FI111026B - Roll support system with hydrostatic bearings - incorporating hydraulic bearing control systems - Google Patents

Abstract

A roll has axles (3) supported in bearing blocks on hydrostatic bearing elements (10,20,30). A principal bearing element (10) acts in the principal vertical loading direction with a back-up bearing element (20) acting in the opposite direction. The bearing elements have shoes to support axles for rotation relative to the bearing blocks. The hydraulic system is controlled to feed a hydraulic fluid to said shoes so that they can float freely around the neck of axles.

Description

111026

Bearing control system for a hydrostatic roller bearing

Styrsystem av lager för en hydrostatiskt lagrad Vals. 5

BACKGROUND OF THE INVENTION The present invention relates to a bearing control system for a hydrostatic-bearing roll, in particular a nip forming roll with a counter roll, such as a calender roll, press roll or the like, which is axially mounted on a bearing or the main bearing element acts in the main load direction and one bearing element, the so-called a counter-bearing element against the main loading direction, and which bearing elements are provided to rotatably support the shaft of the roll shaft relative to the frame member, such as a bearing bracket, around the axis of the roll axle neck.

The rolls of paper machines and paper finishing apparatuses, with the exception of rollers in which the roll shell is movable relative to the roll axis, have hitherto, as a rule, been provided with rolling bearings in machine frame structures. Especially with nip forming rollers such as calender rolls and sofl calender rolls, such rolling-bearing problems have given rise to certain problems, which have required special arrangements e c. In nip rolls, nipple loads need to be measured, but in some cases measurement has become problematic. The soft calender has to be driven in a line loading area where the bearing load is close to zero, which is a rolling bearing, since under zero load, the rolling bearings of the rolling bearing can slide considerably relative to the bearing frames, resulting in a rather fast bearing. In heated rollers, such as calender rolls and the like, the problem is mainly in the successful lubrication of the bearings, rather than in the control of bearing play. This has also led to the use of special glide paths for rolling bearings. A further problem with roller bearings is the speed limitation. Such limitations with respect to rotational speed include, for example, the heat generation of the bearing 2 111026 and the associated possibility of bearing cooling. As a result, rolling bearing sets a rotational speed limit that the bearing manufacturer does not allow to be exceeded. As mentioned earlier, the bearing has a limited rolling accuracy due to e.g. inaccuracies in rolling bodies and orbits. Even if the conventional roll 5 is made with all the components as accurately as possible, the errors caused by inaccuracies are added up in the assembled roll.

In the prior art, reference is made, for example, to European Patent Publication No. 158,051, which relates to the bearing of a washing drum. The bearing in this solution is implemented as sliding bearing 10 by means of hydrostatic bearing segments. However, the bearing method of this publication cannot be used, for example, in nip contact calender rolls, since the bearing method of this EP publication does not include any radial movement capability in the bearing itself. Also, the bearing load cannot be altered in any way, and separate spring elements are required to dampen the vibrations. In the prior art, reference is also made to Fl patent application No. 942756, which discloses a heated roller bearing. A disadvantage of this type of bearing is that there is no radial displacement possibility because at least one hydrostatic bearing segment of the plain bearing solution is radially fully fixed and rigid to the frame structure and thus the roller bearing with the 20 bearings is not able to move e.g.

It is an object and object of the present invention to provide a completely new type of control system for a sliding bearing roller which avoids the drawbacks of prior art solutions and further provides a significant improvement over the prior art solutions. To accomplish this, the control system of the invention is essentially characterized in that the control system is arranged to supply the pressures of the hydraulic pressure medium to the hydrostatic bearing elements so that the bearing shoes 30 of the bearing elements are freely positioned around the shaft neck.

3, 111026

The invention provides significant advantages to prior art solutions. and of these advantages may be highlighted e.g. following. First of all, the bearing is not directly connected to the main load direction, especially in the nip plane direction, but 4 is hydraulically loaded by a piston-cylinder type device towards the shaft.

This allows the control system according to the invention to move and move the roll in said main loading direction. Furthermore, the invention allows for accurate nip roll adjustment of nip rolls and measurement of nip load without special alignment. The hydrostatic bearing shoes included in the bearings may, when using the system of the invention, be completely freely positioned around the neck of the roll shaft, e.g. the increase in shaft diameter caused by shaft heating does not cause any problem, as does the inclination due to shaft deflection or other position errors. The bearing and its control system can be intentionally used, for example, to close and open the roll nip, since all hydrostatic bearing shoes are movable. The nipple load can be directly calculated from the oil pressures of the bearing shoes because these shoes provide a load on the roll nip. The bearing movement itself is receptive and damping from vibration. Other advantages and features of the invention will become apparent from the following detailed description of the invention.

The invention will now be described, by way of example, with reference to the figures in the accompanying drawings.

Figure 1 shows a schematic and partially sectional side view of a second side of the roll slide bearing.

25 'Figure 2 schematically illustrates an embodiment p of a hydraulic control system for a bearing.

Figure 3 schematically shows another embodiment of a hydraulic steering system for bearings.

4, 111026

Figure 4 schematically shows another embodiment of a hydraulic steering system for bearings.

Figure 5 schematically illustrates an alternative embodiment of a hydraulic control system for a bearing.

Figure 6 shows a further variation of the hydrostatic-bearing roller bearing hydraulic control system of the invention.

Figure 7 is an alternative embodiment of a pre-steered hydraulic roller bearing control system.

Figure 7A shows a detail of Figure 7.

Figure 8 schematically illustrates the principle of operation of an electro-hydraulic control system.

Figure 1 of the drawing shows, schematically, a hydrostatic bearing for the roll support which forms a nip with the counter roll. The roll itself is illustrated by dashed dots and is denoted by reference numeral 1 and the axis of the roll is denoted by reference numeral 3. Reference numeral 2 illustrates a backing roll forming a nip • 'N with roller 1, for example a calendering nip. The nipple level is indicated by reference marks A to A.

The hydrostatic bearing according to Fig. 1 consists of bearing elements 10,20,30, 40, 25 mounted on a bearing bracket 4 and supported against a roll axis 3. The bearing arrangement comprises a main bearing 10,30 acting in the main loading direction, i.e. nip plane A-A, The main bearing is divided into three parts in the representation 30 of FIG. 1, comprising a first bearing element 10 acting in a nip plane, and secondly arranged second angularly spaced bearing members 30. However, according to Figure 1, the main bearing divided into three parts 5111026 is merely an exemplary representation of the bearing structure which. can be accomplished in other ways. The first bearing element 10 and the second bearing elements 30 are of a similar construction in that they comprise a cylinder 11,31 mounted on the bearing bracket 4 with a movably mounted load piston 5 13,33. Under the piston 13.33, a cylinder 11.31 is formed in the cylinder 11.31 at pressures 12.32, whereby a pressure piston 13.33 is provided to cause the piston 13.33 to be loaded towards the shaft 3. A bearing shoe 16.36 is provided at the shaft-end 3 lubricating oil pockets 15.35. The loading piston 13.33 is provided with through capillary bores 14.34 which thereby connect the cylinder pressure space 12.32 to the lubricating oil pockets of the bearing shoe 15.35. Through the capillary bores 14.34, the pressure medium thus enters the lubricating oil pockets 15.35 such that an oil film is formed between the bearing shoe 16.36 and the shaft 3, through which the bearing shoe 16.36 is supported against the shaft 3.

The counter-bearing 20 is similar in structure to the main bearing bearing elements 10,30 in that the counter-bearing element 20 consists of a cylinder 21 mounted on the bearing bracket 4 and a load piston 23 movably fitted to the cylinder. along which the oil from the cylinder pressure chamber 22 below the loading piston 23 passes into the lubricating oil level 25 to form an oil film between the bearing shoe 26 and the neck of the shaft 3.

According to Figure 1, the counter bearing element 20 is mounted on the nip plane A-A so that its direction of action is parallel to the nip plane but opposite to that of the main bearing.

25

Figure 1 further shows that the shaft 3 is supported by bearings transverse to the nip plane A-A, i.e. by side bearings 40. In Figure 1, the side bearings 40 consist of a body member 42 mounted on a bearing bracket 4 and a bearing shoe 41 mounted on the bearing member against the roll axis 3. The actual purpose of the side bearings 40 is merely to hold the shaft 3 in its proper position and to dampen the vibrations in the transverse direction.

6 111026

In the arrangement of Figure 1, the loading of the roll nip N is achieved by the bearing elements 10, 30 of the main bearing. This allows the nip load to be directly calculated from the oil pressures of the bearing elements 10.30. Thus, precise control of the nip load can be accomplished in 5 sequentially simple steps in accordance with the invention. As shown in Figure 1, the main bearing is divided into three separate bearing elements 10,30, whereby, if necessary, the bearing is subjected to quite high loading forces. In current soft calendars, in many cases, a nipple load must be obtained which corresponds to the bearing load of the bearings. With the oldest and traditional bearing solutions of the Ta-10, this can be done quite swiftly. However, the zero-load situation does not pose any problem in the invention, since the adjustment of the load to the zero-load situation is easily accomplished and achieved by the main bearing 10,30 and the counter-bearing element 20. As has been pointed out previously, in the main loading direction, i.e. in the nip plane A-A direction 15, the bearing elements 10,20,30 are not fixedly bonded to the bearing bracket 4, but each of the bearing elements is movable. This provides effective vibration damping. Further, thanks to this feature, and especially when the play elements of the bearing elements are sized sufficiently large, the opening and closing of the nip N can be carried out by means of the bearing elements. For example, for soft calenders, the bearing element 20 can then also be used for quick opening of the nip.

♦ *

According to the invention, the bearing control system of the roller bearing mounted as shown in Figure 1 can be implemented as either a mechanical hydraulic or an electro-hydraulic solution. The mechanical hydraulic stiffeners are shown in Figures 2-7 and are based mainly on the determination of the position of the main shoe bearings 10,30 by means of a hydraulic element. The hydraulic tracking system aims to drive the guided bearing element to a predetermined position and maintain that position constant, regardless of load variations. The electro-hydraulic control system, schematically depicted in Figure 6, is based on electrical measurement of the position of the central bearing element 10 of the main bearing 10,30 30 and adjustment of the loading pressure based on the measurement data.

7 111026

Figure 2 schematically shows a part of the hydraulic control system of the bearing. from which the bearing control is implemented directly by the controlled valve 50. The operation of the rigid system shown in Figure 2 is as follows. When the main bearing elements 10,30 are in the lower position, the upward movement 5 of the main bearing elements 10,30 and thus the roller shaft 3 is achieved by first applying a supply pressure ps to the line A. When the main bearing elements 10,30 are in their lower position lowered against the force of the spring 53 because the pin 52 formed in the spindle rests against the piston 13 of the central main bearing element 10, thereby allowing the pressurized oil to flow from line A to lines B and C because the flow path in valve 50 is open From the pressure lines B and C, the pressurized oil can continue to flow to the lines E and F and from there further under the load pistons 13.33 of the main bearing elements 10, whereby the shaft 3 rises up under the action of the load pistons 13.33. The movement of the shaft 3 stops as the flow narrows in the valve 50 when the valve stem 51 has moved to a position corresponding to the load. The spindle 51 is then moved to a position where the valve 15 is open at a distance x 0, as shown in Figure 2. The counter-bearing element 20 is continuously exposed to constant pressure.

With the roll in the load state and with the roll load increasing so much that the shaft 3. ; presses down, moving the load piston 13,33 with it down, valve 50 opens 20 more, allowing the pressurized oil to exert a greater pressure under the load piston 13,33. The opening of the valve 50 thus compensates for the increase in load.

When the roll nipple is to be opened, for example in a quick release situation, the supply pressure ps is removed from the pressure line A. The shaft 3 and the load pistons 13,33 move downwardly and close to the lower position of the valve 50 50 through the piston 13.33. The braking pressure pj is lower than the supply pressure ps and has been • «engaged all the time.

Thus, in the arrangement of Figure 2, the bearing flow and hence the pressure 30 is adjusted based on the load of the bearing elements 10,30 on the bearing shoe. As the load increases, the position of the loading pistons 13.33 tends to change, whereby the system compensates for the change by increasing the flow. Following this principle, 8 111026 seeks to stabilize the bearing oil film thickness. The hydrostatic bearing, whose oil film thickness remains constant regardless of load, has infinite rigidity. This is a common feature sought by hydrostatic bearings.

* 5 Figure 3 illustrates a control system in which the hydrostatic bearing is controlled by a pre-controlled valve. Advantages of pre-steering compared to direct steering include: greater pressure / force gain and better stiffness.

The valve's zero point is still more adjustable. The pressure supply of the three bearing elements 10, 30 with one main valve is easier to implement than, for example, the direct-controlled valve of Figure 2 10.

Figure 3 illustrates a control system with the roll in the loaded condition, with the pilot control valve 60 in the center position shown in Figure 3. The directional valve 80 is then also in the position shown in Figure 3, whereby the supply pressure ps passes through said directional valve 1580 to the main valve 70 and thereafter through an opening xoo to the main bearing elements 10,30. The biasing pressure pe flows through the boreholes B1 and B2 at each end of the main valve 70 (via bores formed in spindle 71) and thence along lines A1 and A2 to the biasing valve 60 and the borehole B2 has equal pressure springs 72,73 '· center the mandrel 71 in the position shown in Fig. 3. Thus, in the situation illustrated in the figure, through the flow opening xoo of the main valve 70, the lubricating flow required by the desired oil film thickness at the designed pressure difference flows to the bearing shoes 10,30 of the main bearing elements.

25

As the load on the roll increases, the force is transmitted through the roll shaft 3 and the central main bearing member 13 to the spindle 61 of the pilot valve 60 via pin 62, whereby spindle 61 moves downwardly as flow increases from line A2 and correspondingly decreases from line A1 to tank duct G. , thereby increasing the leftward force on the spindle 71 of the main valve 70 in Fig. 3. The spindle 71 of the main valve 70 then begins to move to the left in the figure, thereby opening the flow opening X02 controlling the supply of bias pressure pe, and closing the flow opening xoi 9111026 accordingly. The spindle 71 of the main valve 70 moves in Figure 3. to the left until a new balance of power is achieved. The opening of the flow port X02 corresponds to the increased opening of the spindle 61 of the pilot valve 60. At the same time, the flow opening xoo of the spindle 71 of the main valve 70 feeds 10.30 more oil to the bearing elements, which means that the main valve 70 has compensated for the increase in the load on the roll.

First, when opening the roll nip, the directional valve 80 is set to a second position, whereby the pressure below the load piston 13,33 of the main bearing elements 10,30 is coupled to the line G of the tank-10 and the supply pressure ps is applied to the braking pre-control pressure. The shaft 3 and the load pistons 13,33 then move to their lower position and close to said lower position the pilot control valve 60 opens the connection between the braking pilot control pressure pej and the right valve spring chamber B1 of the main valve 70 through the flow line AI. Thus, in Fig. 3, spindle 71 of main valve 70 immediately moves to its leftmost position 15 and closes communication with channels below load pistons 13,33 to tank line G. As main valve spindle 71 moves as described above, the braking pressure pi from line to load pistons 13.33 opens simultaneously. Under the piston 13,33, a braking pressure pi is then applied, which dampens the downward movement of the shaft 3 and the piston 13,33, while also supplying the lubricating oil 20 to the bearings. As the roll is lifted again, at the beginning of lifting, the directional valve 80 again returns to the position shown in Figure 3, whereby the spindle 71 of the main valve 70 moves slightly to the right in Figure 3 and the pressurized oil supply ps is restarted through the flow port xoo.

The bearing control provided by the pre-controlled valve shown in Figure 4 is very much in line with what was described in connection with Figure 3. Fig. 4 differs from Fig. 3 only by the fact that the mode of supplying the bias pressure is different in the present embodiment. The solution of Fig. 4 is shown here as merely an alternative embodiment, since the embodiment of Fig. 3 may be preferred to the present solution, in particular because the stiffness achieved by the arrangement of Fig. 4 is at least somewhat inferior to that of Fig. 3.

10 111026

Referring to Figures 3 and 4, and with regard to positioning the piston 13,33, it may be noted that the pre-controlled main valve 70 enables fine adjustment of the piston 13,33. By tracking the excess leakage to the preconditioning line AI, the equilibrium condition of the spindle 71 of the main valve 70 is realized at various positions of the preconditioning valve 60.

Figure 5 is a fully schematic diagram of yet another embodiment of the invention. In the solution of Figure 5, valve 60A is a main control valve for supplying bearing elements 10,30 10, through which the supply pressure ps is applied under the bearing pistons 13,33 of the bearing elements. The control of the valve 60A may be provided either manually or electrically by means of a control member 62A connected to the valve stem 61A. The system of Figure 5 further includes an auxiliary valve 70A which responds to changes in load so that as the pressure of the bearing elements 10.30 increases, said auxiliary valve 70A increases the pressure oil flow in an effort to maintain the oil film thickness constant. As already pointed out above, FIG. In addition to the components shown in Fig. 5, the system needs a braking auxiliary valve. the solution of Fig. 5 is particularly well suited for bearing elements 10,30 whose movement is very limited.

Fig. 6 is a fully schematic diagram of a variation of the solution of the invention. In the system of Fig. 6, a pin 92 is coupled to a load piston 13 of the central main bearing element 10 which controls a valve 90 underneath said bearing element 10. With the load piston 13 in its down position, the pin 92 presses valve 90 to its lowest position. . When the supply pressure ps is connected to the pressure line, the oil flow is directed through the valve 90 from the valve opening 93 under the load pistons 13,33 of the main bearing elements 10,30, whereby the shaft 3 moves upwards. The counter-bearing element 20 is affected by a constant pressure.

A control pressure branch p0 is taken from the pressure line of the supply pressure ps beneath the valve stem 91 so that as the shaft 3 and the middle load piston 13 move upward, the valve stem 91 111026 also rises upwards. Through the central loading piston 13 and the pin 92, the upward movement of the spindle 91 of the valve 90 reduces the opening of the valve 90 and restricts the flow. The flow orifice 93 decreases until equilibrium between the compressive forces and the mass forces (weight of shaft 3 and possible external load 5) is reached. Increasing the load disrupts the equilibrium achieved and opens valve 90 spindle 91, thereby increasing the oil flow below the loading pistons 13.33 and increasing the pressure to compensate for the increase in load. However, a slight adjustment deviation is created, i.e. an increase in the opening of the valve which is not compensated by the rigid system, the compensation operation is based on the occurrence of this adjustment deviation.

10

Figures 7 and 7A further illustrate an alternative embodiment of a pre-steered bearing control system. The bearing control in this embodiment is implemented by a commercially available control valve 120, the "spindle" of which is used in the so-called. A Rego ™ cartridge 121, as well as a pilot control valve 110, which is shown in greater detail in Figure 7A. The position of the load piston 13 of the central main bearing element 10 is measured by the spindle of the pilot control valve 110, the pin 113 of which is connected thereto and moves the spindle 112 of the pilot control valve 110 accordingly. The control valve 120, or so-called. The Rego ™ valve 120 acts as the master valve and is controlled by the pilot control valve 110. The body 111 of the pilot control valve 110 is provided with a 20 elongated opening 114 in the direction of movement of the mandrel 112 to control the pilot flow according to the position of the load piston 13. Below the spool 112 of the pilot valve is a spring 116 which forces the spindle 112 to follow the movement of the load piston 13.

With the loading piston 13 in its lowered position, the opening 114 machined into the body 111 25 of the pilot valve 110 is fully open and the effective pressure in the chamber 122 above the cartridge 121 of the Rego ™ valve 120 is at its minimum value. When the supply pressure is applied to the ps Rego ™ valve 120, the pressure begins to act on the chamber 123 underneath or to the right of the cartridge, thereby raising the cartridge 121 to the upper position and fully opening the valve 120 to supply the ps to the load piston 13.33. The side of the Rego ™ cartridge 121 is known to have an elongated groove 124, and as valve 120 opens, i.e. as the cartridge 121 rises up through that groove 124, oil can flow from lower chamber 123 to upper chamber 12111026 and thereafter through opening 114 to bias valve 110.

As the spindle 3 and thus the spindle of the pilot control valve 110 rises into the body 111 of the pilot control valve 5 110, the surface area of the machined opening 114 begins to decrease, thereby increasing the pressure above the Rego ™ cartridge and closing the Rego ™ valve 120. As the cartridge 121 closes, the flow area of the groove 124 formed therein is reduced, whereby the flow of the pilot circuit from the Rego ™ valve 120 to the pilot valve 110 is also reduced. The spindle 121 of the Rego ™ valve 120, which acts as the master valve, closes until a steady state is reached between the flow through the groove 124 to the pre-control circuit and the flow from the pre-control circuit 110 through the pre-control valve 110.

7, a differential pressure valve 125 is fitted between the Rego ™ valve and the pilot control valve 110 to provide a constant differential pressure across the pilot control member 15, the flow being dependent solely on the position of the spindle 112 on the pilot control valve 110 and not on the supply pressure. The adjustable throttle 126, which is coupled to the pre-control valve 110, allows the position of the shaft 3 to be externally precisely controlled.

Finally, Fig. 8 is a fully schematic representation of the electrohyd; operating principle of roller steering system. With the system of Fig. 8, the shaft 3 of the rolled bearing can be controlled by hydrostatic bearing elements 10,20,30,40, either in a position-based or a force-dependent driving mode. In the case where the counter roll of the plain bearing roller is a movable diaphragm deflection roll 25, the plain bearing roller is controlled in position. In this case ... the position of the bearing shaft 3 remains constant regardless of the load caused by the back roll • · 4. In the case where the counter roll is a fixed-shell deflection compensated roll, the shaft 3 of the plain bearing roll is controlled by force. The pressures of the bearing elements 10.30 are adjusted in relation to the external load. The higher the line load caused by the 30 fixed counter rolls, the greater the bearing pressure is applied to the bearing elements 10.30. The reference pressure for the load is obtained from the main control system performing line pressure calculation.

13 111026

In Fig. 8, the drive side of the roll is denoted by reference D and the treatment side by reference s. In Fig. 8, the position of the roll shaft 3 is measured vertically, i.e. in the main load direction, indirectly connected to the load piston of main bearing elements. The position sensor is a non-contact magnetostrictive sensor 106. The load force, in turn, is measured by measuring the load pressure. The drive side D and the S-side position and the pressure bearing messages are passed to control a erillislogiikkayksikölle 109, which is connected to the main control. Separate logic controls the position and load force of the bearing shaft 10 3. The master control system provides setpoints for stand-alone logic.

The supply pressure PS is supplied to bearing elements 10,30 directional valve, which operates the control system of the main valve 100, which is controlled either by the feed pressure or the tank pressure medicine-15 polymerization guides in the drive side D and the front side S control valves 101. Päälaakerielementtien 10.30 load pistons is led further 102 through use of mechanical varmistuspaineena relief pressure small individual directional valve . The magnitude of the relief pressure pk is controlled by the mechanical pressure relief valve 103 following the directional valve 102. When the relief pressure pk is lower than the pressure set by the control valves 101 20, the relief pressure pk can influence the load pistons. In the event of a failure, the shaft 3 is guided to the support of this mechanically adjusted relief pressure pk. The relief pressure pk can be removed by changing the position of the directional valve 102.

25 In the pressure passage to the bearing elements 10,30, after the control valves 101, there are ... non-return valves 104. which act as lock valves. The operation of these lock valves 104 is controlled by a directional valve 110. The control pressure is applied to the lock valves 104 via hydraulically controlled directional valves 105. In the driving situation, the lock valves 104 have no control pressure and the oil can only flow from the control valve 100 to the load-bearing elements 10.30 and not in the opposite direction. In the quick-release situation, the lock valves are opened by means of directional valves 105. Simultaneously, the control valves 101 connect the controlled pressure connection to the tank line, whereby the bearing shaft 3 14 111026 moves rapidly to its lower position. Upon reaching the lower position, the pins on the bottom of the central main bearing element 10 open the check valves 107,108 located under the load piston and cylinder. The second check valve 107 engages the braking pressure pi under the load pistons 13 while the second check valve switches the pressure under the load piston 5 as the set pressure for the hydraulically controlled directional valves 105 which change their position so as to release the control valves exit through the control valve to the tank line. The shaft 3 remains rotating in the down position to support the braking pressure.

10

A constant pressure pv supporting the shaft 3 is applied to the counter-bearing element 20. This pressure causes a flow through the bearing shoe of the counter-bearing element 20. This flow is controlled by flow switches. If the load caused by the bearing shoes of the main bearing elements 10,30 becomes too high, the flow through the counter bearing element 20 is prevented and the flow switches detect an overload condition.

The pressures of the side bearing elements 40 are controlled in a known manner such that the shaft 3 remains laterally concentric with the bearing housing.

The invention has been described above by way of example in the figure of the accompanying drawing ;;; by reference. However, the invention is not limited to the example shown in the figure, but different embodiments of the invention may vary within the scope of the inventive idea defined in the appended claims.

Claims (16)

15 111026 1. A bearing control system for a hydrostatic-bearing roll of a paper machine, paper finishing device or the like, in particular with a counter roll (2) forming a nip (N) 5, such as a calender roll, press roll or the like, mounted on a bearing (4) ) or a corresponding frame member by means of a plurality of hydraulically-loaded hydrostatic bearing elements (10,20,30,40), of which at least one bearing element, so-called. the main bearing element (10) acts in the main load direction (A-A) and one bearing element, so-called. the counter-bearing element (20) against the 10 main loading directions and the bearing elements are provided with a bearing shoe (16,26,36,41) positioned around the neck of the roll (1) axis (3) to rotatably support the roll (1), such as bearing bracket (4), characterized in that the control system is arranged to supply the pressures of the hydraulic pressure medium to the hydrostatic bearing elements (10,20,30,40) such that each bearing shoe (16,26, 36,41) is freely positioned around the neck of the shaft (3). Control system according to claim 1, characterized in that the control system is arranged to supply the pressures of the hydraulic pressure medium to the hydrostatic bearing elements (10,20,30,40) so that the roll (1) is in the main loading direction (A-A), e.g. movable in the nip plane direction by moving the shaft (3) by said bearing elements (10,20,30,40). Control system according to Claim 1 or 2, characterized in that the control system 25 is adapted to supply to the main bearing elements (10) acting in the main load direction (A-A) the pressure required to close the roll nip (N) and achieve the desired line load on the nip (N). Control system according to one of the preceding claims, characterized in that the control system, when the bearing elements are loaded, is tracked to maintain the flow of lubricant through the bearing elements. 16 111026 Control system according to one of the preceding claims, characterized in that, in the case where the counter roll (2) of the hydrostatic-bearing roll (1) is a fixed-shell roll, the control system is arranged to control the hydrostatic bearing elements (10,20,30) in a controlled manner. 5 i Control system according to one of claims 1 to 4, characterized in that, in the case where the counter roll (2) of the hydrostatic-bearing roll (1) is a movable-shell roll, the control system is arranged to control the hydrostatic bearing elements (10,20,30) so that the hydrostatic bearing The position of the 10 rolls remains constant regardless of the external load applied to it. Control system according to any one of the preceding claims, characterized in that, when the nipple (N) is opened, e.g. in a quick-release situation or the like, the control system is tracked to release pressure from the bearing elements (10,30) acting in the main loading direction. Control system according to claim 7, characterized in that the control system is provided to ensure a lubricant flow to the bearing elements 20 (10,20,30) even when the nip (N) is open. • t l «i Control valve system according to one of the preceding claims, characterized in that the control valve system is arranged to continuously supply a constant pressure pressure medium flow to the counter-bearing elements (20). 25 Control system according to any one of the preceding claims, characterized in that the control system is provided with a pre-control system. Control system according to one of the preceding claims, characterized in that the control system is mechanically hydraulic. 17 111026 A steering rig according to claim 11, characterized in that the rigid steering rig is arranged to determine the position of the shaft (3) of the hydrostatically-mounted roller hydraulically by means of bearing elements. Control system according to one of Claims 1 to 10, characterized in that the control system is electro-hydraulic. Control steering system according to claim 13, characterized in that the steering suspension system is electrically tracked to measure the position of the main bearing element (10) and to adjust the loading pressure based on the measurement data. 15. claimed in claim 13 or 14 ohjausjäijestelmä, characterized in that the ohjausjäijestelmä is adapted to synchronize and to synchronize the roll (1) the drive side (D) and side (S) movements. 15 A steering system according to any one of the preceding claims, characterized in that the steering system is arranged as a constant independent of the load on the roll of the hydrostatic bearing oil film. «<• φ« • ♦ «18 111026