ROLLER BEARING WITH ENHANCED ROLLER-END AND FLANGE CONTACT

Abstract

A roller bearing (10) defines a bearing axis (34) and a radial plane (52) oriented parallel with the bearing axis. The roller bearing (10) includes an inner ring (42) having an inner raceway (44) and an inner flange (46) extending from the inner raceway. The inner flange (46) includes an inner guide surface (48). The roller bearing (10) also includes a plurality of rollers (22) in rolling engagement with the inner raceway (44) about the bearing axis (34). Each roller (22) includes a first end surface (24a) in engagement with the inner guide surface (48) of the inner flange (46) as the plurality of rollers (22) move relative to the inner ring (42). The first end surfaces (24a) of each roller (22) define a curvature such that a ratio of a first principal effective curvature (Rx) radius in a plane perpendicular to the radial plane (52) and a second principal curvature radius (Ry) in the radial plane (52) is no less than 3.0.

Description

BACKGROUND

Engagement between a roller end and a bearing race flange is commonly seen in tapered roller bearings and in cylindrical roller bearings. The contact serves to guide the rollers and bear axial load. The conditions of the contact are determined by a flange and roller end geometry, which in turn, dictates the axial load carrying capacity of the bearing as well as a flange torque and a wear rate at the surfaces of the roller end and the flange.

For tapered roller bearings, the contact portion of the roller ends is often machined spherical in shape. Spherical roller ends are also used in cylindrical roller bearings (as disclosed within U.S. Pat. No. 6,530,693). This results in a contact ellipse between the inner raceway flange and roller end whose major axis is arranged in the radial direction of the bearing. As the contact load increases, the contact ellipse may be truncated by an undercut or an outer diameter of the bearing flanges. To reduce the risk of truncation of the contact footprint between the roller end and the flanges, and thus reduction of edge loading, cylindrical roller ends are often machined to a torus (as disclosed within U.S. Patent Application Publication No. 2003/0091255) to produce an ellipse whose major axis lies along a circumferential direction.

SUMMARY

The present invention relates to roller bearings, in general, and to a roller bearing with an enhanced roller end and flange contact. The enhancement provides an optimal contact footprint between the roller end and bearing flange that results in reduced bearing rib torque for given surface finishes and anticipated operation conditions.

In one aspect, a roller bearing defines a bearing axis and a radial plane oriented parallel with the bearing axis. The roller bearing includes an inner ring having an inner raceway and an inner flange extending from the inner raceway. The inner flange includes an inner guide surface. The roller bearing also includes a plurality of rollers in rolling engagement with the inner raceway about the bearing axis. Each roller includes a first end surface in engagement with the inner guide surface of the inner flange as the plurality of rollers move relative to the inner ring. The first end surfaces of each roller define a curvature such that a ratio of a first principal effective curvature radius in a plane perpendicular to the radial plane and a second principal curvature radius in the radial plane is no less than 3.0.

Other aspects of the invention will become apparent by consideration of the detailed description and accompanying drawings.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a perspective view of a roller bearing according to an embodiment of the invention including an outer ring, an inner ring, and rolling elements.

FIG. 2 is a cross sectional view taken along a radial plane of the roller bearing of FIG. 1.

FIG. 3A is an enlarged, partial detailed view of FIG. 2 illustrating an engagement between one of the rolling elements and the inner ring.

FIG. 3B is an enlarged, partial detailed view of FIG. 2 illustrating an engagement between one of the rolling elements and the outer ring.

FIG. 4 illustrates the engagement between one of the rolling elements and the inner ring in a plane that is perpendicular to the radial plane.

FIG. 5 illustrates a relationship between friction coefficient and the lambda ratio of the roller bearing at the roller-end and flange contact.

FIG. 6 illustrates an example of an optimal contact curvature ratio as a function of a geometry of the roller bearing and the lambda ratio.

DETAILED DESCRIPTION

Before any embodiments of the invention are explained in detail, it is to be understood that the invention is not limited in its application to the details of construction and the arrangement of components set forth in the following description or illustrated in the following drawings. The invention is capable of other embodiments and of being practiced or of being carried out in various ways.

FIGS. 1-4 illustrate an embodiment of a cylindrical roller bearing 10. The illustrated roller bearing 10 includes an outer ring 12, an inner ring 42, rolling elements 22, and a roller retainer 32 concentric with a rotational bearing axis 34. The outer ring 12 includes an outer raceway 14 and at least one outer raceway flange 16 positioned at an axial end of the outer raceway 14 with the outer raceway flange 16 extending radially inwardly toward the inner ring 42. In other embodiments, the outer ring 12 may include two opposing outer raceway flanges 16. The illustrated inner ring 42 includes an inner raceway 44 and at least one inner raceway flange 46 positioned at an axial end of the inner raceway 44 with the inner raceway flange 46 extending radially outwardly toward the outer ring 12. In other embodiments, the inner ring 42 may include two opposing inner raceway flanges 46. The illustrated rolling elements 22 are annularly separated evenly by the roller retainer 32 and are arranged between the inner and outer raceways 44, 14. Each roller element 22 is in rolling contact with the inner and outer raceways 44, 14 about a rolling element axis 23. In other embodiments, the cylindrical roller bearing 10 may be another type of bearing (e.g., a tapered roller bearing or a spherical roller bearing).

With reference to FIGS. 2 and 3A, when axial loads F are applied to the roller bearing 10 (e.g., the axial loads F act on the inner and/or outer rings 42, 12 to compress the rolling elements 22), a contact surface 48 of the inner raceway flange 46 and a first end face 24a of the roller elements 22 abut at a contact area or point C_P (see FIG. 3A). A similar abutment occurs with a second end face 24b of the roller elements 22 and the outer raceway flange 16 (see FIG. 3B). In one embodiment, a contact surface 48′ of the inner raceway flange 46 includes a linear cross sectional profile within a radial plane 52 (see FIG. 1) of the roller bearing 10 so that the contact surface 48′ is oriented at a contact angle β₀ (also referred to as the nominal contact angle) relative to a line 49 perpendicular to the rotational bearing axis 34. The radial plane 52 is parallel with the rotational bearing axis 34. In another embodiment, the contact surface 48 may be crowned to produce a convex or a concave curvature radius R_rib2 with a center of the curvature located at P₂ that is within the radial plane 52 of the roller bearing 10. For a crowned flange surface 48, an actual contact angle β, defined as the angle between a tangent line T through the contact point C_p and the perpendicular line 49 to the bearing axis 34 may not be equal to the nominal contact angle β₀. The illustrated roller end faces 24a, 24b are profiled in the radial plane 52 at least in the areas where contact between the raceway flanges 16, 46 and the roller elements 22 occur. A curvature radius R_rol2 of the first end face 24a is defined near the contact point C_P and includes a center of the curvature at O₂. Thus, an effective curvature radius R_y of the contact between the rolling elements 22 and the inner raceway flange 46 within the radial plane 52 is:

$\begin{array}{cc}{R}_y=\frac{R_{\mathrm{rib}2}·R_{\mathrm{rol}2}}{R_{\mathrm{rib}2}+R_{\mathrm{rol}2}}& \left(1\right)\end{array}$

R_y is often referred to as the second effective principal curvature radius at the contact between the rolling elements 22 and the inner raceway flange 46.

With drawing a line P₁O₁ perpendicular to the tangential line T through the contact point C_P, the line intersects the rolling element axis 23 of the rolling elements 22 at O₁ and intersects the rotational bearing axis 34 of the roller bearing 10 at P₁. P₁C_P represents a principle radius R_rib1 of the flange surface 48 in a plane perpendicular to the radial plane 52, likewise, O₁C_p represents a principle radius R_rol1 of a roller end face 26 in a plane perpendicular to the radial plane 52. The principle radii R_rib1, R_rol1 are defined by an inner raceway diameter D_inr of the inner raceway 44, a rolling element diameter D_rol of the rolling elements 22, a contact height h_cnt from the inner raceway 44 to the contact point C_p, and the contact angle β.

$R_{\mathrm{rol}1}=\frac{D_{\mathrm{rol}}/2-h_{\mathrm{cnt}}}{\sin \beta }$ $R_{\mathrm{rib}1}=\frac{D_{\mathrm{inr}}/2+h_{\mathrm{cnt}}}{\sin \beta }$

Thus, for the inner raceway flange 46, an effective curvature radius R_x of the contact between the rolling elements 22 and the inner raceway flange 46 in a plane perpendicular to the radial plane 52 is:

$\begin{array}{cc}{R}_x=\frac{\left(D_{\mathrm{rol}}-2h_{\mathrm{cnt}}\right)\left(D_{\mathrm{inr}}+2h_{\mathrm{cnt}}\right)}{2\left(D_{\mathrm{rol}}+D_{\mathrm{inr}}\right)\sin \beta }& \left(2a\right)\end{array}$

R_x is also referred to as the first effective principal curvature radius at the contact between the rolling elements 22 and the inner raceway flange 46.

For the outer raceway flange 16, a first effective principal curvature radius R_x at the contact between the roller end 24b and the flange surface of the outer raceway flange 16 in a plane perpendicular to the radial plane 52 is:

$\begin{array}{cc}{R}_x=\frac{\left(D_{\mathrm{out}}-2h_{\mathrm{cnt}}\right)\left(D_{\mathrm{rol}}-2h_{\mathrm{cnt}}\right)}{2\left(D_{\mathrm{out}}-D_{\mathrm{rol}}-4h_{\mathrm{cnt}}\right)\sin \beta }& \left(2b\right)\end{array}$

As such, the first effective principal curvature radius R), of the outer raceway flange 16 is defined by a diameter D_out of the outer raceway 14, the rolling element diameter D_rol, the contact height h_cnt from the outer raceway 14 to the corresponding contact point C_p, and the contact angle β between the roller end 24b and the outer raceway flange 16.

With reference to FIG. 4, under a bearing axial load F, the footprint at the contact area C_P assumes an ellipse 50 with a semi major dimension a and a semi-minor dimension b. As the load F increases, the footprint 50 increases in size, with the periphery of the ellipse 50 moving closer to a boundary of an undercut 45 between the flange surface 48 and the inner raceway 44. A similar undercut is located between the outer flange 16 and the outer raceway 14. The shape of the contact footprint 50, defined as the ellipticity k=a/b of the contact ellipse 50, is determined uniquely by the ratio of the first and second effective principal curvature radii k_r=R_x/R_y and vice versa. The relationship between k and k_r can be expressed by elliptical integrals of first and second kinds, K(e) and E(e).

$\begin{array}{cc}{k}_r=\frac{k^2E\left(e\right)-K\left(e\right)}{K\left(e\right)-E\left(e\right)}& \left(3\right)\\ e=\sqrt{1-\frac{1}{k}}& \end{array}$

Sliding contact between the roller ends 24a, 24b and the corresponding raceway flange 16, 46 causes power loss and material wear. It is known to those skilled in the art, bearing torque mainly comes from raceway torque which is governed by rolling resistance between the rollers 22 and the inner and outer raceways 14, 44, and flange torque which is caused by the sliding frictions between the roller ends 24a, 24b and the inner and/or outer flanges 16, 46. Under low speed and thrust load operation conditions, the flange torque usually dominates. Therefore, to reduce bearing torque, and thus wear and power loss, it is desirable to minimize a friction moment M (see FIG. 4) between the roller ends 24a, 24b and the raceway flanges 16, 46.

With continued reference to FIG. 4, the friction moment M is defined at the contact ellipse 50 between the roller ends 24a, 24b and the corresponding raceway flanges 16, 46, which is an important component of bearing flange torque and can be expressed as a product of a geometry function ƒ_GM, a load function ƒ_LD, and a curvature ratio function ƒ_CR.

M=ƒ _GM·ƒ_LD·ƒ_CR  (4)

where ƒ_GM is a function of the contact angle β and bearing geometry parameters, which define the first principal effective curvature radius R_x; ƒ_LD is a function of contact load and material properties; and ƒ_CR is a function of a non-dimensional contact height H_cnt and the ratio of the effective principal curvature radii k_r, which in turn, is a function of the contact ellipticity k.

$\begin{array}{cc}{f}_{\mathrm{CR}}\left(k,H_{\mathrm{cnt}}\right)=k_r^{-\left(\frac{1}{6}\right)}·F\left(k\right)·I_{\mu }\left(k,H_{\mathrm{cnt}}\right)& \left(5\right)\end{array}$

In the above equation (5), F(k) is the shape factor of the contact ellipse 50 and is defined by the first and second elliptical integrals K(e) and E(e)

$\begin{array}{cc}F\left(k\right)={\left(\frac{4}{\pi }\right)}^{\frac{1}{3}}\frac{{\left\{\left[k^2E\left(e\right)-K\left(e\right)\right]·\left[K\left(e\right)-E\left(e\right)\right]\right\}}^{\frac{1}{6}}}{k^{\frac{1}{2}}·e^{\frac{2}{3}}}& \left(6\right)\end{array}$

I_μ(k, H_cnt) is an integral of friction moment and is defined as:

$\begin{array}{cc}{I}_{\mu }\left(k,H_{\mathrm{cnt}}\right)={\int }_{-1}^1{\int }_{1-\sqrt{1-X^2}}^{1+\sqrt{1-X^2}}\mu \left(k\right)\sqrt{\left({\mathrm{kX}}^2+\frac{Y^2}{k}\right)·\left[1-X^2-{\left(Y-H_{\mathrm{cnt}}\right)}^2\right]}\mathrm{dYdX}& \left(7\right)\end{array}$

where X and Y are non-dimensional coordinates, X=x/a, Y=y/b, and H_cnt is the non-dimensional contact height measured in the tangential plane of the contacting surfaces, H_cntt=h_cnt/(b*cos(β)). μ(k) is the friction coefficient which is a function of the lambda ratio λ defined as the ratio of lubrication (e.g., oil) film thickness to the composite surface roughness of the contacting surfaces 24a, 48. Since the lubrication film thickness between the contact surfaces 24a, 48 varies with the ellipticity k, μ in turn can be expressed as a function of k for a given lubrication condition characterized by a nominal lambda ratio λ₀.

FIG. 5 illustrates an example friction coefficient μ as a function of lambda ratio λ for typical engineering surfaces. The following regression relationship between friction coefficient μ and lambda ratio λ can be obtained:

μ=0.0207exp[2.197exp(−0.894λ^0.539)]  (8)

where the lambda ratio λ is related to the ellipticity k:

λ=λ₀[1−exp(−0.68/k)]  (9)

As can be appreciated by those skilled in the art, the contact profile in the radial plane 52, which in turn is determined by the crown of the flanges 16, 46 and the curvature of the roller ends 24a, 24b in the radial plane 52, hardly affect the geometry function ƒ_GM and the load function ƒ_LD. Thus, contact profile optimization to minimize bearing's friction moment M, and thus flange torque is equivalent to minimization of the curvature ratio function ƒ_CR as suggested by equation (4). That is to say,

ƒ_{CR_min}(k_m,H_cnt)=min_k=kmƒ_CR(k,H_cnt)  (10a)

where k_m refers to the contact footprint ellipticity k that minimizes ƒ_CR for a given H_cnt.

Since the curvature ratio of the contact k_r is uniquely determined by the ellipticity k of the contact footprint 50, and H_ucut=H_cnt−1.0, thus ƒ_CR can also be expressed as a function of the curvature ratio k_r and H_ucut. Correspondingly, the minimization of ƒ_CR can be obtained in terms of k_rm and H_ucut:

ƒ_{CR_min}(k_rm,H_ucut)=min_kr=krmƒ_CR(k_r,H_ucut)=min_k=kmƒ_CR(k,H_cnt)  (10b)

FIG. 6 illustrates an example of the optimal contact curvature ratio k_rm as a function of H_ucut and λ₀, where

$H_{\mathrm{ucut}}=\frac{h_{\mathrm{ucut}}}{b\cos \beta }$

h_ucut represents the maximum allowable height of the undercut 45 measured in a plane perpendicular to the raceways 14, 44.

The optimal contact curvature ratio k_rm increases as H_ucut increases, and decreases as lubrication condition, characterized by λ₀, improves. Under ideal lubrication conditions with essentially no undercut 45, a minimum contact curvature ratio k_rm of 3.401 was suggested. As a practical matter, for all things considered, a minimum contact curvature ratio k_rm between about 3.0 and about 3.5 is recommended.

Various features of the invention are set forth in the following claims.

Claims

1. A roller bearing defining a bearing axis and a radial plane oriented parallel with the bearing axis, the roller bearing comprising: an inner ring including an inner raceway and an inner flange extending from the inner raceway, the inner flange including an inner guide surface; anda plurality of rollers in rolling engagement with the inner raceway about the bearing axis, each roller including a first end surface in engagement with the inner guide surface of the inner flange as the plurality of rollers move relative to the inner ring;wherein the first end surfaces of each roller define a curvature such that a ratio of a first principal effective curvature radius in a plane perpendicular to the radial plane and a second principal effective curvature radius in the radial plane is no less than 3.0;wherein the first principal effective curvature radius at a contact area between the inner flange and the first end surfaces of the plurality of rollers is defined as:

2. (canceled)

3. (canceled)

4. The roller bearing of claim 1, further comprising an outer ring including an outer raceway and an outer flange having an outer guide surface, wherein second end surfaces of the plurality of rollers are in engagement with the outer guide surface of the outer flange as the plurality of rollers move relative to the outer ring.

5. The roller bearing of claim 4, wherein the first principal effective curvature radius at a contact area between the outer flange and the second end surfaces of plurality of rollers is defined as:

6. A method for design of the roller bearing of claim 1, including determining the ratio of the first and second principal effective curvature radii at the contact area between the first end surfaces and the corresponding inner guide surface by minimizing a friction moment between the first end surfaces and the corresponding inner flange.

7. A method for design of the roller bearing of claim 5, including determining the ratio of the first and second principal effective curvature radii at the contact area between the first and second end surfaces and the corresponding inner and outer guide surfaces by minimizing a friction moment between the first and second end surfaces and the corresponding inner and outer flanges.