FR3096072A1 - Turbomachine comprising a variable viscance shaft bearing damper - Google Patents

Abstract

Turbomachine (10) comprising a rotary shaft (16), a bearing (18) supporting the rotary shaft (16), a bearing support (20) supporting the bearing (18), and an intermediate ring (24) disposed between the bearing (18) and the bearing bracket (20), the intermediate ring (24) being configured to slide axially relative to the bearing (18), a first seal (36) axially coupled with the bearing (18) and a second seal (38) axially coupled with the intermediate ring (24), the first seal (36), the second seal (38), the bearing (18) and the intermediate ring (24) defining an axial length chamber (40) (L ) variable and forming, when filled with oil, a variable stiffness damper. Figure for the abstract: Fig. 2.

Description

Turbomachine including variable stiffness shaft bearing damper

This presentation concerns a turbomachine comprising a variable viscance shaft bearing damper.

The term "turbomachine" designates all the gas turbine devices producing motive power, among which a distinction is made in particular between turbojet engines providing the thrust necessary for propulsion by reaction to the high-speed ejection of hot gases, and turboshaft engines in which the motive power is provided by the rotation of a motor shaft. For example, turbine engines are used as an engine for helicopters, ships, trains, or even as an industrial engine. Turboprops (turbomotor driving a propeller) are also turboshaft engines used as an aircraft engine.

Different turbomachine shaft bearing dampers are known, for example from EP 2 753 844 or FR 2 629 537. However, these dampers are more or less effective depending on the operating range of the turbomachine. So there is a need for that.

One embodiment relates to a turbomachine extending in an axial direction and in a radial direction, comprising at least one rotary shaft, at least one bearing supporting the rotary shaft, a bearing support supporting the bearing, a spring element mechanically connecting the bearing carrier and the bearing, and an intermediate ring radially disposed between the bearing and the bearing carrier, the intermediate ring being configured to slide axially relative to the bearing, a first seal radially disposed between the bearing and the intermediate ring is coupled axially with the bearing while a second seal disposed radially between the bearing and the intermediate ring is axially coupled with the intermediate ring, the first seal, the second seal, the bearing and the intermediate ring thereby defining a chamber configured to receive damping oil, the axial length of the chamber being variable as a function of the axial position of the intermediate ring with respect to the bearing, thus forming, when filled with oil, a shock absorber with variable stiffness (viscance).

In general, the axial direction corresponds to the direction of the axis of the shaft of the turbomachine, and a radial direction is a direction perpendicular to the axial direction. The azimuthal direction corresponds to the direction describing a ring around the axial direction. The three directions axial, radial and azimuthal correspond respectively to the directions defined by the coast, the radius and the angle in a cylindrical coordinate system. Finally, unless otherwise specified, the adjectives interior and exterior (or internal and external) are used in reference to a radial direction so that the interior part (ie radially interior) of an element is closer to the axis of the shaft than the outer part (ie radially outer) of the same element.

It is understood that the intermediate ring is configured to slide in the axial direction whereby the axial length of the chamber varies. Indeed, the first seal is coupled axially to the bearing while the second seal is coupled axially with the intermediate ring. Thus, by moving the intermediate ring axially with respect to the bearing, the first seal is moved axially with respect to the second seal. Thus, the chamber being delimited radially by the bearing on the one hand and by the intermediate ring on the other hand, and axially by the first seal on the one hand and by the second seal on the other hand, by moving the intermediate ring axially relative to the bearing, the axial length separating the first seal and the second seal, and therefore the axial length of the chamber, is modified.

The inventors have found that in operation, by varying the axial length of the chamber, the volume of oil present in the chamber is varied, and therefore the damping thus provided (viscance). The damping can therefore be adapted to different turbomachine speeds, and therefore make this system effective over a wide range of turbomachine speeds. In other words, the damping devices of the state of the art make it possible to damp only one natural mode of vibration of the shaft while by varying the viscance, it is possible to damp several natural modes. distinct from shaft vibrations.

In some embodiments, the first seal is received in an annular groove made in an outer wall of the bearing while the second seal is received in an annular groove made in an inner wall of the intermediate ring.

Such a configuration makes it possible to ensure reliable and efficient axial coupling of the joints. For example, the first seal also cooperates with a uniform portion of the inner wall of the intermediate ring while the second seal cooperates with a uniform portion of the outer wall of the bearing.

In certain embodiments, the turbomachine comprises an actuator configured to move the intermediate ring axially relative to the bearing.

For example, the actuator comprises a hydraulic cylinder, but any other type of actuator known to those skilled in the art can be envisaged, for example electric, or pneumatic. For example, the actuator is configured to have a response time of less than one second (1.0 s), or even less than half a second (0.5 s), or even less than three tenths of a second (0.3 s). Such a configuration makes it possible to be able to control the axial displacement of the intermediate ring in a satisfactory manner, in particular during transients in the speeds of the turbomachine.

In certain embodiments, the turbomachine comprises at least one radial stop configured to block the intermediate ring radially relative to the bearing in at least one predetermined axial position of the intermediate ring relative to the bearing.

It is therefore understood that the turbomachine may have a single radial thrust bearing, or else several radial thrust bearings. Hereafter, and unless otherwise indicated, the term "the stop" means "the at least one radial stop".

It is understood that in one or more predetermined axial positions of the intermediate ring with respect to the bearing, the stop radially blocks the intermediate ring with respect to the bearing. This makes it possible to inhibit the damping carried out by the damping oil within the chamber, and makes it possible, in certain turbomachine speeds, to further improve the vibratory behavior of the shaft.

In some embodiments, the at least one radial abutment is arranged on an outer wall of the bearing and/or on an inner wall of the intermediate ring.

Such locations make it possible to limit the overall size.

In some embodiments, the bearing carrier and the intermediate ring cooperate via an axially extending sliding connection.

Such a slide connection makes it possible to guide the movement of the intermediate ring axially, which avoids jamming, reduces friction and wear, and improves the response time of the axial movement of the intermediate ring.

In certain embodiments, the sliding link comprises a system of grooves/ribs (or system of splines) and/or a system of ball splines and/or a magnetic bearing.

Such configurations are particularly effective and reliable for forming a sliding connection between the bearing support and the intermediate ring.

In some embodiments, the spring element is a variable stiffness spring element.

The combination of a variable stiffness spring element and the variable stiffness damper makes it possible to be able to damp an even wider range of shaft eigenmodes / a wider range of turbomachine speeds. An example of a spring element with variable stiffness is described in document FR 17 62885.

This presentation also relates to a method for controlling a turbomachine according to any one of the embodiments described in this presentation, in which the axial position of the intermediate ring is adjusted with respect to the bearing as a function of the speed of the turbomachine. and/or vibrations undergone by the turbomachine.

For example, a speed sensor makes it possible to determine the effective speed of the turbomachine while the intermediate ring is positioned according to the determined speed. The excited eigenmodes of the shaft being a function of the speed, it is possible to determine the optimal axial position of the intermediate ring, and therefore the optimal viscance, according to each speed.

According to another example, a vibration sensor makes it possible to determine the vibrations to which the turbomachine is subjected while the intermediate ring is positioned according to the determined vibrations. The measurement of the vibrations at a point of the turbomachine makes it possible to determine the eigenmode excited by the shaft. It is therefore possible to determine the optimum axial position of the intermediate ring, and therefore the optimum viscance, as a function of the vibrations undergone by the turbomachine.

According to yet another example, the axial position of the intermediate ring is adjusted with respect to the bearing according to the speed of the turbomachine and according to the vibrations undergone by the turbomachine.

The object of this presentation and its advantages will be better understood on reading the detailed description given below of various embodiments given by way of non-limiting examples. This description refers to the pages of appended figures, on which:

Figure 1 is a schematic sectional view of a turbomachine,

FIG. 2 represents a bearing of the turbomachine of FIG. 1,

Figure 3 shows a first variant of Figure 2,

Figure 4 shows a second variant of Figure 2, and

FIG. 5 represents different steps of a method for controlling the turbomachine of FIG. 1.

FIG. 1 schematically illustrates a turbomachine 10, in this example a gas turbine engine, and more specifically a turbofan engine. The turbomachine 10 extends along an axial direction X, a radial direction R and a circumferential or azimuthal direction C. The turbomachine 10 comprises a fan 11, a compressor 12, a combustion chamber 13, a first turbine 14, and a second turbine 15, arranged successively in the direction of flow M of the air and the combustion gases therethrough. However, the teaching of this presentation is also applicable to other types of gas turbine engines, such as single-flow turbojets (“turbojets”), turboprops, or turboshafts, as well as to other types of turbomachines and rotating machines.

In the turbine engine 10, the fan 11 is coupled in rotation to the second turbine 15, and the compressor 12 to the first turbine 14 by, respectively, first and second coaxial rotary shafts 16, 17. These coaxial rotary shafts 16, 17 are supported in a casing 19 of the turbine engine 10 by corresponding bearings 18.

The structure of the turbomachine around the bearings 18 is described with reference to Figure 2. This description refers to the structure around only one of the bearings 18, but is applicable to all the bearings 18.

The bearing 18 shown in Figure 2 supports the shaft 16. The bearing 18 is itself supported by a bearing support 20. A spring element 22 mechanically connects the bearing 18 and the bearing support 20. An intermediate ring 24 is arranged radially between the bearing 18 and the bearing support 20.

The bearing 18 is in this example a roller bearing 18C, and comprises an inner ring 18A which supports the shaft 16 and an outer ring 18B. The outer ring 18B is mechanically connected to the bearing support 20 via the spring element 22. In this example, the outer ring 18B, the spring element 22 and the bearing support 20 form one and the same part, but can alternatively each be formed by a separate piece, these separate pieces being assembled to each other by means known elsewhere, for example a flange.

In this example, the spring element 22 is a spring element with variable stiffness. More specifically, in this example, the spring element 22 comprises a flexible cage having a first elastically flexible member 22A extending in the axial direction X. This first member 22A has a first axial end 22A1 connected to the bearing 18, and more specifically to the outer ring 18B of the bearing 18, and a second axial end 22A2. The flexible cage also has a second elastically flexible member 22B extending in the axial direction X. This second member 22B has a first end 22B1 connected to the bearing support 20 and a second end 22B2 connected to the second end 22A2 of the first member 22A . Thus, a radial force F can be transmitted between the bearing 18 and the bearing support 20 by bending the first and second elastically flexible members 22A, 22B.

In order to allow a variation of the radial stiffness of the spring element 22, the latter may comprise a movable element 27 with support points 28, 29 in radial contact with, respectively, the first and second members 22A, 22B elastically flexible. Thus, the radial force F can be transmitted between the first and second members 22A, 22B via this movable element 27, whereby the portions extending between the support points 28, 29 and the second ends 22A2, 22B2 do not are subjected to no effort or relatively little effort. Moving the movable element 27 in the axial direction thus makes it possible to vary the radial stiffness of the spring element 22 by increasing or decreasing the overhangs over which the radial force F is transmitted by bending of the first and second members 22A, 22B between their first ends 22A1, 22B1 and the respective support points 28, 29. The displacement of the mobile element 27 is carried out in this example via an actuator 23, in this example a hydraulic jack, which can for example be supplied via a hydraulic circuit of the turbomachine 10. However, other types of actuators, for example electric or pneumatic, are also possible.

Actuator 23 can be connected to a control unit 60, possibly connected to one or more sensors not shown. The control unit 60, which can in particular be an electronic computer, can be configured or programmed to control the actuator 23 in order to move the mobile element 27 in the axial direction, for example when approaching a speed of rotation of the rotary shaft carried by the bearing 18 having been identified, by calculations and/or preliminary tests, as corresponding to a predetermined resonance mode of the rotary shaft 16, 17, and possibly stored in a memory of the control unit 60.

It is noted that in this example the spring element 22 is a spring element with variable stiffness, but that this element is optional, any other configuration being possible.

The intermediate ring 24 is arranged radially between the bearing support 20 and the bearing 18, more particularly the outer ring 18B of the bearing 18, and is configured to slide axially relative to the bearing 18. The outer wall 25E of the intermediate ring 24 cooperates with the inner wall 21I of the bearing support 20 while the inner wall 25I of the intermediate ring 24 cooperates with the outer wall 19E of the outer ring 18B of the bearing 18. More precisely, the intermediate ring 24 cooperates with the bearing support 20 via a sliding connection 30 extending axially, in this example a system of axial grooves/ribs 32 engaged with each other. It is noted that in this example the internal wall of the bearing support 20 is provided with annular grooves each receiving a seal 34 which cooperate with the external wall of the intermediate ring 24. It is noted that the system of axial grooves/ribs 32 is arranged axially between the two seals 34. The intermediate ring 24 cooperates with the outer ring 18B via a first seal 36 and a second seal 38. The first seal 36 is received in an annular groove 18B1 provided in the outer wall of the bearing 18, whereby it is axially coupled with the bearing 18. The second seal 38 is received in an annular groove 24A formed in the inner wall of the intermediate ring 24, whereby it is axially coupled with the intermediate ring 24. Thus, the first seal 36, the second seal 38, the intermediate ring 24, and more particularly in this example the internal wall of the intermediate ring 24, and the bearing 18, and more particularly ement in this example the outer wall of the outer ring 18B of the bearing 18, form a chamber 40. The intermediate ring 24 being axially movable relative to the bearing 18, the first seal 36 being axially coupled to the bearing 18 while the second seal 38 is coupled to the intermediate ring 24, the axial length L of the chamber 40 varies according to the relative axial position of the intermediate ring 24 with respect to the bearing 18.

Chamber 40 has an oil supply orifice 24C formed in intermediate ring 24. The internal wall of intermediate ring 24 has a concavity 24D forming an oil accumulation cavity. It is noted that an oil accumulation chamber 42 is provided between the intermediate ring 24, the bearing support 20, and the two seals 34. This chamber 42 makes it possible to form an accumulation of oil to supply the chamber 40 via orifice 24C (which in this example fluidically connects chamber 40 and chamber 42). Note that in this example, the rib/groove system 32 is disposed in the chamber 42, whereby it is lubricated. The chamber 42 is supplied with oil by means known elsewhere via the supply port 20A. Thus, in operation, the oil follows the flow represented by the broken lines in figure 2.

The intermediate ring 24 has a radial abutment 24E configured to cooperate radially with an axial end 18D of the bearing 18, in this example the outer ring 18B. In this example, the radial stop 24E projects radially inwards from the inner wall 25I of the intermediate ring 24 while the radial end 18D belongs to the outer ring 18B. The bearing 18 has a radial abutment 18E configured to cooperate radially with an axial end 24F of the intermediate ring 24. In this example the radial abutment 18E projects radially outwards from the outer wall 19E of the outer ring 18B. It is noted that the stop 24E is arranged in the vicinity of an axial end of the intermediate ring 24 opposite the axial end 24F. In this example, when the intermediate ring 24 is in a predetermined axial position with respect to the bearing 18, the stop 24E cooperates with the end 18D while the stop 18E simultaneously cooperates with the end 24F. In other words, in this example, two radial stops radially block the intermediate ring 24 relative to the bearing 18 in a predetermined axial position of the ring 24 relative to the bearing 18. This makes it possible to inhibit the damping effect of the film of oil while maintaining solid contact between the intermediate ring 24 and the bearing 18. In this example, the stops 18E and 24E each have an end inclined with respect to the axial direction, this end being configured to facilitate engagement with the axial end 24F and end 18D respectively.

An actuator 44, in this example a hydraulic cylinder, makes it possible to move the intermediate ring 24 axially with respect to the bearing 18. The actuator 44 is for example powered via a hydraulic circuit of the turbine engine 10. In this example, the actuator 44 comprises a cylinder body 44A coupled to the bearing support 20 (fixation not shown) and a piston 44B coupled to the intermediate ring 24. However, other types of actuators, for example electric or pneumatic, are also possible. In this example, the actuator 44 also serves as an axial retaining element holding the intermediate ring 24 between the bearing 18 and the bearing support 20, whereby the ring 24 cannot disengage axially and remains engaged between the bearing 18 and bearing bracket 24.

Thus, by actuating the actuator 44, the intermediate ring 24 is moved axially with respect to the bearing 18/support 20, whereby the axial length L of the chamber 40 is varied, and therefore the viscance of the damper formed when chamber 40 is filled with oil.

Actuator 44 can be connected to control unit 60 (or another control unit), possibly connected to one or more sensors. The control unit 60 can be configured or programmed to control the actuator 44 in order to move the intermediate ring 24 in the axial direction, for example when approaching a rotational speed of the rotary shaft carried by the bearing 18 having been identified, by calculations and/or prior tests, as corresponding to a predetermined resonance mode of the rotary shaft 16, 17, and possibly stored in a memory of the control unit 60.

In this example, the control unit 60 is connected to two sensors, namely a sensor 62A of vibrations, a sensor 62B of the speed of the turbomachine 10.

FIG. 5 represents a method for controlling the turbine engine 10. In this method, the vibration level to which the turbine engine 10 is subjected as well as the speed of the turbine engine 10 are detected in step E1. In step E2, if the vibratory level is greater than a predetermined limit, then the axial position of the intermediate ring 24 is adjusted in step E3. This adjustment can for example also be a function of the detected speed. According to a variant, the axial position of the intermediate ring is a function solely of the speed sensed or solely a function of the vibratory level sensed.

FIG. 3 represents a first variant of the device represented in FIG. 2, which differs from FIG. 2 only by the sliding connection between the intermediate ring 24 and the bearing support 20. In this example, the sliding connection 30' is arranged in outside the oil accumulation chamber 42, and includes a ball spline system 33.

Figure 4 shows a second variant of the device shown in Figure 2, which differs from Figure 2 only by the sliding connection between the intermediate ring 24 and the bearing support 20. In this example, the sliding connection 30'' is arranged outside the oil accumulation chamber 42, and comprises a magnetic bearing 35. The figures, and in particular FIG. 4, being schematic representations, it is understood that, the intermediate ring 24 moving axially, the outer ring of the bearing magnet 35 can be more or less long axially so as to be able to cover an axial travel extending over all the possible axial positions of the intermediate ring 24.

Although the present invention has been described with reference to specific embodiments, it is obvious that modifications and changes can be made to these examples without departing from the general scope of the invention as defined by the claims. In particular, individual features of the different illustrated/mentioned embodiments can be combined in additional embodiments. Accordingly, the description and the drawings should be considered in an illustrative rather than restrictive sense.

It is also obvious that all the characteristics described with reference to a method can be transposed, alone or in combination, to a device, and conversely, all the characteristics described with reference to a device can be transposed, alone or in combination, to a method.

Claims (9)

Turbomachine (10) extending in an axial direction (X) and in a radial direction (R), comprising at least one rotary shaft (16, 17), at least one bearing (18) supporting the rotary shaft (16, 17), a bearing bracket (20) supporting the bearing (18), a spring element (22) mechanically connecting the bearing bracket (20) and the bearing (18), and an intermediate ring (24) disposed radially between the bearing (18) and the bearing support (20), the intermediate ring (24) being configured to slide axially relative to the bearing (18), a first seal (36) disposed radially between the bearing (18) and the intermediate ring (24) is axially coupled with the bearing (18) while a second seal (38) disposed radially between the bearing (18) and the intermediate ring (24) is axially coupled with the intermediate ring (24), the first seal (36), the second seal (38), the bearing (18) and the intermediate ring (24) thus defining a chamber (40) configured to receive d e damping oil, the axial length (L) of the chamber (40) being variable depending on the axial position of the intermediate ring (40) relative to the bearing (18), thus forming, when it is filled with oil, a variable stiffness damper. Turbomachine (10) according to claim 1, wherein the first seal (36) is received in an annular groove (18B1) formed in an outer wall (19E) of the bearing (18) while the second seal (38) is received in an annular groove (24A) formed in an internal wall (25I) of the intermediate ring (24). Turbomachine (10) according to claim 1 or 2, comprising an actuator (44) configured to axially move the intermediate ring (24) relative to the bearing (18). Turbomachine (10) according to any one of claims 1 to 3, comprising at least one radial stop (24E, 18E) configured to radially block the intermediate ring (24) relative to the bearing (18) in at least one predetermined axial position of the intermediate ring (24) relative to the bearing (18). Turbomachine (10) according to Claim 4, in which the at least one radial stop (24E, 18E) is arranged on an external wall (19E) of the bearing and / or on an internal wall (25I) of the intermediate ring (24) . Turbomachine (10) according to any one of claims 1 to 5, in which the bearing support (20) and the intermediate ring (24) cooperate via a sliding connection (30, 30 ', 30' ') extending axially . Turbomachine (10) according to Claim 6, in which the sliding link comprises a system of grooves / ribs (32) and / or a system of ball splines (33) and / or a magnetic bearing (35). A turbomachine (10) according to any one of claims 1 to 7, wherein the spring element is a variable stiffness spring element (22). A method of controlling the turbomachine (10) according to any one of claims 1 to 8, in which the axial position of the intermediate ring (24) is adjusted (E3) with respect to the bearing (18) as a function of the speed of the turbomachine and / or vibrations undergone by the turbomachine (E1).