RU159639U1 - DEVICE FOR SUPPLYING OIL TO A BEARING OF A GAS-TURBINE ENGINE ROTOR SUPPORT - Google Patents

Abstract

1. A device for supplying oil to a bearing of a rotor support of a gas turbine engine, comprising nozzles with output nozzles mounted on the stator parts of the engine, having the ability to connect to the engine oil line, and an oil catch visor designed to direct oil flows from nozzle nozzles to the bearing, characterized in that the oil catch visor is made in the form of a ring mounted on a support stator at the end of the outer bearing ring, the inner surface of the ring has a conical shape with in the direction of the bearing, there is an annular protrusion on the periphery of the ring opening on the outside of the visor, and the nozzles are installed in such a way that their nozzles are directed to the conical surface of the ring. 2. The device according to claim 1, characterized in that the taper angle (γ) of the conical surface of the ring is in the range of: direction; m is the radial clearance between the treadmill of the outer ring and the separator; e is the bevel width in the axial direction on the separator; f is the axial distance from the end of the outer ring of the bearing to the end of the roller; c is the thickness in the radial direction of the outer ring of the bearing; b is the width of the visor without a protrusion.

Description

The invention relates to devices for supplying oil to bearings of bearings of rotary gas turbine machines and can be used to lubricate and cool bearings, as well as to reduce contact stresses on their rolling bodies.

It is known that one of the important problems of reliable operation of gas turbine engines is the proper condition of the bearings of the rotor shaft bearings, which directly depends on the lubrication and cooling conditions of the bearings, as well as on the contact loads between the rolling bodies and rings during engine operation.

It is known from the prior art that in the process of supplying oil to the bearing support at certain feed rates and flow directions of the oil flow, an oil wedge forms between the bearing rings and rolling elements, which reduces stresses between the bearing rings and rolling bodies, similarly to sliding bearings where they are completely eliminated and there is no contact between the moving and stationary ring.

Using this effect allows, without deterioration of the lubrication and cooling conditions of bearings, to significantly increase their service life, especially for high-speed gas turbine engines, in which the contact stresses on the rollers in contact with the outer ring increase significantly compared to the contact on the inner ring due to centrifugal forces rollers, which largely determines the durability of the bearing.

Nozzle devices for lubricating and cooling bearings, in particular, gas turbine engines, containing jet-type oil-injecting nozzles mounted on gas turbine engine units in such a way that their nozzles are directed to the bearing end face and connected to the gas turbine engine oil system, are widely known from the prior art. the nozzle cross-section has an estimated size in accordance with the specified requirements for lubrication conditions (see, for example, Shtoda A.V. et al. “Design of aircraft engines”, ed. VVIA, 1959, p. 114-117).

As a result of the analysis of the known device, it should be noted that during its operation, oil is sprayed through the nozzle nozzles at the end of the bearing simultaneously by several nozzles, which does not allow for guaranteed delivery of oil to all contacting parts of the bearing parts. In addition, part of the oil due to splashing when reflected from the ends of the rollers and the bearing cage does not enter the bearing at all. The use of such devices does not provide effective lubrication and cooling of bearings of bearings.

A device for lubricating and cooling the shaft bearing of a gas turbine engine is known, comprising an oil nozzle and mounted on a shaft pin between the bearing and the nozzle, an oil collecting annular visor forming an oil cavity communicating with the shaft of the bearing through channels and holes made respectively in the pin and the inner race bearing, and the visor is equipped on the inside with radial blades located on the side of the nozzle (see RF patent No. 1130014, CL F02C 7/06, 2004).

With this oil supply scheme, it, compared to that used in the solution given above, is more efficiently delivered to the contacting parts of the bearing due to the action of centrifugal forces, since the oil is fed through the channels on the shaft and through the holes in the inner ring and additionally receives a peripheral speed from rotation of the shaft, which allows more efficient cooling of the contacting surfaces of the bearing.

The disadvantage of this solution is that the oil flow is supplied discretely through the channels in the bearing ring and does not create a continuous flow under all the rollers, which does not provide uniform cooling of the bearing. In addition, oil is fed into the gap between the rolling bodies and the inner ring and does not solve the problem of reducing contact stresses on the outer ring

A device for lubricating the rotor shaft bearing of a gas turbine engine is known, consisting of an oil catch ring fixed with a pin in a predetermined position relative to the rotor shaft, and oil supply means including an oil dispensing groove, an oil supply hole and a flow nozzle.

The bearing of the rotor shaft has a cut inner ring, consisting of two half-holes, at the ends of which oil-dispensing holes are made, connected to the supply cavity, which is connected to the storage cavity by means of oil supply channels. The device contains oil-catching visors, made as a single unit with the oil-collecting ring and profiled so that the spraying of the oil stream is minimal. The pickup holes of the visors are paired with the receiving holes in such a way that the axis of each pickup hole is parallel to the tangent to the inner surface of the receiving hole mated to it. Each receiving hole through a local sampling is communicated with a storage cavity.

During the operation of the device, oil from the oil system of the engine enters from the oil dispensing groove into the oil supply hole, and then is ejected through the nozzle into the oil cavity. An oil catch ring with visors rotates towards the movement of the jet. A grown jet in 1/2 turn of the oil catching ring begins to be caught by the hole until its absorption is completed, after which a slight spraying of the jet occurs. Then, in the area before the approach of the diametrically located catching hole, an oil jet grows, which is absorbed by the second catching hole. The absorbed oil, through receiving wells and local samples having a conical surface to create a centrifugal pressure, enters the accumulation cavity, from where it enters the consumable cavity through the oil-supplying channels. From the discharge cavity through the oil dispensing holes, the oil washes the working surface of the inner ring, the balls of the bearing and under the influence of inertial forces falls on the treadmill of the outer ring, washes it, cools and flows out through the gap between the separator and the outer ring of the bearing (see RF patent No. 2349776, class F02C 7/06, 2009) is the closest analogue.

As a result of the analysis of the known solution, it should be noted that with this oil supply scheme, it is effectively and without loss delivered to the rubbing parts of the bearing, because the visors eliminate splashing and oil loss. However, the disadvantage of this device is that the oil under the rolling elements and the inner ring is supplied discretely through the channels, which does not allow for uniform cooling of the bearing and reduces the influence of the hydrodynamic effect on the contact stresses of the rollers. In addition, the oil is fed into the gap between the rolling bodies and the inner ring and does not solve the problem of reducing contact stresses on the outer ring, where the contact stresses due to the centrifugal forces of the rolling bodies are much larger than on the inner ring. It should also be noted that the finalization of the bearing in terms of performing in its ring of slots and grooves reduces its bearing capacity. The above reduces the life of the bearings of the bearings.

The technical result of this utility model is to increase the life of bearings of the bearing bearings of the rotor shafts of gas turbine engines by creating optimal conditions for their lubrication and cooling, as well as by reducing contact stresses between the bearing parts during engine operation.

, толщина козырька в радиальном направлении, примыкающего к внешнему кольцу подшипника составляет

, а противоположного

, где γ - угол наклона внутренней поверхности козырька в осевом направлении; m - зазор в радиальном направлении между беговой дорожкой внешнего кольца и сепаратором; e - ширина фаски в осевом направлении на сепараторе; f - расстояние в осевом направлении от торца внешнего кольца подшипника до торца ролика; c - толщина в радиальном направлении внешнего кольца подшипника; b - ширина козырька без выступа.The specified technical result is ensured by the fact that in the device for supplying oil to the bearings of the rotor support of the gas turbine engine, comprising, mounted on the oil supply lines of the engine, mounted on the stator parts of the engine, nozzles with output nozzles having the ability to connect to the engine oil line, and an oil catch visor, designed to direct the flow of oil from the nozzles of the nozzles to the bearing, the new is that the oil catch visor is made in the form of a ring, fixed On the stator of the bearing support at the end of the outer ring of the bearing, the inner surface of the ring has a conical shape with an inclination towards the bearing, there is an annular protrusion on the periphery of the ring opening on the outside of the visor, and the nozzles are mounted so that their nozzles are directed to the conical surface of the ring, while the angle of inclination of the inner surface of the visor in the axial direction is within:

, the thickness of the visor in the radial direction adjacent to the outer ring of the bearing is

, but the opposite

where γ is the angle of inclination of the inner surface of the visor in the axial direction; m is the radial clearance between the treadmill of the outer ring and the separator; e is the bevel width in the axial direction on the separator; f is the axial distance from the end of the outer ring of the bearing to the end of the roller; c is the thickness in the radial direction of the outer ring of the bearing; b is the width of the visor without a protrusion.

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

- in FIG. 1 - a device for supplying oil to a bearing of a rotor support of a gas turbine engine in a section;

- in FIG. 2 is a view A of FIG. 1 (stator and visor removed);

- in FIG. 3 is a section B-B of FIG. 2;

- in FIG. 4 is a design diagram for determining the pressure during oil flow in the gap between the rolling body and the stationary outer bearing ring;

For mounting the rotor shaft of a gas turbine engine, as a rule, they use a radial roller bearing containing the outer (outer) 1 and inner 2 rings, between which rolling elements (rollers) 4 are located in the separator 3.

The installation of the bearing is as follows. The outer ring 1 of the bearing is fixed on the stator 6 (that is, it is stationary), and the shaft 5 of the rotor of the gas turbine engine is mounted in the inner ring 2 of the bearing (that is, the inner ring of the bearing rotates together with the shaft). In the process of operation of the bearing unit, the cage 3 is centered on the inner ring 2 of the bearing and the gap between the cage and the inner ring is minimal, while with the outer ring 1 the gap m is much larger.

To supply oil to the bearing, oil nozzles 7 are used, for example, of the jet type, equipped with nozzles (jets) 8. The number of nozzles may be different. So in FIG. 3 shows three nozzles, which does not mean that their number cannot be different. The number of nozzles depends on the necessary uniformity of oil supply to the bearing and, in principle, the more there are, the more uniform the oil supply.

The nozzles 7 are connected by pipelines of the oil manifold to the oil line of the engine and mounted on the stator parts of the engine.

On the stator 6 of the bearing support at the end of the bearing fixed oil catch visor 9, having the shape of a ring. The inner annular surface “A” of the visor is conical in shape, with an inclination towards the bearing, reducing the thickness of the visor, at an angle γ. At the outer end of the visor in the area of its hole there is an annular protrusion 10, designed to eliminate leakage of oil from the visor in the opposite direction from the bearing.

When installing nozzles, they are installed so that their nozzles are directed to the conical surface of the ring. It is most preferable to mount the nozzles in such a way that they are located in the plane normal to the bearing axis, and the nozzle axis of each nozzle is parallel to the tangent to the conical surface of the ring at a point located on a radial line coinciding with the exit section of the nozzle. It is also preferred that the nozzles are deployed by nozzles in the direction of rotation of the inner race of the bearing.

The claimed device operates as follows.

In the process of operation of a gas turbine engine, its shaft is driven into rotation, therefore, is driven into rotation and the inner bearing ring 2 attached to it. Outer ring 1 remains stationary. In this case, the rollers 4 together with the separator 3 perform a rotational movement, providing minimal friction between the rings — rolling friction.

Oil from the engine oil system is fed to the nozzles 7 and, under nozzles 8, is supplied with pressure at a speed V to the conical surface A of the oil catching visor 9. Reflecting from the conical surface, the oil flows onto the end of the bearing, lubricating its contacting parts and cooling them. Given that the oil flows continuously and across the entire end surface of the bearing, a guaranteed oil penetration between the contacting surfaces of all its parts is ensured. It is also very important that the ingress of oil on almost the entire end of the bearing contributes to its uniform cooling. The presence of an annular protrusion 10 on the visor prevents leakage and loss of oil, which ensures its more efficient use. All this increases the bearing life.

To further increase the bearing life due to the creation of an oil wedge between its contacting elements and the creation of hydrodynamic forces, the direction of oil flows reflected from the conical surface of the ring 9 should be organized so that a part of the oil is guaranteed to fall into the end gap between the rollers and the outer bearing ring.

To ensure this condition, it is necessary that the oil catching visor has the appropriate parameters, namely: the conic angle of the conical surface “A”; thickness of the ends of the visor; the width of the visor must be “linked” with the parameters of the bearing, such as the thickness of its outer ring, the distance from the end of the bearing to the ends of the rollers.

Naturally, these parameters can be obtained both by calculation and experimentally.

When determining the parameters by calculation, the thickness of the visor 9 at the end adjacent to the end of the bearing “k” should be equal to

,

,

and the thickness “a” of the opposite end is

,

,

where γ is the angle of inclination of the visor in the axial direction, rad; b - visor width without protrusion, m; f is the distance in the axial direction from the end of the outer ring of the bearing to the end of the roller, m; c is the thickness of the outer ring of the bearing, m

Here, the relation was accepted at small angles of inclination of the visor sinγ = γ.

Since the axial velocity of the oil flow under the action of centrifugal forces during rotation of the oil flow in the circumferential direction with the oil feed rate from the nozzle V is determined by the slope of the visor, its angle should be maximum. However, to eliminate the reflection of the oil flow from the separator, the angle of inclination should be limited to:

,

,

where m is the width of the gap in the radial direction between the outer ring and the bearing cage, m; e - bevel width on the separator, m

With a larger value of the angle γ, the oil flow will be reflected from the separator and the effect of hydrodynamic forces will weaken.

If the minimum angle is close to zero, the axial flow velocity will also be zero and there will be no hydrodynamic force effect. We take the minimum peak angle equal to γ _min = 0.05 rad., At which the effect of hydrodynamic forces, which is of practical importance, will be realized. Visor angle must be in the range

.

.

As studies have shown, the above steps, providing oil from the nozzle with a speed V equal to the speed of the rollers in the circumferential direction along the outer ring, provide the maximum effect.

The speed of movement of the roller relative to the outer ring will be (See Nazarenko Yu.B. Liquid friction in bearings and the influence of hydrodynamic forces on the contact stresses of rolling elements // Engine. - Moscow. - 2015, No. 2. - P. 10-11)

,

,

f is the rotor speed, Hz; d _In , d _N , d _W - respectively, the diameter of the treadmill of the inner, outer ring and roller, m

The rate of oil outflow from the nozzle for an incompressible fluid and with the cross-sectional area of the oil channel for supplying oil to the bearing far exceeding the nozzle nozzle area is determined from the Bernoulli dependence (see L.D. Landau, E.M. Lifshits. Hydrodynamics, Nauka, 24 pp. )

,

,

where P is the oil pressure in the nozzle feed channel, N / m ² ; ρ is the oil density, kg / m ³ .

To ensure maximum hydrodynamic forces, it is necessary that the flow rate in the circumferential direction be equal to the speed of the rollers.

In this case, the hydrodynamic pressure will be determined only by the axial velocity of the oil flow, and it will be maximum. As studies have shown, the effectiveness of the hydrodynamic forces will be much greater in the axial direction of the oil flow than in the circumferential. In addition, it will act both on the left and on the right side of the roller relative to its axis. In the case of a lower or greater oil velocity in the circumferential direction and with the oil supply vector directed at a certain angle to the axis of the roller, the oil pressure will act only on one side of the roller, and its value will be less.

The axial oil flow rate is determined from the condition that the oil jet will be directed to the middle part of the conical surface of the visor, and the centrifugal force will be determined by rotating the oil in the circumferential direction with a speed V, the angle of inclination of the inner surface of the visor in the axial direction γ and the width of the visor without a protrusion b.

The centrifugal force of the oil layer under the visor of mass m is determined from the expression

,

,

where R _H is the inner radius of the outer ring, m; V is the peripheral speed of the oil flow, m / s; b - visor width, m; γ is the angle of inclination of the inner surface of the visor in the axial direction, rad.

The axial force acting on the oil layer under the visor will be equal to F _O = F _C · γ.

The oil acceleration will be

.

.

The transit time of the oil from the middle of the visor to its end and entering the gap between the roller and the outer ring is determined from the condition

.

.

In this case, the oil flow rate at the entrance to the gap between the roller and the outer ring will be

An example of such a calculation is given below.

An example of the implementation of the device with the above technical result was carried out on a model simulating a bearing 5-2272917 P with overall dimensions 85 × 120 × 18 mm with a diameter of the treadmill of the outer ring D _N = 0.112m and a roller D _P = 0.008 m, the diameter of the running the track of the inner ring d _B = 0.096 m, the length of the roller L = 0.009 m. The rotational speed of the rotor shaft n = 14000 rpm or f = 233.3 Hz.

The speed of the roller with these bearing parameters will be:

.

.

To realize the same flow rate in the circumferential direction, the oil pressure in the oil system and in the nozzle should be P = 0.575 MPa.

м/сек,

m / s

where ρ is the oil density ρ = 800 kg / m ³ .

The axial flow rate of oil is determined from the condition that the centrifugal force will be determined by rotating the oil in the circumferential direction with a speed of V = 37.9 m / s, a visor angle of inclination γ = 0.2 rad and a visor width (without protrusion) b = 0.04 m.

The acceleration of oil when moving along the visor will be

,

,

where V is the oil velocity in the circumferential direction, V = 37.9 m / s; b - visor width, b = 0.04 m; R _N is the radius of the treadmill of the outer ring, R _N = 0.056 m; γ is the angle of inclination of the inner surface of the visor in the axial direction, γ = 0.2 rad.

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

.

.

If the distance in the axial direction from the end face of the outer ring of the bearing to the end of the roller is f = 0.004 m, the radial clearance between the outer ring and the bearing cage is m = 0.00075 m and the chamfer width on the cage is e = 0.0015 m, the condition is not exceeded maximum permissible visor angle γ = 0.2 <0.3

, рад.

, glad.

The hydrodynamic pressure in the gap between the roller and the outer ring is determined by the method (See Nazarenko Yu.B. Liquid friction in bearings and the effect of hydrodynamic forces on the contact stresses of rolling elements // Engine. - Moscow. - 2015, No. 2. - P. 10 -11), for which we divide the gap in the axial direction between the roller and the outer ring in the form of separate sections A, B, C and F, which can be represented as flat elements without curvature and which form a gap between two plates (see Fig. 4 )

The clearance between the roller and the ring in zones A, B, C and F with the X coordinate in the center of the roller is determined from the expression (See .. Nazarenko, Yu.B. Liquid friction in bearings and the effect of hydrodynamic forces on contact stresses of rolling elements // Engine. - Moscow. - 2015, No. 2. - S. 10-11)

,

,

where R _N is the radius of the treadmill of the outer ring, R _N = 0.056 m; R _P is the radius of the roller, R _P = 0.004 m; α is the angle between the contact point of the roller and the point on the roller in the circumferential direction, where the gap between it and the outer ring is determined; ζ - equal parameter, ζ = R _Н -R _Р = 0.052 m.

With the width of the zones A, B, C, and F equal to ρ = R _P -sin (Δα) = R _P · Δα = 0.1 mm, with an arc interval Δα = 0.025 rad. and parameter ζ = R _H -R _P = 0.052 mm, the gap in the middle section of the roller in the middle of each zone will be h = 0.3 μm (t. 7), h = 2.6 μm (t. 5), h = 7.3 μm (t. 3), h = 14.2 μm (v. 1).

The gaps for the considered zones at the end of the roller will be greater by the value of the bombinivariness of the roller, which is taken to be b = 10 μm (see Fig. 4) and they will be h = 10.3 μm (t. 8), h = 12.6 μm (t. 6), h = 17.3 μm (v. 4), h = 24.2 μm (v. 2).

In the absence of the movement of the two plates of the upper and lower axial direction and only the oil flow in the axial direction between the plates that form the rollers and the outer bearing ring, the hydrodynamic pressure in the middle of each zone A, B, C and F is determined from the expression

,

,

where h ₁ is the initial value of the gap; h _cf - the gap in the middle of the plate for zones A, B, C and F, respectively, is 5.3, 7.6, 12.3 and 19.2 μm; V is the oil flow rate; V = 14.9 mm / 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 plate, β = 0.0022 rad.

The average hydrodynamic pressure in zones F, C, B, A will be: P _F = 29.6 MPa; P _C = 15.2 MPa; P _B = 6.2 MPa; P _A = 2.7 MPa.

The area of each of the considered zones will be

м²,

m ²

where L is the length of the roller, L = 9 · 10 ^-3 m; ρ is the width of the zone, ρ = 0.1 · 10 ^-3 m.

We summarize the forces of each of the zones we get

_{F = S · (P F +} P C + P B + P A) = 0.45 · 10 ^-6 · (6.29 · 10 ⁶ + 15.2 · 10 ⁶ + 6.2 10 ⁶ + 2.7 × 10 ⁶⁾ = 24.2 N.

We establish the full force acting on the roller, provided that the force acting on both sides of the roller, both in the transverse and in the axial direction, is the same. Then we will have

With a radial load on the bearing F _R = 5600 N, the force per one roller will be (see Chernevsky L.V., Korostashevsky R.V. Rolling bearings: reference book - catalog - M .: Mashinostroenie, 1977. - 205 p.) .

,

,

where Z is the number of rollers, Z = 28.

The decrease in the force acting on the roller and the outer ring in the zone of their contact will be 10.5%. And this increases the service life of the bearing.

Thus, the axial flow of oil into the gap between the rollers and the outer ring of the bearing at a speed equal to the speed of the rollers relative to the stationary outer ring provides a guaranteed reduction in contact stress between the rollers and the outer ring. This effect was obtained experimentally and confirmed by calculations. (See Nazarenko Yu.B. Liquid friction in bearings and the influence of hydrodynamic forces on the contact stresses of rolling elements // Dvigatel. - Moscow. - 2015, No. 2. - P. 10-11).

Claims (2)

1. A device for supplying oil to a bearing of a rotor support of a gas turbine engine, comprising nozzles with output nozzles mounted on the stator parts of the engine, having the ability to connect to the engine oil line, and an oil catch visor designed to direct oil flows from nozzle nozzles to the bearing, characterized in that the oil catch visor is made in the form of a ring mounted on a support stator at the end of the outer bearing ring, the inner surface of the ring has a conical shape with in the direction of the bearing, on the periphery of the ring opening on the outer side of the visor there is an annular protrusion, and the nozzles are mounted in such a way that their nozzles are directed to the conical surface of the ring. 2. The device according to p. 1, characterized in that the taper angle (γ) of the conical surface of the ring is within:

, the thickness of the visor in the radial direction adjacent to the outer ring of the bearing is

, but the opposite

where γ is the angle of inclination of the inner surface of the visor in the axial direction; m is the radial clearance between the treadmill of the outer ring and the separator; e is the bevel width in the axial direction on the separator; f is the axial distance from the end of the outer ring of the bearing to the end of the roller; c is the thickness in the radial direction of the outer ring of the bearing; b is the width of the visor without a protrusion.

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