EP2090786B1 - Housing structure for stabilising the flow in a flow work machine - Google Patents

Description

The invention relates to a housing with at least one housing structuring for stabilizing the flow in the region of the blade tips of the rotor blades in a turbomachine. Furthermore, the invention relates to a use of the housing in a compressor of a gas turbine. Moreover, the invention relates to a method for stabilizing the flow in the region of the blade tips of the blades in a fluid flow machine by means of the housing.

In a fluid flow machine, especially in a compressor, the pressure of a fluid is continuously increased by means of a rotor with blades and a stator with vanes. The stability of the flow of the fluid in the compressor has a significant influence on the efficiency of the compressor and the life of the blades. Therefore, it is an important goal in the design of compressors to reduce flow instabilities, such as those that occur particularly in the blade tip flow around the blades (slit flow), in order to increase the stability limit of the compressor.

To improve the compressor stability, there are basically two approaches, the active and the passive influencing.

To actively influence the compressor stability include e.g. adjustable guide wheels.

In Fig. 1 is a compressor 1 of a jet engine, not shown, with a compressor housing 2, a compressor duct 3, blades 4 and adjustable vanes 5 with adjusting devices 6 according to the prior art shown schematically. The air 7 flows into the compressor and leaves it as compressed air 8. The functioning of the adjustable guide vanes 5 is characterized in that the angle of attack of the moving blades 4 is tracked with a change in the speed of the compressor 1 in order to adjust the flow conditions, that the stability of the housing and profile boundary layers on the rotor blades 4 is maintained.

However, adjustable vanes are very complicated in their construction. There are many items needed, making the compressor heavier and more expensive. In particular, in jet engines, however, an increase in weight through additional facilities is to be avoided. In addition, the adjustment are susceptible to failure. This increases both the maintenance and maintenance costs.

Furthermore, as an active influence on the compressor stability, recirculation of fluid from the rear stages of the compressor and its injection into the area of the blade tips of the front rotor blades is known.

Fig. 2 shows a compressor 1 with a channel 10 for returning a partial flow from a rear compressor stage to a front compressor stage, which is known in practice. The compressor 1 of a jet engine, not shown, essentially comprises a compressor housing 2, a compressor duct 3, blades 4 and vanes 5. The air 7 flows into the compressor and leaves it as compressed air 8. The channel 10 is arranged between the compressor housing 2 and the bypass inner housing 9 of the jet engine. Behind a downstream compressor stage is a tapping point 11, which opens into the channel 10, which leads to a Einblasestelle 12 in front of an upstream compressor stage. According to the mode of operation of the injection of fluid in front of or above the tips of the leaves 40 of the rotor blades 4 of the first compressor stage energy is introduced into the region of the blade tips 40 of the blades 4 and thus positively influenced the gap flow between the blades 4 of the first compressor stage and the compressor housing 2.

However, the return of fluid can only be realized as a function of speed if control is provided via valves. This is very complex and unreliable. The feedback itself causes hot fluid to be fed from the rear of the compressor to the front. This increases the temperature level in the compressor, and the efficiency drops.

An injection of fluid into the vicinity of the vanes of a fluid flow machine is known, for example, from US Pat DE 103 55 241 A1 known. In the document, individual nozzles are described, which are selectively attached to the housing and through which the blade near areas at different locations air is supplied. Furthermore, the document describes channels which lead through feed chambers and open into the housing in the area of the blade tips of the rotor blades. Through the feed chambers fluid is supplied to the blade row. The supply of the fluid takes place either from a foreign source or is fed back from locations of the fluid flow machine or the entire plant enclosing the turbomachine.

Passive influences on compressor stability include casing treatments in the form of small recesses mounted in front of or above the blade tips of the blades on the circumference of the compressor housing and affecting the blade tip flow.

In Fig. 3 is shown such a passive influence. The compressor 1 of a jet engine, not shown, comprises a compressor housing 2, a compressor duct 3, Blades 4 and Vanes 5. The air 7 flows into the compressor and exits it as compressed air 8. At the leading edge 41 of the blade tip 40 of the first blade 4 is a recess 13. The influence of the flow in the area of the blade tip 40 is characterized achieved that the flow at the downstream end of the recess 13 enters the recess 13 and exits at the upstream end of the recess 13 again and thus circulated. This happens because the pressure at the downstream end of the recess 13 is greater than at the upstream end. This pressure difference causes the local recirculation of the flow. As a result, a small amount of energy is transported into the front region of the blade tip 40. The interaction of the flow recirculation with the blade tip overflow leads to a stabilization of the gap flow and thus of the compressor.

With trained in the axial direction of channels housing are for example from the GB 2 408 546 A , of the EP 0 614 014 A1 , of the EP 0 046 173 A1 , of the US 2003/0152455 A1 and the GB 2 245 312 A known. Furthermore, from the WO 95/34745 A1 a housing with distributed circumferentially wells known.

The recesses are not speed dependent, but can be optimally designed only for one operating point. This means that they are insufficient for stability enhancement under all operating conditions.

The invention is therefore based on the object to provide a housing that increases the compressor stability, is simple in design, has a low weight, works reliably and does not heat the fluid in the fluid flow machine.

This object is achieved according to the invention with a housing having at least one housing structure for stabilizing the flow in the region of the blade tips of the blades in a flow machine according to claim 1 solved. Furthermore, the object is achieved with a method for stabilizing the flow in the region of the blade tips of the rotor blades in a fluid flow machine according to claim 11. Advantageous embodiments of the invention are contained in the subclaims.

According to the invention, the object is achieved in a housing with at least one housing structuring for stabilizing the flow in the area of the blade tips of the rotor blades in a fluid flow machine, wherein the housing structuring is arranged at least in one stage on the inner circumference of the housing. The housing structure is formed as a channel comprising a first end and a second end, wherein the first end opens in the region of the blade tips of a blade ring in the interior of the housing and the second end is closed. The length 1 of the channel at the second end is in a range between a minimum length l _min and a maximum length l _max speed-dependent adjustable.

Static pressure fields, which form on the blades, pull past the channel and cause the air column in the channel to vibrate. At a certain speed, a standing wave is formed in the duct. This creates at the mouth of the channel a pulsating mass flow which stabilizes the flow between the blade tips of the blades and the housing.

Since the length l of the channel at the second end in a range between a minimum length l _min and a maximum length l _{max is} speed-dependent adjustable, the natural frequency of the air column in the channel can be set to any operating state of the fluid flow machine. The housing according to the invention thus combines the advantages of the passive housing structures, which are formed by recesses in the housing (simple design, low weight, no return of hot fluid) with the advantages of active flow control by Verstellleiträder (speed-dependent control). The l length adjustable channel allows future compressors to be designed with higher loaded rotor tips, which can be achieved, for example, by reducing rotor blade counts. This leads to a weight and cost reduction.

The arrangement is simple in construction and works reliably. The weight of the compressor is not increased by the channel. Also, the temperature of the flow in the compressor is not increased, as for example in a return of fluid.

Preferably, the channel has a taper at the first end. The rejuvenation enhances the effect of the pulsating mass flow.

In a preferred embodiment of the invention, the channel is rectilinear at least in the range between l _min and l _max and has a constant cross-section in this region, wherein at the second end of the channel in the longitudinal direction of the channel in the range between l _min and l _max movable Piston is arranged. The movable piston makes it easy to adjust the length l of the channel. The arrangement of the piston is very easy to implement, requires few parts and has a lower weight than a Verstellleitradsystem according to the prior art.

The position of the piston may be controllable by means of an electric, hydraulic or pneumatic drive. As an electric drive, for example, a stepper motor comes into question. These drives are reliable and can be easily installed in the fluid flow machine.

In a preferred embodiment, the channel is arranged substantially radially to the inner periphery of the housing. Such a channel can be easily made, for example, by a core during casting or by subsequent drilling.

In an alternative embodiment, the channel is arranged at an angle to the longitudinal axis of the housing. Also such a channel can be easily made by a casting core or drilling.

In a further alternative embodiment, the channel is arcuately curved outside the range between l _min and l _max . This shape allows a length of the channel that exceeds the thickness of the housing wall.

In a further alternative embodiment, the channel is arcuately curved in the region of the first end and in the region between l _min and l _max parallel to the longitudinal axis of the housing. This arrangement is advantageous when the piston is to move in the axial direction.

According to the invention, the position of the first end of the channel is between the trailing edge of the blade and a distance of 1.3 times the axial chord length l _ax of the blade at the blade tip as measured from the trailing edge of the blade. This position range is optimal for stabilizing the flow between the blade tips of the blades and the housing.

Preferably, the housing is used in a compressor of a gas turbine. In a gas turbine, compressor stability is particularly important. The compressor is exposed to high pressures due to pressure and temperature and should not be additionally burdened by flow instabilities.

Furthermore, the solution of the problem in a method for stabilizing the flow in the area of the blade tips of the blades in a fluid flow machine by means of the housing, which forms a static pressure field on each blade. The static pressure field travels past the first end of the channel during rotation of the blade and causes the fluid column in the channel to vibrate, creating a standing wave in the channel. which creates a pulsating mass flow at the first end of the channel.

This method is based on a simple principle and is very reliable. The method results in that an increase in the compressor pumping limit can be achieved without adversely affecting the compressor efficiency, and that the increase in the pumping limit for the entire speed range of the compressor can be optimally utilized.

Preferably, the standing wave is generated by tuning the natural frequency of the fluid column to the blade follower frequency so that the natural frequency of the fluid column is a multiple of the blade follower frequency of the blades. The tuning of the natural frequency of the fluid column enables an improvement of the stability in all operating areas of the fluid flow machine.

Furthermore, the natural frequency of the fluid column can be adjusted as a function of speed by adjusting the length l of the channel. By adjusting the length l of the channel, a simple adjustment of the natural frequency of the air column in the channel is possible.

berechnet werden, wobei

• l die Länge des Kanals,
• k eine beliebige natürliche Zahl,
• κ der Isentropenexponent,
• R die spezifische Gaskonstante,
• n die aerodynamische Drehzahl des Verdichterrotors
  und
• z Schaufelzahl des Laufschaufelkranzes
  sind.

The length l of the channel can be calculated using the formula $$l\left(n\right)=\left(\frac{1}{2}k+\frac{1}{4}\right)\frac{\sqrt{\mathrm{\kappa }R}}{\mathit{nz}}$$

be calculated, where

• l the length of the channel,
• k is any natural number,
• κ isentropic exponent,
• R is the specific gas constant,
• n the aerodynamic speed of the compressor rotor
  and
• z Number of blades of the blade ring
  are.

This formula allows a precise determination of the optimal length 1 of the channel for each operating range. The adaptation of the length l of the channel as a function of the aerodynamic speed of the compressor leads to the fact that always forms a defined flow state in the channel, whereby the channel is maximally effective and the compressor stability is maximally increased. Since the aerodynamic speed is available to the engine computer, the control of the length l of the channel is very easy and reliable to implement. Since the length l of the channel is optimally adapted to all speeds, an improvement in the compressor efficiency is also to be expected.

berechnet werden, wobei

l_min

die Mindestlänge des Kanals,

k_min

eine beliebige natürliche Zahl,

κ

der Isentropenexponent,

R

die spezifische Gaskonstante,

n_max

die maximale aerodynamische Drehzahl des Verdichterrotors und

z

Schaufelzahl des Laufschaufelkranzes

sind.In particular, the minimum length l _{min of} the channel with the formula $$l_{\min }=\left(\frac{1}{2}{k}_{\min }+\frac{1}{4}\right)\frac{\sqrt{\mathrm{\kappa }R}}{{\mathit{n}}_{\mathrm{Max}}z}{\mathrm{with\; K}}_{\min }\le \mathrm{k}$$

be calculated, where

l _min

the minimum length of the channel,

k _min

any natural number,

κ

the isentropic exponent,

R

the specific gas constant,

n _max

the maximum aerodynamic speed of the compressor rotor and

z

Blade number of the blade rim

are.

By determining the minimum length, it can be ensured that the channel length is not set too short.

berechnet werden, wobei

l_max

die maximale Länge des Kanals,

k

eine beliebige natürliche Zahl,

κ

der Isentropenexponent,

R

die spezifische Gaskonstante,

n_min

die minimale aerodynamische Drehzahl des Verdichterrotors und

z

Schaufelzahl des Laufschaufelkranzes

sind.In addition, the maximum length l _{max of} the channel with the formula $$l_{\mathrm{Max}}=\left(\frac{1}{2}k+\frac{1}{4}\right)\frac{\sqrt{\mathrm{\kappa }R}}{{\mathit{n}}_{\min }z}$$

be calculated, where

l _max

the maximum length of the channel,

k

any natural number,

κ

the isentropic exponent,

R

the specific gas constant,

n _min

the minimum aerodynamic speed of the compressor rotor and

z

Blade number of the blade rim

are.

Determining the maximum length of the channel ensures that the channel is not set too long.

Fig. 1

ein Verdichtergehäuse mit verstellbaren Leitschaufeln nach dem Stand der Technik,

Fig. 2

ein Verdichtergehäuse mit einem Kanal zur Rückführung von Fluid nach dem Stand der Technik,

Fig. 3

ein Verdichtergehäuse mit einer Gehäusestrukturierung (Vertiefung) nach dem Stand der Technik,

Fig. 4a

eine schematische Ansicht eines ersten Ausführungsbeispiels eines Kanals in einem erfindungsgemäßen Gehäuse,

Fig. 4b

eine schematische Ansicht eines zweiten Ausführungsbeispiels eines Kanals in einem erfindungsgemäßen Gehäuse,

Fig. 4c

eine schematische Ansicht eines dritten Ausführungsbeispiels eines Kanals in einem erfindungsgemäßen Gehäuse,

Fig. 4d

eine schematische Ansicht eines vierten Ausführungsbeispiels eines Kanals in einem erfindungsgemäßen Gehäuse,

Fig. 5

eine vergrößerte schematische Ansicht des dritten Ausführungsbeispiels und

Fig. 6

eine vergrößerte schematische Ansicht des vierten Ausführungsbeispiels.

In the following, several embodiments of the invention will be explained in more detail with reference to figures. Show it:

Fig. 1

a compressor housing with adjustable guide vanes according to the prior art,

Fig. 2

a compressor housing with a channel for the return of fluid according to the prior art,

Fig. 3

a compressor housing with a housing structuring (recess) according to the prior art,

Fig. 4a

a schematic view of a first embodiment of a channel in a housing according to the invention,

Fig. 4b

a schematic view of a second embodiment of a channel in a housing according to the invention,

Fig. 4c

a schematic view of a third embodiment of a channel in a housing according to the invention,

Fig. 4d

a schematic view of a fourth embodiment of a channel in a housing according to the invention,

Fig. 5

an enlarged schematic view of the third embodiment and

Fig. 6

an enlarged schematic view of the fourth embodiment.

The Fig. 4a, 4b, 4c and 4d each show a part of a housing, which is designed as a compressor housing 2 in a jet engine, not shown, a rotor blade 4 and a channel 20 with a piston 30th

The compressor housing 2 encloses the compressor duct 3, which is circular in cross-section. The compressor duct 3 contains rotor blades arranged radially on a shaft or rotor disk, not shown. In the Fig. 4a-d only one blade 4 is shown in each case. The rotor blade 4 has a blade tip 40, an upstream leading edge 41 and a downstream trailing edge 42. Between the blade tip 40 of the blade 4 and the compressor housing 2, there is a gap 43 in the compressor duct 3.

In the compressor housing 2, the channel 20 is arranged in the region of the blade tip 40 of the blade 4. The channel 20 has a first end 21 and a second end 22. The first end 21 of the channel 20 opens in the region of the blade tip 40 of the blade 4 in the compressor passage 3 and in the gap 43. The second end 22 of the channel 20 is clearly spaced from the first end 21 and closed by the adjustable piston 30. This channel 20 can be arbitrary in number, extent and shape in the axial direction and in the circumferential direction. It can be any number of channels 20 of the four embodiments, which in the Fig. 4a-d shown are arranged on the circumference of the compressor housing 2. In addition, further channels 20 may be arranged on the moving blades of further compressor stages.

In Fig. 4a the first embodiment of the channel 20 is shown in the compressor housing 2. The channel 20 extends in a straight line and radially to the inner periphery of the compressor housing 2. The first end 21 of the channel 20 opens in the downstream region of the blade tip 40 of the blade 4 in the gap 43 between the blade tip 40 and the compressor housing.

In Fig. 4b the second embodiment of the channel 20 is shown in the compressor housing 2. The channel 20 is rectilinear and is inclined at an acute angle to the non-illustrated longitudinal axis of the compressor housing 2, wherein the apex of the angle points in the flow direction. The first end 21 of the channel 20 opens in the upstream region of the blade tip 40 of the blade 4 in the gap 43 between the blade tip 40 and the compressor housing. 2

In Fig. 4c the third embodiment of the channel 20 is shown in the compressor housing 2. The channel 20 extends straight at the second end 22 and radially to the inner periphery of the compressor housing 2. The first end 21 of the channel 20 extends arcuately, tapers in the direction of the compressor passage 3 and flows upstream of the leading edge 41 of the blade 4 just before the gap 43rd between the blade tip 40 and the compressor housing 2 in the compressor duct 3rd

In Fig. 4d the fourth embodiment of the channel 20 is shown in the compressor housing 2. The channel 20 extends only at the second end 22 in a straight line and parallel to the longitudinal axis of the compressor housing 2, not shown. The first end 21 of the channel 20 extends arcuately, tapers in the direction of the compressor passage 3 and flows upstream of the leading edge 41 of the blade 4 just before the gap 43 between the blade tip 40 and the compressor housing 2 in the compressor passage 3rd

In Fig. 5 is the third embodiment according to Fig. 4c shown enlarged for the channel 20 in the compressor housing 2. The compressor housing 2, with the compressor channel 3, the rotor blade 4, a guide blade 5 and the channel 20 with the piston 30 can essentially be seen again. The air flow 7 is fed to the compressor stage formed by the rotor blade 4 and the guide blade 5. The compressed air stream 8 exits the compressor stage.

The channel 20 includes the first end 21 and the second end 22 in which the piston 30 is located. The blade 4 includes the blade tip 40, the leading edge 41 and the trailing edge 42. Between the blade tip 40 and the compressor housing 2 is the gap 43. The axial distance between the leading edge 41 and the trailing edge 42 at the blade tip 40 is the chord length l _ax , The position of the channel 20 may be between the trailing edge 42 of the blade 4 and 1.3 times the axial chord length l _ax measured from the trailing edge 42. This area is in Fig. 5 characterized by l _pos .

In Fig. 6 is the fourth embodiment according to Fig. 4d shown enlarged for the channel 20 in the compressor housing 2. Essentially, the compressor housing 2, with the compressor duct 3, the rotor blade 4 and the duct 20 with the piston 30 are again visible. The duct 20 comprises a center line 23, the first end 21 and the second end 22, in which the piston 30 is located. The blade 4 comprises the blade tip 40, the leading edge 41 and the trailing edge 42. The gap 43 is located between the blade tip 40 and the compressor housing 2.

mit

l_max

maximale Länge des Kanals 20,

k

beliebige natürliche Zahl,

κ

Isentropenexponent,

R

spezifische Gaskonstante,

n_min

minimale aerodynamische Drehzahl,

z

Schaufelzahl des Laufschaufelkranzes.

In the radial direction, the shape of the channel 20 may be arbitrary (see. Fig. 4a-d ). The length l of the channel 20 is against not arbitrary. The maximum length l _{max of} the channel 20 is defined by the minimum aerodynamic speed n _{min of} the compressor at which the channel 20 is to act, cf. Equation (1). Equation (1) is chosen so that the maximum length l _{max of} the channel 20 is formed such that a standing wave is generated in the channel 20. The maximum length l _max lies on the center line 23 of the channel 20. $$l_{\mathrm{Max}}=\left(\frac{1}{2}k+\frac{1}{4}\right)\frac{\sqrt{\mathrm{\kappa }R}}{{\mathit{n}}_{\min }z}$$

With

l _max

maximum length of the channel 20,

k

any natural number,

κ

isentropic exponent,

R

specific gas constant,

n _min

minimum aerodynamic speed,

z

Blade number of the blade rim.

The aerodynamic speed n results from the mechanical compressor speed divided by the root of the compressor inlet temperature. This aerodynamic speed n is available to the engine computer. k is any natural number (0, 1, 2, ...) over which the length l of the channel 20 can be increased without negatively affecting its effect. κ is the isentropic exponent, R is the specific gas constant, and z is the blade number of the blade ring on which the channel 20 acts on the flow.

mit

l

Länge des Kanals 20,

k

beliebige natürliche Zahl,

κ

Isentropenexponent,

R

spezifische Gaskonstante,

n

aerodynamische Drehzahl,

z

Schaufelzahl des Laufschaufelkranzes.

When changing the compressor speed, the length l of the channel 20 is varied according to equation (2) in dependence on the aerodynamic speed n. $$l\left(n\right)=\left(\frac{1}{2}k+\frac{1}{4}\right)\frac{\sqrt{\mathrm{\kappa }R}}{\mathit{nz}}$$

With

l

Length of the channel 20,

k

any natural number,

κ

isentropic exponent,

R

specific gas constant,

n

aerodynamic speed,

z

Blade number of the blade rim.

mit

l_min

Mindestlänge des Kanals 20,

k

beliebige natürliche Zahl,

κ

Isentropenexponent,

R

spezifische Gaskonstante,

n_max

maximale aerodynamische Drehzahl,

z

Schaufelzahl des Laufschaufelkranzes.

The minimum length l _{min of} the channel 20 is dependent on the maximum aerodynamic speed n _max , with which the compressor is operated, see. Equation (3). It should be noted that k _min ≤ k. $$l_{\min }=\left(\frac{1}{2}{k}_{\min }+\frac{1}{4}\right)\frac{\sqrt{\mathrm{\kappa }R}}{{\mathit{n}}_{\mathrm{Max}}z}$$

With

l _min

Minimum length of the channel 20,

k

any natural number,

κ

isentropic exponent,

R

specific gas constant,

n _max

maximum aerodynamic speed,

z

Blade number of the blade rim.

The adjustment of the length l of the channel 20 is realized by means of the piston 30, which moves in the part of the channel 20 which lies between the minimum length l _min and the maximum length l _{max of} the channel 20. The piston 30 serves to change the length l of the channel 20 in such a way that, according to equation (2), the length l appropriate to the current aerodynamic rotational speed n is set. The travel of the piston 30 is s = l _max -l _min . In the region of the travel s of the piston 30, the channel 20 is designed such that the piston 30 fits, ie the channel 20 is rectilinear in this region and has a constant cross section. The method of the piston 30 is controlled by the aerodynamic speed n of the compressor and realized by means of a suitable mechanism, for example electrically (eg by a stepping motor), hydraulically or pneumatically. In operation, the length l of the channel 20 must be selected so that a standing wave is generated in it. In order to set the length l, the movable piston 30 is moved between the minimum length l _min and the maximum length l _{max of} the channel 20. The travel s of the piston 30 is, as described above, dependent on the aerodynamic speed n. In order for the flow to be optimally influenced, two quantities must be matched to one another. These are the blade rate of the blade ring to be controlled and the volume of the channel 20. Each blade 4 of the blade ring surrounds a static pressure field. This moves past the first end 21 of the channel 20 and vibrates the air column in the channel 20. The piston 30 makes it possible to vary the volume of the channel 20. Thus, the natural frequency of the air column in the channel 20 is changed.

If the volume is now adjusted to the compressor speed in such a way that the blade repetition frequency coincides with a multiple of the natural frequency of the air column in the channel 20, a resonance occurs and a standing wave of maximum amplitude is formed in the channel 20. At the second end 22 of the channel 20, the standing wave on the piston 30 shows a node, and the speed is zero. At the first end 21 of the channel 20, the standing wave has a belly. Thus, here the air column oscillates maximally. A pulsating mass flow, which stabilizes the flow in the region of the blade tips 40 of the rotor blades 4, is formed at the first end 21 of the channel 20.

LIST OF REFERENCE NUMBERS

1

compressor

2

compressor housing

3

compressor duct

4

blade

5

vane

6

adjustment

7

air

8th

air

9

Bypass inner housing

10

channel

11

tap

12

injection point

13

deepening

20

channel

21

first end

22

second end

23

center line

_pos

position range

30

piston

l _min

minimum length

l _max

maximum length

s

Piston adjuster track

40

blade tip

41

leading edge

42

trailing edge

43

gap

l _ax

chord length

Claims (16)

1. Casing (2) with at least one casing structuring for stabilizing the flow in the area of the blade tips of the blades (4) in a fluid-flow machine, where the casing structuring is provided at least in one stage on the inner circumference of the casing (2), where the casing structuring is designed as a duct (20), which includes a first end (21) and a second end (22), with the first end (21) issuing into the interior of the casing (2) in the area of the blade tips of a blade ring and with the second end (22) being closed, characterized in that the length l of the duct (20) at the second end (22) is speed-dependably adjustable in a range between a minimum length l_min. and a maximum length l_max.
2. Casing (2) in accordance with Claim 1, characterized in that the duct (20) is provided with a diminution at the first end (21).
3. Casing (2) in accordance with Claim 1 or 2, characterized in that the duct (20) is rectilinear at least in the range between l_min. and l_max. and has a constant cross-section in this range, with a piston (30) which is movable in the longitudinal direction of the duct (20) in the range between l_min. and l_max. being provided at the second end (22) of the duct (20).
4. Casing (2) in accordance with Claim 3, characterized in that the position of the piston (30) is controllable by means of an electric, hydraulic or pneumatic drive.
5. Casing (2) in accordance with one of the Claims 1 to 4, characterized in that the duct (20) is arranged essentially radially to the inner circumference of the casing (2).
6. Casing (2) in accordance with one of the Claims 1 to 4, characterized in that the duct (20) is arranged angularly to the longitudinal axis of the casing (2).
7. Casing (2) in accordance with one of the Claims 1 to 4, characterized in that the duct (20) is arcuately curved outside of the range between l_min. and l_max.
8. Casing (2) in accordance with one of the Claims 1 to 4, characterized in that the duct (20) is arcuately curved in the area of the first end (21) and parallel to the longitudinal axis of the casing (2) in the range between l_min. and l_max.
9. Casing (2) in accordance with one of the Claims 1 to 8, characterized in that the position of an intersection point of a duct axis of the duct (20) and the inner circumference of the casing (2) is between the trailing edge (42) of the blade (4) and a distance measured from the trailing edge (42) of the blade (4) which is 1.3-times the axial chord length l_ax of the blade (4) at the blade tip (40).
10. Use of the casing (2) in accordance with one of the Claims 1 to 9 in a compressor (1) of a gas turbine.
11. Method for stabilizing the flow in the area of the blade tips (40) of the rotor blades (4) in a fluid-flow machine by means of the casing (2) in accordance with one of the Claims 1 to 10, where a static pressure field forms on each blade (4), and where the static pressure field moves past the first end of the duct (20) during rotation of the blade (4) and excites vibrations of the fluid column in the duct (20), with a standing wave being produced in the duct (20) by which a pulsating mass flow is created at the first end (21) of the duct (20).
12. Method in accordance with Claim 11, characterized in that the standing wave is produced in that the natural frequency of the fluid column is matched such to the blade passing frequency that the natural frequency of the fluid column concurs with a multiple of the blade passing frequency of the blades (4).
13. Method in accordance with Claim 12, characterized in that the natural frequency of the fluid column is speed-dependently set by adjusting the length l of the duct (20).
14. Method in accordance with Claim 13, characterized in that the length l of the duct (20) is calculated using the formula $$\mathrm{l}\left(\mathrm{n}\right)=\left(\frac{\mathrm{1}}{\mathrm{2}}\mathrm{k}+\frac{\mathrm{1}}{\mathrm{4}}\right)\frac{\sqrt{\mathrm{\kappa }\mathrm{R}}}{\mathrm{nz}},$$

    

    with

    l being the length of the duct,

    k any natural number,

    κ the isentropic exponent,

    R the specific gas constant,

    n the aerodynamic speed of the compressor rotor, and

    z the number of blades of the blade ring.
15. Method in accordance with Claim 14, characterized in that the minimum length l_min. of the duct (20) is calculated using the formula $${\mathrm{l}}_{\min \mathrm{.}}=\left(\frac{\mathrm{1}}{\mathrm{2}}{\mathrm{k}}_{\min \mathrm{.}}+\frac{\mathrm{1}}{\mathrm{4}}\right)\frac{\sqrt{\mathrm{\kappa }\mathrm{R}}}{{\mathrm{n}}_{\max }\mathrm{.}\mathrm{z}},{\mathrm{with\; k}}_{\min \mathrm{.}}\le \mathrm{k},$$

    

    and with

    l_min. being the minimum length of the duct,

    k_min. any natural number,

    κ the isentropic exponent,

    R the specific gas constant,

    n_max. the maximum aerodynamic speed of the compressor rotor, and

    z the number of blades of the blade ring.
16. Method in accordance with Claim 14, characterized in that the maximum length l_max. of the duct (20) is calculated using the formula $${\mathrm{l}}_{\max \mathrm{.}}=\left(\frac{\mathrm{1}}{\mathrm{2}}\mathrm{k}+\frac{\mathrm{1}}{\mathrm{4}}\right)\frac{\sqrt{\mathrm{\kappa }\mathrm{R}}}{{\mathrm{n}}_{\min \mathrm{.}}\mathrm{z}},$$

    

    with

    l_max. being the maximum length of the duct,

    k any natural number,

    κ the isentropic exponent,

    R the specific gas constant,

    n_min. the minimum aerodynamic speed of the compressor rotor, and

    z the number of blades of the blade ring.

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