WO2008087670A1

WO2008087670A1 - Turbogas system multistage compressor

Info

Publication number: WO2008087670A1
Application number: PCT/IT2007/000035
Authority: WO
Inventors: Stefano Cecchi; Massimiliano Maritano; Andrea Silingardi
Original assignee: Ansaldo Energia S.P.A.
Priority date: 2007-01-17
Filing date: 2007-01-17
Publication date: 2008-07-24
Also published as: EP2126367A1; PL2126367T3; EP2126367B1

Abstract

A turbogas system multistage compressor (1) extends along a longitudinal axis (A), and has an inlet (4), and an outlet diffuser (5), between which are located in succession a first stage (10), a number of intermediate stages (11), and a final stage (12); the final stage (12) has an array of fixed blades (15a) on which an airflow impinges; and the fixed blades are substantially equally spaced about the axis (A) of the compressor (1), and are designed to produce a deflection, substantially ranging between approximately 15° and 30°, of the airflow impinging on the fixed blades (15a).

Description

TURBOGAS SYSTEM MULTISTAGE COMPRESSOR

TECHNICAL FIELD The present invention relates to a turbogas system compressor.

BACKGROUND ART

As is known, gas-turbine electric power systems, or turbogas systems, normally comprise a drive (turbo) assembly comprising a compressor, a combustion chamber, a gas turbine, and a generator connected mechanically to the same shaft as the turbine and compressor, and connected to an electric distribution network by a main switch. The compressor provides for drawing in a given amount of air, adequately increasing its pressure, and feeding it to the turbogas system combustion chamber. The combustion chamber burns a certain amount of fuel mixed with the air compressed by the compressor, and feeds the burnt fluid to the turbine, which expands the burnt fluid to convert the potential thermal energy of the fluid to work. Part (roughly half) of the work produced by the turbine is used to drive the compressor, and the rest to drive the generator. The power absorbed by the compressor must therefore be minimized to produce the maximum useful work.

Multistage axial-flow compressors are known comprising an outlet diffuser duct, or so-called diffuser, at the exhaust, the purpose of which is to effectively recover part of the kinetic energy of the airflow from the final stage of the compressor, and convert it to potential pressure energy prior to supplying the combustion chamber. The more efficient the above conversion is, the less power is absorbed by the compressor, which, as stated, is equivalent to an increase in power produced by the turbogas system.

The diffuser of compressors of this sort is normally characterized by an increase in diameter between the inlet section, close to the final stage of the compressor, and the outlet section, close to the combustion chamber, and by a curved shape giving rise to mixed flow. Moreover, close to the outlet section, the diffuser comprises an array of straightening blades for deflecting airflow so that air flows out of the diffuser with a slight tangential velocity component to minimize kinetic energy and maximize potential pressure energy. The blades, however, produce a pressure loss proportional to deflection of the airflow. The increase in section along the diffuser mainly produces a reduction in velocity of the airflow evolving inside the diffuser. More specifically, as the fluid flows along the diffuser, the meridian component of velocity is reduced by the increase in the flow area, and the tangential component is reduced by the increase in the radius of the channel defined by the diffuser (free- vortex law) . The effect of the reduction in the meridian component of velocity is generally predominant, thus resulting in a significant increase in the flow angle (the angle between the airflow direction and the compressor axis) along the channel.

The airflow normally strikes the fixed blades of the final stage at a roughly 40-60° angle, and, as it flows through the fixed blades, undergoes a limited amount of deflection of at most 15° (the flow angle between the inlet and outlet gets smaller) , so that air flows into the straightening blade array ■ at a particularly large, roughly 60-70°, angle. As a result, significant flow- deflecting action is required of the straightening blades, which results in phenomena, such as airflow separation or swirl, which greatly increase pressure losses . DISCLOSURE OF INVENTION

It is an object of the present invention to provide a compressor designed to eliminate the aforementioned drawbacks of the known art. More specifically, it is an object of the invention to provide a compressor designed to minimize pressure losses as air flows inside the diffuser.

According to the present invention, there is provided a turbogas system multistage compressor extending along a longitudinal axis and comprising an inlet, and an outlet diffuser, between which are located in succession a first stage, a number of intermediate stages, and a final stage; the final stage comprising an array of fixed blades substantially equally spaced about the axis of the compressor, and on which an airflow impinges; the compressor being characterized in that the fixed blades of the final stage of the compressor are designed to produce a deflection, substantially ranging between approximately 15° and 30°, of the airflow impinging on the fixed blades .

BRIEF DESCRIPTION OF THE DRAWINGS

A non-limiting embodiment of the present invention will be described by way of example with reference to the accompanying drawings, in which:

Figure 1 shows a schematic section, with parts removed for clarity, of a compressor in accordance with the present invention;

Figure 2 shows a larger-scale schematic section, with parts removed for clarity, of a detail of the Figure 1 compressor;

Figure 3 shows a graph of radial distribution of the flow angle in the compressor according to the present invention; Figure 4 shows a larger-scale, schematic view in perspective, with parts removed for clarity, of a detail of the Figure 2 compressor;

Figure 5 shows a graph of the pressure distribution on a fixed blade of the final stage of the compressor according to the present invention.

BEST MODE FOR CARRYING OUT THE INVENTION Number 1 in Figure 1 indicates a multistage, axial- flow compressor of a substantially known gas-turbine electric power system comprising compressor 1, inside which air flows; a silo-type combustion chamber; a gas turbine; and a generator connected mechanically to the same shaft as the turbine and compressor, and connected to an electric distribution network.

More specifically, compressor 1 extends along a longitudinal axis A, and comprises a rotor 2 rotating about axis A; a blade housing 3; an inlet 4; and an outlet diffuser 5. Blade housing 3 comprises a substantially conical wall 8, and is fixed to inlet 4 and diffuser 5.

A first stage 10, a number of intermediate stages 11, and a final stage 12 are located between inlet 4 and outlet diffuser 5, and each comprise an array of movable blades 14 and an array of fixed blades 15.

Movable blades 14 are equally spaced about axis A to form a ring about a corresponding ring 17 connected directly to rotor 2; while fixed blades 15, also known as stator blades, are equally spaced about axis A, and extend between an outer ring 18 and an inner ring 19.

With reference to Figure 2, inner ring 18 of fixed blades 15a may be housed indifferently inside a seat formed in wall 8 of blade housing 3, or inside a seat formed in an outer shell 26 of diffuser 5, as in the example shown in the drawing.

Blades 15a of final stage 12 of compressor 1 are designed to produce a deflection in airflow substantially ranging between approximately 15° and 30°, and more specifically between approximately 20° and 25°. The term "deflection" is intended to mean the absolute value of the difference between the outflow angle from the array of fixed blades 15a with respect to axis A, and the inflow angle into the array of fixed blades 15a with respect to axis A. Airflow into the array of fixed blades 15a - normally characterized by an inflow angle of 40° to 50° with respect to axis A - is therefore deflected by fixed blades 15a to produce an outflow angle with respect to axis A of around 20°.

With reference to Figure 3, line Ll indicates radial distribution of the flow angle with respect to axis A at a first section I upstream from the array of fixed blades 15a of final stage 12; and line L2 indicates radial distribution of the flow angle with respect to axis A at a second section 11 downstream from the array of fixed blades 15a of final stage 12. The difference between the flow angle at line Ll and the flow angle at line L2 indicates the deflection imposed by fixed blades 15a, which, in the Figure 3 example, is about 20°.

With reference to Figure 4, each fixed blade 15a of final stage 12 comprises a so-called leading edge 21 struck by the airflow; a so-called trailing edge 22, opposite leading edge 21, where the airflow leaves blade 15a; a so-called suction side or outer side 23; and a so- called pressure side or inner side 24.

Each fixed blade 15a has a large camber angle ranging between approximately 45° and 60°, and more specifically between approximately 48° and 54°. The camber angle represents the difference between the leading edge angle and trailing edge angle of fixed blade 15a, where the leading edge angle is intended to mean the angle between a mid-line M or so-called camber line of the profile of fixed blade 15a and axis A of compressor 1 at leading edge 21, and the trailing edge angle is intended to mean the angle between mid-line M of the profile of fixed blade 15a and axis A of compressor 1 at trailing edge 22.

Each blade 15a of final stage 12 also has a limited deflection angle, preferably of less than 10°, where "deflection angle" is intended to mean the difference between the trailing edge angle of fixed blade 15a and the outflow angle with respect to axis A at trailing edge 22.

So designed, fixed blades 15a of final stage 12 guarantee the stability of compressor 1, by preventing it from reaching the limit, so-called pumping, condition. Each fixed blade 15a of final stage 12 also has a so-called front-loaded profile, which means local loading - that is, the absolute value of the local difference between the air pressure on pressure side 24 and the air pressure on suction side 23 - is maximum in a region of blade 15a substantially close to leading edge 21.

Figure 5 shows a graph of pressure distribution on the profile of a fixed blade 15a, where the x axis shows the profile coordinate nondiraensional with respect to the axial chord (projection along axis A of the segment joining leading edge 21 and trailing edge 22) . The 0 value on the x axis therefore corresponds to leading edge 21 of fixed blade 15a, and the 1 value on the x axis corresponds to trailing edge 22 of fixed blade 15a. Line A shows the pressure pattern along suction side 23, line B shows the pressure pattern along pressure side 24, and the local difference between the two curves represents local loading, which is maximum at leading edge 21. With reference to Figure 2, diffuser 5 comprises two shells 26 and 27, which define a curved diverging annular channel 28 of increasing cross section. More specifically, diffuser 5 comprises an inlet section 29 connected to blade housing 3 and close to final stage 12; and an outlet section 30 having a larger inner annular area than inlet section 29, and connected to a cavity 31 in turn connected to one or more preferably silo-type combustion chambers (not shown for the sake of simplicity) . Diffuser 5 also comprises, substantially at outlet section 30, an array of straightening blades 33, each of which is connected to shells 26 and 27, and provides for deflecting airflow so that air flows out of diffuser 5 with a small amount of kinetic energy and a large amount of potential pressure energy.

The increase in section along diffuser 5 mainly produces a reduction in velocity of the evolving airflow inside channel 28. More specifically, the flow velocity inside channel 28 can be divided into a meridian component, by which is meant the vector sum of the radial component and axial component; and a tangential component, by which is meant the velocity component projected onto a plane perpendicular to the meridian plane (containing the meridian component) . As fluid flows along diffuser 5, the meridian component of velocity is reduced by the increase in the flow area, and the tangential component is reduced by the increase in the radius of channel 28 (free-vortex law) . The effect of the reduction in the meridian component of velocity is generally predominant, thus resulting in a significant increase in the flow angle along channel 28.

With reference to Figure 3, line L3 shows radial distribution of the flow angle with respect to axis A at a third section III immediately upstream from the inlet to the array of straightening blades 33. The difference between the flow angle at line L3 and the flow angle at line L2 indicates the variation in flow angle imposed by diffuser 5, and which, in the Figure 3 example, is roughly 25°.

The angle, with respect to axis A, at which air flows into the array of straightening blades 33 is at most 45-50°, to avoid significant deflection action by straightening blades 33. Conversion of the kinetic energy of the airflow into potential pressure energy is therefore improved, by the improved inflow angle into straightening blades 33 eliminating the risk of pressure losses caused by the formation of swirl and separation of the airflow.

The present invention has the following advantages. Firstly, providing a compressor 1, in which fixed blades 15a of final stage 12 are designed to impose significant deflection of the evolving airflow inside compressor 1, improves the efficiency with which kinetic energy is converted to potential pressure energy. The deflection imposed by the fixed blades of the final stage, in fact, improves the angle at which air flows into the array of straightening blades 33, thus minimizing pressure losses caused by the formation of swirl and separation of the airflow.

Secondly, compressor 1 requires only minor, low-cost structural alterations, which provide for a considerable improvement in performance in terms of efficiency of the turbogas system.

Thirdly, designed in accordance with the invention, fixed blades 15a of final stage 12 guarantee stability of compressor 1, in particular by preventing it from reaching the limit, so-called pumping, condition.

And lastly, the profile of fixed blades 15a combines superior aerodynamic performance with construction features enabling them to be installed in currently marketed compressors with no additional machining required. This enables application of the invention not only to new machines as original components, but also to existing machines, with no need to alter the geometry of the compressor, by simply changing the final-stage blades and possibly the inner and outer rings 19 and 18.

Clearly, changes may be made to the compressor as described and illustrated herein without, however, departing from the scope of the accompanying Claims.

Claims

1) A turbogas system multistage compressor (1) extending along a longitudinal axis (A) and comprising an inlet (4), and an outlet diffuser (5), between which are located in succession a first stage (10) , a number of intermediate stages (11) , and a final stage (12) ; the final stage (12) comprising an array of fixed blades (15a) substantially equally spaced about the axis (A) of the compressor (1) , and on which an airflow impinges; the compressor (1) being characterized in that the fixed blades (15a) of the final stage (12) of the compressor are designed to produce a deflection, substantially ranging between approximately 15° and 30°, of the airflow impinging on the fixed blades (15a) .

2) A compressor as claimed in Claim 1, characterized in that the diffuser (5) comprises an array of straightening blades (33) .

3) A compressor as claimed in Claim 2, characterized in that the diffuser (5) comprises an inlet section (29) close to the final stage (12) , and an outlet section (30) connected to one or more combustion chambers; the array of straightening blades (33) being carried by the diffuser (5) substantially at the outlet section (30) . 4) A compressor as claimed in any one of the foregoing Claims, characterized in that the diffuser (5) comprises, internally, a curved diverging annular channel (28) increasing in cross section. 5) A compressor as claimed in any one of the foregoing Claims, characterized in that each fixed blade (15a) of the final stage (12) has a camber angle ranging between approximately 45° and 60 °, and more specifically between approximately 48° and 54°.

6) A compressor as claimed in any one of the foregoing Claims, characterized in that each fixed blade (15a) of the final stage (12) has a deflection angle of less than 10°. 7) A compressor as claimed in any one of the foregoing Claims, characterized in that each fixed blade (15a) of the final stage (12) comprises a leading edge (21), a trailing edge (22) opposite the leading edge (21) , a suction side (23) , and a pressure side (24) ; each fixed blade (15a) being designed so that local blade loading is maximum in a region of the fixed blade (15a) substantially close to the leading edge (21) .

8) A compressor as claimed in any one of the foregoing Claims, characterized in that the fixed blades (15a) of the final stage (12) are designed to prevent the compressor (1) from reaching the limit pumping condition.