GB2138891A - Mounting rotor blades - Google Patents

Abstract

Rotor blades 13 in an axial fluid flow machine, such as a turbine or compressor, are fixed to the rotor 3 by means of pins 15 which pass through bores 19a, 19b provided in the root portions 11 of the blades. In order to avoid point contact between the pins and the bores during operation of the machine when the blades twist on the pins under a combination of centrifugal and aerodynamic forces, the pins are made effectively barrel- shaped (or the bores are made a waisted shape) so that the contacts between the pins and the bores are lines which are substantially helical about the centre-lines 33 of the bores. This reduces stress per unit area of the bores. <IMAGE>

Description

SPECIFICATION Mounting rotor blades The present invention relates to the mounting of aerofoil-shaped blades on the rotors of axial fluid flow machines. The invention has particular, but not exclusive, relevance to the compressors of gas turbine engines.

Methods of securing rotor blades to gas turbine engine compressor rotors vary with different designs. One well known method is to provide each rotor blade with a root portion comprising one or two lugs, which lugs are held between flanges on the compressor rotor drum or disc by means of cylindrical pins which pass through cylindrical bores in the lugs and the flanges. The pins are a clearance fit in the bores provided in the lugs and in the bores provided in the flanges.

It has been found that the combined effects of centrifugal forces and aerodynamic twisting forces on the blades during operation of the engine cause skewing of the blades on the pins and hence point contacts instead of line contacts between the pins and the ends of the bores in the lugs, leading to very high direct stress concentrations and a consequent shortening of service life of the blades due to cracks in the lugs originating at the point contacts. This problem is emphasised if the lugs are subject to corrosion, since even very small corrosion pits act as stress multipliers at the point contacts.

The present invention provides a bladed rotor for an axial fluid flow machine in which rotor blades having root portions are mounted on the rotor by means of pins which pass through bores provided in the root portions and through bores provided in portions of the rotor which cooperate with the root portions, the pins being a clearance fit in the bores in the root portions and the pins and the bores in the root portions being shaped such that when the bores in the root portions are skewed with respect to the longitudinal centre-line of the pins under the influence of centrifugal forces and aerodynamic forces acting on the blades during operation of the machine, the contacts between the pins and the bores in the root portions are lines which are substantially helical about the centrelines of the bores.

Such helical line contacts can be achieved by making the pins effectively barrel-shaped and the bores substantially cylindrical.

Preferably the root portions of the rotor blades comprise lugs and the cooperating portions of the rotor comprise flanges, the lugs being held between and connected to the flanges by means of said pins.

The root portion of each rotor blade preferably comprises either a single lug or a pair of spaced-apart lugs.

Embodiments of the invention will now be described by way of example only and with reference to the accompanying drawings, in which: Figure 1 is a partly exploded sectional perspective view of part of a prior art gas turbine engine compressor rotor stage showing the main components; Figure 2 is an enlarged diagrammatic sectional plan view of the interaction of the root portion of a rotor blade with its securing pin; and Figure 3 is an enlarged diagrammatic view on Arrow A in Fig. 2, illustrating dimensions to be taken into account in producing an embodiment of the present invention.

The drawings are not to scale.

Referring to Fig. 1, a compressor rotor stage 1 is part of a compressor drum structure, a portion of the drum itself being shown at 3. A ring 5 is welded to the drum 3, and has been machined to provide three radially and circumferentially coextensive flanges 7, 8 and 9. Root portions 11 of rotor blades 1 3 comprise twin lugs 1 2a and 1 2b and are secured between and connected to the flanges 7, 8 and 9 by means of pins 15 (only one of which is shown) which pass through the three bores 17a, 17b, 17cin flanges 7,8,9, and the two bores 1 9a, 19b in lugs 1 2a and 1 2b.

The pins 1 5 themselves are retained in the bores by means of heads 22 and locking rings 23,25 which engage grooves 27 in the ends of the pins. Both the pins and the bores are substantially cylindrical, the pins being a clearance fit in the bores.

When the rotor is rotating at normal operational speeds, centrifugal force acts on the blades and ideally this would cause the radially inmost part of each pin 1 5 to engage the radially inmost parts of bores 19a, 19b in lugs 12a,12b, the engagement being a straight line of contact between two cylindrical surfaces, one cylinder being of slightly smaller radius and nested inside the other so that the longitudinal centrelines of the cylinders are offset from, but parallel to, each other. However, it has been found that aerodynamic forces on the blades cause the blades to twist on the pin, and this places the pins in point contact with the bores in the lugs instead of in line contact.This situation is shown in exaggerated manner in Fig. 2, where it will be seen that point contacts between pin 1 5 and lugs 9 2a. 12b occur at 29 in the front face 20 of the frong lug 1 2a and at 31 in the rear face 22 of the rear lug 1 2b, the longitudinal centre-line 33 of bores 1 9a, 1 9b being skewed in relation to the longitudinal centreline 35 of pin 15. Note that in Fig. 2 the difference between the diameter of bores 1 9a, 1 9b and the diameter of the pins 1 5 has been exaggerated for the sake of clarity. The exaggeration is even greater in Fig. 3.

In Fig. 3, it is shown on the rear face 22 of lug 1 2b that whereas under centrifugal forces alone, line contact between the pin 1 5 and bore 1 9b would occur at position 37 (the ' "6 o'clock" position), and under aerodynamic twisting forces alone, point contact would occur at position 39 (the "3 o'clock posi tion"), the actual point contact is at the position 31 between positions 37 and 39, which is at an angular position 8 degrees from position 37. The angle 8 results from the combined effects of the centrifugal forces and the twisting forces and is dependent upon their relative magnitudes.

The situation on the other side of root portion 11, ie on the front face 20 of front lug 12a, is not illustrated, but it will easily be seen that still looking in the direction of arrow A, the line contact due to centrifugal force alone would again be in the 6 o'clock position on the bore 1 9a, the point contact due to twisting forces alone would be in the 9 o'clock position, and the point contact 29 resulting from the combined effect of the forces is at B degrees from the 6 o'clock position between the 6 o'clock position and the 9 o'clock position.

Although the position 31 and its corresponding position 29 in the front face 20 of front lug 1 2a have so far been thought of as fixed positions, the magnitudes of the aerodynamic forces fluctuate appreciably even at constant rotational speeds, and this causes corresponding fluctuations in the magnitude B, i.e. the lugs 1 2a and 1 2b effectively rock back and forth on the pins 1 5. The frequency at which this occurs is called the "pin rolling frequency" and it varies with the rotational speed of the rotor, being highest at high speeds.

Position 31 therefore represents a mean position, and the actual point of contact oscillates about it.

Since, during operation of the rotor stage 1, only point contacts occur between the pins 1 5 and bores 1 9a, 19b, portions of the edges of bores 1 9a, 1 9b, centred on positions 29 and 31 in the faces 20 and 22 respectively, experience high stress concentrations, which especially when combined with the effects of cyclic fatigue and corrosion, can produce cracks in the lugs originating at the edges of the bores on faces 20 and 22.

By applying the invention, the magnitudes of the stresses per unit area of contact in the bores 1 9a, 1 9b are lowered to an acceptable level, the invention involving shaping the pins 15 and the bores 1 9a 1 9b so that under operational conditions line contacts tend to occur between them rather than point contacts, these line contacts being substantially helical about the centre-line 33 of the bores 19a, 19b (Fig. 2).

The most convenient means of achieving such helical line contacts is to leave the bores cylindrical in shape, but to make the pins barrelled in longitudinal profile.

The degree to which pins 1 5 should be barrel-shaped (assuming the bores are cylindrical) is calculated by first ascertaining (by means of tests if necessary) the angle 8 (see Fig. 3) for a cylindrical pin making point contact with the edge of the bore 1 2b at position 31 in face 22, the position 31 now representing the mean position of contact at the rotor speed considered most important in terms of its effect on the fatigue life of the lugs. The amount, X, by which the radius of the pin at its mid-span should be greater than the radius of the pin at its ends in order to produce the correct barrel shape is given by X = R - R - r) cos 0+ r], where R = radius of the cylindrical bore, and r = radius of the pin at its ends.

The pin is barrelled over the portion which extends from the front face 20 of front lug 1 2a to the rear face 22 of rear lug 1 2b.

Typical clearances between the pins and the bores would be about 127 to 254 ,um on diameters of around 12.7 mm. Thus X, the amount of barrelling, is only of the order of 37 to 76.2 ym. This can give rise to difficulties in machining the pins accurately to the correct shape, since the radius of the barrel curve is up to about 3.6 metres. In order to facilitate production of the pins, the barrelshape can be approximated by machining a double taper on the pin, this being easy to machine accurately.Due to the very small differences in dimensions between the doubletapered shape and the barrel-shape, the double-tapered shape is in fact effectively barrelshaped in its mode of interaction with the bores, lines of contact being produced between the tapered pins and the bores which are substantially helical during operation of the compressor due to slight deformation of the tapers and the bores under working loads.

Although the invention is being described in relation to a blade whose root portion has two lugs as shown in Fig. 1, it is equally applicable to blades fixed to the rotor by means of a single lug spanning the gap between twin flanges.

Helical line contacts between the bores and the pins 1 5 could also be achieved by forming the bores as the inverse of a barrel-shape, i.e. as a waisted shape, the pins being cylindrical, though this would be more difficult and costly to produce. The above-mentioned analytical and dimensional considerations would apply anologously. It should be realised that in the case of a blade with a single lug, the entire bore therethrough would be formed as a waisted shape, but in a case where twin lugs are utilised, the bores in the lugs would form only opposing ends of a waisted shape whose imaginary mid-portion would occupy the gap between the lugs.

The invention has a further advantage in that the barrel-shaped pins (or the waisted bores) have the effect of raising the pin-rolling frequency relative to that produced by cylindrical pins (or bores). This is partly because barrel-shaped pins or waisted bores have a slightly different effective radius from cylindrical pins or bores with the same end radius, and partly because the greater area of contact between the pins and the bores in the blade root portions increases the frictional forces between the pins and the bores. A high pinrolling frequency can be advantageous in avoiding unwanted resonance with other vibrations present in the compressor.

The invention is applicable to turbines as well as to compressors.

Although in the foregoing description, the points of contact 29 and 31 (Figs. 2 and 3) have been described as such, they are of course small areas of contact. Likewise the lines of contact described herein are narrow strip-like areas of contact. Further, although points 29 and 31 have been described as lying in the faces 20 and 22 of lugs 1 2a and 1 2b respectively, they are in fact liable to be a very short distance down bores 1 9a, 1 9b, since the edges of the bores in faces 20 and 22 will be radiused to avoid sharp corners.

Claims (4)

1. A bladed rotor for an axial fluid flow machine in which rotor blades having root portions are mounted on the rotor by means of pins which pass through bores provided in the root portions and through bores provided in portions of the rotor which cooperate with the root portions, the pins being a clearance fit in the bores in the root portions, and the pins and the bores in the root portions being shaped such that when the bores in the root portions are skewed with respect to the longitudinal centre-lines of the pins under the influence of centrifugal and aerodynamic forces acting on the blades during operation of the machine, the contacts between the pins and the bores in the root portions are lines which are substantially helical about the centre-lines of the bores.

2. A bladed rotor according to claim 1 in which the helical line contacts are achieved by the pins being effectively barrel-shaped and the bores in the root portion being substantially cylindrical.

3. A bladed rotor having blades with root portions which are fixed to the rotor by means of pins passing through bores in said root portions and said rotor, said pins and bores being specially shaped to reduce stress concentrations substantially as described in this specification.

4. A gas turbine engine incorporating a bladed rotor according to any one of claims 1 to 3.