1 / 38

ASME Turbo Expo 2011: Power for Land, Sea and Air June 6-10, 2011, Vancouver, BC

ASME Turbo Expo 2011: Power for Land, Sea and Air June 6-10, 2011, Vancouver, BC. Rotordynamic Force Coefficients of Bubbly Mixture Annular Pressure Seals. Luis San Andres Mast-Childs Tribology Professor Turbomachinery Laboratory Texas A&M University. ASME GT2011-45264.

declan
Download Presentation

ASME Turbo Expo 2011: Power for Land, Sea and Air June 6-10, 2011, Vancouver, BC

An Image/Link below is provided (as is) to download presentation Download Policy: Content on the Website is provided to you AS IS for your information and personal use and may not be sold / licensed / shared on other websites without getting consent from its author. Content is provided to you AS IS for your information and personal use only. Download presentation by click this link. While downloading, if for some reason you are not able to download a presentation, the publisher may have deleted the file from their server. During download, if you can't get a presentation, the file might be deleted by the publisher.

E N D

Presentation Transcript

  1. ASME Turbo Expo 2011: Power for Land, Sea and Air June 6-10, 2011, Vancouver, BC Rotordynamic Force Coefficients of Bubbly Mixture Annular Pressure Seals Luis San Andres Mast-Childs Tribology Professor Turbomachinery Laboratory Texas A&M University ASME GT2011-45264 Accepted for publication J Eng. Gas Turb. Power Presentation available at http://rotorlab.tamu.edu Supported by TAMU Turbomachinery Laboratory (Prof. D. Childs)

  2. Annular Pressure Seals Inter-stage seal Impeller eye or neck ring seal Balance piston seal Radial seals (annular, labyrinth or honeycomb) separate regions of high pressure and low pressure and their principal function is to minimize the leakage (secondary flow); thus improving the overall efficiency of a rotating machine extracting or delivering power to a fluid. Seals in a Multistage Centrifugal Pump or Compressor

  3. Annular Pressure Seals The dynamic force response of pressure seals has a primary influence on the stability response of high-performance turbomachinery. Annular seals, although geometrically similar to plain journal bearings, show a flow structure dominated by turbulence and fluid inertia effects. Operating characteristics unique to seals are the * large axial pressure gradients, * large clearance to radius ratio (R/c) < 500, while * the axial development of the circumferential velocity determines the magnitude of cross-coupled (hydrodynamic) forces. Childs, D., 1993, Turbomachinery Rotordynamics: Phenomena, Modeling, and Analysis, John Wiley & Sons, Inc., New York, New Yor, Chap. 4.

  4. Seals and rotordynamics Due to their relative position within a rotor-bearing system, seals modify the system dynamic behavior. Seals typically "see" large amplitude rotor motions. This is particularly important in back-to-back compressors and long-flexible multiple stage pumps Straight-Through and Back-to-back Compressors and 1st Mode Shapes Childs, D., 1993, Turbomachinery Rotordynamics: Phenomena, Modeling, and Analysis, John Wiley & Sons, Inc., New York, New Yor, Chap. 4.

  5. Force Coefficients in Annular Seals Seal reaction forces are functions of the fluid properties, flow regime, operating conditions and geometry. For small amplitudes of rotor lateral motion: forces are linearized with stiffness, damping and inertia force coefficients:

  6. Annular Pressure Seals Intentionally roughened stator surfaces (macro texturing) reduce the impact of undesirable cross-coupled dynamic forces and improve seal stability. Annular seals acting as Lomakin bearings have potential as support elements (damping bearings) in high speed compressors and pumps. Childs, D., and Vance, J., 1997, “Annular Gas Seals and Rotordynamics of Compressors and Turbines”, Proc. of the 26th Turbomachinery Symposium, Texas A&M University, Houston, TX, September, pp. 201-220

  7. Bubbly Mixture Annular Pressure Seals Justification Seals operate with either liquids or gases, but not both…… As oil fields deplete compressors work off-design with liquid in gas mixtures, mostly inhomogeneous. Similarly, oil compression station pumps operate with gas in liquid mixtures The flow condition affects compressor or pump overall efficiency and reliability. Little is known about seals operating under 2-phase conditions, except that the mixture affects seal leakage, power loss and rotordynamic force coefficients; perhaps even inducing random vibrations that are transmitted to the whole rotor-bearing system.

  8. Annular Seals Background literature Experimental – Seals (two phase) Iwatsubo, T., and Nishino, T., 1993, “An Experimental Study on the Static and Dynamic Characteristics of Pump Annular Seals,“ 7th Workshop on Rotordynamic Instability Problems in High Performance Turbomachinery. Hendricks, R.C., 1987, "Straight Cylindrical Seals for High Performance Turbomachinery," NASA TP-1850 Computational – Seals (two phase) Beatty, P.A., and Hughes, W.F., 1987, "Turbulent Two-Phase Flow in Annular Seals," ASLE Trans., 30, pp. 11-18. Arauz, G., and San Andrés, L., 1998, “Analysis of Two Phase Flow in Cryogenic Damper Seals, I: Theoretical Model, II: Model Validation and Predictions,” ASME J. Tribol., 120, pp. 221-227, 228-233 Arghir, M., Zerarka, M., Pineau, G., 2009 "Rotordynamic analysis of textured annular seals with mutiphase (bubbly) flow, “Workshop : “Dynamic Sealing Under Severe Working Conditions” EDF – LMS Futuroscope, October 5,

  9. Annular Seals Background literature Experimental – Seals (two phase) Mxx Iwatsubo, T., and Nishino, T., 1993, “An Experimental Study on the Static and Dynamic Characteristics of Pump Annular Seals,“ 7th Workshop on Rotordynamic Instability Problems in High Performance Turbomachinery. NO description of water lubricated seal (L, D, c) or gas type….. Tests conducted at various speeds (1,500-3,500 rpm) and supply pressures=1.2 - 4.7 bar. Air/liquid volume fraction b=0, 0.25, 0.45, 0.70 Cxx Kxx b, gas volume fraction increases

  10. Squeeze film dampers Background literature Experimental & Physical Modeling Tao, L., Diaz, S., San Andrés, L., and Rajagopal, K.R., 2000, "Analysis of Squeeze Film Dampers Operating with Bubbly Lubricants" ASME J. Tribol., 122, pp. 205-210 Diaz, S., and San Andrés, L., 2001, "Air Entrainment versus Lubricant Vaporization in Squeeze Film Dampers: An Experimental Assessment of their Fundamental Differences,” ASME J. Eng. Gas Turbines Power, 123, pp. 871-877 Diaz, S., and San Andrés, L., 2001, “A Model for Squeeze Film Dampers Operating with Air Entrainment and Validation with Experiments,” ASME J. Tribol., 123, pp. 125-133. Diaz, S., and San Andrés, L., 2002, “Pressure Measurements and Flow Visualization in a Squeeze Film Damper Operating with a Bubbly Mixture,” ASME J. Tribol., 124, pp. 346-350. Sponsored by National Science Foundation and TAMU Turbomachinery Research Consortium, June 1998- May 2002

  11. Squeeze film dampers Background literature Effect of bubbly mixtures and air ingestion on SFD forced performance CCO L=31.1 mm D=129 mm c=0.254 mm Sponsored by National Science Foundation and TAMU Turbomachinery Research Consortium, June 1998- May 2002

  12. Bubbly SFD Background literature Diaz, S., Beets, T., and San Andrés, L., 2000, “Pressure Measurements and Flow Visualization in a Squeeze Film Damper Operating With a Bubbly Mixture” b=0.540 30o Minimum Pressure Zone: Film Thickness Increasing Onset of Air Ingestion Incoming gas from Discharge Maximum Pressure Zone: Film Thickness Decreasing Minimum Gas Volume Fraction Uniform Mixture Uniform Pressure Zone: Maximum Film Thickness Onset of Positive Squeeze Maximum Gas Volume Fraction Non-Uniform Streaks (fingering) See digital videos at http://rotorlab.tamu.edu SFD (CCO): c=0.254 mm, e=0.180 mm, 500 rpm, ISO VG 68

  13. A simple model for bubbly mixtures Gas volume fraction (known at inlet) Ideal gas • Homogenous mixture of 2-components; isothermal & static equilibrium • Both components move with same speed & occupy same volume Mixture density Quasi-static model – ignores bubble dynamics For oil, PV~0.010 bar and S=0.035 N/m, and with c=0.152 mm, PV+2S/c=0.0146 bar Diaz, S., and San Andrés, L., 2001, “A Model for Squeeze Film Dampers Operating with Air Entrainment and Validation with Experiments,” ASME J. Tribol., 123, pp. 125-133

  14. A simple model for bubbly mixtures Mixture viscosity All liquid All gas McAdams model b Realistic model, not depending on mass fraction McAdams, W.H., Woods, W.K., and Heroman, L.C., Jr., 1942, “Vaporization inside Horizontal Tubes- II -Benzene-Oil Mixtures,” ASME Trans., 64, p.193

  15. Bulk-flow Analysis of Annular Seals Flow Continuity Circumferential Momentum transport Axial momentum transport Ps z W U W Pa • Turbulent flow with fluid inertia effects • Mean flow velocities – average across film (h) • No accounting for strong recirculation zones • Includes round-hole and honeycomb pattern (textured surface seal) San Andrés, L., and Soulas, T., 2007, “A Bulk Flow Model for Off-Centered Honeycomb Gas Seals,” ASME J. Eng. Gas Turbines Power, 129, pp. 185-194

  16. Wall shear stress differences Shear stresses Friction factors Other Ps z W U W Pa • Moody’s friction factor • Not affected by flow condition (single or two component) • Actual to be determined Salhi, A., Rey, C., and Rosant, J.M., 1992, “Pressure Drop in Single-Phase and Two-Phase Couette-Poiseuille Flow,” ASME J. Fluids Eng., 114, pp.80-84

  17. Bulk-flow Analysis of Annular Seals Boundary Conditions Numerical Solution Ps z Vz Vx Anti swirl brake at inlet or pressure seal W • Inlet pressure loss due to fluid inertia (Lomakin effect) • Inlet swirl determined by upstream condition (swirl-brake) • Exit pressure without recovery loss, typically. Numerical solution for realistic geometries use CFD technique (staggered grids, upwinding, etc) and predict (4) K,C,Mforce coefficients.

  18. Air in Oil Mixture SFD Model validation Diaz, S., and San Andrés, L., 2001, “A Model for Squeeze Film Dampers Operating with Air Entrainment and Validation with Experiments,” ASME J. Tribol., 123, pp. 125-133. Tangential force Circular Centered orbit b, mixture volume fraction Quasi-static bubbly flow model adequate for whole range of gas volume fractions (b=0.0-1.0) Lines: predictions, Symbols: experiments Radial force SFD (CCO): c=0.254 mm, e=0.120 mm, 1000 rpm, ISO VG 68

  19. Example of analysis MIX OIL with N2 Mixture volume fraction b varies (0.0-1.0) Based on available test rig Predict seal performance Based on a proposed test rig Centered seal (e=0): No static load ~ smooth surfaces; L/D=0.75, c/R=0.002 Table 1 Geometry and operating conditions of seal with mixture

  20. Seal Flow rate vs. inlet gas volume fraction All liquid All gas Figure 2 Mixture N2 in ISO VG 2 oil (DP=71 bar, 10 krpm) Leakage decreases continuously as gas content increases

  21. Gas Mass fraction vs. inlet gas volume fraction All liquid All gas Figure 3b Mixture N2 in ISO VG 2 oil (DP=71 bar, 10 krpm) Gas/liquid mass content increases exponentially with gas volume content

  22. Exit gas volume fraction vs. inlet volume fraction All liquid All gas Figure 3b Mixture N2 in ISO VG 2 oil (DP=71 bar, 10 krpm) Gas volume fraction at exit plane increases quickly because of large pressure drop

  23. Axial pressure drop as gas fraction increases inlet Exit Figure 4 Mixture N2 in ISO VG 2 oil (DP=71 bar, 10 krpm) All liquid: linear pressure drop. All gas: nonlinear with rapid changes near exit plane

  24. Drag power loss vs. inlet volume fraction All liquid All gas Figure 5 Mixture N2 in ISO VG 2 oil (DP=71 bar, 10 krpm) Steady decrease in power; but in region of flow transition

  25. Max. Reynolds # vs. inlet volume fraction All liquid All gas Figure 6 Mixture N2 in ISO VG 2 oil (DP=71 bar, 10 krpm) Axial flow dominates at high volume fractions. Circumf. flow Re# decreases.

  26. Rotordynamic coefficients – lateral motions Model for centered operation KXX = KYY, KXY = -KYX CXX = CYY, CXY = -CYX MXX = MYY, MXY = -MYX Assumes: No static load Y KXY : measure of rotordynamic stability Whirl frequency ratio WFR ~ CXX W Seal reaction forces: - Force coefficients are functions of frequency (w) for gases, and also for a two-component (gas/liquid) mixture. X

  27. Seal stiffnesses vs. inlet volume fraction All liquid All gas Figure 7a Mixture N2 in ISO VG 2 oil (DP=71 bar, 10 krpm) SYNCHRONOUS SPEED Liquid seal (oil) has large cross-coupled stiffness. Gas seal shows strong direct stiffness Mixture viscosity decreases KXY=-KYX KXY=-KYX KXX=KYY Synchronous speed force coefficient (w=W)

  28. Seal damping vs. inlet volume fraction All liquid All gas Figure 7b Mixture N2 in ISO VG 2 oil (DP=71 bar, 10 krpm) SYNC SPEED Direct damping decreases as gas content increases, but in flow transition zone Cross-damping small. Mixture viscosity decreases N-s/m CXX=CYY CXY=-CYX

  29. Whirl frequency ratio – Stability indicator • WFR always 0.50 for inlet swirl = 0.50 – Stable operation up to 2 x critical speed Mixture N2 in ISO VG 2 oil (DP=71 bar, 10 krpm) SYNC SPEED

  30. force coefficients – frequency dependency Y Seal reaction forces (centered seal): Force coefficients are functions of frequency (w) for gases, and also for a two-component (gas/liquid) mixture. X

  31. Seal direct stiffnesses vs. whirl frequency Figure 8a Mixture N2 in ISO VG 2 oil (DP=71 bar, 10 krpm) All liquid shows added mass effect (K-w2M). All gas (b=1) has large KXX. Note increase (*) in KXX for small b=0.1 KXX=KYY K w/W (*) b=0.1: Stiffness hardening is typical in textured gas damper seals (= negative added mass)

  32. Seal cross-stiffnesses vs. whirl frequency Figure 8b Mixture N2 in ISO VG 2 oil (DP=71 bar, 10 krpm) All liquid shows largest k. Cross-stiffness decreases with gas content. Small effect of frequency KXY=-KYX k w/W

  33. Seal direct damping vs. whirl frequency Figure 9a Mixture N2 in ISO VG 2 oil (DP=71 bar, 10 krpm) All liquid shows largest C. Same as cross-K. Small effect of frequency CXX=CYY C w/W Cross damping coefficients are one order of magnitude lower

  34. Equivalent force coefficients (Ke,Ce) Y Seal reaction forces (circular orbits): Ft Radial and tangential components of force e X wt Fr

  35. Seal equivalent stiffness vs. whirl frequency Figure 10a Mixture N2 in ISO VG 2 oil (DP=71 bar, 10 krpm) Cross damping small. All liquid shows added mass effect . All gas (b=1) has large Ke. Note increase (*) in Ke for small b=0.1 w/W

  36. Seal equivalent damping vs. whirl frequency Figure 10a Mixture N2 in ISO VG 2 oil (DP=71 bar, 10 krpm) All liquid shows largest Ce. Steady decrease of Ce with gas content. Note Ce=0 at w/W=0.5 w/W

  37. Conclusions GT2011-45264 Rotordynamic force coefficients of bubbly mixture annular pressure seals Advanced (simple) computational physics bulk-flow model for prediction of seal performance static and dynamic. Assumed homogenous mixture of two components (liquid and gas). Mixture N2 in ISO VG 2 oil (DP=71 bar, 10 krpm) • Leakage and power loss decrease with the gas in liquid volume content – except in transition region from laminar to turbulent flow • Seal force coefficients show strong dependency on whirl frequency. Cross-coupled stiffnesses and direct damping coefficients decrease steadily as gas volume fraction raises. • Direct stiffness coefficients show atypical behavior, in particular a mixture of gas volume fraction bS=0.1 produces stiffness hardening as the excitation frequency increases. • Predictions justify an experimental program to quantify the static and dynamic forced performance of annular seals operating with (bubbly) mixtures

  38. http://rotorlab.tamu.edu Learn more at Rotordynamic force coefficients of bubbly mixture annular pressure seals GT2011-45264 Questions (?) © 2011 Luis San Andres

More Related