Root digging type gear and gear box

Abstract

The invention provides a root digging type gear and a gearbox, belongs to the technical field of gearboxes, and aims to solve the technical problem that the bending strength of a gear in a wind power gearbox is limited and the contact strength is rich. The area between every two adjacent teeth is a tooth groove, each tooth groove comprises a rear side tooth surface line located on the front tooth, a front side tooth surface line located on the rear tooth and a tooth root transition line connected to the bottom of the rear side tooth surface line and the bottom of the front side tooth surface line, and the rear side tooth surface line and the front side tooth surface line are both cut with the tooth root transition line. The technical effects of reducing the stress concentration coefficient of the tooth root and improving the comprehensive bending strength are achieved.

Description

Technical Field

[0001] This invention relates to the field of gearbox technology, specifically to a hollowed-out gear and a gearbox. Background Technology

[0002] In wind turbine gearboxes, gear design faces unique challenges, with insufficient bending strength being a major bottleneck. Due to the high power demands of wind turbines, gear size and weight increase significantly, especially in planetary gear systems where the numerous planetary gears lead to a complex overall structure. This large-scale design subjectes gears to enormous bending stress during operation, becoming a primary strength limiting factor. In contrast, the contact strength of gears is relatively abundant, failing to fully utilize the material's potential. Therefore, optimizing the gear structure to improve bending strength has become a key issue in wind turbine gearbox design.

[0003] In current wind turbine gearboxes, the planetary gear system exhibits significant stress concentration at the tooth root region due to the multiple interactions involved, limiting the gears' load-bearing capacity. Simultaneously, the relatively abundant contact strength fails to translate into overall performance advantages. This strength mismatch leads to premature gear failure, impacting gearbox reliability. Summary of the Invention

[0004] To address the shortcomings of existing technologies, this invention proposes a root-cutting gear and gearbox to solve the technical problem of limited bending strength but relatively abundant contact strength of gears in wind turbine gearboxes.

[0005] The technical solution adopted in this invention is a root-cutting gear and a gearbox.

[0006] One type of root-cutting gear has a tooth groove between adjacent teeth. The tooth groove includes a rear tooth surface line of the preceding tooth and a front tooth surface line of the following tooth, as well as a tooth root transition line connecting the bottom of the rear tooth surface line and the front tooth surface line. Both the rear tooth surface line and the front tooth surface line intersect the tooth root transition line.

[0007] Optionally, the tooth root transition line is an arc, and the center of the arc is located on the line of symmetry between the rear tooth surface line and the front tooth surface line.

[0008] Optionally, the tooth root transition line comprises multiple interconnected smooth curves.

[0009] Optionally, the root fillet radius of the basic rack of the root-cutting gear is greater than 0.39m, where m is the module.

[0010] Optionally, the rear tooth surface line and the front tooth surface line are involutes, and the intersection of the tooth root transition line with the rear tooth surface line and the front tooth surface line exceeds the starting point of the involute.

[0011] Optionally, the gear is made of carburized and quenched steel.

[0012] One type of gearbox includes at least one gear employing a root-type gear as described above.

[0013] Optionally, it includes an input shaft, an output shaft, a first planetary speed regulating unit, a second planetary speed regulating unit, and a cylindrical gear speed regulating unit; the first planetary speed regulating unit includes a first internal gear ring, a first planet carrier, and a first sun gear; the second planetary speed regulating unit includes a second internal gear ring, a second planet carrier, and a second sun gear; the first internal gear ring and the second internal gear ring are fixed, the input shaft is connected to the first planet carrier, the first sun gear is connected to the second planet carrier, and the cylindrical gear speed regulating unit is disposed between the output shaft and the axle of the second sun gear.

[0014] Optionally, the first planetary speed regulating unit, the second planetary speed regulating unit, and the cylindrical gear speed regulating unit are arranged in parallel. The first sun gear and the second planetary carrier are connected by the same component. The first sun gear has a central hole, and the axle of the second sun gear also has a central hole. The two central holes are coaxial, and the hollow tube passes through the central hole to penetrate the gearbox.

[0015] Optionally, the number of planetary gears in the first planetary speed regulating unit and / or the second planetary speed regulating unit is greater than four.

[0016] As can be seen from the above technical solution, the beneficial technical effects of the present invention are as follows: This root-cutting gear significantly reduces the stress concentration coefficient in the tooth root region by optimizing the tooth groove structure, resulting in a more uniform stress distribution and effectively mitigating the risk of tooth root fatigue cracks. Simultaneously, the rear tooth surface line intersects the front tooth surface line at the tooth root transition line, forming a non-smooth transition, further reducing local stress peaks and improving the overall bending strength of the gear. This design, while maintaining advantages in contact strength, specifically strengthens weak points, enabling the gear to possess superior load-bearing capacity and durability under the high-load conditions of wind turbine gearboxes, extending its service life and improving reliability. Attached Figure Description

[0017] To more clearly illustrate the specific embodiments of the present invention or the technical solutions in the prior art, the accompanying drawings used in the description of the specific embodiments or the prior art will be briefly introduced below. In all the drawings, similar elements or parts are generally identified by similar reference numerals. In the drawings, the elements or parts are not necessarily drawn to scale.

[0018] Figure 1 This is a cross-sectional schematic diagram of a wind turbine gearbox.

[0019] Figure 2 This is a schematic diagram comparing tooth profiles.

[0020] Figure 3This is a schematic diagram of the tooth profile of the root-cutting gear in this scheme.

[0021] Figure 4 This is a schematic diagram of the tooth shape in the existing root excavation process.

[0022] Figure 5 This is a schematic diagram of the tooth profile of a finished product using existing technology.

[0023] Reference numerals: Rear tooth surface line 31, Front tooth surface line 32, Tooth root transition line 33, Involute starting point 34, Input shaft 4, Output shaft 5, First planetary speed regulating unit 6, First internal gear ring 61, First planetary carrier 62, First sun gear 63, Center hole 631, Second planetary speed regulating unit 7, Second internal gear ring 71, Second planetary carrier 72, Second sun gear 73, Cylindrical gear speed regulating unit 8. Detailed Implementation

[0024] The embodiments of the technical solution of the present invention will now be described in detail with reference to the accompanying drawings. These embodiments are merely illustrative of the technical solution of the present invention and are therefore intended to limit the scope of protection of the present invention.

[0025] It should be noted that, unless otherwise stated, the technical or scientific terms used in this application should have the ordinary meaning as understood by one of ordinary skill in the art to which this invention pertains.

[0026] One type of root-cutting gear, please refer to the appendix. Figure 3 One possible implementation is as follows: the area between adjacent teeth is a tooth groove, which includes: a rear tooth surface line 31 located on the preceding tooth and a front tooth surface line 32 located on the following tooth, and a root transition line 33 connecting the bottom of the rear tooth surface line 31 and the front tooth surface line 32. Both the rear tooth surface line 31 and the front tooth surface line 32 intersect the root transition line 33. The root transition line 33 is higher than the root of the standard tooth profile. The root circle diameter remains unchanged but increases towards the tooth tip, approaching the forming circle diameter dNf and the involute starting point dFf, thus exhibiting a root-digging characteristic relative to the standard tooth profile. Figure 3 As shown, the root transition line 33 does not smoothly transition with the tooth surface lines on both sides, but rather forms a non-tangential intersection point. See also... Figure 2 , Figure 2 The image shows a portion of a gear, including one tooth and two tooth grooves on either side. Taking the right tooth groove as an example, the standard tooth profile in the prior art uses a smooth tangential transition line 33A at the root. However, this design uses a non-tangential transition, i.e., the transition line shown in the image as tooth root transition line 33. This means that some material needs to be removed from the root of the existing gear.

[0027] It should be noted that, Figure 3This image shows the tooth profile of the final gear in this design. While existing technologies, such as the "GBT1356-2001 Standard for Basic Rack Tooth Profiles of Cylindrical Gears for General Machinery and Heavy Machinery," specify tooth profiles... Figure 4 The gear shown is a cut-out gear, but it is not the final tooth profile. Instead, it represents the cut-out allowance (uFP) left during the manufacturing process. During finishing, this allowance (uFP) must be removed to achieve the desired tooth shape. Figure 5 The conventional shape shown is used in existing technologies to allow finishing tools to operate without interference. This tooth profile is suitable for high-precision gears that transmit high torque; therefore, the tooth profile is finished using grinding or shaving. During finishing, it avoids the formation of dents at the tooth root fillets, thus preventing stress concentration caused by dents.

[0028] In this solution, Figure 3 The tooth profile shown is that of the final finished gear, not the intermediate tooth profile during manufacturing. The purpose of this root-cutting design is not to facilitate subsequent surface finishing, but to minimize the root stress coefficient. Increasing the diameter of the root transition line (33) reduces the root stress concentration coefficient; although the root thickness is also reduced, the overall bending strength of the root is improved. This is particularly suitable for multi-planetary gear structures in wind turbine gearboxes. In multi-planetary gear structures, the contact strength is relatively abundant, while the bending strength is relatively low.

[0029] Contact strength is the ability of a gear tooth surface to resist surface fatigue damage. During meshing, the tooth surfaces come into contact under high pressure and experience rolling and sliding, generating enormous alternating shear stress beneath the surface. Prolonged exposure to this stress can lead to micro-fatigue cracks. Crack propagation causes metal to peel off, forming small pits on the tooth surface, a phenomenon known as pitting. Contact strength measures a gear's ability to resist this pitting damage.

[0030] Bending strength is the ability of gear teeth to resist root fracture. When meshing, gear teeth act like a cantilever beam, with the greatest bending stress occurring at the tooth root. This stress is alternating, causing fatigue cracks to initiate at the stress concentration point at the tooth root and gradually propagate, eventually leading to tooth breakage. Bending strength measures a gear's ability to resist this bending fatigue fracture.

[0031] The stress concentration factor at the tooth root directly affects, and only directly affects, the bending strength of the gear, while having virtually no direct impact on the contact strength. This is because during gear transmission, the gear teeth act like a cantilever beam, with the tooth root subjected to alternating bending stress. The geometry at the junction of the tooth root transition curve and the bottom of the tooth space undergoes abrupt changes, leading to a sharp increase in stress at this point—a phenomenon known as stress concentration. A higher stress concentration factor results in a higher calculated bending stress at the tooth root, lower bending fatigue strength, and a greater susceptibility to fatigue tooth breakage.

[0032] In this design, a unique tooth profile is innovatively used to address the issue of relatively high gear contact strength but low bending strength in wind turbine gearboxes. The tooth root transition line 33 is an arc, with its center located on the line of symmetry between the rear tooth surface line 31 and the front tooth surface line 32. The root fillet radius of the basic rack is greater than 0.39m, where m is the module. That is, the root fillet radius is greater than 0.39m; further increasing the root fillet radius is described in the following section. Figure 2 / Figure 3 Theoretically, as long as the involute starting point 34 is not exceeded, the end face overlap will not decrease. In special cases, the root fillet can also exceed the involute starting point 34, thereby further increasing the root fillet by reducing the end face overlap. That is, in some designs, the rear tooth surface line 31 and the front tooth surface line 32 are involutes, and the intersection of the root transition line 33 with the rear tooth surface line 31 and the front tooth surface line 32 exceeds the involute starting point 34. It should be noted that the starting point of the involute usually refers to the point where the mating gear begins to mesh, which is calculated based on the tip circle diameter of the mating gear and is called the involute starting circle dFf, not the base circle itself. See [reference needed]. Figure 2 db is the base circle, df is the root circle, and d is the pitch circle. (See also...) Figure 3 In this design, the overall tooth height coefficient remains unchanged, so the overall tooth height is constant, and the lowest point of the tooth root also remains unchanged. Increasing the fillet radius occupies part of the tooth surface. That is, the fillet of the original tooth root transition line 33A is tangent to both tooth surfaces, forming the tooth profile of a typical working gear. However, in this design, the fillet of the tooth root transition line 33 is a circle determined by the intersection of dFf with the two tooth surfaces and the lowest tooth root point, therefore it is tangent.

[0033] As an alternative to the above embodiment, in this embodiment, the tooth root transition line 33 includes multiple interconnected smooth curves. The gear is made of carburized and quenched steel.

[0034] One possible implementation of a gearbox is as follows: it includes at least one root-type gear as described above.

[0035] The gearbox includes: an input shaft 4, an output shaft 5, a first planetary speed regulating unit 6, a second planetary speed regulating unit 7, and a cylindrical gear speed regulating unit 8; the first planetary speed regulating unit 6 includes a first internal gear ring 61, a first planet carrier 62, and a first sun gear 63; the second planetary speed regulating unit 7 includes a second internal gear ring 71, a second planet carrier 72, and a second sun gear 73; the first internal gear ring 61 and the second internal gear ring 71 are fixed, the input shaft 4 is connected to the first planet carrier 62, the first sun gear 63 is connected to the second planet carrier 72, and the cylindrical gear speed regulating unit 8 is located between the output shaft 5 and the axle of the second sun gear 73.

[0036] The first planetary speed regulating unit 6, the second planetary speed regulating unit 7, and the cylindrical gear speed regulating unit 8 are arranged in parallel. The first sun gear 63 and the second planetary carrier 72 are connected by the same component. The first sun gear 63 has a central hole 631, and the axle of the second sun gear 73 also has a central hole 631. The two central holes are coaxial, and a hollow tube passes through the central holes to penetrate the gearbox. The first planetary speed regulating unit 6 and / or the second planetary speed regulating unit 7 have more than 4 planetary gears, preferably 5-15.

[0037] In the above embodiment, power is input from the first-stage planetary carrier 62. The input shaft 4 (connected to the propeller) is integrated with the first planetary carrier 62. The first internal gear ring 61 is fixed and cannot rotate. When the input shaft 4 drives the first-stage planetary carrier 62 to rotate, the first sun gear 63 will rotate on its own axis, and the first planetary gears will rotate on their own axis and revolve around the sun. Since the first sun gear 63 and the second planetary carrier 72 are connected to the same component, the power reaches the second planetary speed regulating unit 7. The second internal gear ring 71 is fixed. When the second planetary carrier 72 rotates, the second sun gear 73 will rotate on its own axis. Since the shaft of the second sun gear 73 is transmitted to the output shaft 5 through the cylindrical gear speed regulating unit 8, the input power reaches the output shaft 5 (connected to the generator) after passing through two stages of planetary reduction and one stage of cylindrical gear reduction. The overall speed ratio can reach 40 to 60.

[0038] The following example, a 10MW NGW wind turbine gearbox, will further illustrate this. IEC 61400-4 requires that, when calculated according to ISO 6336-2006, SH ≥ 1.25 and SF ≥ 1.56. SH is the contact strength safety factor. SF is the bending strength safety factor. The failure mode for insufficient contact strength (SH) is pitting. The key to improving SH in existing technology lies in increasing the combined radius of curvature of the tooth surface, thereby reducing Hertzian contact stress. The failure mode for insufficient bending strength (SF) is tooth breakage. The key to improving SF in existing technology lies in increasing the tooth root thickness and reducing stress concentration at the tooth root. The table below shows the basic parameters of a planetary gear train in this gearbox: In existing technologies, the tooth profile is designed according to conventional tooth profiles, resulting in a tooth profile such as... Figure 5 As shown, the parameters are as follows: In this scheme, the root fillet radius coefficient ρfP is increased so that the intersection of the root circle and both sides is at the starting point 34 of the involute (dNF in the table), with the following parameters: Continue to increase the root fillet radius coefficient ρfP so that the intersection of the root circle and both sides exceeds the starting point of the involute by 34. For the planetary gears, use 0.65, and for the internal gear ring, use 0.5. The parameters are as follows: As we can see, dFf increases to dNf, but this is difficult to control in actual manufacturing. We need dNf ≥ dFf with a certain difference, which can be achieved by actually increasing the chamfer of the meshing tooth tip. However, this will reduce the actual contact ratio of the end faces, which is generally considered by increasing the values ​​of the safety factors SH and SF. The optimal value is when dNf = dFf, at which point the theoretical contact ratio does not decrease, which is reflected in the SH value, meaning the SH value remains unchanged. dNf ≤ dFf is not impossible; theoretically, it is possible. However, the impact of the decrease in contact ratio on gear strength is complex, and there is currently no standard procedure to calculate it.

[0039] Building upon the previous step, reduce the tooth width (the overall width of the gear, distinct from the root width of a single tooth), balancing SH and SF. It's important to consider that the actual end face might reduce this width. Consider SH ≥ 1.3 (> 1.25, IEC requirement), SF ≥ 1.6 (> 1.56, IEC requirement), with the following parameters: As can be seen, the tooth width can be reduced to 440, which is about 14% less. The reduction in tooth width is also limited by the planetary gear bearings. In the specific design, it is also necessary to consider the selection of appropriate bearing types.

[0040] In summary, when NGW planetary gears employ multiple planetary gears, SH and SF do not change equally; SF changes more slowly. This necessitates effectively increasing SF, even at the cost of sacrificing face overlap. Increasing the root fillet can effectively reduce the root stress concentration factor and increase SF, thereby balancing SH and SF.

[0041] The root-cutting gear provided in this solution and its application in wind turbine gearboxes have produced significant beneficial effects, mainly reflected in two aspects: improved gear performance and system-level optimization. The core benefit of the root-cutting gear is a significant improvement in bending fatigue strength. The fundamental innovation of this design lies in using a non-smooth, secant large circular arc or multiple smooth curves as the tooth root transition line. This increases the root fillet radius coefficient to over 0.39m, even exceeding the involute starting point. This directly and drastically reduces the root stress concentration factor, thereby fundamentally improving the gear's resistance to bending fatigue fracture (i.e., the bending strength safety factor SF). Although the root thickness may be slightly reduced, the benefits of reducing stress concentration far outweigh the losses, achieving a net increase in the overall bending strength of the tooth root.

[0042] This solution offers numerous advantages in multi-planetary gear scenarios for wind turbine gearboxes: To precisely address the issue of strength imbalance, this design provides a targeted solution to the specific pain point of abundant contact strength (SH) but relatively insufficient bending strength (SF) in multi-planetary gear transmissions in wind turbine gearboxes. By prioritizing and significantly enhancing the weaker bending strength, SH and SF are effectively balanced, resulting in a more rational and balanced strength configuration for the transmission system.

[0043] Achieving lightweight design and cost reduction, along with a significant improvement in bending strength, provides design flexibility to reduce gear tooth width and thus lower contact strength. Examples show that, while meeting IEC safety standards (SH>1.3, SF>1.6), tooth width can be reduced by approximately 14%. This directly leads to a significant reduction in gearbox volume, weight, and material costs, which is of great significance for the wind power industry's efforts to reduce costs and increase efficiency.

[0044] To ensure highly reliable operation, this design gives gears, especially the vulnerable planetary gears in multi-planetary gear sets, greater resistance to tooth breakage, directly improving the reliability and service life of the entire gearbox and reducing maintenance costs.

[0045] In summary, this solution, through its innovative "functional" tooth profile design, not only achieves a significant improvement in the bending strength of the gear body, but also realizes strength balance, lightweighting, and high reliability at the system application level, providing a high-performance and high-economic optimization path for large wind turbine gearboxes.

[0046] Finally, it should be noted that the above embodiments are only used to illustrate the technical solutions of the present invention, and not to limit them; although the present invention has been described in detail with reference to the foregoing embodiments, those skilled in the art should understand that modifications can still be made to the technical solutions described in the foregoing embodiments, or equivalent substitutions can be made to some or all of the technical features; and these modifications or substitutions do not cause the essence of the corresponding technical solutions to deviate from the scope of the technical solutions of the embodiments of the present invention, and they should all be covered within the scope of the claims and specification of the present invention.

Claims

1. A root-cutting type gear, characterized in that, The area between adjacent teeth is a tooth groove, which includes: a rear tooth surface line (31) located on the preceding tooth and a front tooth surface line (32) located on the following tooth, and a tooth root transition line (33) connecting the bottom of the rear tooth surface line (31) and the front tooth surface line (32). The rear tooth surface line (31) and the front tooth surface line (32) both intersect the tooth root transition line (33).

2. The root-digging gear as described in claim 1, characterized in that: The tooth root transition line (33) is an arc line, and the center of the arc line is located on the line of symmetry between the rear tooth surface line (31) and the front tooth surface line (32).

3. A root-digging gear as described in claim 1, characterized in that: The tooth root transition line (33) includes multiple interconnected smooth curves.

4. A root-digging gear as described in claim 2, characterized in that: The root fillet radius of the basic rack of the root-cutting gear is greater than 0.39m, where m is the module.

5. A root-digging gear as described in claim 1, characterized in that: The rear tooth surface line (31) and the front tooth surface line (32) are involutes, and the intersection of the tooth root transition line (33) with the rear tooth surface line (31) and the front tooth surface line (32) exceeds the starting point (34) of the involute.

6. A root-digging gear as described in claim 1, characterized in that: The gears are made of carburized and quenched steel.

7. A gearbox, characterized in that: It includes at least one root-digging gear as described in any one of claims 1-6.

8. A gearbox as described in claim 7, characterized in that, include: Input shaft (4), output shaft (5), first planetary speed regulating unit (6), second planetary speed regulating unit (7) and cylindrical gear speed regulating unit (8); The first planetary speed regulating unit (6) includes a first internal gear ring (61), a first planet carrier (62), and a first sun gear (63); The second planetary speed regulating unit (7) includes a second internal gear ring (71), a second planet carrier (72), and a second sun gear (73); The first internal gear ring (61) and the second internal gear ring (71) are fixed. The input shaft (4) is connected to the first planetary carrier (62). The first sun gear (63) is connected to the second planetary carrier (72). The cylindrical gear speed regulating part (8) is located between the output shaft (5) and the second sun gear (73).

9. A gearbox as described in claim 8, characterized in that: The first planetary speed regulating unit (6), the second planetary speed regulating unit (7) and the cylindrical gear speed regulating unit (8) are arranged in parallel. The first sun gear (63) and the second planetary carrier (72) are connected to the same component. The first sun gear (63) is provided with a central hole (631). The axle of the second sun gear (73) is also provided with a central hole. The two central holes are coaxial. The hollow tube passes through the central hole and penetrates the gearbox.

10. A gearbox as described in claim 8, characterized in that: The first planetary speed regulating unit (6) and / or the second planetary speed regulating unit (7) have more than 4 planetary gears.

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