WO2021060389A1 - Double row tapered roller bearing - Google Patents

Double row tapered roller bearing Download PDF

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Publication number
WO2021060389A1
WO2021060389A1 PCT/JP2020/036072 JP2020036072W WO2021060389A1 WO 2021060389 A1 WO2021060389 A1 WO 2021060389A1 JP 2020036072 W JP2020036072 W JP 2020036072W WO 2021060389 A1 WO2021060389 A1 WO 2021060389A1
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Prior art keywords
load
row
tapered roller
double
bearing
Prior art date
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PCT/JP2020/036072
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French (fr)
Japanese (ja)
Inventor
偉達 顔
智仁 伊藤
Original Assignee
Ntn株式会社
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Publication date
Priority claimed from JP2020092067A external-priority patent/JP7456851B2/en
Application filed by Ntn株式会社 filed Critical Ntn株式会社
Priority to CN202080067729.4A priority Critical patent/CN114514382A/en
Publication of WO2021060389A1 publication Critical patent/WO2021060389A1/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

Definitions

  • the present invention relates to a double-row tapered roller bearing having an asymmetrical design in the left-right row, for example, a double-row tapered roller bearing such as a bearing for a spindle of a wind power generator.
  • Patent Document 1 As a tapered roller bearing with a centering ring, it has been proposed to increase the load capacity of the row that loads the axial by making the roller length, roller diameter, etc. of the left and right rows of the double-row tapered roller bearing different from each other (for example). , Patent Document 1).
  • An object of the present invention is to balance the load loads in both rows under load conditions in which the axial load acts mainly from one direction and the axial load acts in the opposite direction by optimizing the contact angle. It is an object of the present invention to provide a double-row tapered roller bearing that can achieve a long life as a whole bearing.
  • the double-row tapered roller bearing of the present invention is a front-combination double-row tapered roller bearing.
  • the contact angle of the load-bearing row which is the row on which the axial load is mainly applied, is larger than the contact angle of the non-load side row, which is the opposite row, and the difference between the contact angles of both rows is 15 ° or more.
  • the term "mainly” refers to the direction in which the axial load is applied if the direction of the axial load is determined, and the direction in which the axial load acts for many hours when the direction of the axial load may fluctuate.
  • the contact angle of the load-bearing row is larger than the contact angle of the non-load side row
  • the loading capacity of the axial load of the load-bearing row is increased.
  • the load capacity of the radial load of the load load train is reduced.
  • the acting load is generally larger than the axial load, and the radial load is applied in both rows. Therefore, for double-row tapered roller bearings, it is necessary to balance the load capacity of the axial load and the radial load of the bearing as a whole.
  • the contact angle of the non-loaded side row can be sufficiently reduced while the contact angle of the loaded row is increased to some extent.
  • the load capacity of the radial load of the non-load side row is increased, and the reduction of the radial load capacity of the load load row can be compensated.
  • either or both of the roller lengths and diameters of the rollers in the unloaded side row may be larger than the roller lengths and diameters of the load-loaded row.
  • Double-row tapered roller bearings may be used under conditions where the direction and magnitude of the load fluctuate significantly. For example, when used as a bearing for a spindle of a wind power generator, the wind load of the wind turbine changes significantly, so that a large load may be applied to the non-load side row.
  • the distance M between the action points of both rows on the bearing center axis of both rows is symmetrically designed. It is about 50% to 75% of the double row tapered roller bearings. Therefore, the load ratio of the radial load in the non-load side row is higher than that in the double row tapered roller bearing of the symmetrical design.
  • the ratio of the roller pitch circle diameters (PCD A / PCD B ) of both examples is 0.9 ⁇ (PCD A / PCD B ) ⁇ 1.1 It may be.
  • Different contact angles in both rows result in a difference in the required mid-brimmed height between the two rows.
  • PCDA A / PCD B the ratio of the pitch circle diameters
  • the thickness TH of the center brim is determined. It is preferable that TH ⁇ 0.15 ⁇ inner ring width.
  • the middle brim needs a certain thickness TH to suppress the axial movement of the roller when an axial load is applied, but when the thickness TH of the middle brim becomes larger than the range of 0.15 ⁇ inner ring width, The middle brim reduces the track surface width and roller length unnecessarily.
  • the bearing for the spindle of the wind power generator of the present invention is a double-row tapered roller bearing having any of the above configurations of the present invention.
  • the wind load of the wind turbine fluctuates greatly from time to time, so that the action and effect of the double-row tapered roller bearing configuration of the present invention can be effectively exhibited.
  • the double row tapered roller bearing 1 is a left- right asymmetric type in which the contact angles ⁇ A and ⁇ B of the left and right rows A and B are different from each other, and is a front combination type.
  • the double row tapered roller bearing 1 is a double row interposed between the inner ring 2 and the outer ring 3 arranged in the left and right rows A and B, respectively, and the raceway surfaces 4 and 5 of the double rows of the inner and outer rings 2 and 3, respectively. It is composed of rollers 6 and 7 and two cages 8 and 9 that hold rollers 6 and 7 in each row, respectively.
  • Rollers 6 and 7 in each row are tapered rollers.
  • the rollers 6 and 7 in each row have a large diameter on the center side in the bearing width direction.
  • the inner ring 2 is composed of a single member having a double row of raceway surfaces 4 and 4
  • the outer ring 3 is a front combination composed of two single row outer rings 3A and 3B formed separately. It is a type.
  • the raceway surfaces 4 and 4 in both rows of the inner ring 2 are tapered surfaces having a large diameter on the center side in the bearing width direction.
  • a middle brim 11 is provided between the raceway surfaces 4 and 4 of both rows.
  • the middle brim 11 has a tapered surface shape in which the outer peripheral surface portion in the range from the vicinity of the center in the width direction to the end of the unloaded row side row gradually becomes lower toward the unloaded row side row B side.
  • End brims 12 and 12 are provided adjacent to each of the raceway surfaces 4 and 4 on the bearing end side.
  • the raceway surfaces 5 and 5 of both rows of the outer ring 3 are tapered surfaces having a large diameter on the center side in the bearing width direction.
  • the inclination angles of the raceway surfaces 5 and 5 of both rows of the outer ring 3 are different from the inclination angles of the raceway surfaces 4 and 4 of the inner ring by the inclination angles of the outer peripheral surfaces of the rollers 6 and 7.
  • the outer ring 3 does not have a brim.
  • An outer ring spacer 3C is interposed between the outer rings 3A and 3B in both rows.
  • the left column A in the figure is a load-bearing row
  • the right row B is a non-load side row. Due to the difference in the inclination angles of the raceway surfaces 4 and 5 of the inner and outer rings 2 and 3 and the taper angles of the rollers 6 and 7, the contact angle ⁇ A of the load-bearing row A is larger than the contact angle ⁇ B of the non-load side row B. It is formed large.
  • the load load row A is a row on the side where the axial load F acting on the inner ring 2 in the case of inner ring rotation or the axial load G acting on the outer ring 3A in the case of outer ring rotation is mainly loaded.
  • the non-load side row B is a row opposite to the load-load row A.
  • the difference between the contact angle theta B contact angle theta A and the non-load side column B of the applied load column A is a 15 ° or more. Further, the contact angles ⁇ A and ⁇ B of the columns A and B are within a range satisfying the following equation. 25 ° ⁇ (contact angle ⁇ A of load-bearing column A) ⁇ 35 ° 5 ° ⁇ (contact angle ⁇ B of unloaded side row B) ⁇ 15 °
  • roller length and roller diameter rollers of the roller 6, 7 B, both, roller length L B of the roller 7 of the non-load side column B found the following (Fig. 2) and roller diameter D B, It is larger than the roller length of the load-bearing column a L a and roller diameter D a (not shown). Note that only one of the non-load side roller length of the roller 7 of column B L B and the roller diameter D B may be larger than the load-bearing column A.
  • roller length L A for a comparison of L B, even when compared with a length between including the width of the chamfer surface Lengths that do not include the width of the chamfer may be compared.
  • Roller diameters D A and D B are the maximum diameters of rollers 6 and 7 in each row.
  • a center hole 6-1 (see Fig. 3) is provided in the center of the large end surface of row A roller 6.
  • a center hole 7-1 is provided in the center of the large end surface of the B row roller 7, and an annular identification mark 7-2 is provided on the small end surface.
  • the thickness TH of the middle brim 11 between the raceway surfaces 4 and 4 of both rows A and B of the inner ring 2 is TH ⁇ 0.15 x inner ring width W It is said that.
  • HA ⁇ HB (HA HB is desirable) It is said that.
  • roller length L B of the roller 7 of the non-load side column B is greater than the roller length L A of the rollers of the load-bearing column A. Therefore, the following advantages can be obtained. Double-row tapered roller bearings may be used under conditions where the direction and magnitude of the load fluctuate significantly. For example, when used as a bearing for a spindle of a wind power generator, the wind load of the wind turbine changes significantly, so that a large load may be applied to the non-load side row.
  • roller length L B of the roller 7 of the non-load side column B is the formed larger than the roller length L A of the roller 6 in the load-bearing column A, the non-load side column B Even when a large load is applied, it becomes easy to satisfy the safety factor standard.
  • the contact angles ⁇ A and ⁇ B of both rows not only make the difference 15 ° or more, but also the above-mentioned conditions. 25 ° ⁇ load load column A contact angle ⁇ A ⁇ 35 ° 5 ° ⁇ contact angle of non-load side row B ⁇ B ⁇ 15 ° Therefore, under the usage conditions in which the axial load acts unevenly in one direction, the load equalization of both rows can be realized as compared with the symmetrical design.
  • the contact angle theta A load load column A is at 25 ° or more, non the contact angle of the load side column is at 15 ° or less, both rows A, the point of application of both rows at the bearing axis on O of B P A ,
  • the distance M between P and B is about 50 to 75% of the symmetrically designed double row tapered roller bearing. Therefore, the load ratio of the radial load of the non-load side row B is higher than that of the double row tapered roller bearing of the symmetrical design.
  • the contact angle ⁇ A angle of the load-bearing row A exceeds 35 °, the load capacity of the radial load of the load-load row A is reduced, which is not preferable. Further, the contact angle theta B of the non-load-side column B is less than 5 °, when the axial load is applied in the opposite direction, insufficient load carrying capacity of the axial load.
  • the ratio of the pitch circle diameters PCD A and PCD B in both rows ( It is preferable to quantify PCD A / PCD B).
  • the ratio of the roller pitch circle diameters (PCD A / PCD B ) of both examples A and B is determined. 0.9 ⁇ (PCD A / PCD B ) ⁇ 1.1 It is said that. Therefore, it is possible to prevent the height difference between the two rows A and B from becoming too large on both sides of the middle brim 11.
  • TH ⁇ 0.15 x inner ring width W Is preferable.
  • the middle brim 11 needs a certain thickness TH in order to suppress the axial movement of the rollers 6 and 7 when an axial load is applied, but the thickness TH of the middle brim 6 and 7 is 0.15 ⁇ inner ring width W. When it becomes larger than the range of, the raceway surface width and the roller length are unnecessarily reduced by the middle brim 11.
  • Table 1 compares the safety factor and the basic rated life of each design with the same size.
  • the asymmetrical designs (1) and (2) are the products of the embodiments shown in FIGS. 1 and 2, and the asymmetrical design (3) is that the difference between the contact angles ⁇ A and ⁇ B of both rows A and B is less than 15 °.
  • a comparative example is shown.
  • the values of the symmetrically designed product are used as the roller length L, roller diameter D W , and safety factor S0A, respectively.
  • each value of the asymmetrical design (1) and (2) is shown by the magnification of each value of the conventional product.
  • the value of row A (axial load load row) of the symmetrical design product is shown by LA
  • the basic rated life and total life of each row in each design are shown by the magnification of LA.
  • the rolling element load distribution curve (FIGS. 5B and 5C) in the axial non-load side row B is slightly larger than the rolling element load distribution curve (FIG. 5A) in the axial load-bearing row A. Since the diameter of the rolling element load distribution curve (FIG. 4A) in the axial load load row A of the symmetric product is significantly larger, the embodiment product (asymmetric product) rolls as a whole in both rows A and B. The moving body load distribution curve is small.
  • FIG. 6 shows another embodiment of the present invention.
  • the inner ring 2 is composed of a single member having a double row of raceway surfaces 4, 4, but in the embodiment of FIG. 6, the inner ring 2 is formed as a separate body. It consists of inner rings 2A and 2B.
  • the middle brim 11 is composed of two single-row middle brims 11A and 11B.
  • the outer ring 3 is composed of two single row outer rings 3A and 3B as in the first embodiment.
  • the double-row tapered roller bearing 1 is composed of two single-row tapered roller bearings 1A and 1B.
  • the same cages as the cages 8 and 9 in FIG. 1 are provided, but these cages are not shown. Even in this configuration, each action and effect described in the first embodiment can be obtained.
  • FIG. 7 shows an example of a bearing device in which a double-row tapered roller bearing 1 and a cylindrical roller bearing 15 are combined.
  • This bearing device is applied to support spindles of wind turbines and various industrial machines.
  • the front and rear portions of the spindle 16 are supported by the housing 17 via the double row tapered roller bearing 1 and the cylindrical roller bearing 15.
  • the double-row tapered roller bearing 1 uses the embodiment shown in FIG. 6, it may be the double-row tapered roller bearing 1 according to the first embodiment shown in FIG.
  • the cylindrical roller bearing 15 has an inner ring 18, an outer ring 19, a cylindrical roller 20, and a cage (not shown).
  • the housing 17 is composed of one cylindrical member in this example, the double-row tapered roller bearing 1 and the cylindrical roller bearing 15 that support the shaft 1 are installed in separate housings (not shown). May be.
  • FIG. 8 shows an example of a wind power generation device using the double row tapered roller bearing 1 according to the embodiment of the present invention.
  • the casing 23a of the nacelle 23 is horizontally swivelly installed on the support base 21 via the swivel bearing 22.
  • the spindle 26 is rotatably installed via the bearing 25 for the main shaft of the wind power generator installed in the bearing housing 24, and the swivel blade is formed in a portion of the spindle 26 protruding outside the casing 23a.
  • a blade 27 is attached.
  • the other end of the spindle 26 is connected to the speed increaser 28, and the output shaft of the speed increaser 28 is coupled to the rotor shaft of the generator 29.
  • two bearings 25 for the spindle of the wind power generator are installed side by side, but one may be used.
  • the first embodiment of FIGS. 1 and 2 or the second embodiment shown in FIG. 6 is attached to each of the wind power generator spindle bearings 25.
  • the double-row tapered roller bearing 1 according to the embodiment is used.
  • the double row tapered roller bearing 1 may be used as any of the bearings.

Abstract

In a double row tapered roller bearing (1) of face-to-face duplex type, the contact angle θA of a loaded row (A) to which the axial load is applied is greater than the contact angle θB of a non-loaded row (B), and the difference between the contact angles θA and θB of the respective rows (A, B) is 15° or more. One or both of the length and the diameter of a roller of the non-loaded row (B) may be greater than the length and the diameter, respectively, of a roller of the loaded row (A). The contact angles of the respective rows may satisfy: 25° ≤ contact angle θA of loaded row A ≤ 35°; and 5° ≤ contact angle θB of non-loaded row B ≤ 15°.

Description

複列円すいころ軸受Double row tapered roller bearings 関連出願Related application
　本出願は、２０１９年９月２６日出願の特願２０１９－１７５９０９および２０２０年５月２７日出願の特願２０２０－０９２０６７の優先権を主張するものであり、その全体を参照により本願の一部をなすものとして引用する。 This application claims the priority of Japanese Patent Application No. 2019-175909 filed on September 26, 2019 and Japanese Patent Application No. 2020-92067 filed on May 27, 2020, which is a part of the present application by reference in its entirety. Quoting as something that makes up.
　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。 The present invention relates to a double-row tapered roller bearing having an asymmetrical design in the left-right row, for example, a double-row tapered roller bearing such as a bearing for a spindle 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）。 As a tapered roller bearing with a centering ring, it has been proposed to increase the load capacity of the row that loads the axial by making the roller length, roller diameter, etc. of the left and right rows of the double-row tapered roller bearing different from each other (for example). , Patent Document 1).
特開２００６－１７７４４６号公報Japanese Unexamined Patent Publication No. 2006-177446
　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。 The arrangement in which a double-row tapered roller bearing and a cylindrical roller bearing in a front combination are combined is often used as a bearing device for a wind turbine spindle.
One of the causes of insufficient strength of double-row tapered roller bearings is that an unbalanced load is generated between the rollers in both rows due to the axial load, and when the specifications of both rows are the same, the axial load is mainly applied. It is considered that the load-load row, which is a row, reaches the fatigue limit first. Therefore, in the usual design concept, the load capacity of the load load train is increased.
　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, since the wind load of the wind turbine is not constant, a large load may be applied to the non-load side row, and a high load capacity of the non-load side row is also required. Therefore, in the conventional design concept, 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 standard.
　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。 An object of the present invention is to balance the load loads in both rows under load conditions in which the axial load acts mainly from one direction and the axial load acts in the opposite direction by optimizing the contact angle. It is an object of the present invention to provide a double-row tapered roller bearing that can achieve a long life as a whole bearing.
　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 front-combination double-row tapered roller bearing.
The contact angle of the load-bearing row, which is the row on which the axial load is mainly applied, is larger than the contact angle of the non-load side row, which is the opposite row, and the difference between the contact angles of both rows is 15 ° or more. Is.
The term "mainly" refers to the direction in which the axial load is applied if the direction of the axial load is determined, and the direction in which the axial load acts for many hours when the direction of the axial load may fluctuate.
　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。 According to this configuration, since the contact angle of the load-bearing row is larger than the contact angle of the non-load side row, the loading capacity of the axial load of the load-bearing row is increased. On the other hand, the load capacity of the radial load of the load load train is reduced. In a double-row tapered roller bearing, the acting load is generally larger than the axial load, and the radial load is applied in both rows. Therefore, for double-row tapered roller bearings, it is necessary to balance the load capacity of the axial load and the radial load of the bearing as a whole.
In this case, since the difference between the contact angles of the two rows is increased to 15 ° or more, the contact angle of the non-loaded side row can be sufficiently reduced while the contact angle of the loaded row is increased to some extent. By reducing the contact angle of the non-load side row, the load capacity of the radial load of the non-load side row is increased, and the reduction of the radial load capacity of the load load row can be compensated. As a result, when both the axial load and the radial load are taken into consideration, the load loads in both rows are balanced, the life in both rows is balanced, and the life of the entire bearing can be extended.
　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 lengths and diameters of the rollers in the unloaded side row may be larger than the roller lengths and diameters of the load-loaded row.
Double-row tapered roller bearings may be used under conditions where the direction and magnitude of the load fluctuate significantly. For example, when used as a bearing for a spindle of a wind power generator, the wind load of the wind turbine changes significantly, so that a large load may be applied to the non-load side row. In such a case, if either or both of the roller lengths and diameters of the rollers in the non-load side row are formed larger than the roller lengths and diameters of the load-bearing rows, the non-load side rows Even if a large axial load and a radial load are applied to the vehicle, it becomes easy to satisfy the safety factor standard.
　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 not only to make the difference between the contact angles of both rows 15 ° or more, but also to satisfy the following conditions for the contact angles of each row.
25 ° ≤ load column contact angle ≤ 35 °
5 ° ≤ contact angle of non-load side row ≤ 15 °
When the contact angle of the load-bearing row is 25 ° or more and the contact angle of the non-load side row is 15 ° or less, the distance M between the action points of both rows on the bearing center axis of both rows is symmetrically designed. It is about 50% to 75% of the double row tapered roller bearings. Therefore, the load ratio of the radial load in the non-load side row is higher than that in the double row tapered roller bearing of the symmetrical design. By increasing the load ratio of the radial load in the non-load side row in this way, it is possible to realize equalization of the load in both rows as compared with the symmetrical design under the usage condition in which the axial load acts unevenly in one direction.
If the contact angle of the load-bearing train is 35 ° or more, the radial load capacity of the load-bearing train is reduced, which is not preferable. Further, when the contact angle of the non-load side row is 5 ° or less, the load capacity of the axial load is insufficient when the 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 (PCD A / PCD B ) of both examples is
0.9 ≤ (PCD A / PCD B ) ≤ 1.1
It may be.
Different contact angles in both rows result in a difference in the required mid-brimmed height between the two rows. In order to suppress the difference in the height of the middle brim, it is preferable to quantitatively determine the ratio of the pitch circle diameters of both rows (PCDA A / PCD B ), and the ratio of the pitch circle diameters (PCDA A / PCD B ). , Above,
0.9 ≤ (PCD A / PCD B ) ≤ 1.1
By setting the range to be, it is possible to prevent the difference in the height of the middle brim 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 brim between the raceway surfaces of both rows, the thickness TH of the center brim is determined.
It is preferable that TH ≦ 0.15 × inner ring width.
The middle brim needs a certain thickness TH to suppress the axial movement of the roller when an axial load is applied, but when the thickness TH of the middle brim becomes larger than the range of 0.15 × inner ring width, The middle brim reduces the track surface width and roller length unnecessarily.
　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 the spindle of the wind power generator of the present invention is a double-row tapered roller bearing having any of the above configurations of 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 fluctuates greatly from time to time, so that the action and effect of the double-row tapered roller bearing configuration of the present invention can be effectively exhibited.
　請求の範囲および／または明細書および／または図面に開示された少なくとも２つの構成のどのような組合せも、本発明に含まれる。特に、請求の範囲の各請求項の２つ以上のどのような組合せも、本発明に含まれる。 Any combination of claims and / or at least two configurations disclosed in the specification and / or drawings is included in the present invention. In particular, any combination of two or more of each claim is included in the present invention.
　この発明は、添付の図面を参考にした以下の好適な実施形態の説明から、より明瞭に理解されるであろう。しかしながら、実施形態および図面は単なる図示および説明のためのものであり、この発明の範囲を定めるために利用されるべきものではない。この発明の範囲は添付の請求の範囲によって定まる。添付図面において、複数の図面における同一の符号は、同一または相当する部分を示す。
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。 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。 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。 従来の対称品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。 第１の実施形態に係る設計（１）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。 第１の実施形態に係る設計（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。 従来の対称品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。 第１の実施形態に係る設計（１）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。 第１の実施形態に係る設計（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。 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。 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。 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。 The present invention will be more clearly understood from the following description of preferred embodiments with reference to the accompanying drawings. However, the embodiments and drawings are for illustration and description purposes only and should not be used to define the scope of the invention. The scope of the present invention is determined by the appended claims. In the accompanying drawings, the same reference numerals in a plurality of drawings indicate the same or corresponding parts.
It is a partial sectional view of the double row tapered roller bearing which concerns on 1st Embodiment of this invention. It is sectional drawing which shows the dimension of each part of the double row tapered roller bearing. It is a partial fracture front view of the roller of both rows in the same double row tapered roller bearing. It is explanatory drawing which shows the simulation example of the rolling element load distribution by the fatigue load in the load load row of the conventional symmetrical double-row tapered roller bearing. It is explanatory drawing which shows the simulation example of the rolling element load distribution by the fatigue load in the load load row of the double row tapered roller bearing of the design (1) which concerns on 1st Embodiment. It is explanatory drawing which shows the simulation example of the rolling element load distribution by the fatigue load in the load load row of the double row tapered roller bearing of the design (2) which concerns on 1st Embodiment. It is explanatory drawing which shows the simulation example of the rolling element load distribution by the fatigue load in the non-load side row of the conventional symmetrical double-row tapered roller bearing. It is explanatory drawing which shows the simulation example of the rolling element load distribution by the fatigue load in the non-load side row of the double row tapered roller bearing of the design (1) which concerns on 1st Embodiment. It is explanatory drawing which shows the simulation example of the rolling element load distribution by the fatigue load in the non-load side row of the double row tapered roller bearing of the design (2) which concerns on 1st Embodiment. It is a partial sectional view of the double row tapered roller bearing which concerns on other embodiment of this invention. It is sectional drawing of the bearing apparatus which combined the double row tapered roller bearing and the cylindrical roller bearing. It is sectional drawing of the wind power generation apparatus which used the double row tapered roller bearing which concerns on 1st Embodiment as the bearing for the spindle of the wind power generation apparatus.
　＜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、左右の列Ａ，Ｂにそれぞれ配置された内輪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は円すいころである。各列のころ６，７は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. 1 and 2.
The double row tapered roller bearing 1 is a left- right asymmetric type in which the contact angles θ A and θ B of the left and right rows A and B are different from each other, and is a front combination type. The double row tapered roller bearing 1 is a double row interposed between the inner ring 2 and the outer ring 3 arranged in the left and right rows A and B, respectively, and the raceway surfaces 4 and 5 of the double rows of the inner and outer rings 2 and 3, respectively. It is composed of rollers 6 and 7 and two cages 8 and 9 that hold rollers 6 and 7 in each row, respectively. Rollers 6 and 7 in each row are tapered rollers. The rollers 6 and 7 in each row have a large diameter on the center side 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が別体に形成された２個の単列外輪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 a double row of raceway surfaces 4 and 4, and the outer ring 3 is a front combination composed of two single row outer rings 3A and 3B formed separately. It is a type.
　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ている。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 and 4 in both rows of the inner ring 2 are tapered surfaces having a large diameter on the center side in the bearing width direction. On the outer peripheral surface of the inner ring 2, a middle brim 11 is provided between the raceway surfaces 4 and 4 of both rows. The middle brim 11 has a tapered surface shape in which the outer peripheral surface portion in the range from the vicinity of the center in the width direction to the end of the unloaded row side row gradually becomes lower toward the unloaded row side row B side. End brims 12 and 12 are provided adjacent to each of the raceway surfaces 4 and 4 on the bearing end side.
　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る。外輪３の両列の軌道面５，５の傾斜角度は、内輪の軌道面４，４の傾斜角度とは、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はつばを有していない。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 and 5 of both rows of the outer ring 3 are tapered surfaces having a large diameter on the center side in the bearing width direction. The inclination angles of the raceway surfaces 5 and 5 of both rows of the outer ring 3 are different from the inclination angles of the raceway surfaces 4 and 4 of the inner ring by the inclination angles of the outer peripheral surfaces of the rollers 6 and 7. The outer ring 3 does not have a brim. An outer ring spacer 3C is interposed between the outer rings 3A and 3B in 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が非負荷側列である。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, the left column A in the figure is a load-bearing row, and the right row B is a non-load side row. Due to the difference in the inclination angles of the raceway surfaces 4 and 5 of the inner and outer rings 2 and 3 and the taper angles of the rollers 6 and 7, the contact angle θ A of the load-bearing row A is larger than the contact angle θ B of the non-load side row B. It is formed large. The load load row A is a row on the side where the axial load F acting on the inner ring 2 in the case of inner ring rotation or the axial load G acting on the outer ring 3A in the case of outer ring rotation is mainly loaded. The non-load side row B is a row opposite to the load-load 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 theta B contact angle theta A and the non-load side column B of the applied load column A is a 15 ° or more. Further, the contact angles θ A and θ B of the columns A and B are within a range satisfying the following equation.
25 ° ≤ (contact angle θ A of load-bearing column A) ≤ 35 °
5 ° ≤ (contact angle θ B of unloaded 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。 Both rows A, The length and roller diameter rollers of the roller 6, 7 B, both, roller length L B of the roller 7 of the non-load side column B found the following (Fig. 2) and roller diameter D B, It is larger than the roller length of the load-bearing column a L a and roller diameter D a (not shown). Note that only one of the non-load side roller length of the roller 7 of column B L B and the roller diameter D B may be larger than the load-bearing column A. If the ends of the rollers 6 and 7 are chamfered portion is provided on the outer peripheral surface, the roller length L A, for a comparison of L B, even when compared with a length between including the width of the chamfer surface Lengths that do not include the width of the chamfer may be compared. Roller diameters D A and D B are the maximum diameters of rollers 6 and 7 in each row.
　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。 A center hole 6-1 (see Fig. 3) is provided in the center of the large end surface of row A roller 6. A center hole 7-1 is provided in the center of the large end surface of the B row roller 7, and an annular identification mark 7-2 is provided on the small end surface.
　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。 Both rows A, the pitch circle diameter PCD A of the roller array of B, PCD B, for the ratio (PCD A / PCD B),
0.9 ≦ (PCD A / PCD B ) ≦ 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。 Regarding the thickness TH of the middle brim 11 between the raceway surfaces 4 and 4 of both rows A and B of the inner ring 2, the thickness TH is
TH ≤ 0.15 x inner ring width W
It is said that.
Regarding the height HA on the A row side and the height HB on the B row side of the middle brim 11,
HA ≧ HB (HA = HB is desirable)
It is said that.
The height of the contact point between the middle brim 11 and the A-row roller, and the height CB of the contact point between the CA and the B-row roller,
｜ 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 composition>
According to this configuration, since the contact angle θ A (FIG. 1) of the load-bearing row A is larger than the contact angle θ B of the non-load side row B, the load capacity of the axial load of the load-bearing row A is increased. On the other hand, the load capacity of the radial load of the load load row A is reduced. In a double-row tapered roller bearing, the acting load is generally larger than the axial load, and the radial load is applied in both rows. Therefore, for double-row tapered roller bearings, it is necessary to balance the load capacity of the axial load and the radial load of the bearing as a whole. In this case, since the difference between the contact angles θ A and θ B of both rows A and B is increased to 15 ° or more, the contact angle of the unloaded side row B is increased to some extent while the contact angle θ A of the loaded row A is increased to some extent. θ B can be made sufficiently small. By reducing the contact angle θ B of the non-load side row B, the load capacity of the radial load of the non-load side row B is increased, and the reduction of the radial load capacity of the load load row A can be compensated. As a result, when both the axial load and the radial load are taken into consideration, the load loads of both rows A and B are balanced, the life of both rows A and B is balanced, and the life of the entire bearing can be extended.
　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。 Also, roller length L B of the roller 7 of the non-load side column B is greater than the roller length L A of the rollers of the load-bearing column A. Therefore, the following advantages can be obtained. Double-row tapered roller bearings may be used under conditions where the direction and magnitude of the load fluctuate significantly. For example, when used as a bearing for a spindle of a wind power generator, the wind load of the wind turbine changes significantly, so that a large load may be applied to the non-load side row. In this case, roller length L B of the roller 7 of the non-load side column B is the formed larger than the roller length L A of the roller 6 in the load-bearing column A, the non-load side column B Even when a large load is applied, it becomes easy to satisfy the safety factor standard.
　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。 By providing center holes 6-1 and 7-1 and identification marks 7-2 on the roller end faces, it is possible to prevent heterogeneous roller assembly.
　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。 The contact angles θ A and θ B of both rows not only make the difference 15 ° or more, but also the above-mentioned conditions.
25 ° ≤ load load column A contact angle θ A ≤ 35 °
5 ° ≤ contact angle of non-load side row B θ B ≤ 15 °
Therefore, under the usage conditions in which the axial load acts unevenly in one direction, the load equalization of both rows can be realized as compared with the symmetrical design. The contact angle theta A load load column A is at 25 ° or more, non the contact angle of the load side column is at 15 ° or less, both rows A, the point of application of both rows at the bearing axis on O of B P A , The distance M between P and B is about 50 to 75% of the symmetrically designed double row tapered roller bearing. Therefore, the load ratio of the radial load of the non-load side row B is higher than that of the double row tapered roller bearing of the symmetrical design. By increasing the load ratio of the radial load in the non-load side row B in this way, it is possible to realize equalization of the load in both rows as compared with the symmetrical design under the use condition in which the axial load acts unevenly in one direction.
　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。 If the contact angle θ A angle of the load-bearing row A exceeds 35 °, the load capacity of the radial load of the load-load row A is reduced, which is not preferable. Further, the contact angle theta B of the non-load-side column B is less than 5 °, when the axial load is applied in the opposite direction, insufficient load carrying capacity of the axial load.
　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。 Further, in order to suppress the difference in height on both sides of the middle brim 11 in both rows A and B depending on the size of the contact angles θ A and θ B , the ratio of the pitch circle diameters PCD A and PCD B in both rows ( It is preferable to quantify PCD A / PCD B).
In this embodiment, as described above, the ratio of the roller pitch circle diameters (PCD A / PCD B ) of both examples A and B is determined.
0.9 ≤ (PCD A / PCD B ) ≤ 1.1
It is said that.
Therefore, it is possible to prevent the height difference between the two rows A and B from becoming too large on both sides of the middle brim 11.
　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 brim 11, as described above,
TH ≤ 0.15 x inner ring width W
Is preferable. The middle brim 11 needs a certain thickness TH in order to suppress the axial movement of the rollers 6 and 7 when an axial load is applied, but the thickness TH of the middle brim 6 and 7 is 0.15 × inner ring width W. When it becomes larger than the range of, the raceway surface width and the roller length are unnecessarily reduced by the middle brim 11.
　　＜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 result>
Table 1 compares the safety factor and the basic rated life of each design with the same size. In the table, the asymmetrical designs (1) and (2) are the products of the embodiments shown in FIGS. 1 and 2, and the asymmetrical design (3) is that the difference between the contact angles θ A and θ B of both rows A and B is less than 15 °. A comparative example is shown. For the roller length, roller diameter, and safety factor of each row, the values of the symmetrically designed product (conventional product with the same contact angle in both rows) are used as the roller length L, roller diameter D W , and safety factor S0A, respectively. It is shown by S0B, and each value of the asymmetrical design (1) and (2) is shown by the magnification of each value of the conventional product. For the basic rated life, the value of row A (axial load load row) of the symmetrical design product is shown by LA, and the basic rated life and total life of each row in each design are shown by the magnification of LA. From the results shown in Table 1, the asymmetrical designs (1) and (2), which are the embodiments, achieve a basic rated life that is about twice that of the symmetric product, and the safety factor is almost the same as that of the symmetric product. I understand.
　図４Ａ～図４Ｃ、図５Ａ～図５Ｃ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４Ａ）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５Ａ）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４Ｂ，４Ｃ）とアキシアル非負荷側列Ｂの転動体荷重分布曲線（図５Ｂ，５Ｃ）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。 4A to 4C and 5A to 5C show the rolling element load distributions around the entire bearing under fatigue load in the axial load load row A and the axial non-load side row B, respectively. In the symmetrical product, the diameter of the rolling element load distribution curve (FIG. 4A) of the axial load-bearing row A is significantly larger than that of the rolling element load distribution curve (FIG. 5A) in the axial non-load side row B. On the other hand, in the embodiment product (asymmetric product), the rolling element load distribution curve (FIGS. 4B and 4C) of the axial load-bearing row A and the rolling element load distribution curve (FIGS. 5B and 5C) of the axial non-load side row B. It can be seen that there is no large difference in diameter and the rolling element loads of both rows A and B are balanced.
　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５Ｂ，５Ｃ）がアキシアル荷重負荷列Ａの転動体荷重分布曲線（図５Ａ）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４Ａ）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。 In the embodiment product (asymmetric product), the rolling element load distribution curve (FIGS. 5B and 5C) in the axial non-load side row B is slightly larger than the rolling element load distribution curve (FIG. 5A) in the axial load-bearing row A. Since the diameter of the rolling element load distribution curve (FIG. 4A) in the axial load load row A of the symmetric product is significantly larger, the embodiment product (asymmetric product) rolls as a whole in both rows A and B. The moving body load distribution curve is small.
Figure JPOXMLDOC01-appb-T000001
Figure JPOXMLDOC01-appb-T000001
　＜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が別体に形成された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、本実施形態に係る複列円すいころ軸受１は、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と同様の保持器が設けられているが、これらの保持器は、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. In this embodiment, the same as the first embodiment described with FIGS. 1 to 5 except for the matters to be described in particular. In the first embodiment, the inner ring 2 is composed of a single member having a double row of raceway surfaces 4, 4, but in the embodiment of FIG. 6, the inner ring 2 is formed as a separate body. It consists of inner rings 2A and 2B. Along with this, the middle brim 11 is composed of two single-row middle brims 11A and 11B. The outer ring 3 is composed of two single row outer rings 3A and 3B as in the first embodiment. As described above, the double-row tapered roller bearing 1 according to the present embodiment is composed of two single-row tapered roller bearings 1A and 1B. In the figure, the same cages as the cages 8 and 9 in FIG. 1 are provided, but these cages are not shown. Even in this configuration, each action and effect 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。主軸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>
FIG. 7 shows an example of a bearing device in which a double-row tapered roller bearing 1 and a cylindrical roller bearing 15 are combined. This bearing device is applied to support spindles of wind turbines and various industrial machines. The front and rear portions of the spindle 16 are supported by the housing 17 via the double row tapered roller bearing 1 and the cylindrical roller bearing 15. Although the double-row tapered roller bearing 1 uses the embodiment shown in FIG. 6, it may be the double-row tapered roller bearing 1 according to the first embodiment shown in FIG. The cylindrical roller bearing 15 has an inner ring 18, an outer ring 19, a cylindrical roller 20, and a cage (not shown). Although the housing 17 is composed of one cylindrical member in this example, the double-row tapered roller bearing 1 and the cylindrical roller bearing 15 that support the shaft 1 are installed in separate housings (not shown). May be.
　＜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はいずれの軸受として用いられてもよい。 <Wind power generator>
FIG. 8 shows an example of a wind power generation device using the double row tapered roller bearing 1 according to the embodiment of the present invention. The casing 23a of the nacelle 23 is horizontally swivelly installed on the support base 21 via the swivel bearing 22. In the casing 23a of the nacelle 23, the spindle 26 is rotatably installed via the bearing 25 for the main shaft of the wind power generator installed in the bearing housing 24, and the swivel blade is formed in a portion of the spindle 26 protruding outside the casing 23a. A blade 27 is attached. The other end of the spindle 26 is connected to the speed increaser 28, and the output shaft of the speed increaser 28 is coupled to the rotor shaft of the generator 29. In the illustrated example, two bearings 25 for the spindle of the wind power generator are installed side by side, but one may be used.
The first embodiment of FIGS. 1 and 2 or the second embodiment shown in FIG. 6 is attached to each of the wind power generator spindle bearings 25.
The double-row tapered roller bearing 1 according to the embodiment is used. Of the two bearings 25 and 25, the double row tapered roller bearing 1 may be used as any of the bearings.
　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。 Although the embodiments for carrying out the present invention have been described above based on the embodiments, the embodiments disclosed this time are examples in all respects and are not restrictive. The scope of the present invention is shown by the scope of claims rather than the above description, and it is intended to include all modifications within the meaning and scope equivalent to the scope of 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 bearings 1A, 1B ... Single row tapered roller bearings 2 ... Inner ring 2A, 2B ... Single row inner ring 3 ... Outer ring 3A, 3B ... Single row outer ring 3C ... Outer ring spacers 4, 5 ... Track surfaces 6, 7 ... Roller 6-1 and 7-1 ... Roller end face center hole 7-2 ... Roller end face identification mark 8, 9 ... Cage 11 ... Middle brim 11A, 11B ... Single row middle brim 12 ... End brim 15 ... Cylindrical roller bearing 16 ... Main shaft 17 ... Housing 18 ... Inner ring 19 ... Outer ring 20 ... Roller 21 ... Support base 22 ... Swing seat bearing 23 ... Nacelle 23a ... Casing 24 ... Bearing housing 25 ... Main shaft support bearing 26 ... Main shaft 27 ... Blade 28 ... Speed increaser 29 ... generator A ... axial load the load string B ... axial load non-load side column θ A, θ B ... contact angle D A, D B ... roller diameter L B, L B ... roller length PCD A, PCD B ... pitch circle Diameter HA, HB ... Height of middle bearing CA, CB ... Height of contact point at middle bearing

Claims (6)

  1. 　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。     It is a double row tapered roller bearing with a front combination.
    The contact angle of the load-bearing row, which is the row on which the axial load is mainly applied, is larger than the contact angle of the non-load side row, which is the opposite row, and the difference between the contact angles of both rows is 15 ° or more. Double row tapered roller bearing.
  2. 　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。 In the double row tapered roller bearing according to claim 1, one or both of the roller lengths and diameters of the rollers in the non-load side row are larger than the roller lengths and diameters of the load load row. Column tapered roller bearing.
  3. 　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 ° ≤ load column contact angle ≤ 35 °
    5 ° ≤ contact angle of non-load side row ≤ 15 °
    Double row tapered roller bearings.
  4. 　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 (PCD A / PCD B ) of both examples is determined.
    0.9 ≤ (PCD A / PCD B ) ≤ 1.1
    Double row tapered roller bearings.
  5. 　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。     In the double row tapered roller bearing according to any one of claims 1 to 4, the inner ring has a center brim between the raceway surfaces of both rows, and the thickness TH of the center brim is determined.
    Double-row tapered roller bearing with TH ≤ 0.15 x inner ring width.
  6. 　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 spindle of a wind power generator, which is a double-row tapered roller bearing according to any one of claims 1 to 5.
PCT/JP2020/036072 2019-09-26 2020-09-24 Double row tapered roller bearing WO2021060389A1 (en)

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JP2019-175909 2019-09-26
JP2019175909 2019-09-26
JP2020092067A JP7456851B2 (en) 2019-09-26 2020-05-27 Double row tapered roller bearing
JP2020-092067 2020-05-27

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Cited By (2)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
CN113883022A (en) * 2021-10-29 2022-01-04 新疆金风科技股份有限公司 Wind generating set shafting and wind generating set
CN113969876A (en) * 2021-10-29 2022-01-25 新疆金风科技股份有限公司 Wind generating set shafting and wind generating set

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DE102004047881A1 (en) * 2004-10-01 2006-04-06 Fag Kugelfischer Ag & Co. Ohg Self-aligning roller bearing for use in applications where axial loading is predominantly on one side has two rows of rollers which are inclined to vertical in opposite directions, angle being greater on side with higher loading
WO2014129658A1 (en) * 2013-02-25 2014-08-28 アイシン・エィ・ダブリュ株式会社 Dynamic force transmission device
WO2016146115A1 (en) * 2015-03-19 2016-09-22 Schaeffler Technologies AG & Co. KG Roller bearing, for example of a wind power plant
WO2017164325A1 (en) * 2016-03-24 2017-09-28 Ntn株式会社 Double-row spherical roller bearing
WO2018095452A1 (en) * 2016-11-28 2018-05-31 Schaeffler Technologies AG & Co. KG Wind turbine shaft assembly

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Publication number Priority date Publication date Assignee Title
DE102004047881A1 (en) * 2004-10-01 2006-04-06 Fag Kugelfischer Ag & Co. Ohg Self-aligning roller bearing for use in applications where axial loading is predominantly on one side has two rows of rollers which are inclined to vertical in opposite directions, angle being greater on side with higher loading
WO2014129658A1 (en) * 2013-02-25 2014-08-28 アイシン・エィ・ダブリュ株式会社 Dynamic force transmission device
WO2016146115A1 (en) * 2015-03-19 2016-09-22 Schaeffler Technologies AG & Co. KG Roller bearing, for example of a wind power plant
WO2017164325A1 (en) * 2016-03-24 2017-09-28 Ntn株式会社 Double-row spherical roller bearing
WO2018095452A1 (en) * 2016-11-28 2018-05-31 Schaeffler Technologies AG & Co. KG Wind turbine shaft assembly

Cited By (4)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
CN113883022A (en) * 2021-10-29 2022-01-04 新疆金风科技股份有限公司 Wind generating set shafting and wind generating set
CN113969876A (en) * 2021-10-29 2022-01-25 新疆金风科技股份有限公司 Wind generating set shafting and wind generating set
CN113883022B (en) * 2021-10-29 2023-06-16 新疆金风科技股份有限公司 Shafting of wind generating set and wind generating set
CN113969876B (en) * 2021-10-29 2023-07-04 新疆金风科技股份有限公司 Wind generating set shafting and wind generating set

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