A Thermohydrodynamic Analysis of Journal Bearings

Suganami, Takuya; Szeri, A. Z.

doi:10.1115/1.3453273

Cited by 88 publications

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“…Here (Elrod, 1991), turbulent isothermal flow with inertia (King and Taylor, 1977), and turbulent thermal flow without inertia (Suganami and Szeri, 1979), and with the experimental data for a turbulent isothermal bearing (Harada and Aoki, 1976). Given the accuracy and generality of Shyu's model, it was used as the basis to validate EGFLUM.…”

Section: Resultsmentioning

confidence: 98%

An Efficient General Fluid-Film Lubrication Model for Plane Slider Bearings

Shyu

Jeng

2002

Tribology Transactions

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ment is made in this study by combining the original model with a proposed in the former study. Pressure is obtainedfr-om the bulk-B = bearing width (m) B* = slenderness ratio = B l L C,, = specific heat at constant pressure (J / kg K) h(x,zj = film thickness (m) h, = film thickness at inlet (m) h2 = film thickness at exit (m) h* = convergence ratio = hi / h2 i , , = total number of grid points in s-direction ke = effective thermal conductivity = km+ k, (W / m K) km = molecular thermal conductivity (W / m K) k,nar = total number of grid points in z-direction k, = eddy thermal conductivity (W / m K) L = bearing length (m) N = highest degree of Legendre polynomial p = pressure (Pa) po = inlet pressure (Pa) 2 p* -= normalized pressure = (p-p,)h2 /(pp,L) P,,(c) = Legendre polynomial of degree n Pe = Peclet number = ~~~~u , , h~/ k~,~ Re* = modified Reynolds number = Po~,h,2/P,~ Rem = mean Reynolds number = pou,(h,+h2)/2p, T = temperature (K) To = inlet temperature (K) T,, = Legendre coefficient of temperature (K) 2 T* = normalized temperature = Re*C,(T-T,)/U, u = axial velocity component (m / s) u = film-averaged axial velocity component (m / s) u* = normalized axial velocity component = ulu, u, = axial velocity component at the runner surface (m / s) un = Legendre coefficient of u (m / s) I = velocity component through the film thickness (m / s) 3 = v -0.5 ( 5 + 1) [gu+ g w ] ( m / s ) fin = Legendre coefficient of 6 (m / s 2 L .+a/l W* = normalized load = h2& J-B,2 @-pJdz dx/p,u,B L' w = lateral velocity component (m / s) z3 = film-averaged lateral velocity component (m / s) MI,, = Legendre coefficient of w (m / s) x = axial coordinate (m) x* = normalized axial coordinate = x / L y = coordinate through the film thickness (rn) z = lateral coordinate (m) pe = effective viscosity = pm + p, (Pa s) pm = molecular viscosity (Pa. s) po = molecular viscosity at inlet = pn, ( po, To ) (Pa . S) p, = eddy viscosity (Pa. s) p = density (kg / m3) = film-averaged density (kg / m3) 3 po = density at inlet = p ( p,, To ) (kg / m ) 5 = normalized coordinate through film thickness, 5 = (2y -h)/h -5 = coordinate through film thickness with stretching, ( = l n (~) / l n (~) b = runner surface, y = 0 m = molecular quantity n = degree of Legendre polynomial; Legendre coefficient o = inlet quantity (reference quantity)Downloaded by [UNAM Ciudad Universitaria] at 14:51 20 December 2014 S. SHYU AND Y. JENG j7ow model. Comparisons of this model with the original model arc made for plane slider bearings in four flow regimes: laminar isothermal, lanlinar thermal, turbulent isothermal, nnd turbulent thern~al. Res~lts show good agreement between the two models, arid the model developed here is one order of magnitude faster thun the original model for turbulent thernzal flow.

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Section: Resultsmentioning

confidence: 98%

An Efficient General Fluid-Film Lubrication Model for Plane Slider Bearings

Shyu

Jeng

2002

Tribology Transactions

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“…Other experimenters have encountered the same temperature phenomenon at the transistion point [8,9]. Turbulence is not induced solely as a function of shaft speed.…”

Section: Additional Referencementioning

confidence: 92%

Discussion: “A Comparison of Tilting Pad Thrust Bearing Lubricant Supply Methods” (Mikula, A. M., and Gregory, R. S., 1983, ASME J. Lubr. Technol., 105, pp. 39–45)

Elwell

1983

Journal of Lubrication Technology

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operating temperature. The spray-fed bearing proved less adaptable, requiring a change in the number, diameter, or orientation of the jet holes to achieve flow or pressure variations that would result in power loss or temperature changes.All three bearing designs were prone to shoe instabilities and flutter in the unloaded bearings when operating without a restriction to provide suitable backpressure at the discharge. This was especially true at shaft speeds below 7000 rpm. The instability manifests itself as an audible, metallic clicking emanating from the housings. Fatigue failure of instrumentation wires indicates actual movement of the unloaded shoes. This problem was eliminated by the use of the oil control ring with tangential discharge opening. Bearing operating temperatures were reduced as a result of using the oil control ring. However, the overall power loss was adversely affected by this discharge restriction.The maximum measured shoe operating temperatures were found to be very responsive to the choice of lubricant supply method. The leading edge distribution groove proved superior at reducing maximum shoe temperatures. This is attributed to the positive introduction of cool supply oil directly into the oil film wedge between the hot oil carryover adhering to the high speed rotating collar and the shoe working surface. Test work is in progress to further refine and optimize the reduction of both operating temperature and power loss offered by the leading edge distribution groove bearing.The conventional pressurized supply or flooded bearing makes no attempt to minimize the effects of hot oil carryover from the previous shoes, and thus demonstrates high operating temperatures. The spray-fed bearing could not possibly scour away a thin, tenacious oil film moving at high speed with the moderate supply pressures normally en-DISCUSSION R. C. Elwell

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“…At θ = 0 o , the pressure is cavitation pressure and the temperature is specified as a reference temperature T 0 . Because the problems considered here are turbulent flow, the adiabatic journal condition can be applied according to the work by Suganami and Szeri (23). As for the fluid-bearing interface, Safar and Szeri (24) derived a simple algebraic equation for temperature:…”

Section: Methodsmentioning

confidence: 99%

THD Effects of Static Performance Characteristics of Infinitely Wide Turbulent Journal Bearings

Shyu

Jeng

et al. 2010

Tribology Transactions

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Formulas for static parameters were found for infinitely wide turbulent full journal bearings that correlate either load capacity or friction coefficient for thermohydrodynamics (THD) effects in terms of a single THD parameter. The database was built by numerical simulation of turbulent liquid lubricant flows with various eccentricity ratios in a wide range of the Reynolds number for both isoviscous and THD cases. The least-squares method was applied to the groups of parameters yielding the formulas of load capacity, friction coefficient, and attitude angle. The isoviscous attitude angle was fitted as a function of the maximum-to-minimum film thickness ratio, and the variation of attitude angle due to THD is linearly dependent on the THD-to-isoviscous load capacity ratio. With the formulas provided in this study, designers can quickly determine static parameters of turbulent journal bearings without the burden of labor-intensive numerical computation of the governing differential equations.

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A Thermohydrodynamic Analysis of Journal Bearings

Cited by 88 publications

References 0 publications

An Efficient General Fluid-Film Lubrication Model for Plane Slider Bearings

An Efficient General Fluid-Film Lubrication Model for Plane Slider Bearings

Discussion: “A Comparison of Tilting Pad Thrust Bearing Lubricant Supply Methods” (Mikula, A. M., and Gregory, R. S., 1983, ASME J. Lubr. Technol., 105, pp. 39–45)

THD Effects of Static Performance Characteristics of Infinitely Wide Turbulent Journal Bearings

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