Rotary kinetic fluid motors or pumps – Plural runners having different type flow paths
Reexamination Certificate
2000-10-06
2002-12-03
Verdier, Christopher (Department: 3745)
Rotary kinetic fluid motors or pumps
Plural runners having different type flow paths
C415S181000, C415S198100, C415S199200, C415S199600, C415S208200, C415S211200, C060S804000
Reexamination Certificate
active
06488469
ABSTRACT:
TECHNICAL FIELD
The invention relates to a two stage compressor for a gas turbine engine with a mixed flow first stage, a centrifugal second stage, and an intermediate diffusing duct.
BACKGROUND OF THE ART
Although mixed flow, or diagonal flow, compressors are well known to those skilled in the art of gas turbine engine design, the commercial adoption of mixed flow compressors, particularly in aircraft engines, has been very limited. Most aircraft engines to-date utilise axial flow compressor rotors, centrifugal flow compressors or a combination of both.
The configuration and design of axial flow compressors and centrifugal flow compressors is considered very well known in the art and it is only necessary here to present a general outline of their features and advantages or disadvantages.
Large diameter gas turbine engines are generally constructed with multiple exclusively axial stages generally arranged downstream of an intake fan and by-pass duct. In conventional engines the low pressure axial compressor is mounted on the same shaft as the fan and a low pressure turbine, and the high pressure compressor rotors are mounted on a coaxial high pressure shaft driven by a high pressure turbine. Understandably these multiple stage axial compressors are large and complex machines. They can be justified for their high efficiency in high thrust engine constructions.
Smaller engines are sometimes constructed with a centrifugal compressor as the terminal high pressure stage with a series of axial low pressure stages upstream. The centrifugal rotor, together with surrounding diffuser ducts, considerably increases and dictates the maximum diameter of the engine and forward surface area. However, especially in smaller engine designs, the centrifugal compressor provides high efficiency and reduces the axial length of the engine at the expense of an increase in the radial dimension.
A third common compressor structure includes two centrifugal compressor rotors; however, commercial adoption of this compressor design is very limited. The duct work required to convey compressed air from the first stage centrifugal compressor to the second stage centrifugal compressor rotor is very complex, difficult to manufacture accurately and assemble. Multiple centrifugal stages subject the engine to a significant weight and air drag penalty with prohibitively increased diameter. The increased bulk of the engine envelope and loss of compressor efficiency, through the complex ducting required between first and second stages, has severely limited the adoption of a two stage centrifugal compressor.
The common combination of a centrifugal compressor with axial low pressure stage also suffers from several disadvantages that are generally accepted by designers as inevitable. The engine envelope diameter is dictated by the centrifugal compressor and surrounding diffuser. The axial compressor is often constructed of two or more axial rotors with stator blade assemblies between each axial stage. As a result the number of blades and rotors significantly adds cost to the engine and mechanical complexity. In the current economy for gas turbine engines the overall engine price has been dropping relative to inflation, whereas the cost of materials and engine design costs have been rising. For example, the cost of titanium used in the axial compressor blades has tripled in the last ten years. Due to high design costs for such complex machine parts, the expedient of conservative design practices has resulted in heavier, more robust blades to ensure an adequate safety margin. Therefore, although the design and construction of axial compressors is well known, increases in material costs and concern over the high cost of designing these compressors has led to a desire for a less complex and economically efficient compressor design.
In general, the fewer rotor stages and stator stages that are required in a compressor, the better. Multiple stages and highly complex geometries significantly increase the costs of compressors. To-date, however, experiments in adopting diagonal flow or mixed flow compressor blades have been inconclusive. For example there is no production gas turbine engine available with a mixed flow compressor to date, although experimental results are well documented.
It is well recognised that the cost and reliability of modern gas turbine engines is significantly determined by the number of compressor stages, or acceleration/diffusion operations within the compressor section. It is long recognised that reducing the number of compressor stages will have beneficial effect on the cost of this equipment. Although centrifugal compressor stages compared with axial flow compressors offer lower cost and higher static pressure ratio, centrifugal compressors are slightly less efficient and penalise the design with a larger outer engine envelope diameter than a comparable axial flow compressor. The axial flow compressor of course has a longer axial dimension but suffers from a lower resistance to foreign object damage as well as a lower tolerance to distortion and non-uniformity of inlet airflow distribution. On the other hand using multiple centrifugal compressor stages occasions large aerodynamic losses in the duct work required between the stages as well as significant penalties in weight, engine complexity and manufacturing costs.
Mixed flow or diagonal flow, compressor stages have been recognised in the prior art as providing advantages over both the axial flow and centrifugal flow compressors. For example, a mixed flow compressor has a more rugged design which is superior in foreign object damage resistance to an axial flow compressor and the length of blades enable designers to increase the blade width significantly strengthening the mixed flow blades in comparison to axial flow blades. In addition the part speed benefits of a mixed flow compressor with a significant radius change reduces stress, increases part and bearing life when compared with a large diameter centrifugal compressor rotor. The manufacturing of a mixed flow compressor rotor is somewhat simplified in comparison to a centrifugal compressor and the increase in diameter is significantly lessened.
An example of an adoption of a mixed flow compressor rotor is shown in U.S. Pat. No. 4,678,398 to Dodge et al. In this example the mixed flow compressor rotor is positioned as the first stage upstream of a flow splitter by-pass duct and high pressure axial flow compressor. The mixed flow rotor positioned at the engine inlet with relatively rugged blade construction increases the resistance to foreign object damage and fully utilises the centrifugal effect to propel foreign objects radially outwardly through the by-pass duct thereby protecting the high pressure axial compressor sections downstream.
A significant limitation of the Dodge mixed flow compressor however, is the stat ed objectives which inevitably result in transonic/supersonic air flow speeds in the compressor. The design parameters limit the engine envelope to be comparable to that of an axial flow compressor whereas the design objective is to attain the static pressure ratio, cost and inlet resistance to foreign object damage of a centrifugal compressor. In order to obtain these objectives however, Dodge approaches the mixed flow compressor design by requiring transonic velocities and deals with the need to accommodate sonic shock waves within the structure.
Another example of an attempt to replace several axial compressor stages with a single mixed flow compressor stage as a cost reduction is shown in a paper entitled “Mixed-Flow Compressor Stage Design and Test Results with a Pressure-Ratio of 3:1” Musgrave, D. S. and Plehn, N. J. presented at Gas Turbine Conference and Exhibition, Anaheim, Calif. May 31, to Jun. 4, 1967. In this example the mixed flow compressor stage was designed to be the terminal stage behind an upstream multi-stage axial compressor. The mixed flow compressor stage has an advantage over a conventional centrifugal stage compressor in that the envelope radius is signific
Weir Gary
Youssef Nashed
Pratt & Whitney Canada Corp.
Verdier Christopher
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