view of a wind power generator in which a double-row tapered roller bearing according to a first embodiment is used as a bearing for a main shaft of the wind power generator.

＜第１の実施形態＞
この発明の第１の実施形態を図１、図２と共に説明する。
この複列円すいころ軸受１は、左右の列Ａ，Ｂの接触角θＡ，θＢが互いに異なる左右非対称型で、かつ正面組み合わせ型とされている。この複列円すいころ軸受１は、内輪２と、外輪３と、これら内外輪２，３の複列の軌道面４，５間にそれぞれ介在した複列のころ６，７と、各列のころ６，７をそれぞれ保持する２つの保持器８，９とで構成される。各列のころ６，７は円すいころであり、軸受幅方向の中央側が大径とされている。 First Embodiment
A first embodiment of the present invention will be described with reference to FIGS.
This double row tapered roller bearing 1 is an asymmetrical type with different contact angles θ A and θ B for the left and right rows A and B, and is a face-to-face assembled type. This double row tapered roller bearing 1 comprises an inner ring 2, an outer ring 3, double row rollers 6, 7 interposed between double row raceway surfaces 4, 5 of the inner and outer rings 2, 3, and two cages 8, 9 which respectively retain the rollers 6, 7 of each row. The rollers 6, 7 of each row are tapered rollers, with the larger diameter at the center in the bearing width direction.

この実施形態では、内輪２が複列の軌道面４，４を持つ単独の部材で構成され、外輪３が単列外輪３Ａ，３Ｂで構成されている正面組み合わせ型である。 In this embodiment, the inner ring 2 is composed of a single member having double-row raceway surfaces 4, 4, and the outer ring 3 is a front combination type composed of single-row outer rings 3A, 3B.

内輪２の両列の軌道面４，４は、軸受幅方向の中央側が大径となるテーパ面とされ、内輪２の外周面には、両列の軌道面４，４の間に中つば１１が設けられている。中つば１１は、幅方向の中央付近から非負荷列側列端に至る範囲の外周面部が、非負荷列側列Ｂ側に次第に低くなるテーパ面状とされている。各軌道面４，４の軸受端部側には隣接して端つば１２，１２が設けられている。 The raceway surfaces 4, 4 of both rows of the inner ring 2 are tapered with a larger diameter toward the center in the bearing width direction, and a center rib 11 is provided on the outer circumferential surface of the inner ring 2 between the raceway surfaces 4, 4 of both rows. The center rib 11 has a tapered outer circumferential surface portion in the range from near the center in the width direction to the non-loaded row side row end that gradually becomes lower toward the non-loaded row side row B side. End ribs 12, 12 are provided adjacent to the bearing end side of each raceway surface 4, 4.

外輪３の両列の軌道面５，５は、軸受幅方向の中央側が大径となるテーパ面であり、内輪の軌道面４，４とは、ころ６，７の外周面の傾斜角度分だけ相違している。外輪３はつばを有しておらず、両列の外輪３Ａ，３Ｂの間に外輪間座３Ｃが介在している。 The raceway surfaces 5, 5 of both rows of the outer ring 3 are tapered surfaces with a larger diameter toward the center in the bearing width direction, and differ from the raceway surfaces 4, 4 of the inner ring by the inclination angle of the outer peripheral surfaces of the rollers 6, 7. The outer ring 3 does not have a flange, and an outer ring spacer 3C is interposed between the outer rings 3A, 3B of both rows.

この複列円すいころ軸受１は、図の左側の列Ａが荷重負荷列、右側の列Ｂが非負荷側列であり、内外輪２，３の軌道面４，５の傾斜角度の差、およびころ６，７のテーパ角度により、荷重負荷列Ａの接触角θＡが、非負荷側列Ｂの接触角θＢよりも大きく形成されている。荷重負荷列Ａは、内輪回転の場合内輪２に作用するアキシアル荷重F、あるいは外輪回転の場合外輪３Ａに作用するアキシアル荷重Ｇが主に負荷される側の列である。非負荷側列Ｂは荷重負荷列Ａと反対側の列である。 In this double row tapered roller bearing 1, row A on the left side of the figure is the loaded row, and row B on the right side is the non-loaded row, and due to the difference in the inclination angles of the raceway surfaces 4, 5 of the inner and outer rings 2, 3 and the taper angles of the rollers 6, 7, the contact angle θ A of the loaded row A is larger than the contact angle θ B of the non-loaded row B. The loaded row A is the row that is primarily subjected to the axial load F acting on the inner ring 2 when the inner ring rotates, or the axial load G acting on the outer ring 3A when the outer ring rotates. The non-loaded row B is the row opposite the loaded row A.

荷重負荷列Ａの接触角θＡと非負荷側列Ｂの接触角θＢとの差は、１５°以上とされている。また、各列Ａ，Ｂの接触角θＡ，θＢは、次式を満たす範囲とされている。
２５°≦（荷重負荷列Ａの接触角θＡ）≦３５°
５°≦（非負荷側列Ｂの接触角θＢ）≦１５° The difference between the contact angle θ A of the loaded row A and the contact angle θ B of the non-loaded row B is set to be 15° or more. The contact angles θ A and θ B of the rows A and B are set to a range that satisfies the following formula:
25°≦(contact angle θ A of load application row A )≦35°
5°≦(contact angle θ B of non-load side row B)≦15°

両列Ａ，Ｂのころ６，７のころ長さおよびころ径については、いずれも、非負荷側列Ｂのころ７のころ長さＬＢ（図２）およびころ径ＤＢの方が、荷重負荷列Ａのころ長さＬＡおよびころ径ＤＡ（図示せず）よりも大きくされている。なお、非負荷側列Ｂのころ７のころ長さＬＢおよびころ径ＤＢのいずれか一方だけが荷重負荷列Ａより大きくされてもよい。ころ６，７の端部が外周面に面取部が設けられている場合、ころ長さＬＡ，ＬＢの比較について、面取部の幅を含む長さ同士で比較しても、面取部の幅を含まない長さ同士で比較してもよい。ころ径ＤＡ，ＤＢは、各列のころ６，７の最大径である。
Ａ列ころ６の大端面中央にセンター穴６－１(図３参照)を設けている。Ｂ列ころ７の大端面中央にセンター穴７－１、小端面に円環状の識別印７－２を設けている。 Regarding the roller lengths and roller diameters of the rollers 6, 7 in both rows A, B, the roller length L B (FIG. 2) and roller diameter D B of the roller 7 in the non-loaded row B are greater than the roller length L A and roller diameter D A (not shown) of the loaded row A. Note that only one of the roller length L B and roller diameter D B of the roller 7 in the non-loaded row B may be greater than those of the loaded row A. When the ends of the rollers 6, 7 are provided with chamfered portions on their outer circumferential surfaces, the roller lengths L A , L B may be compared either with or without the width of the chamfered portion. The roller diameters D A , D B are the maximum diameters of the rollers 6, 7 in each row.
A center hole 6-1 (see FIG. 3) is provided in the center of the large end face of the A-row rollers 6. A center hole 7-1 is provided in the center of the large end face of the B-row rollers 7, and an annular identification mark 7-2 is provided on the small end face.

両列Ａ，Ｂのころ配列のピッチ円直径ＰＣＤＡ，ＰＣＤＢは、その比（ＰＣＤＡ／ＰＣＤＢ）について、
０．９≦（ＰＣＤＡ／ＰＣＤＢ）≦１．１とされている。 The pitch circle diameters PCDA and PCDB of the roller arrangements in both rows A and B are expressed as follows, with the ratio ( PCDA / PCDB ) being:
It is set to 0.9≦( PCDA / PCDB )≦1.1.

内輪２の両列Ａ，Ｂの軌道面４．４の間の中つば１１の厚みＴＨについては、
ＴＨ≦０．１５×内輪幅Ｗ
とされている。
前記中つば１１のＡ列側の高さＨＡとＢ列側の高さＨＢについては、
ＨＡ≧ＨＢ （ＨＡ＝ＨＢが望ましい）
とされている。
前記中つば１１のＡ列ころとの接触点の高さＣＡとＢ列ころとの接触点の高さＣＢについては、
｜ＣＡ－ＣＢ|≦３mm
とされている。 The thickness TH of the center rib 11 between the raceway surfaces 4.4 of both rows A and B of the inner ring 2 is as follows:
TH≦0.15×inner ring width W
It is said that.
The height HA of the center flange 11 on the A row side and the height HB on the B row side are as follows:
HA≧HB (HA=HB is preferable)
It is said that.
Regarding the height CA of the contact point of the center rib 11 with the A-row rollers and the height CB of the contact point with the B-row rollers,
| CA-CB |≦3mm
It is said that.

＜作用、効果、詳細構成＞
この構成によると、荷重負荷列Ａの接触角θＡ（図１）が非負荷側列Ｂの接触角θＢよりも大きいため、荷重負荷列Ａのアキシアル荷重の負荷能力が高まる。その反面、荷重負荷列Ａのラジアル荷重の負荷能力が低減する。複列円すいころ軸受では、作用する荷重は一般的に、アキシアル荷重よりもラジアル荷重の方が大きく、ラジアル荷重は両列で負荷することになる。そのため複列円すいころ軸受は、軸受全体としてアキシアル荷重とラジアル荷重との負荷能力の均衡化を図る必要がある。
この場合に、両列Ａ，Ｂの接触角θＡ，θＢの差を１５°以上と大きくしたため、荷重負荷列Ａの接触角θＡをある程度大きくしながら、非負荷側列Ｂの接触角θＢを十分に小さくできる。非負荷側列Ｂの接触角θＢが小さくなることで、非負荷側列Ｂのラジアル荷重の負荷能力が高まり、荷重負荷列Ａのラジアル荷重の負荷能力の低減が補える。これにより、アキシアル荷重とラジアル荷重の両方を考慮すると、両列Ａ，Ｂの負荷荷重が均衡化し、両列Ａ，Ｂの寿命が均衡化して軸受全体の長寿命化が達成できる。 <Action, effect, detailed configuration>
With this configuration, the contact angle θ A (FIG. 1) of the loaded row A is larger than the contact angle θ B of the non-loaded row B, so the load capacity of the loaded row A for axial loads is increased. Conversely, the load capacity of the loaded row A for radial loads is reduced. In double-row tapered roller bearings, the applied load is generally greater for radial loads than for axial loads, and the radial load is borne by both rows. For this reason, it is necessary for double-row tapered roller bearings to balance the load capacities of the axial load and radial load for the entire bearing.
In this case, the difference between the contact angles θ A and θ B of rows A and B is set to 15° or more, so the contact angle θ A of the loaded row A can be increased to a certain extent while the contact angle θ B of the non-loaded row B can be made sufficiently small. The smaller contact angle θ B of the non-loaded row B increases the radial load capacity of the non-loaded row B, compensating for the reduced radial load capacity of the loaded row A. As a result, when both axial and radial loads are taken into consideration, the applied loads of rows A and B are balanced, and the lives of rows A and B are balanced, thereby achieving a long life for the entire bearing.

また、非負荷側列Ｂのころ７のころ長さＬＢが、荷重負荷列Ａのころのころ長さＬＡよりも大きい。そのため、次の利点が得られる。
複列円すいころ軸受は、荷重の方向や大きさが大きく変動する条件で使用される場合がある。例えば、風力発電装置主軸用軸受として使用する場合、風車の風荷重は大きく変化するため、非負荷側列に大きな荷重が負荷される場合がある。このような場合に、非負荷側列Ｂのころ７のころ長さＬＢが、荷重負荷列Ａのころ６のころ長さＬＡよりも大きく形成されていると、非負荷側列Ｂに大きな荷重が負荷されても、安全率の基準を満足させることが容易となる。
ころ端面にセンター穴６－１，７－１、識別印７－２を設けることで、異種ころ組みの防止ができる。 In addition, the roller length L B of the rollers 7 in the non-loaded row B is greater than the roller length L A of the rollers in the loaded row A. Therefore, the following advantages are obtained.
Double row tapered roller bearings are sometimes used under conditions where the direction and magnitude of the load vary greatly. For example, when used as a bearing for the main shaft of a wind power generator, the wind load of the wind turbine varies greatly, and a large load may be applied to the non-load side row. In such a case, if the roller length LB of rollers 7 in the non-load side row B is made longer than the roller length LA of rollers 6 in the load side row A, it becomes easy to satisfy the safety factor standard even if a large load is applied to the non-load side row B.
By providing center holes 6-1, 7-1 and identification marks 7-2 on the roller end faces, it is possible to prevent assembling different types of rollers.

両列の接触角θＡ，θＢは、その差を１５°以上するだけでなく、前記の条件、
２５°≦荷重負荷列Ａの接触角θＡ≦３５°
５°≦非負荷側列Ｂの接触角θＢ≦１５°
を充足しているため、アキシアル荷重が一方向に偏って作用する使用条件下で、対称設計よりも両列の荷重の均等化が実現できるなどの効果が得られる。
荷重負荷列Ａの接触角θＡが２５°以上であり、非負荷側列の接触角が１５°以下であると、両列Ａ，Ｂの軸受中心軸Ｏ上における両列の作用点ＰＡ，ＰＢ間の距離Ｍが、左右対称設計の複列円すいころ軸受の５０～７５％程度になる。そのため、非負荷側列Ｂのラジアル荷重の負荷割合が、左右対称設計の複列円すいころ軸受よりも高くなる。このように非負荷側列Ｂのラジアル荷重の負荷割合が高くなることで、アキシアル荷重が一方向に偏って作用する使用条件下で、対称設計よりも両列の荷重の均等化が実現できる。
荷重負荷列Ａの接触角θＡ角が３５°を超えると、荷重負荷列Ａのラジアル荷重の負荷能力が低減するため、好ましくない。また、非負荷側列Ｂの接触角θＢが５°未満であると、アキシアル荷重が反対方向に作用した場合に、アキシアル荷重の負荷能力が不足する。 The difference between the contact angles θ A and θ B of the two rows is set to 15° or more, and the above condition is satisfied.
25°≦contact angle θ A of load application row A ≦35°
5°≦Contact angle θ B of non-load side row B ≦15°
Since this satisfies the above requirements, it is possible to achieve effects such as equalizing the load on both rows more than with a symmetrical design under operating conditions where the axial load acts biased in one direction.
When the contact angle θ A of the loaded row A is 25° or more and the contact angle of the non-loaded row is 15° or less, the distance M between the application points P A and P B of both rows A and B on the bearing center axis O is about 50 to 75% of that of a symmetrically designed double-row tapered roller bearing. Therefore, the load ratio of the radial load of the non-loaded row B is higher than that of a symmetrically designed double-row tapered roller bearing. By increasing the load ratio of the radial load of the non-loaded row B in this way, it is possible to achieve more equal loads on both rows than with a symmetrical design under operating conditions where the axial load acts biased in one direction.
If the contact angle θA of the load bearing row A exceeds 35°, the radial load bearing capacity of the load bearing row A decreases, which is not preferable. Also, if the contact angle θB of the non-load bearing row B is less than 5°, the axial load bearing capacity is insufficient when an axial load acts in the opposite direction.

また、接触角θＡ，θＢの大きさにより両列Ａ，Ｂでの中つば１１の両側の高さの差を抑えるためには、両列のピッチ円直径ＰＣＤＡ，ＰＣＤＢの比（ＰＣＤＡ／ＰＣＤＢ）を定量的にすることが好ましい。
この実施形態では、前記のように両例Ａ，Ｂのころピッチ円直径の比（ＰＣＤＡ／ＰＣＤＢ）が、
０．９≦（ＰＣＤＡ／ＰＣＤＢ）≦１．１
とされている。
このため、両列Ａ，Ｂ間で中つば１１の両側の高さの差が大きくなり過ぎることを抑制できる。 In order to suppress the difference in height between the two sides of the center flange 11 in both rows A and B depending on the contact angles θ A and θ B , it is preferable to make the ratio of the pitch circle diameters PCDA and PCDB of both rows ( PCDA / PCDB ) quantitative.
In this embodiment, as described above, the ratio of the roller pitch circle diameters ( PCDA / PCDB ) of the two examples A and B is
0.9≦( PCDA / PCDB )≦1.1
It is said that.
Therefore, it is possible to prevent the difference in height between the two sides of the center brim 11 between the two rows A and B from becoming too large.

中つば１１の厚みＴＨに関しては、前記のように、
ＴＨ≦０．１５×内輪幅Ｗ
であることが好ましい。
中つば１１は、アキシアル荷重が作用する場合にころ６，７の軸方向の移動を抑えるためにある程度の厚みＴＨ必要であるが、中つば６，７の厚みＴＨが０．１５×内輪幅Ｗの範囲を超えて大きくなると、中つば１１によって軌道面幅、ころ長さが無駄に小さくなる。 Regarding the thickness TH of the middle flange 11, as described above,
TH≦0.15×inner ring width W
It is preferable that:
The center rib 11 needs to have a certain thickness TH to suppress axial movement of the rollers 6, 7 when an axial load is applied, but if the thickness TH of the center rib 6, 7 exceeds the range of 0.15 × inner ring width W, the center rib 11 will unnecessarily reduce the raceway width and roller length.

＜計算結果＞
表１に、同一サイズで各設計での安全率と基本定格寿命を比較して示す。表中の非対称設計(1),(2)は図１，２の実施形態品を、非対称設計（３）は両列Ａ，Ｂの接触角θＡ，θＢの差が１５°未満となる比較例を示す。各列のころ長さ、ころ径、安全率については、対称設計品（両列の接触角が同じである従来品）の値を、それぞれころ長さＬ、ころ径ＤＷ、安全率Ｓ０Ａ、Ｓ０Ｂで示し、非対称設計(1), (2)の各値では、従来品の各値の倍率で示している。基本定格寿命については、対称設計品のＡ列（アキシアル荷重負荷列）の値をＬＡで示し、各設計における各列の基本定格寿命及び総合寿命では、ＬＡの倍率で示している。
表1に示した結果より、実施形態品となる非対称設計(1),(2)では、対称品の約２倍の基本定格寿命を達成し、安全率も対称品とほぼ同程度であることが分かる。 <Calculation results>
Table 1 shows a comparison of the safety factor and basic rating life of each design of the same size. In the table, asymmetric designs (1) and (2) are the embodiments shown in Figs. 1 and 2, and asymmetric design (3) is a comparative example in which the difference in contact angles θ A and θ B between rows A and B is less than 15°. The roller length, roller diameter, and safety factor of each row are shown as roller length L, roller diameter D W , and safety factor S0A and S0B, respectively, for the symmetric design (conventional product in which the contact angles of both rows are the same), and the values of asymmetric designs (1) and (2) are shown as the multiplication factor of the conventional product. For the basic rating life, the value of row A (axial load row) of the symmetric design is shown as LA, and the basic rating life and overall life of each row in each design are shown as the multiplication factor of LA.
From the results shown in Table 1, it can be seen that the asymmetric designs (1) and (2), which are the embodiment products, achieve a basic rated life that is approximately twice that of the symmetric product, and the safety factor is also approximately the same as that of the symmetric product.

図４、図５は、アキシアル荷重負荷列Ａおよびアキシアル非負荷側列Ｂにおける疲労荷重での軸受全周の転動体荷重分布をそれぞれ示す。対称品では、アキシアル荷重負荷列Ａの転動体荷重分布曲線（図４（Ａ））の径が、アキシアル非負荷側列Ｂにおける転動体荷重分布曲線（図５（Ａ））に比べて大幅に大きくなっている。これに対し、実施形態品（非対称品）では、アキシアル荷重負荷列Ａの転動体荷重分布曲線（図４（Ｂ）（Ｃ））とアキシアル非負荷側列Ｂの転動体荷重分布曲線（図５（Ｂ）、（Ｃ））の径とに大きな差が生じておらず、両列Ａ、Ｂの転動体荷重分が均衡化していることが分かる。
なお、実施形態品（非対称品）は、アキシアル非負荷側列Ｂでは転動体荷重分布曲線（図５（Ｂ）、（Ｃ））がアキシアル荷重負荷列Ａの転動体荷重分布曲線（図５（Ａ））よりも若干大きいが、対称品のアキシアル荷重負荷列Ａにおける転動体荷重分布曲線（図４（Ａ））の径が大幅に大きくなっているため、両列Ａ、Ｂの全体として、実施形態品（非対称品）の方が転動体荷重分布曲線が小さくなっている。 4 and 5 respectively show the rolling element load distribution over the entire bearing circumference under fatigue load in the axial load row A and the axial non-load side row B. In the symmetrical product, the diameter of the rolling element load distribution curve of the axial load row A (FIG. 4(A)) is significantly larger than the rolling element load distribution curve of the axial non-load side row B (FIG. 5(A)). In contrast, in the embodiment product (asymmetric product), there is no significant difference in diameter between the rolling element load distribution curve of the axial load row A (FIGS. 4(B)(C)) and the rolling element load distribution curve of the axial non-load side row B (FIGS. 5(B)(C)), and it can be seen that the rolling element loads of both rows A and B are balanced.
In the embodiment product (asymmetric product), the rolling element load distribution curve (Figures 5B and 5C) in the axial non-loaded row B is slightly larger than the rolling element load distribution curve (Figure 5A) in the axial load loaded row A. However, since the diameter of the rolling element load distribution curve (Figure 4A) in the axial load loaded row A of the symmetric product is significantly larger, the rolling element load distribution curve of the embodiment product (asymmetric product) is smaller overall for both rows A and B.

Figure 2024069478000002
Figure 2024069478000002

＜他の実施形態＞
図６は、この発明の他の実施形態を示す。この実施形態において、特に説明する事項の他は、図１～図５と共に説明した第１の実施形態と同様である。第１の実施形態では、内輪２が複列の軌道面４，４を持つ単独の部材で構成されているが、図6の実施形態では、内輪２が２個の単列内輪２Ａ，２Ｂからなる。これに伴い、中つば１１は、２つの単列中つば１１Ａ，１１Ｂで構成される。外輪３は、第１の実施形態と同様に、２個の単列外輪３Ａ，３Ｂで構成されている。このように、２個の単列円すいころ軸受１Ａ，１Ｂで構成されている。同図において、図１の保持器８，９は、図示が省略されている。
このように構成した場合も、第１の実施形態で説明した各作用、効果が得られる。 <Other embodiments>
Fig. 6 shows another embodiment of the present invention. Other than the matters specifically described, this embodiment is similar to the first embodiment described with reference to Figs. 1 to 5. In the first embodiment, the inner ring 2 is composed of a single member having double-row raceway surfaces 4, 4, but in the embodiment of Fig. 6, the inner ring 2 is composed of two single-row inner rings 2A, 2B. Accordingly, the center rib 11 is composed of two single-row center ribs 11A, 11B. The outer ring 3 is composed of two single-row outer rings 3A, 3B, as in the first embodiment. In this way, it is composed of two single-row tapered roller bearings 1A, 1B. In this figure, the cages 8, 9 in Fig. 1 are omitted.
Even in this configuration, the functions and effects described in the first embodiment can be obtained.

＜円筒軸受との組み合わせ例＞
図７は、複列円すいころ軸受１と円筒ころ軸受１５とを組み合わせた軸受装置の一例を示す。この軸受装置は、風車や各種の産業機械の主軸の支持などに適用される。
主体１６の前後部が、複列円すいころ軸受１と、円筒ころ軸受１５とを介してハウジング１７に支持されている。複列円すいころ軸受１は、図6に示す実施形態を用いているが、図１に示す第１の実施形態に係る複列円すいころ軸受１であってもよい。円筒ころ軸受１５は、内輪１８，外輪１９、および円筒ころ２０、および保持器（図示せず）を有する。ハウジング１７は、この例では円筒状の一つの部材で構成されているが、軸１を支持する複列円すいころ軸受１および円筒ころ軸受１５は、それぞれ個別のハウジング（図示せず）に設置されていてもよい。 <Example of combination with cylindrical bearing>
7 shows an example of a bearing device combining a double row tapered roller bearing 1 with a cylindrical roller bearing 15. This bearing device is used for supporting the main shafts of wind turbines and various industrial machines.
The front and rear portions of the main body 16 are supported by a housing 17 via a double row tapered roller bearing 1 and a cylindrical roller bearing 15. The double row tapered roller bearing 1 uses the embodiment shown in Fig. 6, but may be the double row tapered roller bearing 1 according to the first embodiment shown in Fig. 1. The cylindrical roller bearing 15 has an inner ring 18, an outer ring 19, cylindrical rollers 20, and a cage (not shown). In this example, the housing 17 is composed of a single cylindrical member, but the double row tapered roller bearing 1 and the cylindrical roller bearing 15 which support the shaft 1 may each be installed in separate housings (not shown).

＜風力発電装置＞
図８は、この発明の実施形態に係る複列円すいころ軸受１を用いた風力発電装置の一例を示す。支持台２１上に旋回座軸受２２を介してナセル２３のケーシング２３ａが水平旋回自在に設置されている。ナセル２３のケーシング２３ａ内には、軸受ハウジング２４に設置された風力発電装置主軸用軸受２５を介して主軸２６が回転自在に設置され、主軸２６のケーシング２３ａ外に突出した部分に、旋回翼となるブレード２７が取り付けられている。主軸２６の他端は、増速機２８に接続され、増速機２８の出力軸が発電機２９のロータ軸に結合されている。風力発電装置主軸用軸受２５は、図示の例では２個並べて設置してあるが、１個であってもよい。
前記各風力発電装置主軸用軸受２５に図１，２の第１の実施形態、または図6に示す第２の実施形態に係る複列円すいころ軸受１が用いられる。2個の軸受２５，２５のうち、前記複列円すいころ軸受１はどこであってもよい。 <Wind power generation equipment>
8 shows an example of a wind power generator using a double row tapered roller bearing 1 according to an embodiment of the present invention. A casing 23a of a nacelle 23 is installed on a support base 21 via a slewing seat bearing 22 so as to be horizontally rotatable. A main shaft 26 is installed in the casing 23a of the nacelle 23 so as to be rotatable via a wind power generator main shaft bearing 25 installed in a bearing housing 24, and a blade 27 serving as a swivel vane is attached to a part of the main shaft 26 protruding outside the casing 23a. The other end of the main shaft 26 is connected to a speed increaser 28, and the output shaft of the speed increaser 28 is coupled to the rotor shaft of a generator 29. Two wind power generator main shaft bearings 25 are installed side by side in the illustrated example, but only one may be used.
The double row tapered roller bearing 1 according to the first embodiment shown in Fig. 1 and Fig. 2 or the second embodiment shown in Fig. 6 is used for the main shaft bearing 25 of each wind turbine generator. Of the two bearings 25, 25, the double row tapered roller bearing 1 may be located anywhere.

以上、実施形態に基づいてこの発明を実施するための形態を説明したが、今回開示された実施の形態はすべての点で例示であって制限的なものではない。この発明の範囲は上記した説明ではなくて特許請求の範囲によって示され、特許請求の範囲と均等の意味および範囲内でのすべての変更が含まれることが意図される。 The above describes the mode for carrying out the present invention based on the embodiment, but the embodiment disclosed herein is illustrative in all respects and is not restrictive. The scope of the present invention is indicated by the claims, not the above description, and is intended to include all modifications within the meaning and scope of the claims.

１…複列円すいころ軸受
１Ａ，１Ｂ…単列円すいころ軸受
２…内輪、
２Ａ，２Ｂ…単列内輪
３…外輪
３Ａ，３Ｂ…単列外輪
３Ｃ…外輪間座
４．５…軌道面
６．７…ころ
６－１．７－１…ころ端面センター穴
７－２…ころ端面識別印
８，９…保持器
１１…中つば
１１Ａ，１１Ｂ…片列中つば
１２…端つば
１５…円筒ころ軸受
１６…主軸
１７…ハウジング
１８…内輪
１９…外輪
２０…ころ
２１…支持台
２２…旋回座軸受
２３…ナセル
２３a…ケーシング
２４…軸受ハウジング
２５…主軸支持軸受
２６…主軸
２７…ブレード
２８…増速機
２９…発電機
Ａ…アキシアル荷重負荷列
Ｂ…アキシアル荷重非負荷側列
θＡ，θＢ…接触角
ＤＡ，ＤＢ…ころ径
ＬＢ，ＬＢ…ころ長さ
ＰＣＤＡ，ＰＣＤＢ…ピッチ円直径
HA，HB…中つばの高さ
CA，CB…中つばところの接触点の高さ 1... double row tapered roller bearing 1A, 1B... single row tapered roller bearing 2... inner ring,
[0033] 2A, 2B...Single row inner ring 3...Outer ring 3A, 3B...Single row outer ring 3C...Outer ring spacer 4.5...Raceway surface 6.7...Rollers 6-1.7-1...Roller end face center hole 7-2...Roller end face identification mark 8, 9...Cage 11...Middle rib 11A, 11B...Single row middle rib 12...End rib 15...Cylindrical roller bearing 16...Main shaft 17...Housing 18...Inner ring 19...Outer ring 20...Rollers 21...Support base 22...Slewing seat bearing 23...Nacelle 23a...Casing 24...Bearing housing 25...Main shaft support bearing 26...Main shaft 27...Blades 28...Gearbox 29...Generator A...Axial load row B...Axial load non-load row θ A , θ B ...Contact angles D A , D B ...Roller diameters L B , L B ...Roller lengths PCD A , PCD B ...Pitch circle diameter
HA, HB: height of middle brim
CA, CB: Height of contact point between middle rib and roller

Claims (6)

正面組合せの複列円すいころ軸受であって、
アキシアル荷重が主に負荷される側の列である荷重負荷列の接触角が、反対側の列である非負荷側列の接触角よりも大きく、かつ両列の接触角の差が１５°以上である複列円すいころ軸受。 A face-to-face double-row tapered roller bearing,
A double-row tapered roller bearing in which the contact angle of the load-bearing row, which is the row on which the axial load is primarily applied, is larger than the contact angle of the non-load-bearing row, which is the row on the opposite side, and the difference in contact angle between the two rows is 15° or more. 請求項１に記載の複列円すいころ軸受において、前記非負荷側列のころのころ長さところ径のいずれか一方または両方が、前記荷重負荷列のころのころ長さところ径よりも大きい複列円すいころ軸受。 A double row tapered roller bearing as described in claim 1, in which either or both of the roller length and roller diameter of the rollers in the non-load side row are greater than the roller length and roller diameter of the rollers in the load-bearing row. 請求項１または請求項２に記載の複列円すいころ軸受において、
２５°≦荷重負荷列の接触角≦３５°
５°≦非負荷側列の接触角≦１５°
である複列円すいころ軸受。 In the double row tapered roller bearing according to claim 1 or 2,
25°≦Contact angle of the load application row≦35°
5°≦non-load side row contact angle≦15°
A double row tapered roller bearing. 請求項１ないし請求項３のいずれか１項に記載の複列円すいころ軸受において、両例のころピッチ円直径の比（ＰＣＤＡ／ＰＣＤＢ）が、
０．９≦（ＰＣＤＡ／ＰＣＤＢ）≦１．１
である複列円すいころ軸受。 In the double row tapered roller bearing according to any one of claims 1 to 3, the ratio of the roller pitch circle diameters ( PCDA / PCDB ) of the two examples is:
0.9≦( PCDA / PCDB )≦1.1
A double row tapered roller bearing. 請求項１ないし請求項４のいずれか１項に記載の複列円すいころ軸受において、内輪が両列の軌道面の間に中つばを有し，この中つばの厚みＴＨが、
ＴＨ≦０．１５×内輪幅
である複列円すいころ軸受。 5. The double row tapered roller bearing according to claim 1, wherein the inner ring has a center rib between the raceway surfaces of both rows, and the thickness TH of the center rib is:
A double-row tapered roller bearing with TH≦0.15 x inner ring width. 請求項１ないし請求項５のいずれか１項に記載の複列円すいころ軸受である風力発電装置主軸用軸受。 A bearing for a wind turbine main shaft, the bearing being a double-row tapered roller bearing according to any one of claims 1 to 5.
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