GB2407346A

GB2407346A - Turbine shroud seal

Info

Publication number: GB2407346A
Application number: GB0418733A
Authority: GB
Inventors: Timothy Stephen Rice
Original assignee: Alstom Technology AG
Current assignee: General Electric Technology GmbH
Priority date: 2003-10-24
Filing date: 2004-08-20
Publication date: 2005-04-27
Anticipated expiration: 2024-08-20
Also published as: GB2407346B; GB0324843D0; GB0418733D0

Abstract

A turbine comprises shrouded turbine rotor blades (12) and a surrounding static casing (14). Seals (22) acting between the blade shrouds (16) and the casing (14) efficiently restrict leakage of turbine working fluid over the outer surfaces of the shrouds between high and lower pressure (PH, PL) sides of the turbine rotor. The casing 14 includes resonator passages 28 positioned just ahead of the blades 12, in the region where seal leakage fluid passes into a gap G between the blade shrouds 16 and upstream static structure 11. The resonator passages 28 improve turbine efficiency by preventing or reducing cyclic reflux of the leakage fluid through the gap G into the main turbine passage in the low-pressure region between adjacent turbine blades. Applicable to axial flow steam and gas turbines.

Description

IMPROVEMENTS IN OR RELATING TO AXIAL FLOW TURBINES

Field of the Invention

The invention is applicable to axial flow turbines of the type which use a compressible working fluid, such as steam and gas turbines. It provides a way of improving the aerodynamic environment upstream of shrouded turbine rotor blades, particularly in cases where an efficient seal acts between the shrouds and a surrounding turbine casing.

Background of the Invention

In an axial flow steam or gas turbine, a turbine rotor comprises turbine buckets or blades attached to a disc or drum, with the outer periphery of the turbine rotor circumscribed by a turbine casing. There will of course be a small gap defined between the periphery of the turbine rotor and the casing and turbine working fluid from the main turbine flow passage will inevitably leak through this gap, so reducing the efficiency of the turbine. Hewe, it is desirable to minimise the amount of leakage through the gap. In some turbines, the rotor blades are unshrouded, in which case the leakage to be minimised is that which flows over the blade tips from the pressure surfaces to the suction surfaces of the blades. In many other turbines, the rotor blades are shrouded and the leakage to be nunimised is that which exits the main turbine passage and flows OVER the radially Outfit surface of the shroud between the upstream and downstream sides of the turbine rotor. In either case, the gap between the periphery ofthe turbine rotor and the casing is minimized to minimise leakage. In the case of shrouded rotors, such leakage can be reduced by use of a variety of known types of seal, acting between the shroud and the casing. It is with such shrouded turbine rotors that the present invention is particularly concerned.

The velocity of leakage flow passing into a small part of a seal on a shrouded rotor, when considered as a function of time, can be represented in terms of a variable component (or "variable velocity") and a direct component (or "average velocity").

The variable velocity is characterized by a locally increased velocity in the region of the leading edges of the turbine blades, followed by a reduced velocity in the region - 2- between adjacent turbine blades. It therefore varies cyclically relative to the average velocity as the blades pass any particular point on the casing.

The variable velocity of leakage flow into a known type of seal is illustrated s diagrammatically in Figure 1 by the solid curve AA. The yaxis of Figure I represents the velocity, in relative units, of the leakage flow into a small part of the seal. The x-axis represents time, where each unit is the time between successive turbine blades passing a point on the turbine casing. Plainly, the value of each time unit will depend on the number of blades on the rotor and its rotational speed. Hence, a particular operational condition of the turbine is chosen for investigation, this conveniently being the normal full-load condition of the turbine, at which it spends most of its operational life. It can therefore be seen that in this example the variable velocity A-A varies cyclically between a maximum value of about 3.2, when the leading edges of the turbine blades are aligned with the chosen point on the casing, and a minimum value of about 1.8, when the regions mid-way between angularly adjacent turbine blades are aligned with the chosen point.

The average velocity of the leakage flow as a function of time is represented in Figure l by the dashed line B-B. In this example, it is a constant value of about 2.5.

Normally, the overall velocity of the leakage flow into any part of the seal is always positive, as is the case in Figure 1. However, with some technically advanced, improved shroud seals the average velocity of leakage flow can be reduced to such an extent that the local velocity of leakage flow becomes negative in the low-pressure region between adjacent turbine blades. The variable velocity and average velocity for such an improved seal are represented in Figure l by the solid line C-C and the dashed line E-E respectively. Although the average velocity E-E is now only about 0.4, the amplitude of the variable velocity Cog restive to the average velocity E-E remains little changed, such that it varies cyclically from a maximum value of about 1. I to a minimum value of about.3. The latter negative velocity indicates that some seal leakage flow that has already left the main turbine passage is readmitted to the main turbine flow. Furthermore, such readmitted flow will no longer carry the correct - 3 flow angle relative to the blade surface that intercepts it. This can disturb the smooth aerodynamic flow around the radially outer ends of the blades and reduce the efficiency of the turbine, thereby tending to negate the technical benefit of the improved seal. s

It is an object of the present invention to improve the aerodynamic environment upstream of the turbine blades at a predetermined running speed of the turbine by reducing or eliminating the occurrence of negative seal leakage flow velocities at that condition. The invention enables a very efficient seal to be employed to act between a shrouded turbine rotor and a turbine casing while avoiding or at least reducing the aerodynamically disadvantageous, cyclic readmission of leakage flow into the main turbine flow passage.

Summary of the Invention

In its broadest aspect, the invention provides an axial flow turbine having a bladed turbine rotor, a seal means acting between a shroud structure on the turbine rotor and static casing structure circumscribing the shroud structure, and an annular array of generally radially extending resonator passages that communicate with a gap defined between a leading edge of the shroud structure and adjacent static structure, the resonator passages being adapted to resonate to an integer multiple or vulgar fraction of a predetermined blade passing frequency of the bladed rotor, thereby in use to reduce cyclic reflex of seal leakage fluid through the gap into the regions between angularly adjacent turbine blades.

The present invention further provides an axial flow turbine comprising: a turbine passage, a turbine rotor comprising turbine rotor blades, shroud structure on the radially outer ends of the rotor blades, static casing structure circumscribing the shroud structure; seal means acting between the shroud structure and the static casing structure to restrict leakage flow of turbine working fluid from the turbine passage over the radially outer side of the shroud structure; and an annular chamber immediately upstream of the seal means and outboard of the turbine passage, the annular chamber having an inlet upstream of the turbine rotor for admitting to the chamber (a) seal leakage flow from the turbine passage upstream of the turbine rotor, and (b) pressure waves emanating from the rotor blades, wherein the static casing structure includes a circumferential array of generally radially extending resonator passages that are closed-ended within the static structure and in open ended communication with the annular chamber, each resonator passage being dimensioned to accept pressure waves from the chamber, reflect the pressure waves from a distal end of the passage, and expel the pressure waves back into the chamber, the arrangement being such that at a predetermined blade passing frequency, the resulting pressure variations in the chamber between successive blade passing events act to minimise the occurrence of negative velocities in the flow of seal leakage fluid through the inlet into the annular chamber.

The effect of the invention is to maintain a positive velocity in the flow of seal leakage fluid through the inlet passage into the annular chamber, or at least to minimise the time during which a negative velocity occurs in such seal leakage flow.

At the same time, the average velocity of the seal leakage fluid is substantially the same as it would be if the resonator passages were not present. It is important to note that the invention is intended to reduce the occurrence of negative velocity excursions rather than reduce the average velocity of the leakage fluid.

The resonator passages may be arranged as two or more axially successive rows of passages.

In accordance with one embodiment of the invention, and ignoring end effects, the length of a resonator passage is calculated by deriving the wavelength of twice the predetermined blade passing frequency, and setting the length of the resonator passage so that it resonates to the derived wavelength. The blade passing frequency is the frequency at which any point on the turbine casing is passed by a turbine rotor blade and is determined by multiplying the number of turbine blades on the rotor with its frequency of rotation. However, it should be understood that, dependent upon the - 5 dimensions, shape and structure of the turbine, resonator passages having lengths which resonate to a wavelength derived from one or more other multiples or fractions of the predetermined blade passing frequency may also be found to have a beneficial effect on velocity variations in the leakage flow. For example, half the predetermined blade passing frequency, or even the blade passing frequency itself, may be an advantageous choice. The exact choice must be made by the designer of a particular turbine, based on engineering and acoustic analysis of the design and/or routine testing.

The wavelength of twice the predetermined blade passing frequency (or other chosen multiple or suitable vulgar fraction of the predetermined blade passing frequency) is of course determined by dividing the local speed of sound in the leakage fluid for any given pressure by twice the blade passing frequency (or other chosen multiple or vulgar fraction ofthe predetermined blade passing frequency). The inventor currently believes that the best results will be obtained if the length of each resonator passage is equal to about a quarter of the chosen wavelength. If more than one wavelength is found to influence favourably the velocity of the leakage flow through the inlet to the chamber, then adjacent resonator passages may be given appropriately different lengths. For example, using multiple rows of resonator passages, axially successive rows may have different passage lengths. However, variation in passage length as between circumfffentially adjacent resonator passages is also possible.

Including end effects, the presently preferred length L of a resonator passage can be expressed mathematically by the following equation: L= I v __D where 4 (Nxnx f) 4 ' v is the local speed of sound in m/s in the leakage fluid for any given pressure at the predetermined operating condition, N is the number of turbine blades, n is an integer or a vulgar fraction, possible preferred values of n being 2, l and 1/2, - 6- f is the frequency of rotation of the turbine blades in Hz at the predetermined operating condition, and D is the inner diameter ofthe resonator passage in metres.

The seal means between the shroud structure and the turbine casing may, for example, comprise one or more labyrinth-type seals and/or one or more bnsh-type seals. For the purpose of this specification, labyrinth seals include fin seals and brush seals include so-called "foil" or"leaf" seals.

The resonator passages can be formed by a honeycomb cell structure in the turbine casing, or by drilled holes, machined slots, or by spaces between axially spaced fins projecting generally radially inwards from the static casing structure. While it is necessary for the resonator passages to have a radial extent within the turbine casing, they can at the same time be inclined in the circumferential and/or axial directions, so allowing the resonator passages to take up less space in the radial direction.

It is generally preferred that the width OVER length ratio of the resonator passages is as small as possible, e.g., in the range l/6 to l/15. For example, in a particular embodiment of the invention, hexagonal section resonator passages having a width across the flats of about l. Srnm and a length of about 18mm are proposed. This enables attainment of good acoustic resolution in terms of the packing density of the resonator passages and their response to local pressure conditions.

The invention is applicable to axial flow steam turbines and axial flow gas turbines.

Brief Description of the Drawings

Exemplary embodiments of the invention will now be described with reference to the accompanying drawings, in which: Figure l is a graph showing the variable and average velocities for a conventional seal, an improved seal and a turbine according to the present invention; Figure 2 is a diagrammatic broken-away" cross-section of part of a steam turbine, the section being taken in the radial and axial directions of the turbine and - 7 showing a shrouded turbine blade and a surrounding turbine casing in accordance with the present invention.

Description of Exemplar Embodiments

With reference to Figure 2, the arrow S shows the overall direction of steam flow in a main turbine flow passage 1, during operation of a steam turbine stage 2 in an axial flow steam turbine. In flow series, turbine stage 2 comprises an annular array (a Prows) of angularly spaced-apart static blades 10, only the trailing edge of which is shown, and a row of angularly spaced apart-moving rotor blades 12. A turbine casing 14 surrounds the turbine stage 2. The static blades extend between - and are attached to - an outer static blade support ring 11, housed in turbine casing 14, and an inner static ring (not shown). Moving blades 12 are attached to a disc or drum (not shown) at their radially inner ends to form a bladed rotor for rotation within the casing 14. At their radially outer ends, each moving blade 12 has an axially and circumferentially extending shroud portion 16, which may be either formed separately from the blades and then fixed to them, or formed integrally with the blades.

During normal operation of the turbine, there is a gap G between the upstream edge of the shrouds 16 and the confronting side of the static ring 11. Gap G is necessary to allow for axial movement of the bladed rotor due to differential thermal expansion and aerodynamic forces.

On the upstream side of the moving blade row, the steam pressure PH is higher than the pressure PI on the downstream side. Hence, as shown by the broad flow arrows, there is a tendency for some of the steam to leave the main turbine passage 1, flow through the gap G with a velocity U. and pass over the radially outer side of the shroud 16. It will be apparent to the skilled person that such leakage flow over the outside ofthe shroud is undesirable, because it reduces the efficiency ofthe turbine.

To reduce such leakage of steam, the gap 20 between the outer surface of the shroud 16 and the inner surface ofthe casing 14 is provided with a seal 22, or as shown, what is effectively an efficient labyrinth seal comprising a series of seals 22 in the form of - 8- pairs of fins 24 extending from the casing 14 towards lands 26 on the shroud 16. The fins 24 and lands 26 extend circumferentially of the shroud surface and may be radially dimensioned such that during normal operation of the turbine, there is a minimum clearance (say, lmm or less) between the inner tips of the fins and the surfaces of the lands. During running up of the turbine from cold to normal operating temperatures, this clearance may be reduced or eliminated due to differential thermal expansion of the moving blades 12 and the casing 14. Consequently, rubbing of the moving lands 26 against the fins 24 can occur. To preserve the minimum clearance at normal turbine operating temperatures, the lands 26 can be made of a softer material than the fins (or given an abradable coating) so that the tips of the fins can cut matching grooves in the lands.

The number of fins 24 and their opposing lands 26 is to some extent at the option of the designer. Plainly, within the constraints imposed by the structural design of the turbine, the more fin/land seals are provided in the labyrinth seal between the casing 14 and the shroud 16, the less steam will leak over the shroud between PH and Pr.

Furthermore, even more efficient seals, such as so called "brush" seals, can be used in place of, or in addition to, one or more of the fins 24 in the labyrinth seal. However, the use of efficient sealing configurations in itself contributes to the problem tackled by the invention.

As previously explained, with some very efficient shroud seals, the average velocity of the leakage flow through the gap G can be reduced to such an extent that the local velocity U becomes negative in the lowerpressure region between adjacent turbine rotor blades 12.

To help overcome this problem, the invention provides a circumferential array of acoustic or pressure-wave resonator passages 28 in the casing 14. These resonator passages are located outboard of the shrouds 16 and are axially located in the region of the gap G. where the leakage steam passes into the seal 22 with a velocity U. In the configuration shown, the gap G forms an inlet into an annular channel or chamber bounded by static ring 11, turbine casing 14, seal 22, and shroud 16. The resonator - 9 passages 28 are connected to chamber 30 through their open ends and extend generally radially into the turbine casing 14 for a length L, measured to their closed ends. To reduce the radius of the space they occupy in the turbine casing, the resonator passages may lean away from the radial direction in the circumferential and/or axial directions, if desired.

According to one embodiment of the invention, if length L (ignoring end effects for the moment), is made equal to one quarter of the wavelength of twice the rotor blade passing frequency, a second harmonic standing wave will be created in each of the passages. To maximise the effect of the resonator passages around the circumference of the turbine rotor, they should have small internal diameters and be closely adjacent to each other. Close packing can be achieved by using a hexagonal cell "honeycomb" type of construction for the acoustic passages, the internal distance across their flats being equivalent to the diameter of a round section tube ofthe same length L. Each resonator passage 28 is dimensioned to accept, reflect and expel pressure waves emanating from the rotor blades 12 as they pass the location of the resonator passage so that pressure variations in the annular chamber 30 are a modified form of the pressure variations in the turbine passage I caused by passing of the turbine rotor 2Q blades. In fact, relative to the same turbine structure without the benefit of the invention, the invention acts to flatten or reduce the profile of a time-domain plot of momentary pressure in the annular chamber 30.

The effects on the variable velocity U of the thus-modified pressure variations caused by the invention are illustrated in Figure 1, which allows a direct comparison to be made between the solid curve C{:, representing a prior art turbine having a very efficient seal but no acoustic passages, and the dashed curve D, representing a turbine according to the present invention with the same seal. For the prior art turbine, the local velocity of the leakage flow through gap G varies periodically from a maximum value of about I. I when the leading edge of each of the turbine blades passes in front of the portion to a minimum value of about -0.3 when the lower- pressure region between adjacent turbine blades passes in front of the portion. For the - 10 turbine according to the present invention, the local velocity of the leakage flow varies periodically from a maximum value of about 1.4 to a minimum value of about 0.1. The local velocity of the steam into the seal 22 through gap G is therefore always positive and no leakage flow is readmitted to the main turbine flow. The average velocity of 0.4 is illustrated by the straight dashed line E-E in Figure I and is the same for the prior art turbine and the turbine according to the invention.

Assuming the above length L is adopted, the action of the resonator passages 2$ is for each one to accept a pressure wave via the gap G as the leading edge of a rotor blade passes it. The pressure wave passes down the resonator passage, reflects from its distal end, then passes back up the acoustic passage and is expelled into annular chamber 30. The reflection and expulsion of the pressure wave is followed by a rarefaction in the resonator passage, which lowers the pressure in chamber 30. This rarefaction then passes down the resonator passage 28 and reflects from its distal end.

The reflection and expulsion of the rarefaction completes one cycle. Another cycle starts as a further blade pressure-pulse travels down the resonator passage. The length L of the resonator passages is such that midway between successive blades, the pressure in chamber 30 (relative to the pressure without the invention) has a value which prevents the reversal of the velocity of the leakage flow through gap G. In this way, the amplitude of the velocity variation of leakage flow into the seal 22 is modified and the problems associated with negative leakage-flow velocity can be reduced or eliminated.

Allowing for end-effects, the length L required for correct functioning of the invention in respect of twice the blade passing frequency can be calculated directly from the blade passing frequency, the diameter of the acoustic passages, and the local speed of sound in steam (or turbine gas in the case of an axial flow gas turbine) by using the formula L= I V __D, 4 (Nx2x f) 4 where Nis the number ofturbine blades, f is the frequency of rotation of the turbine rotor in Hz, - 11 D is the inner diameter ofthe acoustic passage, and v is the local speed of sound in m/s in the leakage fluid for any given pressure (in this case, the pressure chosen is the average pressure during normal operation of the turbine).

For example, if the turbine rotor has 90 turbine blades and rotates at a frequency of 50 Hz, twice the blade passing frequency will be 9000 Hz. If the local speed of sound is 656 rn/s, then the effective wavelength of twice the blade passing frequency for this particular turbine is about 18.22 mm. As mentioned above, it is advantageous if acoustic passages are narrow, so if their inside diameter is 1.5 mm, the end effect reduces the length L by only 0. 375mm, giving a final value for L of about 17.85 mm.

Although in some circumstances a single circumferentially extending row of acoustic passages 28 may suffice to put the invention into effect, it is preferable that there be two or more rows of passages, the rows being axially adjacent to each other. In the present embodiment, the passages are arranged in a double row. As mentioned above, for the best packing density and acoustic performance, the passages have a hexagonal crosssection and in the present embodiment are 1.5 mm across their flats, though different cross-sections and dimensions may be chosen at the option of the designer.

For example, the passages may comprise slots machined into the turbine casing, or drilled holes.

Furthermore, depending on the characteristics of the turbine, it may be desirable to take into account the effect of fractions or multiples (harmonics) of the blade passing frequency on the velocity U of the seal leakage air through gap G. In this case, the length L of at least some of the acoustic passages, in at least one of the rows, may need adjusting to match the quarter wavelength of a frequency of interest. - 12

Claims

1. An axial flow turbine having a bladed turbine rotor, a seal means acting between a shroud structure on the turbine rotor and static casing structure s circumscribing the shroud structure, and an annular array of generally radially extending resonator passages that communicate with a gap defined between a leading edge of the shroud structure and adjacent static structure, the resonator passages being adapted to resonate to a harmonic of a predetermined blade passing frequency of the bladed rotor, thereby in use to reduce cyclic reflex of seal leakage fluid through the gap into the regions between angularly adjacent turbine blades.

2. An axial flow turbine comprising: a turbine passage, a turbine rotor comprising turbine rotor blades, shroud structure on the radially outer ends ofthe rotor blades, static casing structure circumscribing the shroud structure; seal means acting between the shroud structure and the static casing structure to restrict leakage flow of turbine working fluid from the turbine passage over the radially outer side of the shroud structure; and an annular chamber immediately upstream of the seal means and outboard of the turbine passage, the annular chamber having an inlet upstream of the turbine rotor for admitting to the chamber (a) seal leakage flow from the turbine passage upstream of the turbine rotor, and (b) pressure waves emanating from the rotor blades, wherein the static casing structure includes a circumferential array of generally radially extending resonator passages that are closed-ended within the static structure and in open-ended communication with the annular chamber, each resonator passage being dimensioned to accept pressure waves from the chamber, reflect the pressure waves from a distal end of the passage, and expel the pressure waves back into the chamber, the arrangement being such that at a predetermined blade passing frequency corresponding to a chosen turbine operating condition, the resulting pressure variations in the chamber between successive blade passing events act to minimise the - 13 occurrence of negative velocities in the flow of seal leakage fluid through the inlet into the annular chamber.

3. An axial flow turbine according to claim I or claim 2, in which the lengths of at least some of the resonator passages are such that they resonate at a multiple of the predetermined blade passing frequency.

4. An axial flow turbine according to claim 3, in which the lengths of at least some of the resonator passages are such that they resonate at twice the predetermined blade passing frequency.

S. An axial flow turbine according to claim I or claim 2, in which the lengths of at least some of the resonator passages are such that they resonate at a fraction of the predetermined blade passing frequency.

6. An axial flow turbine according to claim 5, in which the lengths of at least some of the resonator passages are such that they resonate at half the predetermined blade passing frequency.

7. An axial flow turbine according to claim 1 or claim 2, in which the lengths of at least some of the resonator passages are such that they resonate at the predetermined blade passing frequency.

8. An axial flow turbine according to claim I or claim 2, wherein, at the predetermined operating condition, the approximate length of the resonator passages is proportional to the local speed of sound in the leakage fluid of the inlet passage and inversely proportional to the predetermined blade passing frequency of the turbine blades.

9. An axial flow turbine according to claim 8, wherein the length L of the resonator passages is determined in accordance with the following equation: - 14 L 1 v _ 1 D,where 4(Nxnxf) 4 v is the local speed of sound in metres/second in the leakage fluid for any given pressure at the predetermined operating condition, N is the number of turbine blades, n is an integer or vulgar fraction, f is the frequency of rotation of the turbine rotor in Hz at the predetermined operating condition, and D is the inner diameter of the resonator passage in metros.

10. An axial flow turbine according to any preceding claim, wherein the circumferential array of resonator passages comprises at least two axially successive rows of resonator passages.

11. An axial flow turbine according to any preceding claim, wherein adjacent resonator passages have different depths such that they resonate to at least two different frequencies.

12. An axial flow turbine according to claim 11 as dependent from claim 10, wherein the resonator passages in at least two axially successive rows are of different lengths such that they resonate to at least two frequencies.

13. An axial flow turbine according to any preceding claim, wherein the resonator passages comprise a honeycomb structure forming part of the static casing structure.

14. An axial flow turbine according to any one of claims I to 12, wherein the resonator passages comprise round section holes in the static casing structure.

15. An axial flow turbine according to any one of claims 1 to 12, wherein the resonator passages comprise slots in the static casing structure. -

16. An axial flow turbine according to any one of claims I to 12, wherein the resonator passages comprise spaces between axially spaced fins projecting generally radially inwards from the static casing structure.

17. An axial flow turbine according to any preceding claim, wherein the resonator passages extend radially and at the same time are inclined in at least one of the circumferential and axial directions.

18. An axial flow turbine according to any preceding claim, wherein the resonator passages have a width overlengthratioin the range 1/6to 1/lS.

19. An axial flow turbine according to any preceding claim, wherein the seal means between the shroud structure and the turbine casing structure comprises at least one of a labyrinthtype seal and a brush-type seal.

20. An axial flow turbine substantially as herein described with reference to the accompanying drawings. - 16

List of Drawing References A-A. Variable velocity of leakage flow (prior art with low efficiency seal) B-B. Average velocity of leakage flow (prior art with low efficiency seal) C-C. Variable velocity of leakage flow (prior art with high efficiency seal) D-D. Variable velocity of leakage flow (invention with high efficiency seal) E-E. Average velocity of leakage flow (prior art and invention with high efficiency seal) G. Gap L. Length of acoustic passages PH,PL.Steam pressures S. Overall direction of steam flow U. Velocity of seal leakage flow through gap G 1. Turbine passage 2. Steam turbine stage 10. Static blades 11. Static blade support ring 12. Rotor blades 14. Turbine casing 16. Shroud 20. Gap 22. Seals 24. Fins 26. Lands 28. Resonator passages 30. Annular channel or chamber