Gear arrangement

Abstract

A gearbox with a small tooth difference internal and external gear pair, the internal and external gear pair comprising an internal gear arranged around an internal gear shaft and having a first number of internal gear teeth, and an external gear having an external gear shaft arranged to rotate around the internal gear shaft and having a second number of external gear teeth arranged to mesh with the internal gear teeth, the small tooth difference between the first number of external gear teeth and the second number of internal gear teeth is less than one-fifth the number of internal gear teeth, the internal gear teeth and the external gear teeth have involute shapes with no profile shift or with a profile shift less than a value that would cause the working pressure angle of the internal and external gear pair to exceed the nominal pressure angle of the internal and external gear pair by 20% when no profile shift is applied, and at least one of the internal gear teeth and the external gear teeth has a profile shift coefficient less than 0.8. Such a gearbox can provide improved efficiency.

Description

[Technical field]

[0001] The present invention relates to small tooth difference internal and external gear pairs and their use in torque converting gearboxes. [Background technology]

[0002] An internal-external gear pair may have a ring gear with inwardly facing gear teeth and a relatively large pinion with outwardly facing gear teeth located inside the ring gear. For such large pinions, standard gear profiles and moduli can lead to interference between the tips of the teeth of the two gears, known as tooth interference. Known methods of addressing interference include applying a profile shift to the gear tooth profiles, but profile shifts can reduce the efficiency of the gears. Summary of the Invention [Problem to be solved by the invention]

[0003] There is a need for an improved arrangement of small tooth difference internal and external gear pairs. [Means for solving the problem]

[0004] According to a first aspect of the present invention, there is provided a gearbox including an internal and external gear pair with small tooth difference, the internal and external gear pair having any or all of the following features: an internal gear arranged around an internal gear shaft and having a first number of internal gear teeth; an external gear having an external gear shaft arranged to rotate around the internal gear shaft and having a second number of external gear teeth arranged to mesh with the internal gear teeth; the small tooth difference between the first number of external gear teeth and the second number of internal gear teeth is less than one fifth of the number of internal gear teeth, the internal gear teeth and the external gear teeth having an involute shape with no profile shift or less than a value where the pressure angles of the external gear teeth and the internal gear teeth differ by more than 20% of the working pressure angle of the external gear teeth. The internal gear teeth and the external gear teeth may have a profile shift coefficient of less than 0.8. Such a gearbox may provide improved efficiency.

[0005] For any of the disclosed internal and external gear pairs, the working pressure angle of the gear pair formed by the internal and external gears may exceed the nominal pressure angle of the same gear pair if no profile shifting was applied by less than 20%, optionally less than 15%, less than 10%, or less than 5%.

[0006] For any of the disclosed internal and external gear pairs, the pressure angles of the internal and external gear teeth measured at the midpoint of the working portion of the gear tooth flanks may be the same. Optionally, the pressure angles may differ by less than 20% of the pressure angle of the external gear tooth.

[0007] For any of the disclosed gear pairs, the internal gear teeth and the external gear teeth can have an involute profile with no profile shift or with a profile shift less than the value that would cause the working pressure angle of the gear pair to exceed the nominal pressure angle of the gear pair by 20% if no profile shift were applied. The internal gear teeth and the external gear teeth can have a profile shift coefficient of less than 0.8. Such a gearbox can provide improved efficiency.

[0008] The sum of the profile shift coefficients of the internal and external gear teeth of any of the gear pairs disclosed herein may be less than 1.6.

[0009] The theoretical contact ratio of an internal / external gear pair, calculated for a gearbox with the same geometry and no material deformation, can be less than 1. The low contact ratio can be achieved as a result of no profile shift and having a profile shift coefficient of less than 0.8, whereas a contact ratio of greater than 1 after material deformation actually allows the gearbox to transmit torque more consistently.

[0010] The profile modification coefficient may be less than 0.7, optionally less than 0.6, which may provide additional clearance to avoid trochoidal interference when the bearings have play and may be necessary when the difference in the number of teeth between the external and internal gears is very small.

[0011] The addendum coefficient is greater than 0.3, and more preferably greater than 0.5, which can reduce the surface stress on the tooth surface and improve the life of the gear.

[0012] The theoretical contact ratio of the internal and external gear pair, calculated for a gearbox having the same geometry and no material deformation, may be less than 0.8, optionally less than 0.6, and even less than 0.4.

[0013] The profile shift coefficient of the internal gear is less than 0.05, optionally less than 0.04, further optionally less than 0.03, further optionally less than 0.02, further optionally less than 0.01. There may be no profile shift.

[0014] The pitch diameter of the external gear is at least half the pitch diameter of the internal gear. The internal gear can cover or overlap the shaft of the external gear while in contact with the teeth of the external gear.

[0015] The load contact ratio of the gearbox occurring in use may be 100% greater than the theoretical contact ratio calculated for a gearbox having the same geometry and no material deformation, optionally the load contact ratio of the gearbox occurring in use may be 200% greater than the theoretical contact ratio, optionally 400% greater than the theoretical contact ratio, optionally 500% greater than the theoretical contact ratio, and optionally 1000% greater than the theoretical contact ratio. This allows for a gearbox that is free of interference due to unnecessary gear contact while maintaining an optimal pressure angle, and the gearbox may experience reduced wear during use due to the high effective contact ratio effectively distributing the load between the different teeth. Also, the efficiency of the gearbox is improved. The addendum may be less than 0.1.

[0016] An internal / external gear pair may have an instantaneous center of rotation ICR defined as the point where a radius drawn through the internal and external gear axes intersects the pitch circles of the internal and external gears, and all tooth pairs that transmit load between the external and internal gears may be contained within a 20 degree arc about the ICR.

[0017] Each of the internal and external gears has 40 or more teeth.

[0018] The external gear may be a first external gear, the internal gear may be a first internal gear, and the gearbox may further comprise a second external gear coupled to the first external gear and a second internal gear movable relative to the first internal gear, the second external gear being arranged inside and meshing with the second internal gear. In this way, a differential planetary gearbox may be formed, providing a higher gear ratio.

[0019] The second external gear and the second internal gear may form a second small tooth number difference internal / external gear pair having a tooth number difference of less than one-fifth of the number of teeth of the second internal gear, and the teeth of the second internal gear and the teeth of the second external gear have involute shapes with no profile shift or less than a value such that the respective operating pressure angles of the teeth of the second external gear and the teeth of the second internal gear differ by more than 20% of the operating pressure angle of the teeth of the second external gear, and at least one of the teeth of the second internal gear and the teeth of the second external gear may have a profile shift coefficient of less than 0.8.

[0020] In general, the first and second internal and external gear pairs can be designed substantially similarly, except for a difference in the number of teeth between the first and second external gears and / or between the first and second internal gears.

[0021] The first external gear may be fixed to the second external gear, and the second internal gear may be arranged to rotate relative to the first internal gear about the internal gear axis.

[0022] An input can be provided to drive the first external gear to orbit the first internal gear shaft, causing an output as relative rotation between the first and second internal gears about the internal gear shaft.

[0023] The meshing between the first internal gear and the first external gear can be axially aligned with the meshing between the second internal gear and the second external gear. In other words, the two gear pairs can have meshing aligned in a plane perpendicular to the rotational axis of the gears. This can reduce gear twist due to meshing forces and can also provide a more compact gearbox.

[0024] One or both of the first and second external gears may include a stepped portion, with the gear teeth lying in a different plane than a central portion of the gear, which may allow for compact packaging of the gears.

[0025] The meshing between the first internal gear and the first external gear and the meshing between the second internal gear and the second external gear can overlap a common plane P1, preferably perpendicular to one or both of their rotation axes, and can further be coplanar.

[0026] The gearbox may further comprise a third internal gear and a third external gear, the third external gear may be arranged inside and meshing with the third internal gear, the third external gear may be coupled to the first and second external gears such that the second external gear is between the first external gear and the third external gear, and the third internal gear may be fixed relative to the first internal gear. This provides a symmetrical gearbox with a larger torque capacity due to axial balancing of the external gears and reduces twisting of the external gears.

[0027] All gears may be spur gears, i.e. non-helical gears. The gear teeth may be substantially parallel to the gear's axis of rotation. Note that helix angle and contact ratio are interrelated features considered together in gear design, and the contact ratios described herein may not be reproducible with helical gears.

[0028] According to a second aspect of the present invention, there is provided a symmetric gearbox comprising a first gearbox according to the first aspect and a second gearbox according to the first aspect, each of the gearboxes of the first aspect having a plurality of external gears and a plurality of internal gears, the first internal gears being fixed together and the second internal gears being fixed together, and the first gearbox and the second gearbox being arranged symmetrically and coaxially.

[0029] The external gears of the first and second gearboxes may be arranged to rotate 180° out of phase with each other about the internal gear shaft, or in other words, diametrically opposed with respect to the internal gear shaft, providing a more balanced external gear set and a more balanced gearbox overall.

[0030] According to a third aspect of the invention, there is provided an internal and external gear pair comprising any or all of the features of: an internal gear arranged around an internal gear shaft and having a first number of internal gear teeth; and an external gear having an external gear shaft arranged to rotate around the internal gear shaft and having a second number of external gear teeth arranged to mesh with the internal gear teeth, wherein the internal gear teeth and the external gear teeth have respective pressure angles measured at midpoints of the working portions of the gear teeth, the pressure angles differing by less than 20% of the pressure angle of the external gear teeth, and at least one of each of the internal gear teeth and the external gear teeth has a profile shift coefficient of less than 0.8. The internal gear teeth and the external gear teeth may have the same pressure angle.

[0031] According to a fourth aspect of the present invention, there is provided an internal / external gear pair comprising any or all of the features of: an internal gear arranged around an internal gear shaft and having a first number of internal gear teeth; and an external gear having an external gear shaft arranged to rotate around the internal gear shaft and having a second number of external gear teeth and arranged to mesh with the internal gear teeth, wherein the internal gear teeth and the external gear teeth have a standard module and are free of profile shifts, and at least one of each of the internal gear teeth and the external gear teeth has a profile shift coefficient of less than 0.8.

[0032] According to a fifth aspect of the present invention, a gearbox is provided having any or all of the features of a first internal / external gear pair and a second internal / external gear pair, wherein the first internal / external gear pair comprises a first internal gear and a first external gear, the first external gear being arranged inside the first internal gear and meshing with the first internal gear, the second internal / external gear pair comprises a second internal gear and a second external gear, the second external gear being arranged inside the second internal gear and meshing with the second internal gear, the first internal gear and the second internal gear are rotatable relative to each other, the first external gear and the second external gear are fixed to each other, the first external gear has a stepped portion, and the meshing of the first internal gear and the first external gear and the meshing of the second internal gear and the second external gear at least partially overlap in a common plane.

[0033] This allows a more compact gearbox to be provided and reduces the offset between the external gears, thereby reducing the twist of the external gears.

[0034] The first and second internal gears can be rotatable about a first axis, the first and second external gears can rotate about the first axis, and / or the common plane can be perpendicular to the first axis.

[0035] The first and second external gears can form a compound pinion gear, and the compound pinion gear can have a recess for at least partially receiving the second internal gear, which may be entirely received within the recess.

[0036] The meshing of the first internal gear with the first external gear and the meshing of the second internal gear with the second external gear can be coplanar. Overall, the teeth of the first internal gear, the second internal gear, the first external gear and the second external gear can be coplanar.

[0037] The gearbox may further comprise a crank-shaped input shaft arranged to drive the first and second external gears. The input shaft may be used to provide torque to the gearbox. One of the internal gears may be fixed and one of the internal gears may be movable, such that the torque is output from the movable internal gear.

[0038] The gears of the gearbox of the fifth aspect of the invention may have teeth with any or all of the features of the teeth described with reference to any other aspect of the invention.

[0039] Any feature(s) of the third and / or fourth aspect of the invention may be combined with any of the above-mentioned properties and features of the first and second aspects of the invention. Any or all of the features of the fifth aspect of the invention may be combined with any of the above-mentioned properties and features of the first to fourth aspects of the invention. [Brief description of the drawings]

[0040] [Figure 1] 1 shows an internal and external gear pair with a small tooth difference exhibiting trochoidal interference. [Diagram 2] This shows an internal and external gear pair with a small tooth number difference that has no trochoid interference and exhibits profile shifting. [Diagram 3] Shown is a pair of internal and external gears with a small tooth number difference, no profile shift, small profile shift coefficient, and no trochoidal interference. [Figure 4] The change in gear tooth profile due to the change in addendum modification coefficient is shown. [Figure 5a] 1 shows a profile shifted gear with specific geometrical characteristics indicated. [Figure 5b] 1 shows a profile shifted gear with specific geometrical characteristics indicated. [Figure 5c] 1 shows a profile shifted gear with specific geometrical characteristics indicated. [Figure 6a] A profile-free gear with specific geometrical characteristics is shown. [Figure 6b] A profile-free gear with specific geometrical characteristics is shown. [Figure 6c] A profile-free gear with specific geometrical characteristics is shown. [Figure 7] The instantaneous centres of rotation of the internal and external gear pairs are shown, along with the teeth in contact. [Figure 8a] FIG. 2 is a schematic diagram illustrating a first gearbox according to an embodiment of the present invention. [Figure 8b] FIG. 2 is a cross-sectional view showing a first gearbox according to an embodiment of the present invention. [Figure 9] FIG. 4 is a schematic diagram illustrating a second gearbox according to an embodiment of the present invention. [Figure 10a] FIG. 13 is a schematic diagram illustrating a third gearbox according to an embodiment of the present invention. [Figure 10b] FIG. 11 is a cross-sectional view showing a third gearbox according to an embodiment of the present invention. [Figure 11] FIG. 11 is a schematic diagram illustrating a fourth gearbox according to an embodiment of the present invention. DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

[0041] Small tooth difference internal and external gear pairs are useful in making high ratio gearboxes or torque converters. Such gear pairs can be used in two-stage pairs to make high ratio gearboxes that can turn a relatively high speed, low torque input into a low speed, high torque output. The efficiency of such gearboxes varies with the way the tooth profiles are arranged. In addition, the ability to drive a high ratio gearbox backwards, i.e., to apply a low speed, high torque input to the output and create a relatively high speed, low torque output at the input, can be important in certain applications. This can be called backdrivability of the gearbox. Minimizing losses in the gearbox is important to enable backdrivability and overall efficiency of the gearbox in either drive direction. The inventors have identified that the long-standing standard approach to gear design in small tooth difference internal and external gear pairs can have detrimental results on the overall performance of the gear pair.

[0042] When providing such a gearbox having internal and external gears with a relatively small difference between the number of teeth on each gear, problems can arise where the teeth of the two gears interfere, as shown in FIG. 1 .

[0043] FIG. 1 shows a gear arrangement 10 having an internal gear 12 and an external gear 16, the internal gear 12 having 47 teeth 14 and a pitch circle C1, the internal gear teeth 14 being arranged to mesh with the external gear teeth 18 of the external gear 16, the ring gear 16 having 50 external teeth 18 and a pitch circle C2.

[0044] It should be understood that a variety of different terms can be used for the gears. The internal gear 12 may be referred to as a pinion, spur gear, or external gear because the gear teeth face outward. Similarly, the external gear 14 may be referred to as an annular gear, ring gear, or internal gear because the gear teeth face inward. Both the internal gear 12 and the external gear 14 may be straight cut gears, i.e., non-helical gears.

[0045] The internal gear 16 is centered on the internal gear axis A2 and the external gear 12 is centered on the external gear axis A1. It should be understood that in a suitable arrangement, as the external gear 12 rotates, it can rotate about the internal gear axis A2. In some arrangements, the external gear axis A1 and the internal gear axis A2 can remain stationary and each of the external and internal gears 12, 16 can rotate about their respective axes A1, A2. The kinematic configuration of a gearbox using an internal and external gear pair is described in more detail in connection with subsequent figures.

[0046] In the meshing region 24, the teeth 14 of the external gear 12 mesh with the internal gear 18. In this region, the flanks of the respective teeth 14, 18 contact each other and torque is transmitted between the two gears 12, 16.

[0047] The gears in FIG. 1 are designed with a standard module and involute (the standard gear module is in accordance with the well-known Japanese Industrial Standard JIS B 1701 (1973), the details of which are incorporated below for reference). The teeth of the standard gear will have a nominal pressure angle of 20°. JIS B 1701 specifies the tooth profiles and dimensions of involute spur and helical gears for general use. The size of the teeth is expressed in module m. Standard values ​​of module m are given in mm in Table 1 below. Preferably the module is of the series No. 1, No. 2 or No. 3. Table 1 Module standard values TIFF2025512007000002.tif161158

[0048] The tooth profile curve is an involute curve, and the standard pressure angle is α 0 The pressure angle shall be 20° and the dimensions shall be as specified in Table 2. However, no external modifications are specified in this standard. Other standard tooth forms may have pressure angles of 5°, 14.5°, 20°, 22.5°, 25°, and 30°. Table 2 Dimensions of involute spur gears and helical gears TIFF2025512007000003.tif123169

[0049] The profile shift coefficient is determined by the working pressure angle αw, the nominal pressure angle α (as in the other examples above, the nominal pressure angle according to generally accepted standards is usually 20°), the involute function inv of the associated angle, and the number of teeth of the internal gear z 1 and the number of teeth of the external gear z 2 For helical gears, the helical angle β can also be used. The following formula is the profile shift coefficient x of the first gear 1 can be used to determine the profile shift coefficient x of the second gear. 2 It should be appreciated that rearrangement can be used to determine TIFF2025512007000004.tif14170

[0050] The center distance a of the internal and external gear pair can be determined using the following formula, where m is the gear module and y is the center distance increment factor. TIFF2025512007000005.tif17170

[0051] The center distance increment factor is determined by the following formula: TIFF2025512007000006.tif18170

[0052] Use of the described standard gears can cause trochoidal interference, since the tips of the teeth of the external gear 14 can contact the tips of the teeth of the internal gear 18 outside the intended area 24 of contact. This can be seen in the area 22 marked in Figure 1. Trochoidal interference is fatal for the gearbox and in the case shown in Figure 1 can fundamentally prevent the gearbox from being built.

[0053] To avoid trochoidal interference, a common solution is to shift the gears. An example of a shifted gear tooth profile used to avoid trochoidal interference is shown in Figure 2. Gear arrangement 30 shows an external gear 32 with 47 outward gear teeth 34 arranged about the external gear axis A3.

[0054] Outboard of the external gear 32, which may be referred to as a pinion, is an internal gear 36. The internal gear 36 has inwardly facing gear teeth 38 that are positioned to contact the outwardly facing gear teeth 34 at a contact area 44.

[0055] Traditional gear design may require the selection of the number of teeth on each gear, determined by the center distance between the two gears (i.e., the distance between their axes) and the desired gear ratio. For two spur gears, the pitch diameter of each gear is typically selected so that the sum of the pitch diameters is twice the center distance, and the ratio of the two pitch diameters is the same as the ratio of the numbers of teeth (i.e., the gear ratio). For an internal / external gear pair, the difference between the pitch diameters is twice the center distance of the gears. The pitch diameter thus determined is sometimes called the nominal pitch diameter.

[0056] Profile shifting of a gear involves moving a cutting tool toward or away from the gear during cutting to adjust the gear's working pitch diameter. Thus, a profile shifted gear can have a working pitch diameter that differs from its nominal pitch diameter.

[0057] The base circle of the involute that characterizes the tooth flank remains the same during the shift so that the same cutting tool can be used. However, the shift effectively changes the portion of the involute that characterizes the tooth flank, moving part of the involute closer to or away from the base circle.

[0058] It can be seen that there is no trochoidal interference between the external and internal gear teeth 34, 38. This is due to the profile shift applied to the gear teeth. Profile shifting involves modifying a standard gear cutting tool so that the involute profile of the gear teeth remains unchanged and the involute base circle remains in the same position. Thus, profile shifting changes the gear pitch circle from the nominal pitch circle of the standard profile to the profile that has been profile shifted. However, gear profile shifting increases the working pressure angle of the gears, one effect of which is a reduction in the efficiency of the gears.

[0059] As a result, the profile shifted gear of FIG. 2 has a different pressure angle than the nominal pressure angle of 20° for the standard involute profile (in the case of FIG. 2, the pressure angle is 37.42°). Also, the profile shifted teeth of FIG. 2 have a standard addendum coefficient of 1. A high working pressure angle can lead to reduced bearing life and reduced gearbox efficiency due to higher sliding speeds between teeth 34 and 38 while they are carrying the load in meshing area 44.

[0060] FIG. 3 shows a gear arrangement 100 intended to avoid trochoidal interference problems while maintaining high efficiency. The gear arrangement 100 has an internal or ring gear 106. The illustrated example has 50 internal gear teeth 108, but other numbers of teeth can be envisaged. The example also has an external or pinion gear 102 with 47 external gear teeth 104, but other external gears with a small difference in the number of teeth of the internal gear 106 can be envisaged. In general, the interference problems discussed above with reference to FIG. 1, for example, occur when an internal / external gear pair is constructed with a small difference in the number of teeth between the internal and external gears, for example a difference of less than 5 teeth or about one fifth of the number of teeth of the internal gear.

[0061] The internal gear 106 is disposed about the internal gear axis A6 and the external gear 102 is disposed about the external gear axis A5. It should be appreciated that if the internal gear 106 can be held stationary, the external gear 102 can rotate and orbit about the external gear axis A6.

[0062] 3 have a particular number of teeth, with the internal gear 106 having 50 teeth 108 and the external gear 102 having 47 teeth 104, it should be understood that the external and internal gears can have any number of teeth where the difference in the number of teeth is relatively small. For example, the difference in the number of teeth is less than 5, with the internal gear 106 having less than 5 more teeth than the external gear 102. Alternatively, the difference in the number of teeth can be considered as a percentage, where the difference in the number of teeth can be less than 20% or less than a fifth, optionally less than 10% or less than a tenth, of the number of teeth of the internal gears.

[0063] In the gear arrangement 100, trochoidal interference is avoided by changing the profile shift coefficient. The gears do not exhibit profile shifts, which can be determined by the fact that the pressure angle at a point M located in the middle of the active profile is the same for both gears, the pressure angle being measured as the angle between the normal to the tooth flank and the tangent to a circle centered on the center of the gear and passing through this same point M. The pressure angles of the two gears at point M may differ by up to 20% without significantly affecting the working principle of not producing profile shifts. The lack of profile shifts can be determined in some cases due to the use of a standard module and the pressure angles of the gear teeth being standard pressure angles. The standard module may be as defined above with reference to FIG. 1 and Japanese Industrial Standard JIS B 1701 (1973). In particular, the profile shift coefficients of both the external gear teeth 104 and the internal gear teeth 108 are less than 0.8. However, the external gear teeth 104 and the internal gear teeth 108 may have different profile shift coefficients. This is sometimes called an asymmetric profile shift coefficient. Generally, to avoid trochoidal interference, the sum of the profile shift coefficients can be less than 1.6. Furthermore, a small profile shift coefficient undesirably leads to excessive surface pressure under load, so the sum of the profile shift coefficients can be greater than 0.6. Thus, the sum of the profile shift coefficients of the external gear teeth and the internal gear teeth can be between 0.6 and 1.6. More preferably, the sum of the profile shift coefficients is between 1 and 1.6.

[0064] In conventional gear design techniques and related literature, it has long been the prevailing view that reducing the profile shift coefficient of a gear pair below 0.8 results in an undesirably low contact ratio. The contact ratio is the average number of teeth on either gear that are in constant contact with the teeth on the other gear. Thus, if the contact ratio is less than one, the gear arrangement will have periods when the gears are not in contact and the gears cannot effectively mesh correctly, i.e., in a smooth and effective manner. This can lead to inconsistent torque transfer between the gears, severe vibrations, and damage to the gear teeth. For these reasons, contact ratios less than one are generally considered undesirable, and conventional practice and gear design literature teach that contact ratios less than one should be avoided.

[0065] In the arrangement of FIG. 3, the theoretical contact ratio, i.e. the contact ratio that would be calculated based on the undeformed teeth at rest, would be less than 1. However, due to deformation of the teeth under load, i.e. while torque is being transmitted by the gears, the contact ratio may increase to or even exceed 1. This contact ratio may be called the loaded contact ratio. Thus, the gearbox may initially appear not to function, but may function in practical situations. Deformation under load will be readily understood by those skilled in the art, at least in the extreme case where the strain is below the ultimate tensile strain. In most gearboxes with a contact ratio less than 1, the contact ratio will not increase to 1 even if the material is deformed as much as possible without plastic deformation or fracture. However, in the present case, due to the elastic deformation of the gear teeth, the contact ratio can change from less than 1 to more than 1.

[0066] Figures 4a and 4b show the change in gear tooth profile with the change in the profile shift coefficient ac for the internal gear tooth 108 and the external gear tooth 104. From Figures 4a and 4b it can be seen that the base circle and pitch circle of the involute are constant as well as the gear tooth module. Thus, the pressure angle remains a constant nominal pressure angle. In effect, the reduction of the profile shift coefficient truncates the tooth by removing the upper part of the tooth. As shown in Figure 4b, contact with the tooth tip can therefore be avoided without changing the desired pressure angle provided by the tooth profile.

[0067] The profile shift of a gear tooth is the distance the gear tooth protrudes beyond the pitch circle and can be seen as ac=0. The profile shift coefficient ac is a dimensionless number obtained by dividing the profile shift by the tooth module. The conventional profile shift coefficient is 1, but by reducing the profile shift coefficient, the possibility of interference between the tips of the gear teeth can be reduced. Therefore, it can be advantageous to design gears with a profile shift coefficient of less than 0.8, or optionally even less.

[0068] 5a to 6c are presented to more clearly show how the translocation is determined.

[0069] FIG. 5a shows the profile shifted gear of FIG. 2 in the meshing region 44. The working part of the internal gear 38 is at point V I and point U I The range is defined as V I is a point on the internal gear tooth 38 that is on the outer radius of the external gear 32 and that contacts the tip of the external gear tooth 34. I is the tip of the internal gear tooth 38. Similarly, for the external gear tooth 34, point U E and point V E is stipulated. E is the tip of the tooth 34 of the external gear, and V E is the position on the tooth 34 of the external gear where the tip of the tooth 38 of the internal gear comes into contact. E is the point U when the gear teeth 34 and 38 are in the meshing region. I Therefore, the only part of the tooth flanks that come into contact with each other is the point U I , VI , U E , and U E It should be understood that the working portion of the teeth is defined by:

[0070] Referring to FIG. 5b, a further point M I is prescribed. I is U I and V I The pressure angle α I is the circle C around the gear axis M Tangent T M And, M. I Normal line N of the tooth surface at M For completeness, the circle C is defined as the angle between M Radius R M The angle α I is measured at the center of the working part of the tooth and can therefore be described as the working pressure angle.

[0071] FIG. 5c shows the equivalent point on the external gear 32. A further point M E is prescribed. E is U E and V E The pressure angle α is the midpoint between E is the circle C around the gear axis M Tangent T M And, M. E Normal line N of the tooth surface at M For completeness, the circle C is defined as the angle between M Radius R M The angle α E is measured at the center of the working part of the tooth, and therefore can be described as the working pressure angle. As mentioned above, the respective midpoints M of the working parts of the internal and external gears I and M E By comparing the pressure angles of the two gears measured at , it is found that the pressure angles differ by more than 20% of the pressure angle of the external gear. The gears therefore exhibit a profile shift.

[0072] FIG. 6a shows the gears of FIG. 3 in the meshing region 110, where the gears have no profile shifts. The working part of the internal gear tooth 108 is at point VI and point U I The range is defined as V I is a point on the internal gear tooth 108 that is on the outer radius of the external gear 102 and that contacts the tip of the external gear tooth 104. I is the tip of the internal gear tooth 108. Similarly, for the external gear tooth 104, point U E and point V E is stipulated. E is the tip of the paddle tooth 104, and V E is the location on the external gear tooth 104 where the tip of the internal gear tooth 108 makes contact. Thus, the only part of the tooth flanks that make contact with each other is the point U I , V I , U E , and U E It should be understood that the working portion of the tooth is defined by the area between.

[0073] Referring to FIG. 6b, a further point M I is prescribed. I is U I and V I The pressure angle α I is the circle C around the gear axis M Tangent T M And, M. I Normal line N of the tooth surface at M For completeness, the circle C is defined as the angle between M Radius R M The angle α I is measured at the center of the working part of the tooth and can therefore be described as the working pressure angle.

[0074] FIG. 6c shows the equivalent point on the external gear 102. A further point M E is prescribed. E is U E and V E The pressure angle α E is the circle C around the gear axis M Tangent T M And, M. E Normal line N of the tooth surface at MFor completeness, the circle C is defined as the angle between M Radius R M The angle α E is measured at the center of the working part of the tooth and can therefore be described as the working pressure angle.

[0075] As can be seen from Figures 6b and 6c, the pressure angles at the midpoint of the working parts of the internal and external gear teeth are the same.

[0076] As can be seen from the above description, the following becomes clear: the presence or absence of profile shifting can be detected by detecting the difference in pressure angle between the internal gear teeth and the external gear teeth; the pressure angle of the internal gear teeth at the midpoint of the working portion of the internal gear teeth can be determined; the pressure angle of the external gear teeth at the midpoint of the working portion of the external gear teeth can be determined; the difference of the determined pressure angles can be determined; the difference in the working pressure angles of the internal gear and the external gear can be determined; the pressure angle difference can be determined at each midpoint of the working portion of the respective internal gear teeth and external gear teeth.

[0077] For the present disclosure, the measured pressure angle difference between the internal gear teeth and the external gear teeth is advantageously less than 20% of the pressure angle of either the internal gear teeth or the external gear teeth. This represents a degree of profile shift that has minimal impact on the efficiency of the gear pair. It may be even more advantageous to limit the pressure angle difference to a smaller number, for example 15%, 10%, or 5% or less. These smaller numbers further limit the impact of profile shift and / or pressure angle difference on the efficiency of the gear pair.

[0078] Figure 7 shows the line of action l α The gearbox 100 is shown with the mark l α represents the gear teeth that are in contact and transmitting the load at any given moment. Line of action l αis centered on the instantaneous center of rotation ICR of the gearbox, which is the point on the pitch circle about which the external gear 104 can be said to rotate, given the assumption that the internal gear 108 is fixed. ICR is the point at the intersection of the lines represented by the two gear axes and one or the other of the pitch circles. It is also the position where the pitch circles are tangent. TIFF2025512007000007.tif11170 is the line of action l α and the external gear shaft. The operating angle is the maximum operating angle. As shown in TIFF2025512007000008.tif6170, it is preferably less than 20°. It is preferred that the load on the gearbox is within this range to avoid loss of efficiency due to friction.

[0079] 8a to 11 show gearboxes in which various configurations of gear teeth can be implemented. In particular, one or more features of the internal and external gear pairs described in this disclosure can be advantageously implemented in the gearboxes shown.

[0080] 8a and 8b respectively show a schematic and a cross-sectional view of a first gearbox 200 in which the internal and external gear pairs described herein can be advantageously implemented. The illustrated example comprises two internal and external gear pairs, namely a first gear pair 204 and a second gear pair 210. These gear pairs can be of the type described above with reference to FIG. 3, in particular the gear teeth of the gear pairs can have no or a small profile shift and a profile shift coefficient less than 0.8, such that the working pressure angle of the gear teeth does not exceed 20% of the nominal pressure angle of the gear teeth when no profile shift is applied.

[0081] The gearbox 200 has a first internal / external gear pair 204 with an external pinion gear 206 and an internal annular gear 208. The gearbox 200 also has a second internal / external gear pair 210, which also has an external pinion gear 214 and an internal annular gear 216. The external pinion gear 214 of the second internal / external gear pair 210 is fixedly fixed to the external pinion gear 206 of the first internal / external gear pair 204. As a result, the two external pinion gears 214, 206 rotate around the internal gear axis A7 at the same speed and have the same angular velocity with respect to their own axis A8. The two pinions 206, 214 are combined to form a combined pinion 212.

[0082] The gearbox 200 has an input shaft 202 arranged to receive torque from a drive source such as a motor. The input shaft 202 is crank shaped to receive torque about the gearbox axis A7 (i.e. the axis of the internal gears 208, 216) and to provide a force to the external pinion gears offset from the gearbox axis, which is a linear force through the external gear axis A8. The input shaft 202 is supported for rotation about the gearbox axis by a first bearing arrangement 222 and is coupled to the external pinion gears via bearings 224 so that the external pinion gears can rotate about the portion of the input shaft that passes through the external gears. The external gears also mesh with the internal annular gears so that the external pinion gears 206, 214 (i.e. the coupling pinion 212) and the offset portion of the input shaft 202 rotate together about the internal gear axis A7. The input shaft also has a counterweight 203 to balance the off-axis load of the offset portion.

[0083] The first internal / external gear pair 204 and the second internal / external gear pair 210 may have a slight difference in their respective gear ratios. Such a slight difference may be achieved in a number of ways, for example, by having the same tooth count difference between the internal and external gears, but a different total number of teeth on the internal and / or external gears of each gear pair. In other examples, the amount of tooth count difference between each gear pair may be varied to achieve different gear ratios. In one example, the internal gear 208 and the external gear 206 of the first internal / external gear pair 204 may have 60 and 57 teeth, respectively, and the internal gear 216 and the external gear 214 of the second internal / external gear pair 210 may have 90 and 87 teeth, respectively.

[0084] The overall output of the gearbox 200 can be realized as a relative differential rotation between the two internal gears 208 and 216 about a common axis. One of the internal gears 208, 216 can be fixed, i.e., remain stationary, while the other of the internal gears 208, 216 can rotate and be the output gear. The two gears can be coupled via a bearing 226. The fixation of the internal gear 208 is shown in FIG. 5 by the internal gear 208 of the first internal / external gear pair 204 being grounded, indicating that it is a reference gear. Due to one of the internal gears being fixed and the two internal / external gear pairs having similar gear ratios, a high gear ratio can be achieved throughout the gearbox. Although the internal gear 216 of the second internal / external gear pair 210 is shown coupled to a bearing adjacent to the input shaft, it will be appreciated that the output can be taken at the opposite axial end of the gearbox and any suitable bearing configuration can be used.

[0085] In general, it should be understood that the portions of a gearbox described as "fixed," "input," and "output" are so described only to reflect conventional uses of gearboxes, and discussion of input, output, and fixed portions can be a matter of reference frame, e.g., any portion labeled "fixed" can alternatively be used as "input" or "output." In the context of the illustrated example, a drive input provided as the difference in rotation between the input shaft 202 and the internal gear 208 results in an output that is the difference in relative rotation between the internal gear 216 and the internal gear 208.

[0086] FIG. 9 shows a further gearbox 300, which has a symmetrical arrangement. The symmetrical gearbox 300 has a first internal / external gear pair 304 and a second internal / external gear pair 310, each having a respective internal gear 308, 316 and external gear 306, 314, which may be substantially similar to those described above with reference to FIG. 5. The external gears may rotate around an internal gear axis A9 on which the internal gears are arranged and may rotate around their own external gear axis A10. Furthermore, the gearbox 300 comprises a third internal / external gear pair 320, which includes an external gear 324 and an internal gear 326. The internal / external gear pair 320 may be substantially similar to the first internal / external gear pair 304 and may have the same gear ratio. The external gear 324 of the third internal / external gear pair 320 may be fixedly fixed to the other external gears 306, 314, such that the three gears orbit and rotate together to form a coupling pinion 312.

[0087] A symmetrical arrangement may be preferred as it can prevent or reduce twisting of the central external gear due to the axially opposed internal gears providing balanced forces on the external gears about an axis perpendicular to the gearbox axis.

[0088] Additionally, the input shaft 302 may have bearings on both axial ends to avoid a cantilever arrangement due to reaction forces from the planetary gears 306, 314, 324. This may strengthen the input shaft 302.

[0089] 10a and 10b show a third gearbox arrangement 400 which is substantially similar to an axially mirrored, antipodal copy of the gearbox 200 of FIG.

[0090] The gearbox 400 has two separate planetary gears 412, 422, each planetary gear having two external gears fixed together. The first planetary gear 412 has external gears 406, 414 arranged around a first external gear axis A12, which are arranged to mesh with respective internal gears 408, 416, which rotate around the internal gear axis A11. The gears 406, 408, 414, 416 form the first and second internal / external gear pairs 404, 410. The second planetary gear 422 has third and fourth external gears 424, 434 arranged around a second external gear axis A13. The third and fourth external gears 424, 434 are arranged to mesh with respective internal gears 426, 436, which rotate around the internal gear axis A12. Gears 424 , 426 , 434 , 436 form third and fourth internal and external gear pairs 420 , 430 .

[0091] It should be understood that the first and fourth external gears 406, 434 can be identical and the second and third external gears 414, 424 can be identical. In some cases, the first, second, third and fourth external gears 406, 414, 424, 434 can all be identical. Also, the first and fourth external gears 408, 436 can be identical and the second and third internal gears 416, 426 can be identical. In some cases, the second and third internal gears 416, 426 can be integrally formed with common gear teeth and mounted in bearings 452 for rotation relative to the first and fourth internal gears 408, 436. In general, each of the first and fourth internal and external gear pairs 404, 430 may have a slight difference in gear ratio from the second and third internal and external gear pairs 410, 420 to collectively provide the differential planetary gearbox 400.

[0092] The gearbox 400 has a crank-shaped input shaft 402 having diametrically opposed portions each offset the same distance from the internal gear axis A11, with each offset portion lying along a respective external gear axis A12, A13. Each offset portion is coupled via a bearing to a respective first and second planetary gear 412, 422 such that the input shaft 402 provides a linear force to the planetary gears that is offset from the internal gear axis A11.

[0093] In this manner, a symmetric gearbox 400 with center output and improved balance can be provided.

[0094] Figure 11 shows a gearbox 201 having substantially similar features to gearbox 200 of Figures 8a and 8b. For the sake of brevity, like features are similarly numbered and will not be described again here.

[0095] In the gearbox 201, the external pinion gear 214 of the second internal / external gear pair 210 has a stepped portion 214a, which allows the external pinion gear 214 to lie in two separate axial planes perpendicular to the rotation axis(es) of the mated gear pair 212. In particular, the meshing of the second internal / external gear pair 210 and the meshing of the first internal / external gear pair 204 overlap or are coplanar with a common plane P1, which is perpendicular to the rotation axis of the gears A7, A8. The internal gear 208 is at least partially located in a recess or cavity of the mated pinion gear 212. This allows the gear pair 204 to be at least partially located in the external gear 214. This can further allow the gear pair 204 to be partially or completely located in the gear pair 210.

[0096] By arranging the first and second internal / external gear pairs 204, 210 in a common plane, it is possible to reduce the twist of the associated pinion gear 212. Furthermore, it is possible to reduce the axial extent of the gearbox 201, thereby providing a more compact gearbox 201.

[0097] The gearbox 201 shown in Figure 11 may have any type of gear teeth. In a particular example, the gearbox 201 of Figure 11 may have one or more characteristics of the gear teeth as described with reference to Figures 1 to 7, and in particular Figure 3. The gear teeth may have any of the characteristics of the gear teeth described with reference to Figure 3, such as a profile shift coefficient within a specified range and / or a profile shift within a specified range.

[0098] In any or all of the gearbox arrangements described above, the gears may have the characteristics or qualities described above with reference to any of Figures 1 to 7, and in particular Figure 3. In particular, the gears may exhibit small addendums and low addendum coefficients, as described above.

Claims

1. A gearbox comprising a pair of internal and external gears with a small difference in the number of teeth, An internal gear arranged around an internal gear shaft and having a first number of internal gear teeth, An external gear having an external gear shaft arranged to rotate around the internal gear shaft, having a second number of teeth on the external gear, and arranged to mesh with the teeth of the internal gear, Equipped with, The difference in the number of small teeth between the teeth of the first number of external gears and the teeth of the second number of internal gears is less than one-fifth of the number of teeth of the internal gears. The teeth of the internal gear and the teeth of the external gear have an involute shape with no displacement or with displacements such that the operating pressure angles of the teeth of the external gear and the teeth of the internal gear differ by less than 20% of the operating pressure angle of the teeth of the external gear. A gearbox in which at least one of the teeth of the internal gear and the teeth of the external gear has a displacement coefficient of less than 0.

8.

2. The gearbox according to claim 1, wherein the theoretical contact ratio of the internal and external gear pairs, calculated for gearboxes having the same geometric shape and no material deformation, is less than 1.

3. The gearbox according to claim 1 or 2, wherein at least one of the teeth of the internal gear and the teeth of the external gear has a displacement coefficient of less than 0.7, and optionally less than 0.

6.

4. The gearbox according to claim 1, wherein at least one of the teeth of the internal gear and the teeth of the external gear has a displacement coefficient greater than 0.3, and optionally greater than 0.

5.

5. The gearbox according to claim 1, wherein the displacement coefficient of the internal gear is less than 0.1, optionally less than 0.05, optionally less than 0.04, optionally less than 0.03, optionally less than 0.02, optionally less than 0.01, and optionally 0.

00.

6. The gearbox according to claim 1, wherein the theoretical contact ratio of the internal and external gear pairs, calculated for gearboxes having the same geometric shape and no material deformation, is less than 0.8, optionally less than 0.6, and optionally less than 0.

4.

7. The gearbox according to claim 1, wherein the pitch diameter of the external gear is at least half the pitch diameter of the internal gear.

8. The gearbox according to claim 1, wherein the load contact ratio of the gearbox during use is greater than 100% than the theoretical contact ratio calculated for a gearbox with the same geometry and no material deformation, and optionally, the load contact ratio of the gearbox during use is greater than 200% than the theoretical contact ratio, optionally greater than 400%, optionally greater than 500%, and optionally greater than 1000%.

9. The gearbox according to claim 1, wherein the sum of the displacement coefficients of the teeth of the internal gear and the teeth of the external gear is less than 1.

6.

10. The gearbox according to claim 1, wherein the internal and external gear pairs have an instantaneous rotation center ICR defined as the point where the radius drawn through the internal and external gear axes intersects the pitch circles of the external and internal gears, and all gear pairs that transmit load between the external gear and the internal gear are contained within a 20-degree arc centered on the instantaneous rotation center ICR.

11. The gearbox according to claim 1, wherein each of the internal gear and the external gear has 40 or more teeth.

12. The gearbox according to claim 1, wherein the internal gear and the external gear have a helical angle of 0°.

13. The gearbox according to claim 1, wherein the external gear is a first external gear, the internal gear is a first internal gear, and the gearbox further comprises a second external gear coupled to the first external gear and a second internal gear movable relative to the first internal gear, wherein the second external gear is located inside the second internal gear and is arranged to mesh with the second internal gear.

14. The second external gear and the second internal gear form a second small tooth difference internal / external gear pair having a tooth difference of less than one-fifth of the tooth number of the second internal gear. The teeth of the second internal gear and the teeth of the second external gear have an involute shape with no displacement or with displacements such that the operating pressure angles of the teeth of the second external gear and the teeth of the second internal gear differ by less than 20% of the operating pressure angle of the teeth of the second external gear. The gearbox according to claim 13, wherein at least one of the teeth of the second internal gear and the teeth of the second external gear has a displacement coefficient of less than 0.

8.

15. The gearbox according to claim 13, wherein the first external gear is fixed to the second external gear, and the second internal gear is arranged to rotate relative to the first internal gear around the internal gear shaft.

16. The gearbox according to claim 13, which drives the first external gear to supply an input that causes the first internal gear shaft to rotate, thereby producing an output as relative rotation between the first internal gear and the second internal gear around the internal gear shaft.

17. The gearbox according to claim 13, wherein the meshing between the first internal gear and the first external gear and the meshing between the second internal gear and the second external gear overlap in a common plane perpendicular to their axes of rotation.

18. The system further comprises a third internal gear and a third external gear, wherein the third external gear is located inside the third internal gear and is arranged to mesh with the third internal gear. The third external gear is coupled to the first external gear and the second external gear such that the second external gear is located between the first external gear and the third external gear. The gearbox according to claim 13, wherein the third internal gear is fixed to the first internal gear.

19. A symmetrical gearbox comprising a first gearbox according to claim 13 and a second gearbox according to claim 13, The first internal gear is fixed together, and the second internal gear is fixed together, The first gearbox and the second gearbox are symmetrical and coaxially arranged gearboxes.

20. The symmetrical gearbox according to claim 19, wherein the external gears of the first gearbox and the second gearbox are arranged to rotate around the internal gear shaft with a phase difference of 180° from each other.

21. An internal gear arranged around an internal gear shaft and having a first number of internal gear teeth, An external gear having an external gear shaft arranged to rotate around the internal gear shaft, having a second number of teeth on the external gear, and arranged to mesh with the teeth of the internal gear, A pair of internal and external gears, The teeth of the internal gear and the teeth of the external gear each have a pressure angle measured at the midpoint of the operating portion of the gear teeth, and the pressure angle differs from the pressure angle of the teeth of the external gear by less than 20%. An internal and external gear pair, wherein at least one of the teeth of the internal gear and the teeth of the external gear has a displacement coefficient of less than 0.

8.

22. The internal gear pair and the external gear pair according to claim 21, wherein the teeth of the internal gear and the teeth of the external gear have the same pressure angle measured at the midpoint of the respective operating parts of the teeth of the gears.

23. An internal gear arranged around an internal gear shaft and having a first number of internal gear teeth, An external gear is positioned to rotate around the internal gear shaft, has a second number of teeth, and is arranged to mesh with the internal gear. A pair of internal and external gears, The teeth of the internal gear and the teeth of the external gear have a standard module and are free from displacement. An internal and external gear pair, wherein at least one of the teeth of the internal gear and the teeth of the external gear has a displacement coefficient of less than 0.8.

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