Open pocket bearing: calculated values of load carrying capacity versus pocket depth, for different values of minimum film thickness. Textured retention: calculated values of load-carrying capacity versus indentation depth, for different values of minimum film thickness.
Introduction
Bearings and applications
An important parameter related to bearing performance is the relative speed of the bearing rotor and stator. For given bearing geometry and load, increase in relative speed leads to a steeper pressure build-up in the lubricant domain, and to an increased value of .
Literature review
It is shown that the effects of the recess are generally positive, but performance can deteriorate when oil viscosity and bearing speed are high. Detailed analyzes of the effects of manufacturing defects have been reported in the recent literature for aerostatic bearings.
Goals of present study
Furthermore, in [41] the effects of manufacturing defects on a single type of aerostatic bearings with rectangular flat pads with multiple openings are considered; it was found that deviations of the bearing surface from the ideal plane had a significant effect on the load-bearing capacity and stiffness coefficient. For a similar bearing geometry [39], the effects of recess shape, orifice diameter and gas film thickness on performance parameters are investigated.
Comparison between different bearing configurations
Introduction
Bearing Geometries
- Taper-land bearing geometry
- Open and closed pocket bearing geometry
- Textured bearing geometry
Closed and open pocket bearings (Figures 3 and 4) contain a stator pocket, the depth of which varies and has values from 10 to 100 μm. The recesses have a depth of 20 and extend from the leading edge of the groove to an angle of 29o.
Computational model
Representative meshes generated for the different bearing types in this study are presented in Figure 6(a)-(d). a) Tapered land bed: computational grid in the fluid domain, (b) Closed pocket bed: computational grid in the stator domain, (c) Open pocket bed: computational grid in the stator domain, and (d) Textured bed: details of computational mesh in the fluid domain. In fluid-pad and fluid-rotor interfaces, continuity of temperature and heat flux is implemented.
Computational results
At higher values of minimum film thickness (lower compressive loads), the four bearing types have similar performance characteristics. Calculated values of (a) bearing capacity, (b) friction torque, (c) maximum fluid temperature, and (d) maximum pad temperature versus minimum film thickness, for the reference design for the four types of thrust bearings in this study at N=1000 RPM. Calculated values of (a) bearing capacity, (b) friction torque, (c) maximum fluid temperature, and (d) maximum pad temperature versus minimum film thickness, for the reference design for the four types of thrust bearings in this study at N=4000 RPM.
Calculated values of (a) load capacity, (b) friction torque, (c) maximum fluid temperature and (d) maximum brake pad temperature versus minimum film thickness, for the reference design of the four types of thrust bearings of this study, at N=8000 rpm. Calculated values of (a) load capacity, (b) friction torque, (c) maximum fluid temperature and (d) maximum brake pad temperature versus minimum film thickness, for the reference design of the four types of thrust bearings of this study, at N=10,000 rpm. Figure 15 - Figure 18 shows the calculated values of load capacity versus pocket/cone/texture depth, for different values of minimum film thickness, for the.
Cone soil bearing: calculated values of load bearing capacity versus cone depth, for different values of minimum film thickness.
Conclusions
Computational Investigation of Thermoelastohydrodynamic (TEHD) Lubrication in a
Introduction
Bearing geometry
A sketch of the bearing of this study and a top view of the bearing pad.
Computational model
The density of the used lubricating oil (ISO VG 46) is 870 kg/m3, while the temperature-dependent viscosity is taken into account, according to the McCoull and Walther relationship Eq:(5), as used in the previous chapter. The outer surface of the fluid domain is considered as the outlet boundary, with atmospheric pressure and a Neumann boundary condition for temperature. Regarding the thermal boundary conditions, the rotor is assumed stationary (frozen), therefore a constant thermal load of the rotor is calculated.
Rotational periodicity conditions are applied to the leading/trailing edges of the rotor and pad. On the remaining external surfaces of the rotor and pad, suitable combinations of convection coefficient, and ambient temperature, are prescribed, according to the work in [9]; they are shown below.
Computational results
A significant increase in the load-bearing capacity (and the maximum liquid temperature) is observed with decreasing values of the minimum film thickness. Load capacity, maximum fluid temperature and maximum rotor/pad displacement versus rotational speed, for four different values of minimum film thickness. Figure 6(a)-(f) presents the calculated variation of the payload and maximum fluid temperature, and the corresponding variation of the maximum rotor/pad displacement, as a function of the rotor thickness, for three different values of the minimum film thickness; a negligible effect on load capacity, maximum fluid temperature and brake pad displacement has been verified.
Load capacity, maximum liquid temperature and maximum rotor/pad displacement versus rotor thickness, for different values of minimum film thickness. An increase in the dimple area is verified in all cases, with the pressure levels decreasing with increasing minimum film thickness.
Conclusions
Introduction
Bearing geometry
Due to manufacturing defects, the stator surface can be characterized by several types of deviations from the nominal surface, see e.g. In the present work, a number of manufacturing errors have been taken into account, which result in the following deviations from a nominal design, see Figure 2: ii) Discrepancy in normalized indentation depth. iii) Deviation of indentation shape from the orthogonal reference shape. iv) Concavity/convexity of base stator surface. v) Waves on the base stator surface in the direction of flow. Deviation of the indentation shape from the orthogonal reference shape is accounted for by considering trapezoids inscribed in the reference (rectangular) indentation geometry.
Concavity/convexity (in the flow direction) is defined by first considering two points, with a vertical coordinate relative to a local coordinate system, with the origin at the center of the base (plane) stator surface. 46 Waves in the flow direction are considered sinusoidal, defined in terms of wave number, amplitude and phase angle.
Computational model
Here, the density of the networks used is similar to that of the networks validated in Papadopoulos et al. The upper wall is assumed to move at a constant velocity (parallel to the x-axis), , 0,. The inlet and outlet surfaces of the bearing are considered openings: the pressure is assumed constant, with the same value p=0 defined at both boundaries, while a Neumann boundary condition is assumed for the velocity.
At the bearing sides, z=0, B, an exit condition is provided, which prevents the fluid from entering the computational domain. In all cases, steady-state convergence was verified by monitoring the calculated velocity and pressure at a number of representative points of the flow domain.
Geometry parameters of optimized reference textured bearings
For Re=1, the governing equations were integrated for a total non-dimensional time ( ) of 10 with a time step of 0.05. For a given value of B/L, a convergence ratio of corresponds to the global maximum in carrying capacity.
Computational results
In the presence of concavity and convexity, the load-bearing capacity of parallel slides decreases (Figure 32(c)), with a corresponding increase in the coefficient of friction (Figure 32(h)). However, a significant increase and results in a decrease in the pressure rise, due to a decrease in the equivalent (useful) length of the sump. a Wave number=1 Wave number=3 Wave number=5. a Wave number=1 Wave number=3 Wave number=5. Normalized values of load bearing capacity and friction coefficient for different parameters expressing manufacturing errors: (a), (f) discrepancy in convergence ratio; (b), (g) discrepancy in normalized pit depth; (c), (h) concavity (+) / convexity (‐); (d), (i) ripple; (e), (j) discrepancy from the orthogonal shape of the pit. a Wave number=1 Wave number=3 Wave number=5. a Wave number=1 Wave number=3 Wave number=5.
B/L = inf, k=0: (a) Non-dimensional pressure distribution in the moving wall for different values of parameters ba and bb. b) Corresponding streamlines in the region of the first dimple, coded with the velocity magnitude. Figure 41(a),(d) shows that, in the nominal design case, a significant increase in pressure is present, for the textured part as well as for a large part of the untextured part. Finally, Figure 41(c),(d) shows that in the case of a corrugated stator with n=1 and a=0.3, a significant increase in pressure build-up is achieved for a large part of the untextured part; this significantly increases the maximum pressure value as well as the pressure integral (by approximately 20%, see Figure 37(d) for a=0.3).
In the case of 0° and 330°, the introduction of waves results in a significant increase in the pressure build-up in the indentation area of the stator, followed by a sharp decrease in the non-textured area, which generally leads to a significant increase in the load-carrying capacity (pressure integral) for parallel and converging sliders.
Conclusions
Conclusions and future study
The effect of introducing surface texturing in the form of pockets or recesses should be investigated for the case of tapered thrust bearings. Alternative pocket bearing geometries, characterized by an elliptical front face, converging or diverging side faces, and bottom surface inclination, should also be studied to determine optimal design features. A two-way FSI thrust bearing model needs to be developed to predict bearing behavior at high thrust load and rotational speed values.
This model should also be able to calculate thermal expansion of both the rotor and the stator, which can be of the same magnitude as elastic strains. A 3-D THD parametric analysis regarding the effect of manufacturing defects on industrial bearing designs should be performed, to extend the findings of the present study.-.
Literature
Experimental studies of pressure distributions in finite sliding bearings with single continuous surface profiles on pads. Numerical analysis of elasto-hydrodynamically lubricated point contacts with three-dimensional laser micro-texture dexterity. On the shape of the lubricant film for optimal performance of a longitudinally rough sliding bearing.
The effect of indentation shape on performance analysis of the gas-lubricated bearing in optical lithography. 44] Papadopoulos C, Nikolakopoulos P, Kaiktsis L. Evolutionary optimization of micro-thrust bearings with periodic partial trapezoidal surface texturing.