The principle of operation of hydrodynamic bearings. A hydrodynamic bearing is a fluid friction bearing. These bearings are radial and thrust. The radial bearing has three or four segments (shoes) 1 (Fig. 7.6). The support is filled with oil using a hydraulic system. Gravity non-rotating spindle 3 descends into segments. When the spindle is set in rotation, its rough surface draws oil into the gaps between it and the segments. Segment design, in particular the offset position of its support 2 relative to the axis of symmetry, allows it to rotate under the action of oil pressure, resulting in a wedge gap, narrowing in the direction of rotation of the spindle, in this gap there is a hydrodynamic pressure R, holding the spindle in a suspended position. If the spindle rotates on multi-V-bearings with self-aligning segments covering it evenly around the circumference, a slight displacement of it from the middle position under the action of an external load leads to a redistribution of pressure in the wedge gap and the emergence of a resulting hydrodynamic force that balances the external load.

Hydrodynamic bearings are recommended for spindles that rotate at a high constant or little changing frequency and perceive a small load, for example, for grinding machine spindles. The advantages of hydrodynamic bearings are in high accuracy and durability (mixed friction only at the moments of starts and stops), the disadvantages are in the complexity of the design of the oil supply system for bearings, in changing the position of the spindle axis with a change in its rotational speed.

Oil for hydrodynamic bearings. Typically, mineral oil grade L (velocit) is used, which has a dynamic viscosity coefficient y.= (4...5)10~ 3 Pa-s at a temperature of 50 C. Oil (1...3 l / min at a pressure of 0.1 ... 0.2 MPa) is supplied to the bearing using a hydraulic system, including a fine filter and a refrigeration unit.

Designs of radial hydrodynamic bearings. The bearing segments must be able to independently change their position both in the plane perpendicular to the spindle axis and in the plane passing through the axis. The latter eliminates possible high edge pressures in the support, accompanied by overheating of the oil in a thin boundary film and the loss of its lubricating properties. There are a number of bearing designs where the clearance between the shaft and the segments automatically changes depending on the load and the spindle speed.

One of the designs - LON-88, developed by ENIMS, is shown in fig. 7.7. The bearing is made in the form of a separate block, consisting of two rings 2, three segments 1 and spacer ring 3. The outer end surface of the segments is in two-point contact with the conical surfaces of the rings, as a result of which the segments can be installed along the axis of the spindle and in the direction of its rotation. The spacer ring with its protrusions prevents the segments from moving around the circumference. By changing the thickness of the spacer ring, the working clearance in the bearing can be adjusted.

Bearings of a different design - LON-34 - with segments 1 , established as a result of rotation on spherical supports BUT(Fig. 7.8), allow sliding speed up to 60 m/s in the absence of edge pressure* Segment supports are made in the form of screws 2 made of hardened steel with fine thread. By moving them in the radial direction, the radial clearance in the support and the position of the spindle axis are adjusted. To increase rigidity, the gaps in the threaded connections of the support pins with the body are selected with nuts 3, In order to reduce the wear of the segments at the moments of starting and braking the spindle, they are made bimetallic: a layer of bronze Br OF10-0.5, Br 0S10-10 or other antifriction material is applied to the steel base by centrifugal casting. Roughness parameter Ra the working surfaces of the segments should be no higher than 0.32 microns, the spindle necks - no higher than 0.04 ... 0.16 microns. The dimensions of the segments and support screws are given in Table. 7.1 and 7.2.

An example of the design of the spindle assembly. Hydrodynamic bearings are installed in the front and rear supports of the spindle assembly of the grinding machine (Fig. 7.9) 1 type LON-88. Axial loads are taken up by a double-sided thrust bearing formed by discs 2 and 4, Burt is in contact with them 3 spindle. Lubricant is supplied to this bearing through holes B and 5. Throat seals prevent oil from flowing out of the headstock. By channel G oil from the seal cavities drains into the headstock housing.

Structural parameters of bearings. Diameter D spindle necks are selected according to the stiffness conditions. Length I of the bearing for grinding machines - 0.751), for precision lathes and boring machines - (0.85-0.9) D. The length of the arc of coverage of the liner (0.6-0.8)1. Diametral clearance = 0.003 D. Typically, bearings with three or four bushings are used.

Calculation of hydrodynamic radial bearings. The calculation is carried out in order to determine the dimensions of the bearing depending on the given load capacity of the support and its rigidity. In addition, the friction losses in the support are determined.

Below is a method for calculating radial hydrodynamic bearings with three or four self-aligning segments for bearings with sliding speeds up to 30 m/s [67].

Initial data: design parameters of the bearing, spindle speed, maximum radial load, required radial rigidity of the support.

Load capacity (N) of one segment at the central position of the spindle

where is the dynamic viscosity of the oil, Pa-s; n- spindle speed, rpm; D- segment boring diameter, mm; AT- segment arc chord, mm; L- segment length, mm; ; estimated diametrical clearance, mm.

Under the action of the resulting force, the spindle is displaced from its initial position by e millimeters, and its new position is characterized by relative eccentricity If the resulting force is directed along the segment support axis, the load capacity of a three-segment bearing

There are two common ways to create supportive» pressure:

static ( hydrostatic) and hydrodynamic. Accordingly, there are hydrostatic and hydrodynamic fluid friction bearings. AT hydrostatic bearings the pressure in the supporting layer of the lubricant is created by a pump that supplies the material into the gap between the journal and the bearing. These bearings require a complex hydraulic system for normal operation. Hydrodynamic bearings have become more widespread. In them, the lubricant should be supplied only to the low pressure zone, from where it is pumped down by a rotating pin, forming a wedge support layer. Passing through a narrow section of the radial clearance, part of the lubricant is removed into the end gap between the journal and the bearing. Another part of it flows into the end gap over the trunnion, cooling the bearing. Specific bearing load p=F r /(ld).

73. Structures of plain bearings and materials of parts. Plain bearings consist of two main parts: a housing and a bearing sleeve (liner). The use of inserts makes it possible to manufacture housing parts from cheap materials and facilitates repairs. In small-sized and non-critical bearings, liners are sometimes absent; in this case, their purpose is performed by housings. The most common supports with a fixed axis b) and with a movable axis c) The mechanisms use supports on centers and supports on cores d, e) Cores are made in the form of cylindrical axes with a diameter of 0.25 ... 2 mm, their conical ends are rounded along a spherical surface radius rk= 0.01...0.2 mm. The supports of mechanisms and machines can be conditionally divided into standalone and embedded. Autonomous supports are manufactured according to standards in detachable and non-detachable executions. Bearings with a one-piece housing are relatively simple and cheap, but difficult to mount. This limits their scope. Split bearings are widely used in various designs. It consists of: body 1 , covers 2, bushing 3, fixing bolts with nuts 4 and oiler 5. Bearing shells are cylindrical without shoulder for radial load or with shoulder for absorbing axial and radial forces. They are made detachable and detachable It is recommended to split the liner in a plane perpendicular to the radial load, and the housing split should be stepped. The ledge in the stepped connector prevents the cover from moving laterally relative to the bearing housing. Lubrication is carried out with various lubricants using cap or drip oilers.

74. General information Classification of rolling bearings. Rolling bearings are the most common type of bearings for parts of mechanisms and machines. Unlike plain bearings, they implement rolling friction between parts: outer 1 and internal 2 rings, rolling elements 3 located between the rings. To protect the rolling elements from contact with each other, they are separated from each other by a separator 4.

Rolling elements move on carefully machined treadmills BUT made on rings. Advantages of rolling bearings over plain bearings: 1) small axial dimensions, 2) low resistance to start-up and rotation, 3) ease of maintenance, 4) low cost, 5) interchangeability. Flaws: 1) large radial dimensions and complex installation, 2) lower radial rigidity, 3) low durability at high speeds (due to overheating), etc. Classification of bearings. 1) According to the shape of the rolling elements, bearings are divided into ball and roller in the form of rollers a) c short and and long cylindrical rollers, b) c conical c) barrel-shaped G) needle-like di twisted rollers). 2) In the direction of perceived forces, bearings are divided into: a) radial, perceiving predominantly radial loads, b) angular contact, perceive the action of radial and axial loads; c) thrust-radial, perceive the axial load with a slight radial load; G) stubborn, perceiving only axial forces According to the ability to self-align, bearings are divided into not self-aligning and self-aligning allowing rotation of the axis of the inner ring relative to the axis of the outer ring. By the number of rows of rolling elements, bearings are distinguished single row , double row and four-row. Bearings of the same bore diameter are subdivided into series: according to the overall dimensions of the outer diameter ultralight, extra light, light, medium and heavy, and depending on the width, they are divided into: extra narrow, narrow, normal, wide, extra wide.

75. Static bearing capacity. The static load capacity of a bearing is the load So(radial and axial), which causes a total permanent deformation of the most loaded rolling element. C values about for bearings various types and series are given in reference books. If the bearing is loaded simultaneously with radial F r and axial Fa forces, and accept that the axial force is evenly distributed between the rolling elements, then using the loading scheme, we can find the value of the static equivalent load according to the formula F се = x 0 F r + Y 0 F a , where X 0 and At 0 coefficients of radial and axial forces. Coefficient values x o and Y about for bearings of various types are given in reference books. For any bearing, the same static equivalent load can be obtained with different force ratios F r and Fa The bearing is selected from the condition F s ≤C 0 if F s >F r at F s ≤F r accept F s =F r .

76. Dynamic load rating of bearings. Under dynamic load capacity FROM bearings is defined as the permanent radial load (in N) that a bearing with one fixed ring can bear for a nominal life of one million revolutions. Taking into account the condition of strength reliability of the bearing, the durability of the bearing can be represented as L=(C/F) q ≤L p , where L- nominal life of the bearing (million revolutions); FROM- dynamic load capacity (N); q- an indicator of the degree of the bearing fatigue curve; lp= 6 - calculated bearing life, (million revolutions) P- frequency of rotation of the ring, (min-1); lh- estimated bearing life, (hour). Exponent q= 3 - for ball bearings and q= 3.33 - for roller bearings. Dynamic load ratings FROM for bearings of various types and series are given in the reference books.

No. 77 Types of products requirements for them. Machine development stages.

Set of details designed to work together are called assembly unit (node). : bearing, support assembly, gearbox, etc. Despite the difference in machines, the parts and assemblies in them are basically the same: various connections (threaded, welded, etc.), gears (toothed, screw, etc.) shafts, couplings, etc. Product requirements

performance one of the most important requirements criteria: strength( resistance of machine parts to destruction) , rigidity(the ability of parts to resist shape change) , wear resistance(the ability of parts to resist wear, i.e., the process of destruction and separation of material from the surface

solid body). , vibration resistance .

STAGES OF MACHINE DEVELOPMENT

First stage - development terms of reference (TK) - a document containing the name, main purpose, technical requirements, quality indicators, economic indicators and special customer requirements for the product.

Second stage - development of a technical proposal (TP) - aggregates KD substantiating the feasibility of developing a product based on proposals in TK, considering options for solutions. TP approved by the customer and general contractor.

Third stage - draft design development (EP)-collection KD, containing fundamental design solutions, giving an idea of ​​​​the device device, the principle of operation, dimensions and main parameters. This includes an explanatory note with the necessary calculations.

Fourth stage - development of a technical project- aggregates KD- the final decision with a complete understanding of the product design. the issues of reliability of nodes, compliance with safety regulations, storage and transportation conditions are considered and etc.

Fifth stage - development of working documentation (RD) - a set of documents containing drawings so that they can be used to manufacture products and control production and operation. At this stage, optimal designs of parts are developed.

A hydrodynamic, or, as it is often called, a hydraulic bearing is a machine-building unit in which the working fluid that directly perceives the load of the mechanism shaft is a thin layer of insulating lubricating fluid injected into the structure using a lubricated shaft.

History of the invention of the bearing

The history of the invention of the bearing has more than one thousand years. The first primitive plain bearings date back to the Neolithic era. People made them from stones and used them in fire-drilling tools and various spinning tools. With the development of human civilization, primitive plain bearings began to be used in many mechanisms using the wheel principle: in wagons, for making round pottery using a potter's wheel, in windmills for lifting water and driving millstones.

The first information about rolling bearings dates back to 330 BC. During this period, the ancient Greek engineer Diad developed the design of a battering ram for the destruction of fortress walls. In this design, the movable part moved on special rollers along guides.

For the first time, a metal rolling bearing was made in the 111th century in England for a windmill. Structurally, it consisted of two cast-iron rings, which were guides, between which up to forty cast-iron balls were placed.

In the twentieth century, the work of scientists O. Reynolds and N.P. Petrov, who worked independently of each other, led to a remarkable discovery. They found that if the speed of rotation of the machine shaft in a plain bearing filled with lubricant is high enough, then a kind of artificial weightlessness is created on the shaft, at which the shaft ceases to put pressure on the bearing. The technical application of this discovery led to the development of plain bearings with very low coefficients of friction. Further development of the discovery led to the creation of bearings in which the lubricating working medium is injected from the outside with a special pump.

Features of the use of hydrodynamic bearings

Modern hydrodynamic bearings are used in a variety of precision mechanisms, when conventional ball or roller bearings do not meet the necessary requirements for the operation of certain structures and assemblies. For example, if it is necessary to ensure minimum vibration, low noise level, minimum dimensions in cramped operating conditions, and a sufficiently long service life. With further developments and improvements, these bearings become more and more competitive due to decreasing manufacturing costs.

The difference between hydrostatic bearings and hydrodynamic bearings is that in the former, the necessary working pressure of the liquid is created using a special pump, and in the latter, self-lubrication is provided by the working shaft during its rotation. It should be taken into account that the self-lubrication effect is sufficiently effective only when the nominal shaft rotation speeds are reached, otherwise the lubricant layer under the shaft is not thick enough, and this inevitably leads to an increase in friction forces and, as a rule, to premature wear of the mechanism. Therefore, in order to prevent such cases, which can occur quite often, for example, when starting and stopping mechanisms, it is advisable to provide a special “start-up” pump, which will be used only during the above-mentioned transients.

Performance Benefits of Hydrodynamic Bearings

Structurally, hydrodynamic bearings are quite simple and reliable. As a rule, they consist of outer and inner toroidal rings with hermetic seals at the joints. Operating costs are minimal or non-existent. Bearings have a virtually unlimited service life. The requirements for the accuracy of their manufacture are much lower than for the accuracy of the manufacture of ball or roller bearings. The noise level from such bearings is much lower than the noise generated by rolling bearings. Vibrations are minimal. Based design features bearings in some cases have a huge damping capacity.

Disadvantages of hydrodynamic bearings

It is impossible not to note the disadvantages of hydrodynamic bearings.

They have significant energy losses. These losses vary due to outdoor temperature conditions, which greatly complicates the necessary temperature calculations. Hydrodynamic bearings are more prone to sudden failures in emergency situations. Bearings are very sensitive to inaccuracies in the manufacture of shafts and their accessories. Possible leaks working environment during operation. Therefore, it is quite common practice to install two or more journals in bearings to prevent leakage on one side.

Application area

Bearings are used most often in computer installations, for hard drives, for cooling fans personal computer. Application for metalworking machine tools, for nuclear reactors is possible.

The invention relates to mechanical engineering and can be used in thrust and support bearings with a hydrodynamic lubricating layer for machines and, in particular, for bearings of rolling mills, where there are high circumferential speeds and specific loads. The hydrodynamic bearing contains pockets made on one of the working surfaces forming a hydrodynamic lubricating layer. In this case, all pockets are located only in part or throughout the entire area of ​​the layer, where the pressure increases along the length of the layer, and the pockets, starting from the supply one, from which the lubricant enters the layer, are separated from each other along the length of the layer by partitions having pointed tops ending with sealing edges. . Technical result- increasing the minimum thickness of the lubricating layer, reducing heat generation, increasing the bearing capacity, reducing wear. 4 w.p. f-ly, 8 ill.

The invention relates to the field of mechanical engineering and can be used in thrust and thrust bearings with hydrodynamic (liquid or gas) lubrication for various machines, and in particular for bearings of rolling mills, where high circumferential speeds and specific loads occur. Known devices for thrust and support bearings with hydrodynamic lubrication and a viscous lubricating layer, operating according to the Reynolds-Mitchell principle, in which the moving and stationary working surfaces forming the layer are smooth, set at a certain angle between themselves, and the pressure in the liquid (gas) lubricating layer between they are created due to the tightening of the lubricant into a thin, tapering wedge-shaped layer by the forces of viscosity (fluid friction forces) created by the moving working surface. Friction forces from the stationary surface also act on the layer, but they are a reaction to the motion of the layer. With this movement, inertia forces of the mass of the lubricant flow also arise in the layer, caused by a sharp change (including the redistribution over the cross section of the layer) of the velocities of this flow, mainly under the action of fluid friction forces from the stationary working surface in the inlet section of the layer, however, these forces are significant only at the very entrance to the layer on its length (in the direction of movement of the working surface) no more than 2 mm. Further down the layer rapid change speed does not occur and significant inertia forces do not arise. Therefore, in bearings operating according to the Reynolds-Mitchell principle, inertia forces have practically no effect on the formation of pressure in the lubricating layer. Moreover, the forces of inertia arising behind the lubricating layer in its cocurrent flow (in the submerged jet) due to the acceleration of the fluid flowing out of the layer, slowed down in it by a fixed working surface, do not affect. Consequently, in the Reynolds-Mitchel lubricating layer, only viscous forces and the hydrodynamic pressure forces caused by them practically act. The latter push the working surfaces apart and create a lubricant layer of a certain thickness between them. The disadvantage of bearings operating according to the Reynolds-Mitchell principle is that the friction forces acting from the side of a fixed working surface in the region of the layer, where the pressure along its length increases, continuously slow down the lubricant as it moves in the layer. This prevents the lubricant from entering the layer and its further movement there, i.e. reduces the speed and consumption of lubricant in it, which in turn reduces the minimum thickness of the lubricating layer, increases its temperature and reduces the bearing capacity of the bearing. It is impossible to increase the wedge angle (oil gap value) to reduce the specified braking, because any increase in it leads to an increase in lateral leakage of lubricant from the layer, and an increase in the angle of the wedge above a certain size - even to the appearance of a reverse movement of lubricant in the direction of the supply pocket (a recess in the stationary working surface, from where lubricant is supplied to the layer). Known devices are thrust (A. Cameron, "Theory of lubrication in engineering" p. 67, Mashgiz, M., 1962) and support bearings, in which oil pockets in the form of grooves are made on one of the surfaces forming the hydrodynamic lubricating layer, for example, as taken as a prototype of the device according to the author's certificate of the USSR N 796508, class. F 16 From 33/04. In such devices, due to the increase in the thickness of the layer in the oil pockets and the decrease there for this reason of the friction forces from the stationary working surface, the flow in the pockets is accelerated (and swirled) by the moving surface, which improves lubrication in starting modes and, at low specific loads, reduces the release heat. But the inertial forces in these bearing devices also do not contribute to an increase in pressure in the layer, since there pockets along the length of the layer are separated from each other by parts of a fixed working surface, the length of which is much greater than the length of the inlet sections, on which inertial forces are still significant, and they are not able to contribute overcoming the resistance of an extended section of the layer between the pockets and increasing the consumption of lubrication. Consequently, due to the deceleration from these parts of the surface, the inertia forces are completely extinguished and the lubricant flow accelerated in the pockets does not retain the additional speed obtained in the previous pocket until the next pocket. Therefore, occupying the useful area of ​​the working surface where pressure is formed, such pockets at high specific loads reduce the pressure increase in the layer and reduce its minimum thickness. The purpose of the invention is to increase the bearing capacity, reduce energy consumption and wear of bearings. This goal is achieved by the fact that, as in the prototype, on one of the working surfaces that form the hydrodynamic lubricating layer, oil pockets are made that do not communicate with each other. But in addition, according to the invention, all pockets are placed only in part or throughout the entire area of ​​the layer, where the pressure increases along the length of the layer, and the pockets, starting from the supply pocket, from which the lubricant enters the layer, are separated from each other along the length of the layer only by partitions having pointed tops ending with sealing edges. Also, according to the invention, the size of the pockets in the width of the layer is larger than in the length. In addition, there are gaps along the width of the layer between the pockets. Distances along the width of the layer from the edge of the working surface to the pockets increase along the length of the layer. The size of the pockets along the length of the layer and the depth of the sealing edge increase the more, the closer this pocket is to the supply one. The lubrication layer adjacent to the ridge in the pockets, starting from the supply pocket, without experiencing much braking in them from the stationary working surface, is accelerated by the moving working surface and acquires additional speeds throughout its thickness. Further, this layer enters the sealing gap between the pockets (between the sealing edge of the partition and the other working surface). Due to the small length of this gap, the lubricant flow passes through it a path shorter than the length of the inlet section, and the inertial forces in the layer, which are most significant precisely in the initial part of this section, overcoming the friction forces from the side of the edge of the sealing partition and the pressure drop between the pockets on this short path. , to a large extent contribute to the preservation until the next pocket of those values ​​of additional velocities through the layer thickness that were acquired in the previous pocket. Thus, an increase in the consumption of lubricant in the layer is ensured. Due to the fact that, similarly to a tapering wedge, the thickness of the sealing slots at the outlet of the pockets is less than at the inlet, increased lubrication costs with the same layer thicknesses create increased pressures in it, and with the same load on the bearing, increase the layer thickness. Therefore, all other things being equal, in the lubricating layer of the bearing according to the invention, the average lubrication speed, its consumption and the minimum thickness of the lubricating layer (or pressure) will be greater than in the Reynolds-Mitchel layer and in the prototype layer. Since the size of the pocket along the length of the layer is chosen no more than that required to restore in the pocket part of the flow velocity lost to overcome the resistance on the way between the pockets in the sealing gap, the number of pockets along the length of the layer will be optimally large, providing multiple (multi-stage) use of inertial forces to increase the speed of lubrication in the layer. In the region of the layer where the pressure does not increase (has reached a maximum or falls), due to the absence of pockets there, the fixed surface slows down the lubricant flow as much as possible, as is required to reduce the pressure drop. In addition, the location of the pockets outside the zone of maximum wear occurring in the place of the minimum layer thickness, significantly reduces the wear of the thin tops of the partitions between the pockets. The sections of the working surface between the pockets and at the edges of the layer in the region of the pockets mainly serve as seals that reduce lateral leakage, and the formation of pressure in the layer is ensured when the lubricant flow passes through the sealing slots from one pocket to another. Therefore, the deepening of the sealing edges relative to the level of the working surface makes it possible to form different layer thicknesses in the sealing slots and at the working surfaces and create their optimal values ​​both to reduce lateral leaks and to increase lubricant consumption. In addition, providing an increase in the width of the working surface at the edges of the layer, as the pressure increases along its length, reduces lateral leakage. As a result of the general influence of these design factors, the minimum thickness of the lubricating layer increases by more than 2 times. Consequently, the heat emission (energy consumption) decreases by the same amount and the bearing capacity of the bearing increases more than 4 times, and its wear also decreases. In FIG. 1 is an isometric view of a support bearing bushing with running surfaces at intervals separating pockets along the ply width. In FIG. 2 shows a cross section of the bushing shown in FIG. 1, and the section of the shaft. In FIG. 3 shows a section along the length of the Reynolds-Mitchel lubrication layer and the distribution of lubrication rates over the layer thickness. In FIG. 4 shows a section along the length of the lubricating layer of a bearing according to the invention and the distribution of velocities in it over the thickness of the layer. In FIG. 5 shows a plan view of a thrust bearing chock with a variable width of the working surface at the edges of the layer in the region of the pockets. In FIG. 6 shows a section along A-A of the pillow in FIG. 5. In FIG. 7 shows a section along B-B of the pillow in FIG. 5. In FIG. 8 shows a section along A-A of the sleeve in FIG. 2. In FIG. 1 and 2 sleeve 1 support bearing shown: pockets 2, the working surface 3 of the sleeve, located in the area where there are no pockets "baffles 4 between the pockets and sections of the working surface 5 and 6, located respectively along the edges of the sleeve and between the pockets along the width of the sleeve, sealing edges 7, made on the pointed tops of the partitions 4 and having a blunt or rounded size 8. The size of the pockets along the layer width is greater than along the length, and greater than the size along the layer width of the sections of the working surface in the intervals between the pockets. 2, additionally shown: a shaft 9 rotating at a circumferential speed 10 and having a working surface 11, forming with the inner surfaces of the sleeve 1 parts of the lubricating layer 12 and 13, respectively, in the area of ​​the pockets 2 and outside it, and the supply pocket 14. The diagram is also shown 15 pressure distribution in the lubricating layer along its length, angle 16 - the central angle between the location of the maximum pressure in lubricating layer and a partition at the supply pocket and angle 17 - the central angle within which the pockets are located. In FIG. 3 shows a section along the length of the Reynolds-Mitchel lubricating layer formed between the fixed working surface 18 of the thrust pad and the working surface 11 of the thrust bearing moving at a speed of 10. A pressure is formed in the layer, in which the distribution diagram 19 is similar to the diagram in the support bearing layer without pockets. Up to point 20 of diagram 19, the pressure increases, and then it falls. In front of the layer in the space 22 between the thrust pads (or in the supply pocket of the support bearing), from where the lubricant is supplied to the layer, along the flow thickness equal to the maximum thickness 23 of the lubricating layer, the velocity distribution diagram 24 has a rectangular or close to it shape. In the layer, having passed its inlet section 25, the flow acquires a fairly steady (slowly changing along the length of the layer) velocity distribution over the layer thickness, as shown in diagram 26. Such a change in the shape of the diagram in the inlet section (from 24 to 26) occurs due to flow deceleration fixed working surface 18, which changes the diagram to a triangular shape 27, and due to braking by the pressure formed in the layer, additionally changing the diagram to the shape of a concave triangle 26. As can be seen from the comparison of diagrams 24 and 26, the area of ​​diagram 24, and hence the lubricant consumption before entering the layer, is more than 2 times greater than the area of ​​diagram 26 and lubricant consumption in the layer. Consequently, the flow of lubricant with a thickness of 23 does not entirely enter the layer, and most of its flow rate, corresponding to the difference in the areas of velocity diagrams 23 and 26, remains in the supply pocket and is carried away by the vortex 21 circulating there. Further, when the flow moves in the layer, the shape of its velocity diagram , slowly changing, acquires a triangular shape 28 in the place where the pressure reaches its maximum, and then in the area of ​​pressure drop in the layer - the shape of a convex triangle 29, due to the fact that there the pressure accelerates the flow. If we do not take into account the flow in the layer along its width (lateral leakage), then all the areas of diagrams 26, 28, 29 and the corresponding lubricant costs are equal. In the lubricating layer of the prototype (in a bearing with pockets), when the flow enters the layer from each pocket, a process takes place similar to that discussed above when entering the lubricating layer from the supply pocket. There, before entering the lubricating layer, the velocity distribution is the same as in the supply pocket corresponding to plot 24, and in the layer between the pockets, since the length of this layer is greater than the length of the inlet section, the velocity distribution is set corresponding to plot 26. Thus, in the prototype in In all pockets, most of the lubricant of the flow adjacent to the ridge with a thickness equal to the thickness of the layer also does not enter it, but swirls and remains in the pockets. The disadvantage of bearings operating according to the Reynolds-Mitchell principle, including prototype bearings, is that the friction forces acting from the side of a fixed working surface in the region of the layer, where the pressure along its length increases, continuously slow down the lubricant as it moves in the layer. This prevents the lubricant from entering the layer, i.e. reduces the speed and consumption of lubricant in the layer, which in turn reduces the minimum thickness of the lubricating layer, increases its temperature and reduces the bearing capacity of the bearing. It is impossible to increase the wedge angle (oil gap value) to reduce the specified braking, because any increase in it leads to an increase in lateral leakage of lubricant from the layer, and an increase above a certain size - even to the occurrence of a reverse movement of lubricant in the stationary working surface towards the supply pocket. As for the region of the layer where the pressure does not increase (has reached a maximum or falls), then braking from the side of a stationary working surface is useful there, because. it reduces not only lateral, but also end leakage, prevents the lubricant from being carried away from the layer by the working surface. In FIG. 4 in expanded section of the lubricating layer of the support bearing according to the invention shown in FIG. 1 and FIG. 2 (the following is also true for a thrust bearing), shows: a thrust bearing bush 1, non-communicating pockets 2, which are located only in the part 12 of the layer region, where the pressure increases along the length of the layer. In addition, these pockets, starting from the supply pocket 14, from which the lubricant is supplied to the layer, are separated along the length of the layer not by sections of the working surface that inhibit lubricant, but only by partitions 4 having pointed tops, ending with sealing edges 7, made flush with working surface 5 or recessed relative to this level by a value of 30 so that at the inlet of the lubricant into the pocket the thickness of the gap between the sealing edge 7 and the other working surface 11 was greater than this thickness at the exit from the pocket. The size of the oil pockets 31 and 32 along the length of the layer must be no less than the value at which the flow that entered the pocket from the gap between the sealing edge and the other working surface 11 acquires, after passing through the pocket, the average speed is greater than 2/3 of the speed of the moving working surface. This corresponds to diagram 34. The sealing edges have blunt or rounded edges of size 8, which provide minimal flow deceleration due to the fact that this size is minimal, no more than 2 mm and less than the value at which the average flow velocity in the slot decreases at the exit from it to a value not less than 1/2 of the speed of the moving working surface. This corresponds to diagram 33. The size of the pockets along the length of the layer (the distance between the sealing partitions) increases from a value of 31 to a value of 32 at the supply pocket. The depth of the sealing edge increases the more the closer this pocket is to the feeder. It also shows: the working surface 3 of the sleeve, located in the region 13 of the layer, where there are no pockets; a plane 6 connecting the sealing lips and showing the contour of the main laminar flow; working surfaces 5, located along the edges of the sleeve and between the pockets along the width of the sleeve, may coincide with the plane 5, as shown in Fig. 1 and FIG. 2; shaft 9, rotating at peripheral speed 10, and having a working surface 11, forming with the inner surfaces of the sleeve 1 part of the lubricating layer 12 and 13. Also shown is a diagram 15 of the distribution of pressures in the lubricating layer along its length, where the maximum pressure is located at a point given by angle 16. The lubricating layer of a thrust bearing according to the invention would have a similar appearance. If pockets with such partitions are also placed in area 13, where the pressure drops, then this will also reduce the flow deceleration, but will contribute to the removal of the lubricant from the layer, and this is not advisable. Therefore, pockets should be located only in the region of the layer where the pressure increases along its length. The device according to the invention works as follows. The lubricant in the supply pocket, as in the Reynolds-Mitchel layer discussed above, is accelerated by the moving working surface 11 and the adjacent flow with a thickness of 23 equal to the maximum thickness of the lubricant layer acquires additional speeds, as shown in diagram 24. In this case, the process of transferring kinetic energy lubrication from the ridge occurs with maximum efficiency, since the layer over its entire thickness 23 acquires the maximum possible speed (the speed of the moving surface). Further, this flow enters the region 12 (where the pockets are located) of the lubricating layer, which, according to the invention, is a wedge gap between the surface 11 and the surface 5, as well as the plane 6. Then the lubricant enters the pockets 2 and further into the layer of the region 13, where the pockets missing. In region 12, the flow first enters the gap between the sealing lip 7 of the first baffle and the working surface 11 (pocket gap). Due to the influence of this edge, despite its small friction surface (small value 8 of its blunting or rounding), and also due to the pressure drop between the first pocket 2 and the supply pocket 4, the flow rates change in such a way that the diagram 24 of these speeds before the sealing edge is converted to diagram 33 behind the sealing edge. As can be seen from the comparison of these diagrams, in the device according to the invention, the fixed part of the bearing (sleeve or thrust pad) also exhibits some resistance to flow, but this resistance, as can be seen from the comparison of diagram 33 in FIG. 4 and plots 26 in FIG. 3, there is significantly less resistance to the flow of a fixed part in the Reynolds-Mitchel layer and in the prototype layer, since the area of ​​the first diagram at the same speed 10 of the moving working surface 11 is significantly larger than the area of ​​the second diagram. Therefore, the consumption of lubricant introduced from the supply pocket 4 into the bearing layer according to the invention is significantly (more than twice) greater than that of the Reynolds-Mitchell bearing and the prototype. Although not the entire lubricant flow, with a thickness of 23, enters from the supply pocket into the layer, and part of it, corresponding to the difference in the areas of the velocity diagrams 24 and 33, remains in the supply pocket as part of the vortex 21. Further, in the first pocket, the flow, similarly as in the supply pocket, accelerates and along the thickness of the flow (thickness between plane 6 and surface 11), the velocity diagram acquires a shape 34 in front of the second partition. This shape is not a complete rectangle, like the shape of diagram 24, due to the smaller length and depth of pockets 2 than that of the supply pocket. These dimensions of the pocket, and especially its length, must be optimal so that the number of pockets is not very small, but also so that the flow velocity diagram 34 acquires sufficient completeness in the pocket in order to accumulate kinetic energy for it to overcome the resistance of the next gap between pockets without a large loss of flow. This loss still takes place and corresponds to the difference in the areas of velocity diagrams on both sides of the sealing gap. The lubricant that has not entered the sealing slot remains in the pocket and circulates there as part of a vortex, similar to the vortex 21 in the supply pocket. The increase in pressure in the pockets 2 occurs because the gap between the sealing edge 7 and the working surface (the thickness of the sealing gap) at the outlet of the pockets is smaller than at the inlet. Thus, the increase in the consumption of lubricant introduced by the moving surface, and consequently, the increase in pressure in the layer according to the invention compared to the Reynolds-Mitchell layers and the prototype occurs mainly for two reasons: firstly, the size 7 of the blunting or rounding of the sealing edge is significantly smaller than the length of the inlet section, so the hydraulic resistance of the sealing gap between the pockets will be so much smaller that the flow velocity diagram will not yet acquire a steady shape, similar to 26 in Fig. 3 and the inertia forces help to overcome the resistance of this sealing gap; secondly, the dimensions of the pockets along the length of the layer 31 and 32 are such that the flow, as it moves in each pocket, has time to acquire increased speeds throughout the entire thickness of the specified slot to overcome its resistance with a maximum lubricant consumption, but these dimensions should also be as small as possible to increase the number of pockets so that the process of accelerating the flow in the pockets is more multiple throughout the layer where the pressure rises. The considered principle of creating pressure in the lubricating layer according to the invention is similar to the principle of creating pressure in a rotary turbomachine: there, in each stage, the moving rotor transfers kinetic energy to the working fluid, and then, in a stationary guide vane, this energy is converted into pressure energy. Similar to this process, in the lubricating layer according to the invention, in each pocket, along its length, kinetic energy is transferred to the lubricant flow by the moving working surface, and further, in the sealing gaps between the pockets, this kinetic energy is converted into pressure energy in the next pocket, since in this gap the inertial forces flow and hydrodynamic friction forces from the moving surface act against the pressure forces corresponding to the pressure difference between the pockets. Sections 5 of the working surface between the pockets and at the edges of the layer mainly serve as seals that reduce lateral leakage, the formation of pressure in the layer is provided by the difference in the thicknesses of the sealing slots at the inlet and outlet of the pockets. Therefore, the deepening of the sealing edges relative to the level of the working surface makes it possible to form different layer thicknesses in the sealing slots and at the working surfaces and create their optimal values ​​both to reduce lateral leaks and to increase lubricant consumption. For this purpose, the thickness of the lubricating layer between surfaces 5 and 11 is assumed to be minimal, less by 30 than the thickness of the sealing slots. This design measure reduces lateral leakage while increasing the amount of lubricant carried by the moving work surface. In the region of the layer where the pressure does not increase (has reached a maximum or falls), due to the absence of pockets there, the fixed surface slows down the lubricant flow as much as possible, as is required to reduce the pressure drop. In addition, the location of the pockets outside the zone of maximum wear, occurring in the place of the minimum layer thickness, significantly reduces the wear of thin sealing partitions between them. In the pocket area, the width of the working surface at the edges of the layer can increase along the length of the layer, as the pressure in the layer increases, which further reduces lateral leakage. In FIG. 5 shows a plan view of a thrust bearing pad, in which, in the area where the pockets are located, the width of the working surface at the edges of the layer increases along the length of the layer. In FIG. 6 and FIG. 7 shows the cross-sections of this pillow, respectively, along AA and along BB. These figures show: region 12 where pockets 2 are located; area 13 at the exit of the layer, where there are no pockets; diagram 15 of pressure distribution along the length of the layer; the smallest 35 and largest 36 dimensions of the width of the working surface at the edges of the layer; the smallest 37 and the largest 38 pocket dimensions along the length of the layer (pocket length); size 39 of the pocket along the layer width (pocket width), diagram 40 of the pressure distribution along the layer width. In FIG. 8 shows a section along AA (Fig. 2) along the width of the thrust bearing bushing, in which, in addition to sections of the working surface at the edges of the layer, having a size of 41, pockets 2 are separated from each other along the width of the layer by sections of the working surface, having a size of 42. The diagram is also shown there 43 distribution of pressure over the width of the layer. The device according to the invention shown in FIG. 5-8 works as shown in FIG. 4. In addition to the above, it should be noted that an increase in the width of the working surface along the length of the layer at its edges from size 35 to size 36 (Fig. 5) reduces the amount of leakage from the layer, since a greater width is created at the site of greater pressure (see diagram). 15 in Fig. 6). In addition, an increase in the size of the pockets along the length of the layer from a value of 37 to a value of 38 (Fig. 6) at the supply pocket provides optimal conditions for restoring the flow rates in the pockets, which are reduced in the sealing slots at the entrance to the pockets, because the greater the thickness of the slot (the thicker the flow introduced into the pocket), the greater the distance between the sealing slots is necessary to restore the flow rates. From this condition, and also taking into account the actual dimensions of the thicknesses of the sealing slots and the feasibility of forming a larger number of pockets, the dimensions of the pockets 39 (Fig. 7 and Fig. 8) should be greater in the width of the layer than in the length. As for the ratio between the dimensions 39 (Fig. 8) of the pockets and the dimensions 42 of the sections of the working surface in the intervals between the pockets, then given that these sections are intended only to reduce the flow of lubricant across the width of the layer from pocket to pocket, dimensions 32 should be smaller sizes 39. As a result of the general influence of these design factors, the minimum thickness of the lubricating layer increases by more than 2 times. Consequently, the heat emission (energy consumption) decreases by the same amount and the bearing capacity of the bearing increases more than 4 times, and its wear also decreases.

CLAIM

1. Hydrodynamic bearing, in which oil pockets are made on one of the working surfaces forming a hydrodynamic lubricating layer, characterized in that all pockets are placed only in part or throughout the layer area, where the pressure along the length of the layer increases, into the pockets, starting from the feeder, from which the lubricant enters the layer, are separated from each other along the length of the layer by partitions having pointed tops ending with sealing edges. 2. The bearing according to claim 1, characterized in that the size of the pockets in the width of the layer is greater than in the length. 3. The bearing according to claim 1, characterized in that there are sections of the working surface along the width of the layer between the pockets. 4. Bearing according to claim 1, characterized in that the distance across the layer width from the edge of the working surface to the pockets increases along the length of the layer. 5. Bearing according to claim 1, characterized in that the dimensions of the pockets along the length of the layer increase the more the closer this pocket is to the feeder.

In a hydrodynamic bearing, there is no direct contact between the rubbing surfaces, since the gap between them is filled with a lubricating fluid under the action of hydrodynamic forces.

The use of a hydrodynamic bearing makes it possible to replace sliding friction with fluid friction and reduce energy losses.

In a hydrodynamic bearing, a thin layer of fluid takes up the load and transfers it to the support.

Conditions for the occurrence of fluid friction

For the operation of a hydrodynamic bearing, it is necessary to create a hydrodynamic lubrication layer, for this it is necessary to ensure the following conditions:

• lubricating fluid must be kept in a gap (e.g. between shaft and bearing housing)
• sufficient pressure must be maintained in the lubricating fluid to balance the load
• the liquid must completely separate the sliding surfaces, which means that its layer must be higher than the sum of the surface roughness
• the thickness of the liquid layer must be greater than the minimum value

The principle of operation of a hydrodynamic bearing

Consider the scheme of operation of a hydrodynamic bearing.

The shaft is installed in a housing filled with oil with a gap, under the action of a load it is pressed against the lower surface of the housing. It turns out that in the initial position the shaft is located in the housing with eccentricity.

When the shaft rotates, a small layer of liquid begins to move due to adsorption and is carried away after the surface of the shaft. Subsequent layers can also be entrained in rotational motion due to the viscosity of the working oil. It turns out that the shaft acts as a pump, creating a flow of working fluid, and forcing it into the wedge-shaped gap between the housing and the shaft. As a result of the impact of the rotating shaft, the oil tends to fill the wedge-shaped gap and raise the shaft, on the other hand, this is prevented by the load acting on the shaft.

When enough is created to lift the shaft and ensure the flow of oil around the entire circumference, an equilibrium state occurs.

Hydrodynamic bearing with wedge bores

To ensure high anti-vibration properties, a hydrodynamic bearing with wedge bores is used, in which the shaft journal rests on several oil wedges. This reduces the chance of vibrations.

Calculation of a hydrodynamic bearing

The condition for ensuring liquid friction:

H≥1.1(Rz1 +Rz2 +y)

• where h is the thickness of the lubricant layer
• R z1 workpiece roughness 1
• R z2 workpiece roughness 2
• y - deflection arrow of the spike (shaft)

The smallest relative eccentricity ratio can be calculated using the formula:

X=1-(h/0.5s)

• where s is the average clearance
• x - relative eccentricity x = e / 0.5 s

The required viscosity of the fluid, at which it will be possible to achieve the fluid friction mode, can be determined by the formula:

μ=PΨ 2 /ωldFr

• l - shaft length, m
• d - shaft diameter, m
• ω - angular speed of rotation of the shaft
• P - load value
• Ψ - relative clearance Ψ = s/d
• Fr - dimensionless bearing capacity coefficient

During operation of the hydrodynamic plain bearing, the oil will heat up, which means its viscosity will change. The dependence of viscosity on the temperature of the working fluid is reflected in. If the initial oil temperature is unknown, the calculation is carried out by the method of successive approximations, setting the initial value to -50 °C.

Advantages of hydrodynamic bearings

• high resource
• low noise
• small vibrations during operation
• shock damping

Disadvantages of hydrodynamic bearings

• Ability to work only at high speeds
• influence of temperature on the mode of operation, characteristics