The present invention refers to a roller bearing of the type defined in the preamble of claim 1.

Such bearings are known e.g. from US-A-2.595.121, which shows bearings with concave rollers and SE-A-53 256, which swows bearings with convex rollers. As the radii of the profile curve of race tracks and rollers in such bearings are substantially bigger than the corresponding radii e.g. in spherical roller bearings with corresponding radial dimensions, it is possible to make the rollers longer than what is possible in such spherical bearings, and this improves the radial load carrying capacity. Furthermore the friction is lower in a bearing according to the preamble as compared to a spherical bearing with the same axial width, i.e. with equally long rollers. A drawback is that the bearing, when equipped with common flanges and retainer, lacks the self-aligning ability of the spherical bearing, which ability is valuable in many situations. Furthermore it has hitherto been considered necessary to provide a roller mounting opening in one of the race rings, such as shown at reference numeral 8 in SE-A-53 256, for allowing insertion of a desired number of rollers in the bearing. Such an opening means a complication and reduces the usefulness, as the opening may not be situated in a loaded part of the bearing. If a mounting opening is missing in a known bearing of the above defined type, e.g. according to Fig. 8 of the above mentioned US patent, it is necessary to provide the bearing with rollers during the time the outer and inner race rings are positioned eccentrically, which allows filling of the bearing with a comparatively small number of rollers. Such a bearing therefore has a comparatively low load carrying capacity.

In EP-A-0 175 858 is described a roller bearing with cor¬ respondingly curved longitudinal section profiles at rollers and race tracks with a radius of curvature substantially bigger than the biggest distance between the centre axis of

the bearing and the surfaces of the race tracks, whereby the rollers are axially movable between the race tracks without being obstructed by flanges or the like, for relative inclination and axial displaceability for the race tracks. These specifications give good results regarding the load carrying capacity and adjustability, whereby as mentioned above the rollers at increased length, at the same time increase the load carrying capacity. However it has proven itself that very flat rollers, i.e. having a large ratio between the curvature radius of the roller profile and the roller length, will give the bearing a shorter service life, due to high edge pressures, i.e. high pressures at the end of the roller, which means that such bearings, which have very high load carrying capacity, will get unacceptably reduced life span at misalignment.

The purpose of the present invention is to provide a bearing of the type defined in the preamble of claim 1, and in which has been obtained the best possible combination of the criteria; improved radial load carrying capacity, mutual misalignment and axial displacement of the bearing race tracks, and a good service life for the bearing. This is obtained in that the bearing has been given the characteriz¬ ing features defined in the accompanying claims.

Hereinafter the invention will be described with reference to two embodiments shown in the accompanying drawings.

Fig. 1 shows an axial section through a roller bearing according to the invention,

Fig. 2 is a roller forming part of the bearing according to Fig. 1, shown with load and force lines, and Fig. 3 illustrates in an axial section one half of an angular contact roller bearing according to the invention.

Fig. 1 shows schematically a cross section through a bearing 1, incorporating an inner race ring 2 having an outer race track 3, and an outer race ring 4 with an inner race track 5. Between the race tracks 3 and 5 are provided a number of

rollers 6. The inner race ring 2 is fixedly connected to a shaft 7 in a manner not further shown, and which shaft in the position shown, i.a. due to the deflection of the shaft and the direction of the load acting upon the bearing, is misaligned under an angle γ against the centre axis 8 of the bearing. Hereby also the rings 2 and 4 will obtain a mutual misalignment γ, whereby the rollers 6 have a possibility to move axially for compensating this misalignment. The rollers 6, as well as the race tracks 3 and 5 resp. hereby have a longitudinal section profile, with a radius R of curvature bigger than the biggest distance between the centre axis of the bearing and the race track surfaces. The smooth curvature of the surfaces permits that the radial extension of the bearing, required for a given roller length is small, which is space-saving, and that the internal friction appearing in the bearing, due to sliding at contact between rollers and race track, is low as compared to the case in spherical roller bearings with a corresponding load carrying capacity and subjected to corresponding load.

In Fig. 2 is shown a roller 6 forming part of the bearing according to Fig. 1, and having a radius of curvature R for the longituduinal section profile. At misalignment in accordance with Fig. 1 a load Q acts upon the roller under the angle φ, due to the the fact that the roller slides axially at misalignment and the friction force thereby generated must be balanced by the normal forces of the roller.

When the load Q in such a bearing acts against the roller due to the inner and outer ring contact under the angle φ against a line ' through the centre of the roller perpendicularly to the roller axis, it is obtained a radial and an axial force component.

The axial force acting from a race track on the roller will be Q sin cp , which at the small angles, here concerned, can be simplified to Q φ. The roller thereby is subjected to a combined axial force from both race tracks, which can be

expressed 2 Q φ. For the roller in axial equilibrium accord¬ ing to the equilibrium equation is valid

2 Q φ = μ Q (equ. I) wherein μ = the coefficient of friction between the roller 6 and its cooperating race track. This friction coefficient is by experience typically between 0.05 and 0.1, depending on the lubrication conditions.

The force lines in the drawing are drawn just opposite each other, whereas they in practice are displaced axially a small distance relative to each other.

The axial displacement of the force on the roller can be expressed as the distance Δ between a line perpendicular to the centre axis of the roller at the middle of the roller and the point where the load Q attacks the roller. The geometri¬ cal relation for this condition is φ = Δ / R (equ. II)

After reduction of equation I is obtained φ = μ/2 (equ. Ill)

By inserting equation III into equation II is obtained

Δ = μ R/2 (equ. IV)

At the same time it is evident that Δ must be smaller than half the length of the roller, i.e. l _a /2.

Therefrom it follows that μ R/2 < l _a /2 (equ. V)

From this relation it is also obtained that

R/l. < l./μ (equ. VI)

At the above mentioned lowest value for μ of about 0.05, which is established by way of experience, is obtained a limit value for the relation between the radius of curvature R for the roller and the length of the roller l _a , i.e. R/l _a , which is over 20. Over this value the bearing will not

function well, as extreme edge pressures will appear at misalignment, leading to a short life span. The relation therefore must be

R/l _a < 20 (equ. VII)

In Fig. 3 is shown an axial cross section of one half of an angular contact roller bearing with an inner race ring 12 with a curved race track, which is inclined against the centre line 18 of the bearing and an outer race ring 14 likewise having a curved race track inclined against the centre line 18 of the bearing. Between the race tracks in the race rings 12, 14 are provided crowned rollers 6, which are not obstructed from moving axially by flanges or retainer, and which like the race tracks have a longitudinal section profile with a radius of curvature R, which is bigger than the distance between the outer race track and the axis of the bearing, measured perpendicularly to the race track.

The same forces as illustrated in Fig. 2 act upon the rollers in such a bearing, and the same relation reigns between the radius of curvature of the longitudinal section profile and the roller length as in this case.

The invention is not limited to the embodiment shown and described in connection thereto but modifications and variations can appear within the scope of the corresponding claims.