Patent Application
20120272663
This invention relates generally to gas turbine engine turbines and more particularly to gas turbine engine compressors.
A centrifugal compressor is a known type of compressor used in a gas turbine engine. It includes a rotating impeller, to impart kinetic energy to a gas flowing through it, with a high swirl level (i.e. tangential velocity) at the impeller's exit. A stationary diffuser surrounds the impeller and deswirler system follows the diffuser. Collectively, the diffuser and the deswirler system function to convert the high kinetic energy received from the impeller into the more useful form of static pressure, turn the air flow in a meridional plane by 90 degrees or more, from an outward radial direction to a generally axial direction, and reduce the high tangential velocity component to a small value (typically less than 5 degrees).
A compressor's pressure ratio (“PR”) is a ratio of the pressure at the exit to the pressure at the inlet. Increasing PR drives engines to higher thermodynamic efficiency, reducing specific fuel consumption (“SFC”). Increased PR is a function of increased tip speed, which in turn is a function of impeller rotational speed and impeller tip radius. Current technology high pressure ratio and high efficiency designs are often at or near mechanical limits for maximum allowable tip speeds, thus limiting any further increase in PR that could be achieved by increased speed. Further increase in pressure ratio may be obtained by larger radius centrifugal stages, but larger radius stages have an undesirable negative impact on engine weight and cost goals.
Accordingly, there is a need for a centrifugal impeller with high pressure ratios, high efficiency, and adequate stall margin, at a lower impeller tip speed.
This need is addressed by the present invention, which provides a centrifugal compressor stage with an impeller having separate exducer and inducer blades, with a row of stationary stator vanes disposed between them.
According to one aspect of the invention, a compressor assembly for a gas turbine engine is provided. The compressor assembly includes: a rotatable impeller with forward and aft ends, including: an annular hub defining a generally concave-curved cross-sectional inner flowpath surface at its radially outer periphery, the inner flowpath surface extending between an inlet and an exit; an annular array of airfoil-shaped inducer blades extending radially outward from the inner flowpath surface near the forward end; and an annular array of exducer blades extending outward from the inner flowpath surface, the array of exducer blades axially spaced apart from the array of inducer blades; a non-rotating shroud assembly surrounding the impeller and including a convex-curved outer flowpath surface, wherein the inner and outer flowpath surfaces cooperate to define a primary flowpath past the blades and the stator vanes; and an annular array of airfoil-shaped stator vanes extending radially inward from the outer flowpath surface into a space between the inducer and exducer blades.
The invention may be best understood by reference to the following description taken in conjunction with the accompanying drawing figures in which:
Referring to the drawings wherein identical reference numerals denote the same elements throughout the various views,
The centrifugal stage 24 includes an impeller 28 mounted for rotation with the outer shaft 18. The impeller 28 includes a hub 30 with axially forward and aft ends 32 and 34 and a central bore 36 passing axially therethrough. In the illustrated example the hub 30 is constructed from a forward section 40 and an aft section 42 secured together so they rotate as a unit, but they could be part of an integral assembly. An outer portion of the hub 30 defines a generally concave-curved inner flowpath surface 44. The inner flowpath surface 44 extends in a generally longitudinal direction towards the forward end 32 and extends in a generally radial direction near the aft end 34. The impeller 28 includes an inlet 46 and a downstream exit 48.
An annular array of inducer blades 50 extend radially outward from the inner flowpath surface 44, near the inlet 46. Each of the inducer blades 50 is airfoil-shaped (best seen in
An annular array of exducer blades 60 extend radially outward from the inner flowpath surface 44. The exducer blades 60 are axially spaced away from the inducer blades 50 and generally follow the curve of the inner flowpath surface 44 in side elevation view. Each of the exducer blades 60 has a root 62, a tip 64, and leading and trailing edges 66 and 68, respectively. The exducer blades 60 are configured in terms of their dimensions, cross-sectional shape, orientation, spacing, and other parameters to provide an incremental pressure increase to the air flowing past them as the impeller 28 rotates.
An annular shroud assembly 70 surrounds the impeller 28. The shroud assembly 70 defines a generally convex-curved outer flowpath surface 72 that closely surrounds the tips of the inducer 50 and exducer blades 60. Together the inner and outer flowpath surfaces 44 and 72 define a primary flowpath “F” through the centrifugal stage 24.
As depicted in
The cross-sectional area of the flowpath F is variable along the longitudinal length of the impeller 28. Specifically, flowpath F has a first cross-sectional area 80 defined between the inner and outer flowpath surfaces 44′ and 72 at the inlet 46′. The flowpath F has a second cross-sectional area 82 defined between inner and outer flowpath surfaces 44′ and 72 at apex 78. The flowpath F has a third cross-sectional area 84 defined downstream from apex 78, between the second portion 76 and the outer flowpath surface 72. The second cross-sectional area 82 is smaller than first cross-sectional area 80 but equal to or slightly larger than the third area 84. Finally, the impeller exit 48′ has a fourth cross-sectional area 86 defined between the inner and outer flowpath surfaces 44′ and 72 at the exit 48′. This cross-sectional area 86 is smaller than cross-sectional areas 80, 82, and 84.
The impeller work input that produces the required pressure ratio is known to be the product of the wheel linear metal speed and the air, or fluid, turning in the tangential direction. The wheel linear metal speed is the radius multiplied by rotational speed. This relationship applies locally as well globally (i.e. on an average basis from inlet to exit). A higher air, or fluid, turning is a result of a higher blade curvature. The radius of outer flowpath surface 72 is substantially greater than the radius of inner flowpath surface 44′. This is particularity noted at the near the inlet 46′. Therefore, it is common in prior art practice to balance the work input across the span of the inducer blade 50′ at the leading edge by increasing the blade surface curvature (camber) near the inner flowpath surface. However, increased curvature produces a higher adverse pressure gradient (i.e. diffusion) that the flow may not be able to sustain, causing flow separation that may be local or global. If the flow separation is global it may be massive and extend to the exit. Flow separation is known to reduce both efficiency and stall margin (i.e. the safe operating flow range at speed). The inner flowpath surface 44 incorporating the apex 78 (referred to as a “fan-type” flowpath) essentially increases the average inducer metal speed at the location of the apex 78, so that blade curvature can be reduced, allowing production of efficient and higher-pressure ratio inducer blades 50′ than an arrangement with a wholly-concave low hub.
Referring back to
Introduction of the stator vanes 88 between the inducer blades 50 and exducer blades 60 reduces the flow swirl into the exducer blades 60, thereby increasing the amount of work input required by the exducer blades 60. This allows for a lower tip speed (and therefore lower impeller diameter) than a conventional impeller for the same pressure ratio. Reducing the exducer blades' inlet swirl increases their relative inlet flow angle (i.e. the angle between the longitudinal direction and the inlet flow direction).
This in turn increases the allowable blade angle 13 near the exducer blades' leading edges 64 and the wrap angle occupied by each exducer blade 60. This essentially increases the combined physical length of the inducer blades 50 and exducer blades 60 available for optimum diffusion. The result is efficiently increasing the pressure ratio more than what would be accomplished without the stator vanes 88.
Referring back to
Analysis indicates that the concepts described above will allow the construction of high pressure ratio, high specific speed centrifugal compressor stages with lower tip speeds, weight, and cost than would otherwise be required.
The foregoing has described a centrifugal compressor apparatus for a gas turbine engine. While specific embodiments of the present invention have been described, it will be apparent to those skilled in the art that various modifications thereto can be made without departing from the spirit and scope of the invention. Accordingly, the foregoing description of the preferred embodiment of the invention and the best mode for practicing the invention are provided for the purpose of illustration only and not for the purpose of limitation, the invention being defined by the claims.
