WO2012052740A1

WO2012052740A1 - Sealing device for reducing fluid leakage in turbine apparatus

Info

Publication number: WO2012052740A1
Application number: PCT/GB2011/051892
Authority: WO
Inventors: Simon Ian Hogg
Original assignee: University Of Durham
Priority date: 2010-10-18
Filing date: 2011-10-05
Publication date: 2012-04-26

Abstract

A sealing device (122) for reducing fluid leakage between a rotor (104) and a stator (108) of a turbine apparatus (102) is disclosed. The sealing device (122) comprises a series of fluid inlet ports (132) such that a swirl component of fluid flow in a seal inlet region (134) impinges directly into inlet openings of the inlet ports (132), and a series of fluid jets (138) fluid into a leakage region so as to reduce the rate of leakage of fluid through the leakage region. A conduit (136) connects the inlet ports (132) to the jets (138). Kinetic energy resulting from a swirl component of motion of the fluid causes fluid in the conduit (136) to be at a higher pressure than in the seal inlet region (134). A sealing device (122) for reducing fluid leakage between a rotor (104) and a stator (108) of a turbine apparatus (102) is disclosed. The sealing device (122) comprises a series of fluid inlet ports (132) such that a swirl component of fluid flow in a seal inlet region (134) impinges directly into inlet openings of the inlet ports (132), and a series of fluid jets (138) fluid into a leakage region so as to reduce the rate of leakage of fluid through the leakage region. A conduit (136) connects the inlet ports (132) to the jets (138). Kinetic energy resulting from a swirl component of motion of the fluid causes fluid in the conduit (136) to be at a higher pressure than in the seal inlet region (134).

Description

SEALING DEVICE FOR REDUCING FLUID LEAKAGE IN TURBINE

APPARATUS

The present invention relates to a sealing device for reducing leakage of working fluid in a turbine apparatus, and relates particularly, but not exclusively, to a sealing apparatus for minimising leakage of superheated steam between the rotor and stator of steam turbine power generating apparatus .

Figure 1 shows a known type of turbine arrangement used in a high pressure steam turbine power generator. The turbine arrangement 2 has a stator 4 and a rotor 6 mounted for rotation about an axis in order to drive an electricity generator (not shown) . The rotor 6 and stator 4 are provided with opposed sets of blades 8 such that when superheated steam is introduced through an inlet 10 of the stator 4, the heat energy of the steam is transferred to the turbine blades 8, which in turn causes the rotor 6 to rotate about its axis, and the steam is then exhausted from an outlet 12. Any leakage of steam around the external periphery of the turbine blades 8 or between the rotor shaft 14 and shaft glands 16 on the stator 4 fails to transfer its energy to the rotor blades 8 and therefore causes loss of efficiency of the turbine apparatus 2. However, there must also be clearances between the rotor 6 and the stator 4 in order to allow for

dimensional changes caused by temperature variations and mechanical vibrations.

In order to reduce the amount of leakage of steam between the rotor 6 and the stator 4, it is known to provide labyrinth seals at the outer periphery of the turbine blades 8 and at the shaft glands 16. A typical labyrinth seal used as a shaft gland seal is shown in Figures 2A to 2C and consists of a rotating aerodynamic seal used to reduce unwanted leakage flow between the rotor and the stator. The seal consists of a series of tight restrictions 18, 20 between the opposing rotor and stator surfaces and are formed by fins 22 on the stationary and sometimes also on the rotating components, with small clearances (typically less than 1mm) between the tips of the fins 22 and the adj cent sealing surface, as shown in Figure 2C. The leakage flow through each restriction forms a j et which expands into the relatively large volume between the sealing fin 22 forming the jet and the next fin 22 downstream. The rapid

acceleration of the leakage flow to form the jet under each restriction, followed by the uncontrolled expansion into the cavity immediately downstream of the fin 22 (a process in which all of the kinetic energy of the leakage jet is lost) is a very tortuous path for the flow to follow. As a result, a large pressure drop is required to force relatively small quantities of fluid through the labyrinth, thereby

restricting leakage and forming a rotating aerodynamic seal that can operate in the challenging environments found in gas and steam turbines.

Attempts have been made to further reduce leakage flows and thereby improve turbine efficiency by the use of

contacting seal designs such as brush seals as shown in

Figures 3A to 3C. Although brush and leaf seals have good sealing performance, present contacting seal technology is found to be insuff ciently durable for routine long term applications in gas and steam turbine environments. It is found that real turbine phenomena such as high velocity impacts from solid particles carried by the fluid flow, or high levels of swirl in the approaching flow (resulting in chaotic fluttering of seal bristles) are known to cause bristles to break off in use, leading to rapid degradation in sealing performance.

In an attempt to improve seal performance, fluidic seals have been developed in which aerodynamic flow often in a reverse direction in the seal leakage path is used to resist leakage to avoid the use of contact elements such as bristle packs. Such an arrangement is disclosed in US

2009/0297341 and shown in detail in Figure 4. In this arrangement, a et 40 of flow from upstream of a labyrinth seal 48 is introduced part way along the labyrinth leakage path in an upstream direction, in order to provide additional resistance to leakage fluid flow.

The higher the pressure of the fluid supplying the jet in a fluidic seal, the stronger the jet and therefore the greater the potential to create extra resistance against leakage flow. In the arrangement shown in Figure 4, the pressure of the fluid supplying jet 40 and therefore the strength of the jet is limited to the pressure of the flow in the region immediately upstream of the seal.

US 2685429 discloses an arrangement shown in Figures 5A and 5B in which the fluidic sealing jet is provided from a remote pressurised source through inlet port 120 and inclined fluidic jet hole 115. This arrangement enables the sealing jet to be supplied at elevated pressure, but increases the cost and complexity of the turbine apparatus.

Preferred embodiments of the invention seek to overcome one or more of the above disadvantages of known turbine sealing apparatus. According to an aspect of the present invention, there is provided a sealing device for reducing fluid leakage between a rotor and a stator of a turbine apparatus in which at least one rotor is adapted to rotate about a respective axis relative to a stator, the sealing device comprising: - at least one fluid engaging member adapted to engage fluid having a respective first pressure adjacent at least one inlet to a respective leakage region between a said rotor and said stator;

at least one fluid outlet member adapted to direct fluid into said leakage region so as to reduce the rate of leakage of fluid through said leakage region; and

at least one conduit connecting at least one said fluid engaging member to at least one said fluid outlet member, wherein at least one said fluid engaging member is adapted to engage fluid such that kinetic energy resulting from a swirl component of motion of said fluid causes fluid in the corresponding said conduit to be at a respective second pressure, higher than the respective said first pressure.

The present invention is based on the discovery that significantly improved leakage reduction can be achieved at elevated fluidic seal pressures which can be derived from the kinetic energy of flow of working fluid, rather than

requiring an external source of pressurised fluid. By providing at least one said fluid engaging member adapted to engage fluid adjacent an inlet to a leakage region such that fluid in a conduit connected to the fluid inlet member is at a higher pressure than fluid at the fluid engaging member, this provides the advantage that significant improvements in leakage reduction can be achieved without significant

increase in apparatus cost. In addition, by using a swirl component of motion of the fluid to cause the pressure increase, the swirl component of motion is reduced, which in turn provides the advantage of reducing swirl of fluid passing through the leakage path. This in turn reduces destabilising forces on the rotor and further improves turbine performance .

At least one said fluid engaging member may comprise a respective first tube.

At least one said fluid engaging member may comprise a respective concave member.

At least one said fluid engaging member may comprise a respective plate.

At least one said fluid outlet member may be adapted to produce a jet of fluid.

At least one said conduit may comprise fluid reservoir means .

This provides the advantage of enabling the second pressure to be significantly higher than the corresponding first pressure.

According to another aspect of the present invention, there is provided a turbine apparatus comprising :- a stator;

at least one rotor adapted to rotate about a respective axis relative to said stator; and

at least one sealing device as defined above. At least one said sealing device may be mounted to a respective rotor.

At least one said sealing device may be mounted to the stator .

At least one said rotor may comprise a plurality of turbine blades, and at least one said sealing device may be adapted to reduce leakage around the radial periphery of said bl des .

At least one said sealing device may be adapted to reduce leakage between a shaft of at least one said rotor and the stator.

Preferred embodiments of the invention will now be described, by way of example only and not in any limitative sense, with reference to the accompanying drawings, in which : -

Figure 1 is a schematic cross sectional view of a known steam turbine apparatus;

Figure 2A is a perspective view of a stationary gland housing of the turbine cylinder of Figure 1 employing known labyrinth seal technology;

Figure 2B is a perspective part cross sectional view of part of a gland seal ring of Figure 2A co-operating with a rotor (not shown) to form a labyrinth seal; Figure 2C is a schematic cross sectional view (now shown including the rotor) of the labyrinth part of the seal of Figure 2B;

Figures 3A to 3C show a known type of brush seal arrangement ;

Figure 4 is a cross sectional schematic view of a first known fluidic seal;

Figure 5A is a cross sectional view of a second known fluidic seal;

Figure 5B is an enlarged cross sectional view of part of the seal of Figure 5A;

Figure 6 is a schematic cross sectional view of a fluidic seal of a first embodiment of the present invention;

Figures 7A and 7B are detailed views of the fluidic seal of Figure 6;

Figures 8A to 8C illustrate fluid dynamic modelling of the performance of the seal of Figure 6;

Figure 9 is a perspective view, corresponding to Figure 7A, of a second embodiment of the present invention;

Figure 10 is a perspective view, corresponding to

Figure 7A, of a third embodiment of the present invention; and Figure 11 is a partially cutaway perspective view of a fluidic seal of a fourth embodiment of the present invention.

Referring to Figures 6, 7A and 7B, a turbine apparatus 102 has a rotor 104 rotating about a rotation axis 106 relative to a stator 108. The rotor 104 is provided with axially separated sets of rotor blades 110 surrounded by respective shrouds 112, and the stator 108 is provided with stator blades 114 arranged between sets of axially separated rotor blades 110 such that superheated steam entering an inlet (not shown) is directed by the stator blades 114 so as to contact the rotor blades 110 in a direction designed for maximum efficiency of energy transfer. A fluidic seal 116 is provided between a shaft gland 118 on the stator 108 and the rotor shaft 120, and a fluidic seal 122 is provided between the stator 108 and the outer periphery of the shrouds 112 surrounding the rotor blades 110.

The fluidic seal 122 has a three fin non-see-through moving blade tip seal formed by a pair of radially extending fins 124, 126 mounted to the stator 108 and an inclined fin 128 mounted to the stator 108 between the radially extending fins 124, 126, and a series of raised lands 130 on the rotor shroud 112. A fluidic sealing apparatus includes a series of circumferentially arranged fluid inlet ports 132 (Figure 7A) consisting of tubes of generally circular cross section bent through approximately 90° and arranged in a seal inlet region 134 such that a swirl component of fluid flow in the seal inlet region 134 impinges directly onto the inlet openings of the tubes 132. The tubes 132 are connected to a

circumferentially extending reservoir chamber 136 which supplies fluid jets 138 which direct jets of fluid along the surface of the inclined fin 128 of the fluidic seal. The cross sectional area of the ports 132 supplying the reservoir chamber 136 is significantly larger than the cross sectional area of the fluid jets 138, and the circumferential cross sectional area of the reservoir chamber 136 is also large compared with the total cross sectional area of the fluid jets 138 supplied from it. As a result, fluid

velocities within the reservoir chamber 136 are low. The fluid approaching the inlets 132 will be at seal inlet pressure and will have a component of kinetic energy caused by the swirl component of the velocity field in this region. As flow enters an inlet tube 132 and flows along it into the reservoir chamber 136, its velocity is reduced to the low level of the fluid within the reservoir, and as the velocity of the flow reduces, some of the kinetic energy from the swirl is recovered and the pressure of the fluid is

increased. This causes the pressure level within the reservoir chamber 136 to be higher than that in the seal inlet region 134, as will be described in greater detail below .

Figures 8A to 8C show computational fluid dynamic calculations carried out on a three fin labyrinth seal of Figure 6 having a geometry as shown in Figure 8A, with and without a fluidic jet injected along the upstream surface of the inclined fin 138. The calculations were carried out at two separate boundary conditions. These were set to be representative of turbine stage seal flows in high and low pressure steam turbine environments. The simulations used air as the working fluid. The temperature for the air used in the calculations is chosen to give flow densities that are representative of steam at these turbine conditions. For the simulations at high pressure turbine conditions, the pressure at the inlet plane for the calculations (i.e. position 1 in Figure 8A) is set to

and the pressure at the calculation exit plane (position 3 in Figure 8A) is set to P₃=95bar. In the simulations carried out under conditions representative of low pressure steam turbines, the inlet and exit steam pressures are set to

respectively. In the calculations including the fluidic jet 138, the strength of the jet 138 is determined by the value of pressure set at the jet inlet plane, i.e. position 2 in Figure 8A. This pressure is set according to a pressure ratio (PR) defined as PR = (P₂ - P₃) / (Pi - P3) ·

The value PR=1 defines the situation in which the jet inlet pressure P2 is equal to the seal leakage inlet pressure PI, for example as would be the case as shown in Figure 4. If PR is greater than 1, the fluidic jet inlet pressure is greater than the leakage inlet pressure.

The results of the calculations are shown in tables 1 and 2 below. It can be seen that the leakage flow through the labyrinth seal with no fluidic jet (i.e. the data marked by a cross) is compared with the leakage flow with the jet present over a range of jet mass flow rates (defined by the pressure ratio) . The square symbols show the fluidic jet mass flow, and the diamonds indicate the flow calculated at the labyrinth seal inlet 134 (i.e. position 1 in Figure 8A) , and the triangles indicate the total leakage flow {i.e. the fluidic jet flow plus the seal inlet flow) at the exit from the calculation domain, i.e. at position 3 in Figure 8A. It can be seen that the effect of the fluidic jet 138 on seal leakage is compared by comparing the "no jet" crosses with the triangles indicating the total leakage.

Table 1 : Predicted Leakage Flows with and without Fluidic Jets for conditions representative of high pressure steam turbines.

Table 2: Predicted Leakage Flows with and without Fluidic Jets for conditions representative of low pressure steam turbines.

It can be seen that for either high or low pressure steam turbine conditions, for a fluidic jet 138 with the supply pressure equal to the pressure of the fluid upstream of the seal (i.e. PR=1) only relatively modest leakage reductions of just over 5% are achieved by the presence of the jet 138. If, however, the supply pressure of the jet 138 is raised significantly above the seal inlet pressure (PR>1) , larger reductions in seal leakage can be achieved. The highest value of PR calculated was PR=1.4, corresponding to P₂=102bar for the high pressure calculations and P₂=5.1bar for the low pressure calculations, and under this condition, leakage is reduced by 20-25% compared to the "no jet" situation.

A reduction of this order of magnitude has significant commercial benefits. For example, modern high pressure cylinders from steam turbine jets have thermal efficiencies in the region of 90%-95%, and so for an average of 92.5% efficiency, i.e. 7.5% loss, approximately 30% of this 7.5% loss is caused by leakage. In other words, 2.25% of the thermal losses in a typical high pressure cylinder is caused by unwanted leakage flow through the labyrinth seals in the turbine stages and in the shaft gland seals. If this leakage can be reduced by 25% by use of fluidic sealing jets, the thermal efficiency of the cylinder can be improved by

approximately 0.56%. Since a typical steam turbine

generating set {high pressure, intermediate pressure and low pressure turbine cylinders) may generate 600 MW of output, and approximately 30% of this output (i.e. 180 MW) is generated by the high pressure turbine cylinder (for example as shown in Figure 1) , if the efficiency of the high pressure turbine is improved by 0.56% this equates to approximately 1 MW of additional power output.

Referring again to Figures 6, 7A and 7B, the swirl velocity in the tip seal inlet region 134 of a typical steam turbine high pressure stage is usually in the range of 200 m/s to 250 m/s . The density of the steam is also relatively high due to the high pressure level (a typical value of steam density in this region may be 35Kg/m³) . If all of the kinetic energy in the swirl velocity field could be recovered by the system, this would result in reservoir pressures in the region of 5bar to lobar higher than the tip seal inlet pressure, i.e. values of the pressure ratio (PR) in the range 2 to 3. In practice, it is not possible to achieve 100% recovery, but it can be seen that only relatively limited recovery is required in order to achieve values of PR of 1.4 or greater, which enable the significant leakage reduction described above. Even more significant leakage reduction can be achieved by applying fluidic seals embodying the present invention to blade tip seals of low pressure stages, since for these stages absolute values for steam pressure and therefore density are lower than they are for the high pressure stage conditions, but values for swirl velocity at the low pressure tip seal inlet are similar to those for high pressure arrangements. The potential for swirl kinetic energy recovery in relation to pressure drop across the seal is therefore greater for low pressure conditions than for high pressure, and the same percentage of swirl energy recovery in low pressure arrangements will therefore result in even higher values for reservoir pressure ratio (PR) than is the case for high pressure conditions.

Figure 9 shows a further embodiment of the invention applied to a turbine blade tip seal. In the embodiment shown in Figure 9, inlet scoops 232 in the form quarter-spheres are used at the entry to the reservoir 236 instead of the bent tubes shown in Figure 7. In Figure 10, radial planar plates 332 positioned immediately downstream (in the swirl

direction) of the inlets to the reservoir supply ports 340 connected to the reservoir chamber 336 are provided. The flow in the seal inlet region 334 stagnates (i.e. stops) in the swirl direction as it impinges on the plates 332, and this results in increased pressure in the region of the upstream side of the plate 332 because of recovery of some of the swirl kinetic energy. The ports 340 supplying the fluidic seal reservoir 336 draw fluid from this higher pressure region, and this results in elevated pressure in the reservoir 336.

Referring to Figure 11, a fourth embodiment of the invention, applied to the shaft gland seal 116 shown in

Figure 6, is shown. The seal 116 is formed by a stationary gland seal ring 450 having a labyrinth consisting of three long fins 452 and a number of shorter fins 454 cooperating with four raised lands 456 on the rotor, with the same small clearance between stationary and rotating surfaces at all seven of these locations. The first two long fins 452 are inclined radially in the direction of leakage flow, and both of these fins 452 have fluidic jets 458 applied to their upstream surfaces, in a manner similar to the arrangement of Figures 6, 7, 9 and 10. The fluidic jets 458 are supplied through holes through an internal reservoir 460 machined into the stationary gland ring 450, and the reservoir 460 is supplied through ports 462 drilled into the vertical upstream face of the gland seal ring 450. The ports 462 supplying the reservoir 460 are fitted with tubes bent in the opposite direction to the rotating surface of the shaft, so that the swirl energy caused, at least in part, by the rotation of the shaft impinges on the inlet openings of the tubes. In a manner similar to the blade tip seal embodiments described above, this arrangement results in the recovery of some of the kinetic energy from the swirl in the flow in the seal inlet region, resulting in elevated pressures in the

reservoir 460 supplying the fluidic jets. It will be appreciated by persons skilled in the art that the above embodiments have been described by way of example only, and not in any limitative sense, and that various alterations and modifications are possible without departure from the scope of the invention as defined by the appended claims. In particular, although the embodiments described above use a reservoir between the fluidic jet flow inlet ports and the jets, it will be appreciated by persons skilled in the art that these elements could be directly connected to each other. Also, although in the embodiments described above, the fluid jet is applied to the upper surface of an inclined fin, other geometries are possible, for example jets only with no fins, similar to the

arrangement of Figure 5, and/or jets that are supplied through holes drilled internally through the fins exiting at the fin tip such that the jets emerge directly into

the (typically 1mm) clearance between the sealing fin tips and the rotor surface .

Claims

1. A sealing device for reducing fluid leakage between a rotor and a stator of a turbine apparatus in which at least one rotor is adapted to rotate about a respective axis relative to a stator, the sealing device comprising:- at least one fluid engaging member adapted to engage fluid having a respective first pressure adjacent at least one inlet to a respective leakage region between a said rotor and said stator;

at least one fluid outlet member adapted to direct fluid into said leakage region so as to reduce the rate of leakage of fluid through said leakage region,- and

2. A device according to claim 1, wherein at least one said fluid engaging member comprises a respective first tube.

3. A device according to claim 1 or 2 , wherein at least one said fluid engaging member comprises a respective concave member .

4. A device according to any one of the preceding claims, wherein at least one said fluid engaging member comprises a respective plate.

5. A device according to any one of the preceding claims, wherein at least one said fluid outlet member is adapted to produce a jet of fluid.

6. A device according to any one of the preceding claims, wherein at least one said conduit comprises fluid reservoir means .

7. A turbine apparatus comprising :- a stator;

at least one sealing device according to any one of the preceding claims.

8. An apparatus according to claim 7, wherein at least one said sealing device is mounted to a respective rotor.

9. An apparatus according to claim 7 or 8, wherein at least one said sealing device is mounted to the stator.

10. An apparatus according to any one of claims 7 to 9, wherein at least one said rotor comprises a plurality of turbine blades, and at least one said sealing device is adapted to reduce leakage around the radial periphery of said blades .

11. An apparatus according to any one of claims 7 to 10, wherein at least one said sealing device is adapted to reduce leakage between a shaft of at least one said rotor and the stator .