GT2006-90778 A DESIGN TO INCREASE THE STATIC STIFFNESS OF HOLE PATTERN STATOR GAS SEALS

ABSTRACT An analysis is presented which shows that a deep groove located at about 60% along the axial length from the inlet will approximately double the static stiffness of a holepattern-stator, annular gas seal. Test results for a seal using this geometry generally confirm the correctness of this prediction. The groove also produces an increase in leakage by about 4% and a modest decrease in effective damping. INTRODUCTION Injection compressors require comparatively long annular seals with high pressure drops that have a significant impact on rotordynamics. The balance-piston seal for straightthrough compressors usually absorbs the full head rise of the machine. For back-to-back machines, the division-wall seal normally takes about one half of the machine's head rise but deals with higher density gas. Annular seals using smooth rotors and honeycomb (HC) stators have been used since the 1960s in some petrochemical compressors. Conventional aluminum labyrinths were replaced because of the corrosion resistance of a stainless steel honeycomb material. HC surfaces are normally made from high-temperature stainless steels that have been developed as abradable surfaces for tooth-on-rotor labyrinths in aircraft gas turbines. This material is unforgiving in a rubbing condition. In addition, long lead times are frequently involved in securing a custom-manufactured honeycomb seal for a compressor. In response to these circumstances, Yu and Childs [1] tested three aluminum hole-pattern (HP) stator seals. The seal with a hole-area density of 60% performed as well as previously tested honeycomb-stator seals. Moore et al

Rotordynamic Coefficients Measurements Versus Predictions for a HighSpeed Flexure-Pivot Tilting Pad Bearing (Load-BetweenPad Configuration),”

ABSTRACT Experimental dynamic force coefficients are presented for a flexure-pivot-tilting-pad (FPTP), bearing in load-between-pad (LBP) configuration for a range of rotor speeds and bearing unit loadings. The bearing has the following design parameters: 4 pads with pad arc angle 72 o and 50% pivot offset, pad axial length 0.0762 m (3 in), pad radial clearance 0.254 mm (0.010 in), bearing radial clearance 0.1905 mm (0.0075 in), preload 0.25 and shaft nominal diameter of 116.84 mm (4.600 in). Measured dynamic coefficients have been compared with theoretical predictions using an isothermal analysis for a bulkflow Navier-Stokes model. Predictions from two models ---the Reynolds equation and a bulk-flow Navier-Stokes (NS) equation model are compared with experimental, complex dynamic stiffness coefficients (direct and cross-coupled) and show the following results: (i) The real part of the direct dynamic-stiffness coefficients is strongly frequency dependent because of pad inertia, support flexibility, and the effect of fluid inertia. This frequency dependency can be accurately modeled for by adding a direct added mass term to the conventional stiffness/damping matrix model. (ii) Both models underpredict the identified added-mass coefficient (~32 kg), but the bulk-flow NS equations predictions are modestly closer. (iii) The imaginary part of the direct dynamic-stiffness coefficient (leading to direct damping) is a largely linear function of excitation frequency, leading to a constant (frequency independent) direct damping model. (iv) The real part of the cross-coupled dynamic-stiffness coefficients shows larger destabilizing forces than predicted by either model. The direct stiffness and damping coefficients increase with load, while increasing and decreasing with rotor speed, respectively. As expected, a small whirl frequency ratio (WFR) was found of about 0.15, and it decreases with increasing load and increases with increasing speed. The two model predictions for WFR are comparable and both underpredict the measured WFR values. Rotors supported by either conventional tilting PAD bearings or FPTP bearings are customarily modeled by frequency-dependent stiffness and damping matrices, necessitating an iterative calculation for rotordynamic stability. The present results show that adding a constant mass matrix to the FPTP bearing model produces an accurate frequencyindependent model that eliminates the need for iterative rotordynamic stability calculations

GT2006-90374 ROTORDYNAMIC STABILITY PREDICTIONS FOR CENTRIFUGAL COMPRESSORS USING A BULK-FLOW MODEL TO PREDICT IMPELLER SHROUD FORCE AND MOMENT COEFFICIENTS

ABSTRACT An analysis is developed for a compressible bulk-flow model of the leakage path between a centrifugal-compressor impeller's shroud and its housing along the impeller's front and back sides. This development is an extension of analyses performed first by Childs [15] for pump impellers. The bulkflow model is used to predict reaction force and moment coefficients for the impeller shroud. A labyrinth seal code developed by Childs and Scharrer [21] is used to calculate the rotordynamic coefficients developed by the labyrinth seals in the compressor stage and also provides a boundary condition for the shroud calculations. Comparisons between the measured shroud moment coefficients by Yoshida et al. [18] and model predictions show reasonable agreements for the clearance flow and reaction moments. For the conditions considered, low Mach number flow existed in the shroud clearance areas and compressibleflow and incompressible-flow models produced similar predictions. Childs' model predictions for the direct damping and cross-coupled stiffness coefficients of a pump impeller produced reasonable agreement; hence the present model was validated to the extent possible. A rotor model consisting of an overhung impeller stage supported by a nominally cantilevered rotor was analyzed for stability using the present bulk-flow model and an API standard Wachel-formula model Seal rub conditions that arise from surge events and increase the seal clearances are simulated, showing that enlarged clearances increase the preswirl at the seals, thus increasing these seal's destabilizing forces and reducing stability margins. These results are consistent with field experience. Predictions concerning the back shroud indicate that shunt-hole injection mainly acts to enhance stability by changing the flow field of the division wall or balance piston seals, not by influencing the back-shroud's forces or moments. Effective swirl brakes at these seals also serves this purpose

Measurements Versus Predictions for the Rotordynamic Characteristics of a Five-Pad Rocker-Pivot Tilting-Pad Bearing in Load-between-Pad Configuration,”

ABSTRACT Rotordynamic data are presented for a rocker-pivot tilting-pad bearing in a load-between-pad (LBP) configuration for unit loads over the range [345, 3101 kPa] and speeds over the range [4k to 13k rpm]. The bearing was direct lubricated through a leading-edge groove with the following specifications: 5 pads, .282 preload, 60% offset, 57.87° pad arc angle, 101.587 mm (3.9995 in) rotor diameter, .1575 mm (.0062 in) diametral clearance, 60.325 mm (2.375 in) pad length. Dynamic tests were performed over a range of frequencies to investigate frequency effects on the dynamic-stiffness coefficients. Under most test conditions, the direct real parts of the dynamic stiffnesses could be approximated as quadratic functions of the excitation frequency and accounted for with the addition of an added mass matrix to the conventional [K][C] matrix model to produce a frequency-independent [K] [C][M] model. Measured added mass terms in the loaded direction approached 60 kg. At low speeds, "hardening" direct dynamic stiffness coefficients that increased with increasing frequency were obtained that produced negative addedmass terms. No frequency dependency was obtained for the direct damping coefficients. The dynamic experimental results were compared to predictions from a bulk-flow CFD analysis. The static load direction in the tests was y. The direct stiffness coefficients K xx and K yy were slightly over predicted. Measured direct damping coefficients C xx and C yy were insensitive to changes in either load or speed in contrast to predictions of marked C yy sensitivity for changes in the load. Only at the highest test speed of 13000 rpm were the direct damping coefficients adequately predicted. Measurable cross-coupled stiffness coefficients were obtained for the bearings with K xy and K yx being approximately equal in magnitude but opposite in sign ---clearly destabilizing. However, the whirl frequency ratio was found to be zero at all test conditions indicating infinite stability for the bearing