ASME Turbo Expo 2012 GT 2012 68355 June
ASME Turbo Expo 2012 GT 2012 -68355 June 11 -15, 2012, Copenhagen, Denmark On the Effect of Thermal Energy Transport to the Performance of (Semi) Floating Ring Bearing Systems for Automotive Turbochargers Luis San Andrés Mast-Childs Professor, Fellow ASME Texas A&M University Accepted for journal publication
GT 2012 -68355 Thermal Energy Transport to the Performance of (Semi) Floating Ring Bearing Systems for Automotive Turbochargers Authors Luis San Andrés Turbomachinery Laboratory Texas A&M University Vince Barbarie Avijit Bhattacharya Kostandin Gjika Honeywell Turbo Technologies Torrance, CA 90504 Honeywell Turbo Technologies, Thaon-les-Vosges, France Supported by Honeywell Turbocharger Technologies (HTT) 2
Types of bearing supports Fully Floating Bearing Semi Floating Bearing Oil lubricated bearings are economic with longer life span; but prone to harmful subsynchronous whirl & depend on engine oil condition. Ball Bearing Low shaft motion, expensive, limited lifespan The driver: Increased IC engine performance & efficiency demands of robust turbocharging solutions 3
Major challenges extreme operating conditions: • - Low • • - High Oil Temperature (up to 150°C) - Low HTHS (2. 9); Low Oil Pressure (1 bar), • • - Increased Max. Turbocharger Speed 5 k. HZ - Variable Geometry Turbo Technology & Assisted e-power start up • - High • Oil Viscosity, e. g. 0 W 30 or 0 W 20 Engine Vibration Level - More Stringent Noise Requirements Thermal management & reduce thermal loading Need predictive too to reduce costly engine test stand qualification 4
Literature Review: TAMU work • TC linear and nonlinear rotordynamic codes – GUI based – including engine induced excitations • Realistic bearing models: thermohydrodynamic • Novel methods to estimate imbalance distribution and shaft temperatures • NL analysis for frequency jumps (internal & combined resonances) and noise reduction • Measured ring speeds with fiber optic sensors 2004 IMEch. E J. Eng. Tribology 2005 ASME J. Vibrations and Acoustics ASME DETC 2003/VIB-48418 ASME DETC 2003/VIB-48419 2007 ASME J. Eng. Gas Turbines Power ASME GT 2006 -90873 2007 ASME J. Eng. Gas Turbines Power ASME GT 2005 -68177 2007 ASME J. Tribology IJTC 2006 -12001 2007 ASME DETC 2007 -34136 2010 ASME J. Eng. Gas Turbines Power ASME GT 2009 -59108 2010 IFTo. MM Korea Predictive tool for shaft motion benchmarked by test data 5
VIRTUAL TOOL for TC NL shaft motions Tool demonstrated 70% cycle time reduction in the development of new CV TCs. Since 2006, code aids to develop PV TCs with savings up to $1 XX k/year in qualification test time Predicted shaft motion Measured shaft motion ASME DETC 2007 -34136 TC testing is expensive and time consuming Predictive tool saves time and resources 6
Conjugate heat transfer in TCs Thermal energy analysis in TCs complicated b/c of a) Hot gas -work and heat flow from turbine b) Cold gas +work and heat flow from the compressor c) internal heat flow across shaft from T to C and radially thru bearings d) Mechanical drag power in bearings e) Heat flow to/from casing to ambient (convective and radiant) Engine lubricated bearings: (a) low friction-load support (b) oil carries away heat (cooling) 7
Heat flows & energy transfer in a TC Hot air (energy) in Compressed hot air (energy) out Oil in -Hot air (energy) out work turbine Heat conducted (shaft) Heat conducted (casing) Cold air (energy) in Bearing drag power generation compressor Heat conducted (casing) Oil out Heat conducted (casing) 8 Baines, N. , Wygnant, K. , Dris, A. , 2010, J Eng Gas Turb. Power, 042301
Thermal energy analyses in TCs use a) Lumped parameter models with empirical formulas for heat transfer coefficients and simplified formulas for drag power, flow & heat flow in the lubricated bearings b) Current 3 D modeling stresses on solids with over -simplified coupling to the lubricant flows GOAL Engineered thermal management to avoid severe thermal loading with improved reliability of bearing system: avoid oil coking, optimize flow rates, ensure proper clearances, eliminate seizure. 9
SFRB system in engine-oil lubricated TC Lubricant flow paths into bearings on turbine and compressor sides casing Oil supply at Psup, Tsup compressor turbine shaft Turbine side bearing Outer film with ½ moon groove Compressor side bearing oil supply holes to inner and outer films Oil supply holes to inner film Semifloating ring bearing Anti-rotation pin Oil discharges at ambient pressure Pa 10
Geometry & coordinate system for typical SFRB Definitions: X, Y: fixed (inertial) coordinate system g : direction of gravity Y q : Circumferential coordinate Y Oil supply hole LG ½ moon groove Casing outer film inner film DRi 0º DJ X q shaft Shaft rotation W g DRo DB Z Li WR Ring rotation Ring Lo
Kinematics of journal and ring Nomenclature: W : journal rotational speed WR: ring rotational speed Y e q : Circumferential (angular) coordinate = e. R +e J ring rotation WR Journal rotation W e OR e. R q e. J X journal ring Outer film thickness Inner film thickness 12
Axial view of inner and outer films: nomenclature Oil inlet, Psup, Tsup Inlet ½ moon groove Ring, TR(r) TC Outer film, To(q, z) Shaft, TS z Rc TRo Po, To Ro TR M W WR Inner film, Ti(q, z) PA , ambient pressure Feed hole Casing, PA TC Ri TRi Pi, Ti TS z Rs RM
Hydrodynamic pressure generation Nomenclature P : film pressure h : film thickness m : viscosity, fn (T) W journal rotational speed WR ring rotational speed Outer film Inner film Q circumferential coordinate Z axial coordinate Casing TC Po, To TR M Pi, Ti TS shaft RJ: Shaft (journal) radius RB: Bearing casing inner diameter Major assumptions: Laminar flow without fluid inertia effects Average viscosity across film thickness Reynolds equations for inner & outer films 14 Ring
Heat flows & power in a FRB casing Heat flow into casing TC Outer film Heat flow into ring Heat flow from shaft To Mechanical TRo drag power Ring Inner film shaft Flow outer Ti TRi TS Flow inner Heat flow carried by oil Mechanical drag power W 15
Thermal energy transport in inner film Nomenclature Casing TC Ti: inner film temperature TJ, TRi : shaft (journal) and ring ID temperatures h. J, h. Ri : heat convection coefficients Po, To TR M Pi, Ti TS Inner film mechanical energy dissipation: shaft with circ. and axial mass flow rates: 16
Thermal energy transport in outer film Nomenclature Casing TC To: outer film temperature TB, TRo : bearing casing and ring OD temperatures h. B, h. Ro : heat convection coefficients Po, To TR M Pi, Ti TS shaft Outer film mechanical energy dissipation: with circ. and axial mass flow rates: 17
Heat conduction in semi-floating ring Nomenclature TR: ring temperature TJ, TRi : shaft (journal) and ring ID temperatures h. J, h. Ri : heat convection coefficients Major assumptions: Steady state, no heat flow in axial direction, No effect of ring rotation k. R: ring material conductivity q. R : heat flow r Q 18
Heat conduction in semi-floating ring Casing TC Major simplification Radial heat conduction only Po, To TR M Ring r Q Pi, Ti TS shaft Nomenclature TR: ring temperature QR: radial heat flow k. R: ring material conductivity 19
Heat convection Models Ti Ts Heat flow: Q = h A (TS – Ti) A: wetted area for heat transfer h: heat convection coefficient, a function of Nusselt #, oil conductivity and hydraulic diameter (=clearance). Nusselt # =depends on flow conditions (Prandtl # and Reynolds #) 1 Reynolds/Colburn Analogy), Nu=3 Pr 0. 33 2 Kays and Crawford - constant wall temperature, Nu =7. 54 3 Kays and Crawford - constant wall heat flux, Nu =8. 22 4 Haussen - thermally developing constant wall temperature, Nu >3. 657 5 Shah - 6 Stephan - Simultaneous developing, constant wall temp, Nu >3. 66 7 20 Stephan - Simultaneous developing constant wall heat flux, Nu > 4. 364 thermally developing constant wall heat flux, Nu > 4. 364
Numerical method of solution a) Finite element method for Reynolds Eqns. b) Control volume method for energy transport Eqns. Includes balance of drag torques, material properties f(T), bearing clearance changes due to temperature rise, etc. 21
Example Semi FRB for PV turbine bearing Oil inlet CASING Oil: SAE 5 W-30 120 C Shaft (journal) 213 C RING Bearing dimensions Diameter length Cold clearance Inner Film Outer film 7. 9 4. 6 7. 5 14. 1 6. 2 35 mm mm mm 22
Meshes for inner and outer flow domains ½ moon groove Casing =132 º Y Mesh: outer film z groove X NEX=45, NEY=12 journal Z ½ moon groove Circ. groove 132 o ring Mesh: inner film Feed hole x 4 NEX=52, NEY=12 q z Oil supply hole (4 x) Axial groove (4 x) Engineered design to improve flow delivery and reduce temperature rise 23
Predictions for inner film at 240 krpm TS=213 C Feed hole & axial groove q Tsup=120 C z q z (a) Pressure field (bar) (b) Temperature field (C) Oil heats quickly along axial plane 24
Temperatures: maximum in films Shaft Temp inner film exit mixed films Inner film temperature shaft T (< flash T). Outer film relatively cold. outer film 25
Temperatures: average in films Shaft Temp inner film exit mixed films outer film Inner film much hotter than outer film. Exit mixing lubricant temperature nearly constant > 90 krpm 26
Ring Temperatures: ID, OD and mean Shaft Temp Ring. ID (mean radius) Ring. OD Large radial temperature gradient across ring. OD-ID Temperature difference ~ 40 o. C. RING material conductivity is important. 27
Oil viscosity (average): inner & outer films outer film inner film Relative to supply: Inner film viscosity decreases because of increase in film temperature (> Tsup) Thermal effects can not be ignored 28
Oil flow rates: inner & outer film Flow=1 = out+in inner film Oil flow is minimum at top speed. Outer/inner flow decreases/increa ses because of clearance shrinks/grows + lower oil viscosity 29
Heat flows and drag power Heat from shaft Heat to lubricant Drag power Heat to ring (inner film) Low speeds: heat from shaft dominates. High speeds: drag power losses increase. For all conditions lubricant carries more energy that casing soaks. Heat to casing + =1 30
Thermal energy transport and balance Width of boxes denotes intensity of heat flows 36 % 100% 97% Heat to ring Heat to casing Heat to fluid (o) Heat from shaft Heat to fluid (i) Drag Power 3% (inner & outer) 27 % 10 % Heat to fluid (i+o) 74 % 64 % low speed (45 krpm) Lubricant carries away heat from shaft mainly. 31
Thermal energy transport and balance Width of boxes denotes intensity of heat flows 24 % 100% 65% Heat from shaft 35% Drag Power (inner & outer) Heat to ring Heat to fluid (i) 17 % Heat to casing Heat to fluid (o) 8% Heat to fluid (i+o) 83 % 76 % high speed (240 krpm) Drag power losses increase. Lubricant carries away largest portion of heat flow. 32
Drag power and heat from shaft High heat flow High power low power 240 krpm Inlet plane 240 krpm Exit plane 30 krpm Drag power (W) Low heat flow Exit plane 30 krpm Heat from shaft (W) Drag power and heat from shaft are large at inlet because of inlet (cold) lubricant. 33
Conclusions (a) Heat flow from hot shaft into inner film is large; more so at the inlet plane where oil is cold; (b) The inner film temperature increases quickly (viscosity drops) due large heat flow from shaft and drag shear power; (c) The floating ring has a large radial temperature gradient; (d) At all rotor speeds, the lubricant flows carry more than 70 % of the total energy input. The rest soaks into the TC casing. The bearing design must allow for adequate flow paths to cool components. Tool integrated into sponsor engineering design practice to predict thermal loading and mechanical stresses and to ensure lubricant does not overheat (coking) 34
Acknowledgments Thanks to Honeywell Turbocharging Technologies (2002 -2012) Questions (? ) Learn more at: http: //rotorlab. tamu. edu Copyright© 2012 Luis San Andres 35
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