Double-row tapered roller bearing

Abstract

A double-row tapered roller bearing (1) in a face-to-face combination, wherein the contact angle θ of the load row (A) of the axial load A Greater than the contact angle θ of the non-load side row (B) B , and the contact angle θ of the two columns (A, B) A ,θ B The difference is 15° or more. Either or both of the roller length and roller diameter of the rollers in the non-load side row (B) may be greater than the roller length and roller diameter of the rollers in the load side row (A). The contact angle values ​​of each row may also satisfy the following: 25° ≤ contact angle θ of the load side row (A) A ≤35°, and 5°≤contact angle θ of the non-load side row (B) B ≤15°.

Description

[0001] Related applications

[0002] This application claims the benefit of Japanese Patent Application No. 2019-175909, filed on September 26, 2019, and Japanese Patent Application No. 2020-092067, filed on May 27, 2020, the entireties of which are incorporated herein by reference as if they were a part of this application. Technical Field

[0003] The present invention relates to a double-row tapered roller bearing with asymmetrical left and right rows, for example, a double-row tapered roller bearing for a main shaft of a wind turbine generator. Background Art

[0004] As a tapered roller bearing with aligning rings, it has been proposed to increase the load capacity of the row in the axial direction by making the roller lengths and roller diameters of the left and right rows of a double-row tapered roller bearing different from each other (for example, Patent Document 1).

[0005] Prior art literature

[0006] Patent Literature

[0007] Patent Document 1: Japanese Patent Application Laid-Open No. 2006-177446 Summary of the Invention

[0008] Problems to be solved by the invention

[0009] The configuration of a double-row tapered roller bearing and a cylindrical roller bearing in a face-to-face combination is often used as a bearing device for the main shaft of a wind turbine.

[0010] One of the main reasons for the insufficient strength of double-row tapered roller bearings is that axial loads create an unbalanced load on the rollers in the two rows. When the two rows have the same specifications, it is believed that the load-bearing row, which primarily bears the axial load, will reach its fatigue limit first. Therefore, in typical designs, the load capacity of the load-bearing row is increased.

[0011] However, because wind loads on wind turbines are not constant, larger loads may sometimes be applied to the non-loaded rows, requiring a higher load capacity for these rows. Consequently, conventional designs have limited the load capacity of the non-loaded rows, increasing the safety factor of the non-loaded rows relative to the baseline.

[0012] An object of the present invention is to provide a double-row tapered roller bearing that optimizes the contact angle to balance the loads on the two rows under load conditions arising from situations where an axial load acts primarily in one direction and also in the opposite direction, thereby achieving a longer life for the bearing as a whole.

[0013] Technical solutions to problems

[0014] The double-row tapered roller bearing of the present invention relates to a double-row tapered roller bearing assembled face to face, wherein the contact angle of the loaded row, which is the row that primarily bears axial load, is greater than the contact angle of the unloaded row, which is the opposite row, and the difference in the contact angles of the two rows is 15° or more.

[0015] Furthermore, the term “mainly” mentioned above indicates the direction in which the axial load acts at a plurality of times, provided that the direction of the axial load is fixed and when the direction of the axial load varies.

[0016] This structure increases the contact angle of the load-bearing row compared to the non-load-bearing row, improving the load capacity of the load-bearing row for axial loads. Conversely, the load capacity of the load-bearing row for radial loads decreases. In double-row tapered roller bearings, the radial load is generally greater than the axial load, and the radial load is shared by both rows. Therefore, double-row tapered roller bearings must achieve a balanced load capacity for axial and radial loads across the entire bearing.

[0017] In this case, increasing the difference in contact angles between the two rows to 15° or more increases the contact angle of the load-bearing row to a certain extent, while significantly reducing the contact angle of the non-load-bearing row. By reducing the contact angle of the non-load-bearing row, the radial load capacity of the non-load-bearing row is improved, compensating for the reduced radial load capacity of the load-bearing row. This balances the loads and lifespans of the two rows, evenly considering both axial and radial loads, and ultimately extends the life of the entire bearing.

[0018] In the present invention, either one or both of the roller length and the roller diameter of the rollers in the non-load side row may be greater than the roller length and the roller diameter of the rollers in the load side row.

[0019] Double-row tapered roller bearings are sometimes used under conditions where the direction and magnitude of loads vary significantly. For example, when used as main shaft bearings for wind turbines, wind loads from the wind turbine vary significantly, and therefore a heavier load may be placed on the unloaded row. In such situations, if either or both of the roller length and roller diameter of the unloaded row rollers are greater than those of the loaded row rollers, it is easier to meet safety factor criteria even when large axial and radial loads are applied to the unloaded row.

[0020] In particular, it is preferred that, in the present invention, not only the difference in contact angle between the two rows is 15° or more, but also the contact angle of each row satisfies the following conditions:

[0021] 25°≤Contact angle of load column≤35°

[0022] 5°≤Contact angle of non-load side row≤15°.

[0023] If the contact angle of the loaded row is 25° or greater and the contact angle of the unloaded row is 15° or less, the distance M between the two rows' points of contact on the bearing's central axis is approximately 50% to 75% of that of a bilaterally symmetrical double-row tapered roller bearing. Consequently, the radial load ratio of the unloaded row is higher than that of a bilaterally symmetrical double-row tapered roller bearing. This higher radial load ratio of the unloaded row allows for more balanced load distribution between the two rows, compared to a symmetrical design, even under operating conditions where the axial load acts in a biased direction.

[0024] If the contact angle of the load-bearing row is 35° or greater, the radial load capacity of the load-bearing row decreases, so this method is not preferred. In addition, if the contact angle of the non-load-bearing row is 5° or less, the axial load capacity will be insufficient when an axial load acts in the opposite direction.

[0025] In the present invention, the ratio of the pitch diameters of the rollers of the two rows (PCD A / PCD B )satisfy:

[0026] 0.9≤(PCD A / PCD B )≤1.1.

[0027] If the contact angles are different in the two rows, the height of the required center flange will differ between the two rows. In order to suppress the difference in the height of the center flange, it is preferable to quantitatively determine the ratio of the pitch circle diameters of the two rows (PCD A / PCD B ), the ratio of the pitch diameter (PCD A / PCD B ) is set to the above 0.9≤(PCD A / PCD B )≤1.1, thereby preventing the difference in flange height between the two rows from becoming excessively large.

[0028] Preferably, in the present invention, when the inner ring has a center flange between the two rows of raceway surfaces, the thickness TH of the center flange satisfies:

[0029] TH≤0.15×inner ring width.

[0030] When an axial load is applied, the center flange requires a certain thickness TH to suppress the axial movement of the roller. However, if the thickness TH of the center flange exceeds the range of 0.15×the inner ring width and becomes larger, the raceway width and roller length will be unnecessarily reduced due to the center flange.

[0031] The main shaft bearing of a wind turbine generator according to the present invention is a double-row tapered roller bearing of any of the above-mentioned structures according to the present invention.

[0032] In the case of a bearing for a main shaft of a wind turbine generator, the wind load of the wind turbine greatly fluctuates over time, and therefore the function and effect of the structure of the double-row tapered roller bearing of the present invention can be effectively exerted.

[0033] Any combination of at least two solutions disclosed in the claims and / or the specification and / or the drawings is included in the present invention. In particular, any combination of two or more of the claims in the claims is also included in the present invention. BRIEF DESCRIPTION OF THE DRAWINGS

[0034] The present invention can be more clearly understood by the following description of preferred embodiments with reference to the accompanying drawings. However, the embodiments and the accompanying drawings are for illustration and explanation only and are not intended to limit the scope of protection of the present invention. The scope of protection of the present invention is determined by the claims. In the drawings, the same reference numerals in multiple figures represent the same or equivalent parts.

[0035] Figure 1 A partial cross-sectional view of a double-row tapered roller bearing according to a first embodiment of the present invention;

[0036] Figure 2 A sectional view showing the dimensions of various parts of the double-row tapered roller bearing;

[0037] Figure 3 A partially cutaway front view of two rows of rollers in the double-row tapered roller bearing;

[0038] Figure 4A This is an explanatory diagram showing a simulation example of rolling element load distribution under fatigue load in a load row of a conventional symmetrical double-row tapered roller bearing.

[0039] Figure 4B This is an explanatory diagram showing a simulation example of rolling element load distribution under fatigue load in a load row of a double-row tapered roller bearing of design (1) according to the first embodiment;

[0040] Figure 4C This is an explanatory diagram showing a simulation example of rolling element load distribution under fatigue load in a load row of a double-row tapered roller bearing of design (2) of the first embodiment;

[0041] Figure 5A This is an explanatory diagram showing a simulation example of rolling element load distribution under fatigue load in the unloaded row of a conventional symmetrical double-row tapered roller bearing.

[0042] Figure 5B This is an explanatory diagram showing a simulation example of rolling element load distribution under fatigue load in a non-load side row of a double-row tapered roller bearing of design (1) according to the first embodiment;

[0043] Figure 5C This is an explanatory diagram showing a simulation example of rolling element load distribution under fatigue load in a non-load side row of a double-row tapered roller bearing of design (2) of the first embodiment;

[0044] Figure 6 A partial cross-sectional view of a double-row tapered roller bearing according to another embodiment of the present invention;

[0045] Figure 7 is a cross-sectional view of a bearing device combining the double-row tapered roller bearing and the cylindrical roller bearing;

[0046] Figure 8 This is a cross-sectional view of a wind turbine generator in which the double-row tapered roller bearing according to the first embodiment is used as a bearing for a main shaft of the wind turbine generator. DETAILED DESCRIPTION

[0047] <First embodiment>

[0048] according to Figure 1 、 Figure 2 A first embodiment of the present invention will be described.

[0049] The contact angle θ of the double-row tapered roller bearing 1 is θ for the left and right rows A and B. A ,θ B This double-row tapered roller bearing 1 is asymmetrical and features a face-to-face assembly. It consists of an inner ring 2, an outer ring 3, two rows of rollers 6 and 7, and two retainers 8 and 9. The inner ring 2 and outer ring 3 are arranged in left and right rows A and B, respectively. The two rows of rollers 6 and 7 are sandwiched between the raceways 4 and 5 of the inner and outer rings 2 and 3, respectively. The two retainers 8 and 9 retain the rollers 6 and 7 in each row. Each row of rollers 6 and 7 is tapered. The center of each row of rollers 6 and 7 in the bearing width direction has a larger diameter.

[0050] In this embodiment, a face-to-face combination type is formed, in which the inner ring 2 is composed of a single component having double-row raceway surfaces 4, 4, and the outer ring 3 is composed of two single-row outer rings 3A, 3B formed separately.

[0051] The two rows of raceways 4, 4 in the inner ring 2 are tapered surfaces with a larger diameter at the center in the bearing width direction. A center flange 11 is provided on the outer circumference of the inner ring 2, between the two rows of raceways 4, 4. The outer circumference of center flange 11, from near the center in the width direction to the end of the non-load-bearing row, is tapered, gradually decreasing in depth toward the non-load-bearing row B. End flanges 12, 12 are provided adjacent to the bearing ends of each raceway 4, 4.

[0052] The raceways 5, 5 of the two rows of outer rings 3 are tapered surfaces with a larger diameter at the center in the bearing width direction. The inclination angles of the raceways 5, 5 of the two rows of outer rings 3 differ from the inclination angles of the raceways 4, 4 of the inner rings by the inclination angle of the outer circumferential surfaces of the rollers 6, 7. The outer ring 3 does not have a flange. An outer ring spacer 3C is interposed between the two rows of outer rings 3A, 3B.

[0053] In this double row tapered roller bearing 1, row A on the left side of the figure is the load row, and row B on the right side is the non-load 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 of the load row A is A The contact angle θ is greater than that of the non-load side row B. B The load-side row A is the row on the side where the main loads are 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 unloaded row B is the row on the opposite side of the load-side row A.

[0054] Contact angle θ of load column A A Contact angle θ with the non-load side row B B The difference is 15° or more. A ,θ B Set the range to satisfy the following formula:

[0055] 25°≤(Contact angle θ of load column A A )≤35°

[0056] 5°≤(contact angle θ of non-load side row B B )≤15°.

[0057] The roller lengths and roller diameters of the rollers 6 and 7 in the two rows A and B are both equal to the roller length L of the roller 7 in the non-load side row B. B ( Figure 2 ) and roller diameter D B Greater than the roller length L of the load column A A and roller diameter D A (Not shown in the figure) Alternatively, the roller length L of the roller 7 of the non-load side row B may be B and roller diameter DB Any one of them is greater than the load column A. When the ends of the rollers 6 and 7 are provided with chamfered portions on the outer peripheral surfaces, the roller length L A , L B The comparison can be made by comparing the length including the width of the chamfered part or by comparing the length excluding the width of the chamfered part. A 、D B is the maximum diameter of the rollers 6, 7 in each row.

[0058] A center hole 6-1 is provided at the center of the large end surface of the A row roller 6 (refer to Figure 3 A center hole 7-1 is provided in the middle of the large end surface of the B row roller 7, and a circular identification mark 7-2 is provided on the small end surface.

[0059] Pitch diameter PCD of the roller arrangement of two rows A and B A 、PCD B , its ratio (PCD A / PCD B )satisfy:

[0060] 0.9≤(PCD A / PCD B )≤1.1.

[0061] The thickness TH of the center flange 11 between the raceway surfaces 4, 4 of the two rows A and B of the inner ring 2 satisfies:

[0062] TH≤0.15×inner ring width W.

[0063] The height HA of the center flange 11 on the A column side and the height HB of the center flange 11 on the B column side satisfy:

[0064] HA≥HB (best, HA=HB).

[0065] The height CA of the contact point of the center flange 11 with the rollers in row A and the height CB of the contact point with the rollers in row B satisfy the following conditions:

[0066] |CA—CB|≤3mm.

[0067] <Function, Effect, Detailed Structure>

[0068] According to this structure, the contact angle θ of the load column A is A ( Figure 1 ) than the contact angle θ of the non-load side row B BThe load capacity of the axial load of the load row A is large, so the load capacity of the radial load of the load row A is improved. On the contrary, the load capacity of the radial load of the load row A is reduced. In double-row tapered roller bearings, the radial load is usually larger than the axial load, and the radial load is borne by both rows. Therefore, the double-row tapered roller bearing needs to achieve a balance between the load capacity of the axial load and the radial load as the whole bearing. In this case, by setting the contact angle θ of the two rows A and B A ,θ B The difference between the two increases to 15° or more, so the contact angle θ of the load row A can be increased to a certain extent. A , and can fully reduce the contact angle θ of the non-load side row B B By making the contact angle θ of the non-load side row B B By reducing the radial load capacity of the unloaded row B, the radial load capacity of the unloaded row B is improved, compensating for the reduced radial load capacity of the loaded row A. Therefore, considering both the axial load and the radial load, the loads of the two rows A and B are balanced, and the service life of the two rows A and B is balanced, thus achieving a longer life for the entire bearing.

[0069] In addition, the roller length L of the roller 7 in the non-load side row B is B Greater than the roller length L of the roller in load column A A Therefore, the following advantages can be 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 turbine generator, the wind load of the wind turbine varies greatly, so sometimes a larger load is applied to the non-load side row. In such a case, if the roller length L of the roller 7 of the non-load side row B is B Greater than the roller length L of the roller 6 of the load column A A , even when a large load is applied to the non-load side row B, it is easy to meet the safety factor standard.

[0070] By providing the center holes 6-1, 7-1 and the identification mark 7-2 on the roller end faces, it is possible to prevent the roller sets from being of different types.

[0071] The contact angle θ of the two rows A ,θ B Not only does the difference have to be greater than 15°, but the above conditions must also be met:

[0072] 25°≤Contact angle θ of load column A A ≤35°

[0073] 5°≤Contact angle θ of non-load side row B B ≤15°,

[0074] Therefore, under the use condition that the axial load acts in one direction, the effect of equalizing the load of the two rows can be achieved compared with the symmetrical design. A If the contact angle of the non-load side row is less than 15°, the contact angle of the two rows A and B on the bearing center axis O is greater than 25°. A 、P B The distance M between the bearings is approximately 50-75% of that of a bilaterally symmetrical double-row tapered roller bearing. Therefore, the radial load ratio of the unloaded row B is higher than in a bilaterally symmetrical double-row tapered roller bearing. This higher radial load ratio of the unloaded row B allows for more balanced load distribution between the two rows, compared to a symmetrical design, even under conditions where axial loads are biased in one direction.

[0075] If the contact angle θ of the load column A A If the angle exceeds 35°, the radial load capacity of the load-bearing row A will decrease, so this method is not preferred. B If the angle is less than 5°, the axial load carrying capacity will be insufficient when the axial load acts in the opposite direction.

[0076] In addition, in order to pass the contact angle θ A ,θ B The size of the two rows A, B to suppress the height difference between the two sides of the flange 11, preferably, the two rows of pitch circle diameter PCD A 、PCD B The ratio (PCD A / PCD B ) is set as quantitative.

[0077] In this embodiment, as described above, the ratio of the pitch diameters of the rollers of the two rows A and B (PCD A / PCD B ) is 0.9≤(PCD A / PCD B )≤1.1.

[0078] Therefore, it is possible to suppress the difference in height between both sides of the flange 11 between the two rows A and B from becoming excessively large.

[0079] As for the thickness TH of the middle flange 11, it is preferable, as described above, to satisfy:

[0080] TH≤0.15×inner ring width W.

[0081] When an axial load is applied to the center flange 11, a certain thickness TH is required to suppress the axial movement of the rollers 6 and 7. However, if the thickness TH of the center flanges 6 and 7 increases beyond the range of 0.15×the inner ring width W, the track surface width and the roller length will be unnecessarily reduced due to the center flange 11.

[0082] <Calculation results>

[0083] Table 1 compares the safety factor and basic rated life of each design under the same size. Figure 1 、 Figure 2 The embodiment of the product, the asymmetric design (3) represents the contact angle θ of the two rows A and B A ,θ B The difference between the two rows is less than 15°. Regarding the roller length, roller diameter, and safety factor of each row, the values ​​of the symmetrical design product (the existing product with the same contact angle in both rows) are calculated by the roller length L, roller diameter D, and w , safety factors S0A and S0B are expressed as multiples of the values ​​of the conventional products in the asymmetric designs (1) and (2). Regarding the basic rated life, the value of column A (axial load column) of the symmetrical design product is expressed as LA, and the basic rated life and comprehensive life of each column in each design are expressed as a multiple of LA. The results shown in Table 1 show that the asymmetric designs (1) and (2) used in the embodiment products achieve a basic rated life approximately twice that of the symmetrical products, and the safety factor is also approximately the same as that of the symmetrical products.

[0084] Figures 4A to 4C 、 Figures 5A to 5C The rolling element load distribution curves for the axial load column A and the axial unloaded column B under fatigue load over the entire bearing circumference are shown. Figure 4A ) diameter and the rolling element load distribution curve in the axial non-load side row B ( Figure 5A ) is significantly larger than that of the embodiment product (asymmetric product). In contrast, the rolling element load distribution curve of the axial load load column A ( Figure 4B 、 Figure 4C ) and the rolling element load distribution curve of the axial non-load side row B ( Figure 5B 、 Figure 5C ) does not produce a large difference in diameter, which shows that the load on the rolling elements of the two rows A and B is balanced.

[0085] In addition, the rolling element load distribution curve ( Figure 5B 、 Figure 5C ) than the rolling element load distribution curve of axial load row A ( Figure 5A ) is slightly larger, but the rolling element load distribution curve in the axial load column A of the symmetrical product ( Figure 4A ) becomes significantly larger, so as a whole of the two rows A and B, the rolling element load distribution curve of the implementation product (asymmetric product) becomes smaller.

[0086] [Table 1]

[0087]

[0088]

[0089] <Other Implementations>

[0090] Figure 6 In this embodiment, except for matters specifically described, Figure 1 5. In the first embodiment, the inner ring 2 is composed of a single component having double rows of raceway surfaces 4, 4, but Figure 6 In the embodiment, the inner ring 2 is composed of two single-row inner rings 2A and 2B formed separately. In addition, the center flange 11 is composed of two single-row center flanges 11A and 11B. The outer ring 3 is composed of two single-row outer rings 3A and 3B as in the first embodiment. In this way, the double-row tapered roller bearing 1 of this embodiment is composed of two single-row tapered roller bearings 1A and 1B. In this figure, there is a Figure 1 The retainers 8 and 9 are the same as those in the embodiment, but are not shown in the drawings. Even with this configuration, the various functions and effects described in the first embodiment can be obtained.

[0091] <Combination example with cylindrical bearings>

[0092] Figure 7 This figure shows an example of a bearing assembly that combines a double-row tapered roller bearing 1 and a cylindrical roller bearing 15. This bearing assembly is used to support the main shafts of windmills and various industrial machines. The front and rear portions of the main shaft 16 are supported on a housing 17 via the double-row tapered roller bearing 1 and the cylindrical roller bearing 15. The double-row tapered roller bearing 1 is used Figure 6 The embodiment shown, but it can also be Figure 1 The double-row tapered roller bearing 1 of the first embodiment shown in the figure. The cylindrical roller bearing 15 includes 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 supporting the shaft 1 may be provided in separate housings (not shown).

[0093] <Wind power generation device>

[0094] Figure 8 An example of a wind turbine using a double-row tapered roller bearing 1 according to an embodiment of the present invention is shown. A housing 23a of a nacelle 23 is mounted on a support base 21, rotatably horizontally via a swivel bearing 22. A main shaft 26 is mounted within the housing 23a of the nacelle 23, rotatably via a wind turbine main shaft bearing 25 mounted in a bearing housing 24. Blades 27, serving as rotating blades, are attached to the portion of the main shaft 26 that protrudes outward from the housing 23a. The other end of the main shaft 26 is connected to a speed increaser 28, the output shaft of which is coupled to the rotor shaft of a generator 29. While two wind turbine main shaft bearings 25 are provided in the illustrated example, a single one is also acceptable.

[0095] In the above-mentioned main shaft bearings 25 of each wind power generation device, Figure 1 、 Figure 2 The first embodiment or Figure 6 The double row tapered roller bearing 1 of the second embodiment is shown. The double row tapered roller bearing 1 described above in the two bearings 25 and 25 can be used as any bearing.

[0096] While the above-described embodiments have been used to implement the present invention, the embodiments disclosed herein are intended to be illustrative in all respects and are not intended to be limiting. The scope of the present invention is not indicated by the above description but by the claims, and should include all modifications within the meaning and scope of the claims.

[0097] Description of the label:

[0098] Reference numeral 1 indicates a double-row tapered roller bearing;

[0099] Reference numerals 1A and 1B indicate single-row tapered roller bearings;

[0100] Reference numeral 2 indicates the inner ring;

[0101] Reference numerals 2A and 2B indicate single-row inner rings;

[0102] Reference numeral 3 indicates the outer ring;

[0103] Reference numerals 3A and 3B indicate single-row outer rings;

[0104] Reference numeral 3C denotes an outer ring spacer;

[0105] Reference numerals 4 and 5 denote track surfaces;

[0106] Reference numerals 6 and 7 represent rollers;

[0107] Reference numerals 6-1 and 7-1 indicate the center holes of the roller end faces;

[0108] No. 7-2 roller end face identification mark;

[0109] Reference numerals 8 and 9 denote retainers;

[0110] Reference numeral 11 denotes a center flange;

[0111] Reference numerals 11A and 11B denote single-row center flanges;

[0112] Reference numeral 12 denotes an end flange;

[0113] Reference numeral 15 denotes a cylindrical roller bearing;

[0114] Reference numeral 16 denotes a main shaft;

[0115] Reference numeral 17 denotes a housing;

[0116] Reference numeral 18 denotes an inner ring;

[0117] Reference numeral 19 denotes an outer ring;

[0118] Reference numeral 20 denotes a roller;

[0119] Reference numeral 21 denotes a support platform;

[0120] Reference numeral 22 denotes a slewing bearing;

[0121] Reference numeral 23 denotes a cabin;

[0122] Reference numeral 23a denotes a housing;

[0123] Reference numeral 24 denotes a bearing housing;

[0124] Reference numeral 25 denotes a main shaft support bearing;

[0125] Reference numeral 26 denotes a main shaft;

[0126] Reference numeral 27 denotes a blade;

[0127] Reference numeral 28 denotes a speed increaser;

[0128] Reference numeral 29 denotes a generator;

[0129] Symbol A represents the axial load column;

[0130] Symbol B indicates the non-load side of the axial load;

[0131] Symbol θ A ,θ B represents the contact angle;

[0132] Symbol D A 、D B Indicates roller diameter;

[0133] Symbol L B , L B Indicates the roller length;

[0134] Symbol PCD A 、PCD B Indicates the pitch diameter;

[0135] The symbols HA and HB indicate the height of the center flange;

[0136] Symbols CA and CB represent the heights of the contact points between the center flange and the roller.

Claims

1. A double-row tapered roller bearing in a face-to-face combination, wherein: The inner ring of the double-row tapered roller bearing is composed of a single component having a double-row raceway surface. The contact angle of the load-bearing row, which is the row that mainly bears the axial load, is larger than the contact angle of the non-load-bearing row, which is the opposite row. The difference in the contact angles of the two rows is at least 15 degrees. The rollers in the non-load side row have either or both of a roller length and a roller diameter greater than the roller length and roller diameter of the rollers in the load side row. The inner ring has a central flange between the two rows of track surfaces. The outer peripheral surface portion of the center flange ranging from the center portion in the bearing width direction to the non-load side row end has a tapered surface shape that gradually becomes lower toward the non-load row side.

2. The double row tapered roller bearing according to claim 1, wherein: 25°≤Contact angle of load column≤35°; 5°≤Contact angle of non-load side row≤15°.

3. The double row tapered roller bearing according to claim 1 or 2, wherein: The ratio of the pitch diameters of the two rows of rollers (PCD A / PCD B )satisfy: 0.9≤(PCD A / PCD B )≤1.1。 4. The double row tapered roller bearing according to claim 1 or 2, wherein: The thickness TH of the aforementioned middle flange satisfies: TH≤0.15×inner ring width. 5 . A bearing for a main shaft of a wind turbine generator, the bearing being the double row tapered roller bearing according to claim 1 .

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