RU173697U1 - GAS-TURBINE ENGINE ROTOR SUPPORT - Google Patents

Abstract

The utility model relates to rotary gas turbine machines and can be used in the construction of rotor supports, which allow to regulate axial forces on an angular contact bearing. The use of the rotor support allows you to adjust the axial forces on one of the rotor bearings with a radial roller bearing due to the hydrodynamic forces of the radial oil flow passing at a certain speed in the gap between the ring associated with the stator and the end of the bearing inner ring, and allows to increase the life of the angular contact a bearing located on another support by reducing the axial force acting on it without changing the traction characteristics of the engine. The axial force compensation method implemented by the support structure greatly simplifies the installation of the support, does not require additional energy costs and complex structural solutions, but uses the kinetic energy of the oil used to lubricate and cool the bearings, and eliminates the intake of air from the gas-dynamic path and seals between the rotating parts of the rotor and a stator with a pneumatic or hydraulic force compensation approach, which increases the efficiency of the engine. 3 ill.

Description

The utility model relates to the field of mechanical engineering and can be used in the construction of rotor supports of aircraft gas turbine engines (GTE).

It is known from operating experience that during the operation of a gas turbine engine, axial forces due to gas-dynamic forces act on their rotor, namely, the axial force (Fc) directed along the flight and the axial force of the turbine (Ft) directed against the flight act on the compressor rotor.

The axial force acting on the rotor bearings (sometimes its value is very high) limits the resource of the thrust bearings of the rotor bearings.

The solution to this problem by creating a thrust bearing that works reliably under significant load conditions is very difficult, especially for rotors operating at high speeds. This is due to the large centrifugal forces acting on the rolling elements, which limit the use of bearings with large dimensions and higher load capacity.

To solve the problem of increasing the service life of the bearing supports of gas turbine rotors caused by the action of axial forces on them, a rather wide range of solutions is currently used.

One of the ways to reduce the axial force acting on the rotor bearing of the bearing is to connect the compressor rotor shafts and the turbine to each other, and since the axial forces acting on them are directed in opposite directions, they partially compensate each other. The resulting axial force (Fp), usually directed along the flight, is perceived by the thrust bearing of one of the rotors, which is usually used as a radial-thrust ball bearing.

Another way to reduce the axial force acting on the rotor bearing of the bearing is to bring the compressed air axial compressor from the last or one of its last stages to the front end wall of the rotor, as well as isolate the rear end of the compressor from the air compressed in the compressor and bring it to this end atmospheric pressure, as a result of which a force is created acting on the front end of the rotor directed against the flight and the flight force acting on the rear end of the rotor is reduced. This leads to a decrease in the action of axial force on an angular contact bearing and an increase in engine life (see G. Skubachevsky, “Aircraft gas turbine engines. Design and calculation of parts”, Moscow: Mashinostroyenie, 1973, 36 pp.).

As a result of the analysis of the known solution, it should be noted that the selection of air from the compressor flow path leads to a decrease in its efficiency. In addition, a decrease in axial force at the rear end of the compressor, directed in flight, reduces the thrust of the gas turbine engine and the speed characteristics of the aircraft.

It is known to implement a method for controlling axial forces on an angular contact bearing of a turbomachine rotor, the use of which allows one to reduce the effect of axial force on the angular contact bearing of the front part of the composite rotor of the turbomachine by redistributing the excess force on the angular contact bearing of the rear part of the rotor. For this, the rotor is made integral of the front and rear autonomous parts. Each part of the rotor is installed in the stator on angular contact and radial bearings. The rotor parts are connected to each other with the possibility of relative axial movement. The outer rings of the bearings are located in the stator, and the inner rings are fixed to the shafts of the rotor parts. To control the axial forces in the design of the device, there are strain rings connected electrically to the control unit. Strain rings are located between the ends of the outer rings of the angular contact bearings of both parts of the rotor and the stator. An elastic ring is additionally installed between the tensor ring of the angular contact bearing of the front of the rotor and the stator. The axial forces of the rotor are balanced by a power drive with a rod rigidly connected to the inner ring of the angular contact bearing of the rear part of the rotor (see RF patent No. 2474710, class F02C 7/00, 2011).

This device allows you to unload and balance according to a given law the axial forces acting on the angular contact bearings of the rear and front parts of the rotor. However, this device for regulating axial forces on an angular contact bearing is extremely difficult to construct, has a significant mass and dimensions, and is applicable only for the bench version, which limits the scope of its use.

A known axial force control system on a thrust bearing of a gas turbine rotor, comprising an electromagnetic device that provides axial load compensation on a thrust bearing, a control unit for this device, as well as a sensor measuring the speed of rotation of the thrust bearing cage. During the operation of the system, the value of the speed of the bearing cage is controlled and, if the ratio of the speed of the cage to the speed of the turbine rotor falls below the set value, the sensor gives the corresponding signal to the control unit, according to which the electromagnetic device creates a certain additional load on the thrust bearing (see US patent No. 5735666, NKI 415/34, 1998).

As a result of the analysis of the known system, it should be noted that its design allows you to adjust the axial force on the thrust bearing of the rotor and provide the resulting load on it within a given range. Using the system allows you to abandon the mechanical and radial seals used in the installation of hydraulic and pneumatic devices. However, the implementation of this system is very difficult in installation and operation, the presence of an electromagnetic device, its drive, as well as a unit for controlling the operation of the device significantly complicate the design of the system and increase the weight of the engine. It is also very significant that when using a measuring scheme in the system - according to the speed of the separator - the slip of the rolling elements in the bearing of the rotor support is not taken into account and, as a result, the accuracy of measuring the speed of the separator is not high, which does not allow fully compensating the axial force acting on the bearing.

A known automatic device for balancing axial forces acting on the support of the rotor shaft during the operation of the gas turbine engine, containing a hydraulic damper, including a source of fluid pressure (oil) and an annular chamber connected to the source. The GTE case and its rotor are interconnected through an annular chamber coaxial to the rotor. The inner ring of the bearing is mounted on the rotor, and the outer ring is located in the GTE housing. The device also includes a load cell mounted on the outer ring of the bearing, a signal amplifier of the load cell and a control element. The output of the strain gauge through a signal amplifier and a control element is connected to a fluid pressure source. When the axial load on the bearing is higher than the specified one, a signal is supplied from the strain gauge through the amplifier and control element to turn on the fluid pressure source. From the source, liquid under pressure is supplied to the annular chamber and creates an opposing load on the rotor, which unloads the bearing (see US patent No. 4578018, class F01D 3/04, 1986) - the closest analogue.

As a result of the analysis of the known device, it should be noted that its use provides, by balancing the axial forces, an increase in the service life of the bearings of the support. However, equipping the rotor support with such a device greatly complicates its design, requires additional oil pumps and, in addition, significantly increases the weight of the engine. The presence of a chamber in contact with a fixed stator and with a rotating rotor requires seals in contact with high-speed engine parts, which lead to oil leaks.

The technical result of this useful is the significant simplification of the design of the support due to the elimination of additional energy sources, control devices and actuators during its operation, as well as due to the rejection of the use of seals in contact with high-speed engine parts.

The specified technical result is ensured by the fact that in the rotor support of the gas turbine engine, comprising a housing in which a rolling bearing is mounted with an outer ring, the inner ring of which is designed to mount the hollow shaft of the rotor, on which the oil-supplying elements of the jet type are connected, connected by hydraulic channels with a cavity of the shaft, new is that the shaft is mounted in the inner ring of the bearing by means of an intermediate sleeve, an intermediate sleeve and the inner ring of the bearing is pressed from one end by an elastic element, and a ring is attached to the other end of the bearing, having an annular flange in the hole zone, the inner surface of which has a conical shape with expansion towards the bearing, the conical flanging surface, the outer surface of the shaft and the end face of the intermediate sleeve form an axial oil cavity, the outlet of the oil-supplying elements is located in the zone of the axial oil cavity, and between the ends of the ring and the inner ring of the bearings facing each other Single annular groove formed oil gap, whose width decreases from the bearing axis to the periphery, a circumferential groove connected to the transition formed between the intermediate sleeve and the surfaces of the oil rings with axial cavity.

The novelty of the claimed solution, which provides the specified technical result, lies in the fact that the support structure is simplified due to minor improvements to its standard bearing lubrication system, namely, the introduction of an intermediate sleeve and a flange ring into the support structure, the presence of which causes the formation of hydrodynamic forces of radial the flow of oil passing at a certain speed, which he acquired due to centrifugal forces when the flow of oil in the peripheral direction the appearance on the conical surface of the annular flanging of the ring, through a narrowing gap between the ends of the ring and the inner ring of the bearing. As a result of this, hydrodynamic forces are transmitted on the one hand to the inner ring of the bearing against flight and unload the angular contact bearing, and on the other hand, the force acting on the ring connected through the bearing housing to the stator transmits axial force in the direction of flight to the stator. The outgoing stream of oil after passing the gap between the ring and the end face of the inner ring of the bearing (Fig. 1) falls on a wide oil separator (see RF patent No. 2038513, CL F04D 29/04, 1995) and is directed to the end of the bearing for lubrication and cooling of rubbing parts. Thus, the oil flows necessary to create hydrodynamic axial forces on the rotor support are used to lubricate and cool the bearing, which increases the efficiency of GTE operation and does not require additional devices and energy costs.

The essence of the utility model is illustrated by graphic materials on which:

in FIG. 1 - support rotor GTE, axial section;

in FIG. 2 - node C of FIG. one;

in FIG. 3 is a view A of FIG. 1 (parts referred to at 7, 4, 10 removed).

The support of the GTE rotor (Fig. 1) contains a bearing with an outer ring 1, an inner ring 2, and rolling elements 3 (for example, rollers) assembled in a cage 4.

The outer ring 1 of the bearing is installed in the housing 5 of the support, and the inner ring 2 is intended for mounting the hollow shaft 6 of the rotor.

Ring 7 fixed to one of the ends of the bearing is fixed in the housing 5, having an annular flange made in the area of the hole. The inner surface of the flanging has the shape of a cone, expanding in the direction "to the bearing" at an angle γ. At the outer end of the flanging there is an annular protrusion (not indicated by the position), partially covering the hole and designed to eliminate oil leakage from the inner conical flanging surface in the opposite direction from the bearing.

The rotor shaft 6 is mounted in the inner ring 2 of the bearing by means of an intermediate sleeve 8.

From the side of the bearing, opposite the ring 7, to the ends of the inner ring 2 of the bearing and the sleeve 8 is pressed elastic element 9, for example, in the form of a spring ring or an elastic membrane. The elastic element 9 is pressed against the ends of the inner ring 2 of the bearing and the sleeve 8 by means of a locking element 10 mounted on the shaft 6, for example, a fixing nut.

The outer ring 1 of the bearing and the ring 7 are fixed in the housing 5 by the locking elements 10.

To supply oil to the bearing, oil-supplying elements mounted on the shaft 6 are used — an oil manifold 11, for example, of an inkjet type, equipped with nozzles (jets) 14 connected through channels in the collector “K” and the internal shaft “M” with an annular groove 12, to which the oil supplied to the inner cavity of the shaft 6. The number of nozzles (jets) can be different and depends on the necessary uniformity of the oil supply to the bearing in the circumferential direction and, in principle, the more there are, the more uniform the oil supply.

The outlet openings of the nozzles (jets) 14 are located in the flanging zone of the ring 7, are located in one plane normal to the axis of the bearing and at an angle to the axis of the radial channels "K" and "M" (preferably at 90 °). The inner conical surface of the flanging of the ring 7, the outer surface of the shaft 6 and the end face of the sleeve 8 form an axial oil cavity.

Between the end surface of the ring 7, facing the bearing and the end of the bearing ring 2, a radial annular clearance "F" is formed, which is connected with an axial oil cavity by a transitional curvilinear radius R of the channel "P" formed between the sleeve 8 and the ring 7.

Since hydrodynamic forces arise only in the channels when the fluid flows between the plates having a narrowing, the radial annular gap "F" has a variable cross section, namely, its width gradually decreases (b> e) from the bearing axis towards the periphery. The ends of the ring 7 and the intermediate sleeve 8 facing each other are rounded and form a transitional annular groove "P" connecting the annular oil cavity with a radial clearance.

On the end surface of the ring 7, along the periphery of the annular gap "F" there are several (for example, three at an angle of 120 ° relative to each other) protrusions "D" in contact with the end face of the bearing ring 2. The presence of the protrusions "D" eliminates the possibility of overlapping the gap "F" due to contact with each other of the surfaces of the ring 7 and the inner ring 2 of the bearing, forming a gap "F".

The support of the rotor of the gas turbine engine works as follows.

In the process of operation of the gas turbine engine, its shaft 6 is driven into rotation, therefore, the sleeve 8 and the bearing inner ring 2 attached to it are rotated. The oil from the oil system of the gas turbine engine is fed into the cavity of the shaft 6, into the annular groove 12 and through the channels in the collector "K" and the internal shaft "M" enters the nozzles 14, which rotate with speed together with the shaft 6, and, under pressure , with speed V is directed to the conical flanging surface "J". The oil flows in this case have a circumferential direction and under the action of centrifugal forces and the inclination of the cone of the inner flanging surface, they receive a direction “to the bearing” along its axis and are formed into a single stream having the shape of a ring in the circumferential direction.

From the axial oil cavity, the oil flows pressed by centrifugal forces to the annular conical surface of the flanging "J" are directed into the transition curved channel "P" of radius R, change direction to radial and fall into the radial clearance "F" (Fig. 2) between ring 7 and the inner ring 2 of the bearing. In order to guarantee that the oil flow enters the curved channel and to eliminate the resistance when oil enters the end of the intermediate sleeve, it is desirable that the input clearance "a" be larger than the output clearance "b" and not less than 1 mm, and to reduce the resistance of the oil flow when changing direction, it is desirable that the radius of curvature is not less than R = 2 mm.

In addition, the narrowing of the curvilinear channel should provide the thickness of the oil flow at the outlet to values of the order of 0.005–0.1 mm required for entering the radial slotted channel “F” and the emergence of sufficient hydrodynamic forces. The smaller the wedge gap between the plates during the flow of oil, the greater the hydrodynamic forces (see Ivanov MN "Machine Details", Moscow, "Higher School", 1991, 315 S.).

When the oil flow moves at a certain speed in the radial direction between the two plates, narrowing the gap will cause hydrodynamic forces (See Nazarenko Yu.B. Liquid friction in bearings and the influence of hydrodynamic forces on the contact stresses of rolling bodies // Engine. - M. - 2015 , No. 2. - S. 10-11).

The hydrodynamic forces during the movement of the oil flow in the radial clearance will be transmitted to the inner ring 2 of the bearing, rigidly connected by means of the retaining ring 10 with the shaft 6, in the opposite direction to the flight, and will reduce the resulting axial force directed in flight and acting on the angular contact bearing located on one of the compressor rotor supports. A decrease in axial force on a radially thrust bearing will improve its working conditions and increase its service life. But on the other hand, the support of the rotor with an angular contact bearing will transmit less axial force to the stator of the engine, which reduces the traction characteristics of the engine.

However, since the hydrodynamic force in the gap "F" acts on both surfaces forming it, exactly the same force will act on ring 7, which is rigidly connected to the stator and the force will be transmitted to the stator. The decrease in axial force at the front rotor support will be compensated at the rear support with roller bearing. As a result, the engine thrust will not change when the axial force decreases on the angular contact bearing.

In the process of operation of the gas turbine engine, the rotor shaft 6 and the stator 5 have different temperatures and their temperature elongations will be different. Typically, the temperature of the stator, especially on the rear support, where the combustion chamber is located, has the highest temperature. In this case, the elongation of the stator and axial movement of the outer ring 1 of the roller bearing will be greater than the inner 2 associated with the rotor shaft.

The offset of the outer ring 1 and the ring 7 will also move the inner ring of the bearing 2 together with the sleeve 8, which will be implemented due to the elastic element 9, which is able to deform and realize axial displacement.

The proposed design of the support allows you to reduce the axial force on the angular contact bearing of another rotor support without changing the engine thrust, which, in addition to the experiments, is confirmed by the calculations by the method (see RF patent No. 159639, CL F16C 19/22 and F02C 7 / 06, 2015).

The performance of the device was carried out on a support model, with a bearing simulation with overall dimensions of 95 × 145 × 24 mm with an inner ring height of 6.5 mm. The inner radius of the shaft 6 is equal to Rb = 32 mm, and its thickness t is 10 mm. The shaft rotation frequency is n = 10200 rpm, and the circular shaft rotation speed is ω = 1068 rad / s.

The thickness of the sleeve 8 is 5.5 mm. The width "B" of the linear portion of the conical flanging surface of the ring 7 is 40 mm with a taper γ = 0.0375 rad. The largest radius of the inclined surface of the conical part of the flanging of the ring is R ₂ = 48 mm, and the smallest - R ₁ = 46.5 mm. The radius of curvature of the curved channel R = 2 mm

For a length L = 4.5 mm, the minimum and maximum values of the radial annular gap "F" are respectively e = 10 μm and b = 20 μm.

The peripheral velocity of the flow of oil from the nozzle (nozzles) is the sum of the peripheral speed of rotation of the shaft on which the manifold is located and the velocity of the flow of oil from the nozzle.

The peripheral velocity of the output nozzle of the collector is

V = ω⋅ (R _B + t + λ) = 1068⋅ (0.032 + 0.01 + 0.0035) = 48.6 m / s,

where Rв - the inner radius of the shaft,

Rv = 0.032 m; t is the thickness of the shaft, t = 0.01 m; λ is the radial distance from the outer surface of the shaft to the output nozzle λ = 0,0035 m

To determine the rate of oil outflow from the nozzle, set the oil pressure in the manifold at the nozzle level.

The oil pressure in the nozzle at the level of the outlet nozzle is determined by the centrifugal force of the mass of oil in the channels "K" and "M", as well as the height λ and the shaft thickness t, and it is equal to

where m is the mass of the oil column, m = (λ + t) ⋅S⋅ρ = 0.0135⋅S⋅ρ; S is the cross section of the nozzle and shaft channels; R _C is the radius of the center of mass of the oil column, R _C = R _B + (λ + t) / 2 = 0.03875 m; ρ - oil density, 800 kg / m ³ .

The rate of oil flow through the nozzle for an incompressible fluid and with the cross-sectional area of the oil channel for supplying oil to the bearing, which is much larger than the area of the nozzle, is determined from the Bernoulli dependence

The resulting oil flow velocity in the circumferential direction will be 83.2 m / s.

The oil flow rate at the entrance to the gap between the roller and the outer ring will be

where R ₂ is the largest radius of the inclined surface of the conical part of the flanging, R ₂ = 0,048 m; B is the axial length of the conical flanging surface, B = 0.04 m.

In the absence of the movement of two plates in the radial direction and the movement of only the oil flow between the plates, the hydrodynamic pressure in the middle of the plates in the radial direction is determined from the expression

where h ₁ - the initial value of the gap 20⋅10 ^-6 m; h _cp is the gap in the middle of the plate 15⋅10 ^-6 m; V is the oil flow rate, V = 21 m / s; μ is the dynamic viscosity of the oil at a temperature of 100 ° C, μ = 0.0027 Ns / m ^{2 '} ; β is the angle of inclination of the plates of the channel "F"; β = (b-e) / L = (0.02-0.01) / 4.5 = 0.0022 rad; L is the length of the radial annular gap, L = 4.5 mm; e is the minimum value of the channel gap, e = 0.01 mm; b is the maximum gap value of the channel "F", b = 0.02 mm;

When the annular surface area of each of the contacting plates is S = 0.00146 m ² , the decrease in axial force on the angular contact bearing will be

F _O = P⋅S = 6⋅10 ⁶ ⋅0.00146 = 8760H.

Structurally, the bearing support is very simple, its installation is also simple. It is very important that the proposed design solution allows you to abandon the mechanical and radial seals used in the regulation of axial forces using hydraulic and pneumatic devices,

Using this solution also allows you to increase the bearing life of the bearings by reducing the axial force acting on it, without changing the traction characteristics of the engine.

Claims (1)

A rotor support of a gas turbine engine, comprising a housing capable of being coupled to the engine stator, in which a rolling bearing is mounted with an outer ring, the inner ring of which is designed to mount a hollow rotor shaft on which jet type oil-supplying elements are connected, connected by hydraulic channels to the shaft cavity, characterized in that the shaft is mounted in the inner ring of the bearing by means of an intermediate sleeve, the intermediate sleeve and the inner ring of the bearing are elastic m with an element, and a ring is attached to the other end of the bearing, having an annular flange in the hole zone, the inner surface of which has a conical shape with expansion towards the bearing, the conical flanging surface, the outer surface of the shaft and the end face of the intermediate sleeve form an axial oil cavity, the output of the oil supply elements is placed in the zone of the axial oil cavity, and between the ends of the ring and the inner ring of the bearing facing each other, a radial annular oil gap is formed, the width decreases from the bearing axis to the periphery, connected by a transitional annular groove formed between the surfaces of the intermediate sleeve and the ring with an axial oil cavity.

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