SU1280224A1 - Contactless radial thrust plain bearing with outside lubrication pressure - Google Patents

Abstract

The invention relates to the field of contactless, slide bearings, for example, hydro or gas static used in mechanical engineering. The purpose of the invention is to increase the load characteristics under the action of bending, her load. In the bearing bore, the bearing pockets 4 of the radially resistant bearing are distributed around the circumference, and the ends of the bushings have the bearing pockets 5 of the thrust bearing distributed around the circumference. (L kae 00 o yu 4

Description

Pockets A are closed with support plate 7 and section 1, pockets .5 with stop flanges 8 and 9 of spindle 1, between which and the bearing there is a gap. The end faces of rotation 10 and 11 of flanges 8 and 9 and the opposed surfaces of rotation 12 and forming the gap of the thrust bearing.

13 bearings are curved. The angle of inclination of the tangent to the generators of these surfaces to the axis of rotation of the shaft is made increasing in the direction of increasing the radius of the thrust bearing and is not zero at the junction of the forming radial and thrust bearings. hp f-ly, 2il.

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The invention relates to the field of contactless sliding bearings with an external source of lubricant pressure, for example, hydro or gas static bearings, and can be used in machine tool industry, for example, in spindle assemblies of grinding machines.

The aim of the invention is to increase the load characteristics, i.e. increase of rigidity and bearing capacity of the bearing under the action of bending load.

The invention is illustrated by the example of a front hydrostatic radial-thrust bearing of a spindle metal cutting machine.

FIG. 1 shows the proposed angular contact ball bearing, section; in fig. 2 - thrust bearing without shaft, end view.

The spindle 1 of the grinding wheel 2 is located in the front non-contact radial-thrust hydrostatic bearing and rear radial, which is not shown. The bearing pockets 4 of the angular contact bearing are distributed around the circumference in the bore of the hub of the 3 sub-supports, and the ends of the sleeve 3 are distributed around the circumference of the bearing pockets of the 5 thrust bearing. All bearing pockets 4 and 5 are connected to an external source of pressure P lubricant through hydraulic resistances 6. Such a bearing can be gas-static. Pockets 4. are closed by a support collar 7 of spindle 1, located in a bearing with a gap; pockets 5 are closed by stop flanges 8 and 9 of spindle 1, between which and the bearing there is a gap. This is the working clearance of the bearing. The end surfaces 10 and 11 of flanges 8 and 9 and oppositely located surfaces 12 and 13 of the bearing, which form the working gap of the thrust bearing, are made with curvilinear monotonically expanding:; from radial bearing. The figures also show the axis 14 of the bearing and the axis 15 of the spindle 1. The angle (xi of the inclination of the tangents to the generatrix of the thrust bearing to the axis

0 rotation is greater than 0 and the larger, the larger the radius on which the thrust bearing gaps are located, since with increasing radius the relative displacement of the working surfaces increases, the optimum range of angles of inclination of the tangents to the generatrix of rotation to the axis of rotation is between 45 and 90.

0

FIG. 1, for example, shows an angle 0 formed by a tangent drawn at a radius at the thrust bearing.

The bearing works as follows.

In the initial position, when the lubricant pressure Rc is turned on and the load P 0, the spindle 1 is rigidly centered by the lubricant pressure in the bearing cages Hak 4 and 5 of the bearing. The working gaps that form are the bearing surfaces of the bearing and the spindle are generally equidistant, i.e. are equidistant, in particular, equidistant pairs of surfaces 10 and 12, 11 and 13. In this case, axes 14 and 15 are combined.

Under the action of a cantilever positioned load P (in this case, the forces

0 grinding) on the spindle 1, the relative displacement of the working surfaces of the thrust bearing, formed by the end surfaces, occurs.

The operating clearances at the lubricant outlet from the bearing pockets 4 and 5 of the bearing become different and, as a result, the pressures in the bearing pockets become different. The pressure difference in the pockets 4 of the radial bearing balances the radial load, and in the pockets 5 of the multi-pocket thrust bearing - bending moment in the axial plane. Here, each point on the bearing surfaces 11 and 12 of the stop flanges 8 and 9 receives an offset 8 along the axis and about q across the axis. The displacement Sy is formed due to the rotation of the flanges in the axial plane, mainly due to the bending deformations of the spindle body, and the displacement 8i due to the aforementioned rotation and radial displacement of the spindle body.

The expected pattern of displacements and Ojj for each point of the working gap can be obtained using existing programs for calculating the spindle deflection in hydrostatic and gas-static supports. The values of S and 5c are commensurable; moreover, if we consider them as a function of distance from the axis, then Ohms generally remains constant, and Sy grows mostly linearly. The direction of the resultant displacement o is substantially non-parallel to the axis and deviates from it by an angle of 90-oi.

In order that the relative total displacement and at each point of the forming end surfaces 10 and 12 (11 and 13) in the plane of the load P be carried out normal to this surface (namely, in this case the bearing provides the best load characteristics under the action of the bending load ), the opposed surfaces of the rotation 10 and 11 (12 and 13) monotonously expanding in the direction from the radial bearing, are made so that the angle ot of the tangent of the generatrix of the rotation to the axis of rotation of the spindle 1 sun yes O r greater and increases as the radius of the thrust bearing.

Therefore, when bending spindle 1 under the action of load P, the gap along the entire length of the generator varies along the normal to the end surface and the bearing provides maximum stiffness and load-carrying capacity. When the angle o changes (, within 45 - 90, the maximum carrying capacity is ensured, because with a minimum radius of the butt-end, the displacement is S-ii and at the maximum — o, o, and oC — 90

To simplify the manufacture of the bearing, the generators of these surfaces can be made in the form of circular arcs, the centers of which are outside the axis of the rotating shaft and are mirrored (symmetrically) relative to the axis of rotation.

Thus, the technical advantages of the proposed decacting angular contact bearing design with an external source of lubricant pressure are that it allows to significantly increase the load characteristic of the bearing when it is loaded with a bending moment.

In addition, the application of the proposed invention will dramatically increase the technological capabilities of the spindle assemblies of grinding machines by increasing the possible length of the spindle bracket, for example, improving the accuracy and productivity of grinding deep holes, as well as machining using a rotating center.

Claims (2)

Invention Formula T. Non-contact angular contact bearing with an external source of lubricant pressure, containing independent radial and thrust bearings with circumferentially distributed bearing pockets and opposedly made supporting end surfaces, so as to increase load capacity under the action of a bending load, the oror end surfaces are curved, and the angles of inclination of the tangents to the generators of these surfaces to the axis of rotation are increasing in the direction of tim radius of the thrust bearing, and in the zone of the coupling-th constituting radial and thrust bearings of said angle greater than zero degrees. 2. A bearing according to claim 1, characterized in that said inclination angle is in the range of 45-90 °.