Test Results for Liquid “Damper” Seals Using a Round-Hole Roughness Pattern for the Stators

1999 ◽  
Vol 121 (1) ◽  
pp. 42-49 ◽  
Author(s):  
Dara W. Childs ◽  
Patrice Fayolle
Keyword(s):  
Friction Factor ◽  
Round Hole ◽  
Test Results ◽  
Partial Explanation ◽  
Rotordynamic Coefficients ◽  
Stiffness Coefficients ◽  
Friction Factors ◽  
Direct Stiffness ◽  
Factor Data ◽  
Pressure Driven Flow

Test results are reviewed for two annular liquid seals (L = 34.9 mm; D = 76.5 mm) at two clearances (.1 and .12 mm). The seal stators use hole-pattern-roughened stators that are identical except for hole depths of .28 and 2.0 mm. Tests are conducted at three speeds out to 24,600 rpm and three pressures out to 68 bars. Test data consist of leakage rates and rotordynamic coefficients at centered and eccentric positions with static eccentricity ratios out to 0.5. Test results are consistent with expectations in regard to the reduction of cross-coupled stiffness coefficients due to stator roughness. However, the measured direct stiffness coefficients were unexpectedly low. A partial explanation for these results is provided by measured friction factor data which show an increase in the friction factors for pressure-driven flow with an increase in clearance. A prediction model for rotordynamic coefficients, incorporating the friction-factor data, predicted a substantial loss in direct stiffness but could not explain the very low (or negative) values that were measured. The model did explain the measured drop in cross coupled stiffness (k) and provides an alternative explanation to observed reductions in k values; specifically, an increase in the friction factor with increasing clearance causes a reduction in k irrespective of any parallel reduction in the average circumferential velocity.

Rotordynamic Characteristics of a Flexure Pivot Pad Bearing With an Active and Locked Integral Squeeze Film Damper

2012 ◽  
Author(s):  
Jeff Agnew ◽  
Dara Childs
Keyword(s):  
Added Mass ◽  
Squeeze Film ◽  
Test Results ◽  
Squeeze Film Damper ◽  
Rotordynamic Coefficients ◽  
Damping Coefficients ◽  
Stiffness Coefficients ◽  
Direct Stiffness ◽  
Active Damper ◽  
Rotordynamic Coefficient

Measured rotordynamic coefficients are presented for a flexure-pivot-pad journal bearing (FPJB) in a load-between-pad configuration with: (1) an active, and (2) locked integral squeeze film damper (ISFD). Prior rotordynamic-coefficient test results have been presented for FPJBs (alone), and rotor-response results have been presented for rotors supported by FPJBS with ISFDs; however, these are the first rotordynamic-coefficient test results for FPJBs with ISFDs. A multi-frequency dynamic testing regime is employed. For both bearing configurations, quadratic curve fits provide good representation of the real portions of the dynamic-stiffness coefficients yielding a direct stiffness and a direct added-mass coefficient. The imaginary portions are well represented by linear curve fits, implying constant, frequency-independent direct-damping coefficients. Direct stiffness coefficients are ∼50% lower for the active-damper configuration, and direct damping coefficients are only modestly lower. The combination of ∼50% reduction in direct stiffness with a modest drop in direct damping indicates a very effective squeeze-film damper application. Added-mass coefficients are normally lower for the active-damper configuration, and all coefficient trends (for changes in loading and shaft speed) are “flatter” for the active flexure pivot-pad damper bearing. The measured rotordynamic coefficients are used to calculate the whirl frequency ratio and indicate high stability for both bearing configurations.

Rotordynamic Performance of a Negative-Swirl Brake for a Tooth-on-Stator Labyrinth Seal

2014 ◽  
Author(s):  
Dara W. Childs ◽  
James E. Mclean ◽  
Min Zhang ◽  
Stephen P. Arthur
Keyword(s):  
Labyrinth Seal ◽  
Test Results ◽  
Labyrinth Seals ◽  
Rotordynamic Coefficients ◽  
Stiffness Coefficients ◽  
Shaft Rotation ◽  
Direct Stiffness ◽  
The Cross ◽  
Inlet Conditions ◽  
The Stability

In the late 1970’s, Benckert and Wachter (Technical University Stuttgart) tested labyrinth seals using air as the test media and measured direct and cross-coupled stiffness coefficients. They reported the following results: (1) Fluid pre-swirl in the direction of shaft rotation creates destabilizing cross-coupled stiffness coefficients, and (2) Effective swirl brakes at the inlet to the seal can markedly reduce the cross-coupled stiffness coefficients, in many cases reducing them to zero. In recent years, “negative-swirl” swirl brakes have been employed that attempt to reverse the circumferential direction of inlet flow, changing the sign of the cross-coupled stiffness coefficients and creating stabilizing stiffness forces. This study presents test results for a 16-tooth labyrinth seal with positive inlet preswirl (in the direction of shaft rotation) for the following inlet conditions: (1) No swirl brakes, (2) Straight, conventional swirl brakes, and (3) Negative-swirl swirl brakes. The negative-swirl swirl-brake designs were developed based on CFD predictions. Tests were conducted at 10.2, 15.35, and 20.2 krpm with 70 bars of inlet pressure for pressure ratios of 0.3, 0.4, 0.5. Test results include leakage and rotordynamic coefficients. In terms of leakage, the negative-swirl brake configuration leaked the least, followed by the conventional brake, followed by the no-brake design. Normalized to the negative-swirl brake configuration, the conventional-brake and no-brake configurations mass flow rate were greater, respectively, by factors of 1.04 and 1.09. The direct stiffness coefficients are negative but small, consistent with past experience. The conventional swirl brake drops the destabilizing cross-coupled stiffness coefficients k by a factor of about 0.8 as compared to the no-brake results. The negative-swirl brake produces a change in sign of k with an appreciable magnitude; hence, the stability of forwardly-precessing modes would be enhanced. In descending order, the direct damping coefficients C are: no-swirl, negative-swirl, conventional-swirl. Normalized in terms of the no-swirl case, C for the negative and conventional brake designs are, respectively, 0.7 and 0.6 smaller. The effective damping Ceff combines the effect of k and C. Ceff is large and positive for the negative-swirl configuration and near zero for the no-brake and conventional-brake designs. The present results for a negative-brake design are very encouraging for both eye-packing seals (where conventional swirl brakes have been previously employed) and division-wall and balance-piston seals where negative shunt injection has been employed.

Friction-Factor Characteristics for Narrow Channels With Honeycomb Surfaces

1992 ◽  
Vol 114 (4) ◽  
pp. 714-721 ◽  
Author(s):  
T. W. Ha ◽  
G. L. Morrison ◽  
D. W. Childs
Keyword(s):  
Reynolds Number ◽  
Friction Factor ◽  
Reynolds Numbers ◽  
Narrow Channel ◽  
Test Results ◽  
Jump Phenomenon ◽  
Narrow Channels ◽  
Friction Factors ◽  
Factor Data

The experimental determination of friction-factors for the flow of air in a narrow channel lined with various honeycomb geometries has been carried out. Test results show that, generally, the friction-factor is nearly constant or slightly decreases as the Reynolds number increases, a characteristic common to turbulent flow in pipes. However, in some test geometries this trend is remarkably different. The friction factor dramatically drops and then rises as the Reynolds number increases. This phenomenon can be characterized as a “friction-factor jump.” Further investigations of the acoustic spectrum and friction-factor measurements for a broad range of Reynolds numbers indicate that the “friction-factor jump” phenomenon is accompanied by an onset of a normal mode resonance excited coherent flow fluctuation structure, which occurs at Reynolds number of the order of 104. The purpose of this paper is not to present the friction-factor data but to explain the friction-factor-jump phenomenon and friction-factor characteristics.

scholarly journals Annular Honeycomb Seals: Test Results for Leakage and Rotordynamic Coefficients; Comparisons to Labyrinth and Smooth Configurations

1989 ◽  
Vol 111 (2) ◽  
pp. 293-300 ◽  
Author(s):  
D. Childs ◽  
D. Elrod ◽  
K. Hale
Keyword(s):  
Labyrinth Seal ◽  
Test Results ◽  
Rotordynamic Coefficients ◽  
Damping Coefficients ◽  
Stiffness Coefficients ◽  
Shaft Rotation ◽  
Honeycomb Seal ◽  
Cell Dimensions ◽  
Honeycomb Seals ◽  
Rotordynamic Coefficient

Test results are presented for leakage and rotordynamic coefficients for seven honeycomb seals. All seals have the same radius, length, and clearance; however, the cell depths and diameters are varied. Rotordynamic data, which are presented, consist of the direct and cross-coupled stiffness coefficients and the direct damping coefficients. The rotordynamic-coefficient data show a considerable sensitivity to changes in cell dimensions; however, no clear trends are identifiable. Comparisons of test data for the honeycomb seals with labyrinth and smooth annular seals shows the honeycomb seal had the best sealing (minimum leakage) performance, followed in order by the labyrinth and smooth seals. For prerotated fluids entering the seal, in the direction of shaft rotation, the honeycomb seal has the best rotordynamic stability followed in order by the labyrinth and smooth. For no prerotation, or fluid prerotation against shaft rotation, the labyrinth seal has the best rotordynamic stability followed in order by the smooth and honeycomb seals.

Measured Static and Rotordynamic Characteristics of a Smooth-Stator/Grooved-Rotor Liquid Annular Seal

2018 ◽  
Vol 140 (10) ◽  
Author(s):  
J. Alex Moreland ◽  
Dara W. Childs ◽  
Joshua T. Bullock
Keyword(s):  
Radial Clearance ◽  
Test Results ◽  
Dynamic Features ◽  
Axial Pressure ◽  
Pressure Drops ◽  
Rotordynamic Coefficients ◽  
Annular Seal ◽  
Direct Stiffness ◽  
Static Eccentricity ◽  
Dynamic Variables

Electric submersible pumps (ESPs) utilize grooved-rotor/smooth-stator (SS/GR) seals to reduce leakage and break up contaminants within the pumped fluid. Additionally, due to their decreased surface area (when compared to a smooth seal), grooved seals decrease the chance of seizure in the case of rotor-stator rubs. Despite their use in industry, the literature does not contain rotordynamic measurements for smooth-stator/circumferentially grooved-rotor liquid annular seals. This paper presents test results consisting of leakage measurements and rotordynamic coefficients for a SS/GR liquid annular sdeal. Both static and dynamic variables are investigated for various imposed preswirl ratios (PSRs), static eccentricity ratios (0–0.8), axial pressure drops (2–8 bars), and running speeds (2–8 krpm). The seals' static and dynamic features are compared to those of a smooth seal with the same length, diameter, and minimum radial clearance. Results show that the grooves reduce leakage at lower speeds (less than 5 krpm) and higher axial pressure drops, but does little at higher speeds. The grooved seal's direct stiffness is generally negative, which would be detrimental to pump rotordynamics. As expected, increasing preswirl increases the magnitude of cross-coupled stiffness and increases the whirl frequency ratio (WFR). When compared to the smooth seal, the grooved seal has smaller effective damping coefficients, indicative of poorer stability characteristics.

An Annular Gas Seal Analysis Using Empirical Entrance and Exit Region Friction Factors

1990 ◽  
Vol 112 (2) ◽  
pp. 196-204 ◽  
Author(s):  
D. A. Elrod ◽  
D. W. Childs ◽  
C. C. Nelson
Keyword(s):  
Shear Stress ◽  
Friction Factor ◽  
Factor Models ◽  
Exit Region ◽  
Dynamic Coefficients ◽  
Friction Factors ◽  
Direct Stiffness ◽  
Gas Seals ◽  
Rotor Dynamic ◽  
Better Than

Wall shear stress results from stationary-rotor flow tests of five annular gas seals are used to develop entrance and exit region friction factor models. The friction factor models are used in a bulk-flow seal analysis which predicts leakage and rotor-dynamic coefficients. The predictions of the analysis are compared to experimental results and to the predictions of Nelson’s analysis (1985). The comparisons are for smooth-rotor seals with smooth and honeycomb-stators. The present analysis predicts the destabilizing cross-coupled stiffness of a seal better than Nelson’s analysis. Both analyses predict direct damping well and direct stiffness poorly.

scholarly journals Friction-Factor Data for Flat-Plate Tests of Smooth and Honeycomb Surfaces

1992 ◽  
Vol 114 (4) ◽  
pp. 722-729 ◽  
Author(s):  
T. W. Ha ◽  
Dara W. Childs
Keyword(s):  
Flat Plate ◽  
Friction Factor ◽  
Smooth Surface ◽  
Factor Model ◽  
Test Apparatus ◽  
Cell Width ◽  
Plate Test ◽  
Friction Factors ◽  
Line Flow ◽  
Factor Data

Friction-factors for honeycomb surfaces are measured with a flat plate tester. The flat plate test apparatus is described and a method is discussed for determining the friction-factor experimentally. The friction-factor model is developed for the flat plate test based on the Fanno Line Flow. The comparisons of the friction-factor are plotted for smooth surface and twelve-honeycomb surfaces with three-clearances, 6.9 bar to 17.9 bar range of inlet pressure, and 5,000 to 130,000 range of the Reynolds number. The optimum geometries for the maximum friction-factor are found as a function of cell width to cell depth and clearance to cell width ratios.

Test Results for Turbulent Annular Seals, Using Smooth Rotors and Helically Grooved Stators

1990 ◽  
Vol 112 (2) ◽  
pp. 254-258 ◽  
Author(s):  
D. W. Childs ◽  
S. A. Nolan ◽  
J. J. Kilgore
Keyword(s):  
Test Results ◽  
Force Coefficients ◽  
Friction Factors ◽  
Theoretical Predictions ◽  
Direct Stiffness ◽  
Annular Seals

Test results, consisting of leakage data, friction factors, and rotordynamic force coefficients, are presented for seven annular seals using smooth rotors and helically-grooved stators. All seals have the same nominal clearances and groove depths. The helix angles vary from zero (circumferential grooving) to 70 deg. The number of grooves and the leakage rates increase steadily with increasing helix angles. Helically-grooved stators leak more than smooth seals for helix angles greater than 30 deg. The effective direct stiffness of the seals first decreases and then increases with increasing helix angles. Contrary to theoretical predictions (Kim and Childs 1987), the effective net damping was relatively insensitive to changes in the helix angles.

Experimental Study of the Static and Dynamic Characteristics of a Long Smooth Seal With Two-Phase, Mainly Air Mixtures

2017 ◽  
Vol 139 (12) ◽  
Author(s):  
Min Zhang ◽  
James E. Mclean ◽  
Dara W. Childs
Keyword(s):  
Gas Turbines ◽  
Dynamic Stiffness ◽  
Test Fluid ◽  
Liquid Volume ◽  
Test Cases ◽  
Test Results ◽  
Two Phase ◽  
Rotordynamic Coefficients ◽  
Stiffness Coefficients ◽  
Annular Seal

A two-phase annular seal stand (2PASS) has been developed at the Turbomachinery Laboratory of Texas A&M University to measure the leakage and rotordynamic coefficients of division wall or balance-piston annular seals in centrifugal compressors. 2PASS was modified from an existing pure-air annular seal test rig. A special mixer has been designed to inject the oil into the compressed air, aiming to make a homogenous air-rich mixture. Test results are presented for a smooth seal with an inner diameter D of 89.306 mm, a radial clearance Cr of 0.188 mm, and a length-to-diameter ratio (L/D) of 0.65. The test fluid is a mixture of air and silicone oil (PSF-5cSt). Tests are conducted with inlet liquid volume fraction (LVF) = 0%, 2%, 5%, and 8%, shaft speed ω = 10, 15, and 20 krpm, and pressure ratio (PR) = 0.43, 0.5, and 0.57. The test seal is concentric with the shaft (centered), and the inlet pressure is 62.1 bar. Complex dynamic-stiffness coefficients are measured for the seal. The real parts are generally too dependent on excitation frequency Ω to be modeled by constant stiffness and virtual-mass coefficients. The direct real dynamic-stiffness coefficients are denoted as KΩ; the cross-coupled real dynamic-stiffness coefficients are denoted as kΩ. The imaginary parts of the dynamic-stiffness coefficients are modeled by frequency-independent direct C and cross-coupled c damping coefficients. Test results show that the leakage and rotordynamic coefficients are remarkable impacted by changes in inlet LVF. Leakage mass flow rate m˙ drops slightly as inlet LVF increases from zero to 2% and then increases with further increasing inlet LVF to 8%. As inlet LVF increases from zero to 8%, KΩ generally decreases except it increases as inlet LVF increases from zero to 2% when PR = 0.43. kΩ increases virtually with increasing inlet LVF from zero to 2%. As inlet LVF further increases to 8%, kΩ decreases or remains unchanged. C increases as inlet LVF increases; however, its rate of increase drops significantly at inlet LVF = 2%. Effective damping Ceff combines the stabilizing impact of C and the destabilizing impact of kΩ. Ceff is negative (destabilizing) for lower Ω values and becomes more destabilizing as inlet LVF increases from zero to 2%. It then becomes less destabilizing as inlet LVF is further increased to 8%. Measured m˙ and rotordynamic coefficients are compared with predictions from XLHseal_mix, a program developed by San Andrés (2011, “Rotordynamic Force Coefficients of Bubbly Mixture Annular Pressure Seals,” ASME J. Eng. Gas Turbines Power, 134(2), p. 022503) based on a bulk-flow model, using the Moody wall-friction model while assuming constant temperature and a homogenous mixture. Predicted m˙ values are close to measurements when inlet LVF = 0% and 2% and are smaller than test results by about 17% when inlet LVF = 5% and 8%. As with measurements, predicted m˙ drops slightly as inlet LVF increases from zero to 2% and then increases with increasing inlet LVF further to 8%. However, in the inlet LVF range of 2–8%, the predicted effects of inlet LVF on m˙ are weaker than measurements. XLHseal_mix poorly predicts KΩ in most test cases. For all test cases, predicted KΩ decreases as inlet LVF increases from zero to 8%. The increase of KΩ induced by increasing inlet LVF from zero to 2% at PR = 0.43 is not predicted. C is reasonably predicted, and predicted C values are consistently smaller than measured results by 14–34%. Both predicted and measured C increase as inlet LVF increases. kΩ and Ceff are predicted adequately at pure-air conditions, but not at most mainly air conditions. The significant increase of kΩ induced by changing inlet LVF from zero to 2% is predicted. As inlet LVF increases from 2% to 8%, predicted kΩ continues increasing versus that measured kΩ typically decreases. As with measurements, increasing inlet LVF from zero to 2% decreases the predicted negative values of Ceff, making the test seal more destabilizing. However, as inlet LVF increases further to 8%, the predicted negative values of Ceff drop versus measured values increase. For high inlet LVF values (5% and 8%), the predicted negative values of Ceff are smaller than measurements. So, the seal is more stabilizing than predicted for high inlet LVF cases.

Measurements Versus Predictions for Rotordynamic and Leakage Characteristics of a Convergent-Tapered, Honeycomb-Stator/Smooth-Rotor Annular Gas Seal

2008 ◽  
Author(s):  
Daniel E. van der Velde ◽  
Dara W. Childs
Keyword(s):  
Control Volume ◽  
Excitation Frequency ◽  
Test Fluid ◽  
Low Frequencies ◽  
Rotordynamic Coefficients ◽  
Stiffness Coefficients ◽  
Direct Stiffness ◽  
Honeycomb Seals ◽  
Excitation Frequencies ◽  
The Cross

Measured results are presented for rotordynamic coefficients and leakage rates for two honeycomb-stator seal geometries, a convergent-tapered honeycomb seals (CTHC) and a constant-clearance honeycomb seals (CCHC) tested by Sprowl and Childs in 2007. The rotor diameter was 114.3 mm (4.500 in). The CTHC seals had inlet and exit clearances of 0.334 and 0.204 mm, respectively. The CCHC seal had a constant clearance of 0.204 mm. Honeycomb cells had depths of 3.175 mm (0.125 in) and widths of 0.79 mm (0.031 in). Measurements are reported with air as the test fluid, zero preswirl, ω = 20,200 rpm, a supply pressure of 69 bar (1,000 psi) and supply temperature of 18°C (64.4°F) for both seal geometries. The test pressure ratios are 0.5 for the CCHC seal, and 0.46 for the CTHC seal. The tapered seal leaks about 20% more than the constant-clearance seal. Measured and predicted dynamic coefficients are strong functions of excitation frequency. The measured direct stiffness coefficient was higher for the tapered seal at all excitation frequencies, including a projection to zero frequency, where the CCHC seal was on the order of −2MN/m versus roughly +13MN/m for the tapered seal. The CTHC seal had higher cross-coupled stiffness coefficients than the CCHC seal at all excitation frequencies. The CCHC and CTHC seals had comparable direct damping out to ∼80 Hz. For higher excitation frequencies, the CTHC seal had larger direct damping values. The effective damping Ceff combines the positive effect of direct damping and the destabilizing effect of cross-coupled-stiffness coefficients. It is negative at low frequencies and becomes positive for higher frequencies. The frequency at which it changes sign is called the cross-over frequency. The CCHC had a lower cross-over frequency (better from a stability viewpoint) and higher Ceff values out to ∼80 Hz. At higher excitation frequencies from ∼120Hz onward, the tapered seal has higher effective damping values. Kleynhans and Childs’ 1997 two-control-volume model did a generally good job of predicting the direct stiffness coefficients of both seals. It closely predicted the cross-coupled stiffness coefficients for the CCHC seal but substantially under predicted the values for the CTHC seal. It under predicted the direct damping for the CCHC seal at frequencies below ∼120Hz, but did a good job for higher frequencies. It under predicted direct damping for the CTHC seal at all frequencies. For the CCHC seal, the model did a good job of predicting Ceff at all frequencies and also accurately predicted the cross-over frequency. For the CTHC seal, the model accurately predicted the cross-over frequency but over predicted Ceff below the cross-over frequency (the seal was more destabilizing than predicted) and under predicted Ceff at higher frequencies.

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