Dual basket filter — Woven gauze or perforated metal element with changeover valve; elements easily removed for hand cleaning when system in operation. Magnets can be incorporated. Normally 50 µm and above. Volume II 401 The table shows that only large particles will settle within practical time limits. Water settling is hastened by raising the bulk temperature to 70°C after which it can be drained from the bottom of the reservoir. Filtration This is the most universal method and many filter materials and designs are available. Filters should be selected so that under clean conditions and maximum working viscosity the pressure loss does not exceed 0.3 bar (5 psi). Cleaning is normally recommended when pressure loss increases by 1 bar. Filters usually fall within the following types. Line strainer — Woven gauze or perforated metal element; easily removed for hand cleaning when system stopped. Normally 150 µm and above. 395-411 4/10/06 4:54 PM Page 401 Copyright © 1983 CRC Press LLC Mechanically cleaning filter — Interleaved radial plates plough the dirt from the gaps between metal discs when the filter pack is rotated, which can be while the system is operating. May be motorized. Periodically drain contaminant from sump when system is stopped. Normally 150 µm and above. Various other designs of mechanically cleaning filter, such as wire wound and back flushing, are available. 402 CRC Handbook of Lubrication Disposable element filter — Element of materials such as treated paper, felt, and nylon easily replaced when system stopped. Dual versions with changeover valve permit element replacement when system operating. Normally below 50 µm. 395-411 4/10/06 4:54 PM Page 402 Copyright © 1983 CRC Press LLC Centrifuging Installed in a by-pass circuit. Removal of sediment and water can be carried out while the oil system is in operation. For maximum efficiency, the oil should be centrifuged at 70°C with provision of an inline oil heater. COOLERS Constant oil temperature desirably enables constant flow and pressure control as these are both affected by changes in viscosity. Most machine designers recommend that the working temperature of the oil be 40°C. High-ambient temperatures, heat generation from bearings and gears, and machine and oil pump inlet power all transfer heat into the oil. The amount of heat is normally specified by the machine designer or based on experience with similar units. This commonly amounts to an oil temperature increase through the machine in the region of 10°C which needs to be removed by a cooler. Prolonged working temperatures above 60°C shorten oil life. Water and air are the common cooling mediums. When considering water, its cleanliness, corrosion characteristics, hardness, and pressure will affect selection of cooler materials. Temperature, quality, and quantity of cooling medium available are important in obtaining the most efficient cooler. Pressure losses for oil or water through the cooler should not exceed 0.7 bar. Cooler types frequently used are as follows: Volume II 403 Shell and tube — Oil through shell, water through tubes. Requires space for tube removal. Plate — Oil and water between alternate plates. Compact and readily separated for cleaning. 395-411 4/10/06 4:54 PM Page 403 Copyright © 1983 CRC Press LLC Radiator— Oil through tubes, air over tubes motivated by fan. TEMPERATURE CONTROL Along with heaters and coolers, controls are required to cope with inevitable temperature fluctuations. ReservoirHeater The control thermostat or thermostatic valve probe, Figure 3, should be positioned near the pump suction and at about middepth of the oil in the reservoir to sense the average temperature. Ensure that the thermostat is never exposed or placed near any localized hot spot such as a heater. Oil Cooler One control method is to regulate the water/air flow, the other to regulate the flow of oil to be cooled. Water flow can be governed by a hand control valve as its temperature usually only fluctuates on a seasonal basis. Automatic control of air is essential as its temperature fluctuates daily. Automatic control utilizes a direct-acting modulating valve in the cooling water supply line which is controlled by a sensing element in the cooling oil outlet. The effect on oil temperature is not instantaneous but is generally acceptable for industrial systems. Cooling water pressure should be reasonably stable, otherwise a pressure regulating valve will be required. Alternatively, where instantaneous response to control is vital and/or where air is the cooling medium, the cooler should be provided with a bypass line and a control valve to divert flow into the bypass. The valve is of a three-way type with overlapping ports, Figure 4. Mixing of the oil streams within the valve produces an average temperature and a thermostatic element detects any deviation in temperature and corrects the valve position. PRESSURE CONTROL A pressure control valve is necessary to spill-off surplus oil from the pump, to regulate any flow variation due to temperature fluctuations, and to accommodate any changes in demand from the machine being lubricated. The following are typical methods of control. Spring-loaded relief valve — Provides coarse control and is sensitive to viscosity changes. Generally used on smaller, simple systems. 404 CRC Handbook of Lubrication 395-411 4/10/06 4:54 PM Page 404 Copyright © 1983 CRC Press LLC Direct-operated diaphragm valve — The diaphragm chamber is connected so that system pressure is transmitted to the diaphragm. The spring counter-balancing the diaphragm load is adjustable and determines the system pressure. Any change in demand will tend to vary the system pressure and the diaphragm, sensing this, will reposition the valve to adjust the spill-off rate. This valve will maintain the pressure within acceptable limits provided the viscosity remains reasonably stable. Pneumatically controlled diaphragm valve — The diaphragm is air actuated via a control instrument. This valve is normally selected when very accurate control is necessary or if the system operating pressure is too great for the direct-acting valve diaphragm. Header tank — While space requirements often preclude this simple form of control, the procedure is to place a tank at the required height. Filled directly with a line teed off the main pump supply, the tank is fitted with an overflow connection. This method ensures the continual change of tank contents to maintain the oil at system temperature. As an added advantage, if the system pumps fail the tank will discharge its contents via the fill connection to the equipment being lubricated. Pump check valves will prevent oil returning directly to the main reservoir. Providing different pressures within the same system involves the use of pressure-reducing valves. These normally comprise a restricting orifice, or valve opening, which is controlled by imposing the outlet pressure on the valve control diaphragm. EMERGENCY EQUIPMENT AND SYSTEM CONTROL If a machine must continue to run after a failure in the lubrication system main pump, it is imperative to arrange fully automatic starting of a second, and maybe a third pump. This can be done by use of flow or pressure-operated switches, or both. Control panel lights Volume II 405 395-411 4/10/06 4:54 PM Page 405 Copyright © 1983 CRC Press LLC dished end is included in the tank sizing calculations. As some air will be absorbed in the oil, it will be necessary from time to time to add air. Acheck valve should be fitted in the air supply line to prevent oil from entering the air main and an air regulator installed to avoid accidental overpressurizing. SYSTEM PIPING Sizes of interconnecting pipes should be considered in relation to oil viscosity, velocity, and resultant friction losses. Pipes should be large enough to prevent cavitation in pump suction lines, to avoid undue pressure drop in pump supply lines (minimizing pump drive power), and to avoid backup in drain lines. Suction Pipe runs should be short. Right angle bends and tee pieces should be kept to a minimum. Nominal bore of pipe to be one size larger than supply. Supply Friction loss due to viscosity frequently outweighs velocity considerations, particularly with heavier oils. Figure 6 is based on a friction loss of 0.1-m head per m of pipe and restricted to velocities under an acceptable 2 m/ sec. The viscosity at specified operating temperature should be used to determine pipe nominal bore. To determine the friction loss in pipework, multiply the length by 0.1 m head. Friction loss caused by fittings and valves can be determined by converting them to equivalent pipe lengths in Table 2. These, together with pressure losses through the filter and cooler, will determine the system losses. Drain Drain pipes should be sized to run not more than half full so as to encourage escape of entrained air, provide space for any foam and give a margin of safety. Drain pipes should be vented. Flow rate, slope, and viscosity govern their size. Aminimum slope of 1 in 40 is essential but use should be made of all available drop. The pipe nominal bore may be determined from Figure 7. At startup, pipework and any 408CRC Handbook of Lubrication FIGURE 6. Supply line sizing. 395-411 4/10/06 4:55 PM Page 408 Copyright © 1983 CRC Press LLC Design Principles 395-411 4/10/06 4:55 PM Page 411 Copyright © 1983 CRC Press LLC JOURNALAND THRUSTBEARINGS A.A. Raimondi and A.Z. Szeri INTRODUCTION This chapter applies hydrodynamic lubrication theory to the analysis and design of self- acting fluid film journal and thrust bearings, in contrast to earlier chapters which emphasize lubrication theory and solution techniques. Most of the material has been summarized in design charts and tables for estimating performance of a variety of applications. It is not within the scope of this chapter to recommend bearing proportions, allowable temperature rise, etc. These are left to the designer to decide on the basis of experience and test. The charts provided here will serve for performance calculations on many representative bearings; similar information is available in the literature for a variety of designs. Computer programs are also available for studying design parameters for specific applications. LUBRICANTPROPERTIES IN BEARING DESIGN The lubricant property of greatest concern in fluid film bearings is the absolute viscosity, or just viscosity, μ. Its SI unit is Pa · sec (Pascal second), and in English units it is usually expressed in lb F · sec/in. 2 (reyn). The ratio of absolute viscosity to density (ρ) is termed the kinematic viscosity, =μ/ρ.It is measured in m 2 /sec in SI units and commonly in in. 2 /sec in English units. Table 1 contains conversion factors for commonly used viscosity units. Increasing temperature lowers the viscosity of lubricating oils as shown in Figure 1 for typical industrial petroleum lubricants in the various ISO viscosity grades. The viscosity of a number of other fluids is given in Figure 2. Average Viscosity In numerous applications, the temperature rise in the bearing film remains relatively small. However, in estimating bearing performance on the basis of classical (isothermal) theory, the calculations should employ an effective viscosity compatible with the mean bearing temperature rise. 1.19 This calculation might be based on the assumptions that: 1. All heat, H, generated in the film by viscous action is carried out by the lubricant. 2. The lubricant which leaves the bearing by its sides has a uniform average temperature T s = (T i + ΔT/2), where ΔT = T o – T i is the mean temperature rise across the bearing. This mean temperature rise, ΔT, can be calculated from a simple energy balance which gives: ΔΤ ϭ H/[ ρ c(Q Ϫ Q s /2)] (1) For typical petroleum oils, ρc = 112 lb F /in. 2 F (139 N/cm 2 C) and for water ρc = 327 lb F /in. 2 F (406 N/cm 2 C). Rather than assuming a uniform effective viscosity at a mean bearing temperature, using the actual variation in viscosity resulting from temperature changes as the lubricant flows through the bearing will result in considerably improved accuracy in calculating performance. However, this increases the complexity of the solution, usually requires special computer Volume II 413 Copyright © 1983 CRC Press LLC Flow Transition Two basic modes of flow occur in nature: laminar and turbulent. Flow transition from laminar to turbulent in bearings is preceded by flow instability in one of two basic forms: (1) centifugal instability in flows with curved streamlines, or (2) parallel flow instability characterized by propagating waves in the boundary layer. Instability between concentric cylinders was studied by Taylor. 3 He found that when the Taylor number (Ta = Re 2 C/R = Rω 2 C 3 / ) reaches its critical value of 1707.8, laminar flow becomes unstable. The equivalent critical reduced Reynolds number is √ — C/R Re = 41.3. The instability manifests itself in cellular, toroidal vortices that are equally spaced along the axis. As the Taylor number is increased above its critical value, the axisymmetric Taylor vortices become unstable to produce nonaxisymmetric disturbances, and turbulence eventually makes its appearance. 4 If the Reynolds number reaches 2000 before the Taylor number achieves its critical value, turbulence is introduced rapidly 5 without appearance of a secondary laminar flow. Eccentricity plays a role in defining critical conditions as covered in the chapter on Hydrodynamic Lubrication (Volume II). A positive radial temperature gradient in the clear- ance space, such as found in journal bearings at the position of minimum film thickness, is also destabilizing, 6 as is heat generation by viscous dissipation. 7 These statements draw support from experimental journal bearing data. 8 The local critical Reynolds number R h = Rωh/ seems to be in the 400 to 900 range. 2 Accepting R h = 900 for onset of turbulence, the critical value of global Reynolds number is approximately Re = 900 - /(1 − ⑀ ). Here, - is the ratio of the lowest value of the kinematic viscosity in the film to its value at the leading edge. Thus, for a fourfold decrease in viscosity and an 0.8 eccentricity ratio, the critical global Reynolds number is Re = 1125. A global value of 1000 has been used in later examples as a criterion for onset of turbulence. In thrust bearings, it was found 9 turbulent transition takes place within the range 580 < Re < 800, where Re = U a h a / is calculated on average conditions. Reference 10 reports agreement, but after replacing the average film thickness with the minimum film thickness in Re. Turbulence Turbulence is an irregular fluid motion in which properties such as velocity and pressure show random variation with time and with position. Once a relationship is established between the mean flow and Reynolds stresses, averaged equations of motion and continuity can again be combined to yield an equation in the (stochastic) average pressure p - : (3) Calculations in later sections of this chapter make use of a linearized theory 11 for turbulence functions k x and k z . The main contribution of isothermal turbulence to bearing performance is a significant increase in both load-carrying capacity and power loss. Thermal Effects If the bearing is large or if loading conditions are severe, pointwise variation of viscosity in the lubricant film is significant. Assuming negligible temperature variation in the axial direction, thermohydrodynamic (THD) journal bearing lubrication is represented by the following equations of pressure and temperature: 12 (4) Volume II 417 Copyright © 1983 CRC Press LLC [...]... PERFORMANCE INCLUDING INERTIA AND TURBULENCE (L/D = 1.0, β = 160°) CRC Handbook of Lubrication Copyright © 1983 CRC Press LLC 424 CRC Handbook of Lubrication FIGURE 8 Fixed-pad slider bearing (9) Steady state response (x,y, phase angle) to excitation F can then be obtained by solving the above linear equations Equally important, existence of any self-excited vibrations can be investigated by performing a... = 0.44 × 020 = 0.0088 in Position of minimum film thickness — Similarly, from Figure 25, attitude angle φ = 48° Copyright © 1983 CRC Press LLC 438 CRC Handbook of Lubrication FIGURE 27 FIGURE 28 Copyright © 1983 CRC Press LLC Inlet flow for partial arc journal bearing Side flow for partial arc journal bearing Volume II FIGURE 29 439 Film stiffness coefficients for partial arc journal bearing Power... that of a flat pad with its pivot optimally placed Cylindrical and spherically crowned pads can be designed by employing Tables 4 and 5 and Figure 20 These data should be used in conjunction with the following performance equations: (11) (12) (13) (14) (15) Copyright © 1983 CRC Press LLC 434 CRC Handbook of Lubrication Laminar flow solution: from Table 4, load factor CW,L = 0.132; from Equation 11, ... the partial bearings are the two basic configurations of fixed-type hydrodynamic journal bearings (Figure 22) Most other designs found in practice (Figure 23, for example) can be considered as variations of these two The active (load-carrying) bearing arc extends entirely around the journal for the full journal bearing; in the partial bearing, the active arc only partially surrounds the journal The partial... performance calculations of this bearing can be made by relating the rectangular slider bearing (width B, length L) Copyright © 1983 CRC Press LLC 426 CRC Handbook of Lubrication FIGURE 12 FIGURE 13 Spiral groove bearing Pivoted-pad thrust bearing Pivoted-Pad Thrust Bearings This type (Figure 13) employs a supporting pivot at the center of film pressure ( in Figure 8) developed on the surface of a slider bearing... Copyright © 1983 CRC Press LLC Volume II 435 FIGURE 22 Fixed-type journal bearings: (a) full 360° bearing, (b) centrally loaded partial bearing, and (c) offset loaded partial bearing FIGURE 23 lands in top Partial bearing with relief and end as in a pivoted-pad bearing, it can partially respond to changes in speed, load, viscosity, and temperature Advantages include (1) better load distribution than with... (1.06 gal/min) Qs = 1.66 in.3/sec (0.43 gal/min) Temperature rise can then be calculated from Equation 1 with pc = 112 as typical for petroleum oils ΔT = 1.71 × 104/ [112 (4.08 – 1.66/2)] = 47.0 F This temperature rise implies an inlet temperature Ti of Ti = TS – ΔT/2 = 135 – 47.0/2 = 111 .5 F If this did not match the oil inlet temperature, a new effective oil viscosity would be assumed and the steps... active arc, and “offset” or eccentrically loaded” if it does not Design charts are presented here for the most commonly used bearings: the full journal bearing and the centrally loaded partial bearing Centrally Loaded Partial Bearings Figures 24 to 30 are design charts for a 160° centrally loaded partial bearing with a lengthto-diameter (L/D) ratio equal to 1 (see Reference 29 for offset loading) These... the equations of motion are nonlinear when convective inertia is retained, the problem becomes tractable as pointwise lubricant inertia is replaced by its average value, obtained via integration across the film.14,15 The averaged equations of motion and continuity combine in a single equation in lubricant pressure For journal bearings: (7) Copyright © 1983 CRC Press LLC 420 CRC Handbook of Lubrication. .. vibration at exactly one-half (or other exact submultiple) of the running speed known as subharmonic resonance.18 The rotor-shaft configuration of Figure 5 is reduced to a simple dynamical system of springs and dashpots in Figure 7 A mass W/g (one half the rotor weight) can be imagined to be concentrated at Ojs, the steady running position of the journal If some excitation F Copyright © 1983 CRC Press . position of the journal. If some excitation F 420CRC Handbook of Lubrication FIGURE 5.Dynamical elements of rotor-shaft configuration. Copyright © 1983 CRC Press LLC 422 CRC Handbook of Lubrication Table. methods of control. Spring-loaded relief valve — Provides coarse control and is sensitive to viscosity changes. Generally used on smaller, simple systems. 404 CRC Handbook of Lubrication 395- 411 4/10/06. should be made of all available drop. The pipe nominal bore may be determined from Figure 7. At startup, pipework and any 408CRC Handbook of Lubrication FIGURE 6. Supply line sizing. 395- 411 4/10/06