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The fast fracture mechanical strength data at 1200°C (Average as-processed 4-point bend strength σavg=524 MPa, Weibull modulus m=19) was used to predict the life and reliability of the ceramic vane using the CARES/LIFE code [10]. The probability of failure (Pf) was calculated using stresses predicted by FEA. The calculations assumed volume flaws and Shetty's mixed-mode micro-cracking criterion. The Pf of ST5+ vane for average-inlet and hot-streak condition was 7.5x10-14 and 2.6x10-10, respectively. The stress and life predictions indicate that the stresses are lower and the reliability is higher for the new aero-optimized design. Figure 7. Steady state thermal stress (MPa) in ST5+ vane for hot-streak conditions CERAMIC INTEGRALLY BLADED ROTOR Due to high centrifugal loading, the ceramic rotor design is much more challenging than designing ceramic vanes. The high thermal loading in the hot gas path demands heat resistant blades, while high rotating speed requires more mechanical strength at the disk bore. To ensure component reliability, a hybrid ceramic rotor design, using ceramic blades and a metallic disc, has been pursued in the past [11,12]. It offers a good solution for both the challenging design requirement of hot gas path and metallic disk in a thermally less aggressive but mechanically more demanding environment. This is particularly important for medium to large size turbine rotors, as ceramic component reliability is dependent on the stressed volume. Despite the attractiveness of low stressed volume, hybrid ceramic rotors suffer from several major drawbacks. The first is related to the high contact stress at the ceramic blade root where the ceramic blade is attached to the metallic disc. Although a compliant layer material has been developed to act as a buffer between the blade and the disk and hence reduce contact stress, the durability of this compliant layer material has not been demonstrated for the required engine life. The second drawback is the cost associated with the machining of blade roots so that good contact can be achieved between the blade and the disc. A third drawback is the difficulty in controlling the blade tip clearance, which is important for small gas turbine engines, where even a small absolute tip clearance can result in high tip loss due to the relatively small rotor dimension. As mentioned before, one determining factor of whether the microturbine can be successful in the gas turbine market is its cost competitiveness. This important consideration combined with the relative small engine size directed the initial ST5+ compressor turbine design to a ceramic IBR. Such a concept has been pursued in the past [13,14]. During the first phase of rotor design, most of the design efforts have been focused on two areas: disk geometry optimization and blade FOD resistance assessment. Preliminary results of these two efforts are presented next. The blade and outer rim shape were determined by the aerodynamic optimization. The profile of the disk is designed to minimize stress and accommodate an attachment method between the ceramic disk and a metallic shaft. Although for stress considerations, it is desirable to have a solid ceramic IBR, there are other design considerations that require a bored IBR. For example, there is a need for cooling air passage between the front and rear disk surfaces to keep both surfaces at equal temperatures and to create a pressurized cavity so that the disk rear cavity can be air-sealed. This is most conveniently realized by the disk centre bore. To save computing time, a two-dimensional, axisymmetric model was employed to optimize disk geometry. The ST5 disk and thermal boundary conditions were used as a baseline condition and rotor blades were represented by pressure loading at the rim, as aerodynamic design of the blade was done in parallel. The four main control parameters are bore diameter, hub width, web width and the bend radii between the web and the hub. There are two locations of high stress: one at the bore and another at the web. The stress at the bore is primarily circumferential, while the stress at the web is mainly radial. It was found that the maximum stress at the bore is not sensitive to the bore diameter for the range of practical importance. The maximum principal stress during steady state shown in Figure 8 is still high and further optimization is needed. Figure 8. Maximum principal stress (MPa) in turbine disc A 3D finite element (FE) model was created using the current 2D geometry to account for the bending stress induced by the blade. Since the rotor has 27 blades, a 1/27th model 179 161 141 127 110 93 76 59 42 25 8 -10 -27 102 Copyright © 2002 by ASME 496 454 412 369 327 285 242 199 157 115 72 30 -12PDF Image | ADVANCED MICROTURBINE SYSTEMS Final Report for Tasks 1 Through 4 and Task 6
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