Handbook of Lubrication part 11 pps

COOLERSConstant oil temperature desirably enables constant flow and pressure control as these areboth affected by changes in viscosity.. Water flow can be governed by a hand control valv

Dual basket filter — Woven gauze or perforated metal element with changeover valve;

elements easily removed for hand cleaning when system in operation Magnets can beincorporated Normally 50 µm and above

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 thebottom 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 viscositythe pressure loss does not exceed 0.3 bar (5 psi) Cleaning is normally recommended whenpressure 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

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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 isoperating May be motorized Periodically drain contaminant from sump when system isstopped Normally 150 µm and above

Various other designs of mechanically cleaning filter, such as wire wound and backflushing, are available

Disposable element filter — Element of materials such as treated paper, felt, and nylon

easily replaced when system stopped Dual versions with changeover valve permit elementreplacement when system operating Normally below 50 µm

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Installed in a by-pass circuit Removal of sediment and water can be carried out whilethe oil system is in operation For maximum efficiency, the oil should be centrifuged at70°C with provision of an inline oil heater

COOLERSConstant oil temperature desirably enables constant flow and pressure control as these areboth affected by changes in viscosity Most machine designers recommend that the workingtemperature of the oil be 40°C High-ambient temperatures, heat generation from bearingsand 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 similarunits This commonly amounts to an oil temperature increase through the machine in theregion of 10°C which needs to be removed by a cooler Prolonged working temperaturesabove 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 obtainingthe most efficient cooler Pressure losses for oil or water through the cooler should notexceed 0.7 bar

Cooler types frequently used are as follows:

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

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Radiator — Oil through tubes, air over tubes motivated by fan.

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 usuallyonly fluctuates on a seasonal basis Automatic control of air is essential as its temperaturefluctuates daily

Automatic control utilizes a direct-acting modulating valve in the cooling water supplyline which is controlled by a sensing element in the cooling oil outlet The effect on oiltemperature is not instantaneous but is generally acceptable for industrial systems Coolingwater pressure should be reasonably stable, otherwise a pressure regulating valve will berequired

Alternatively, where instantaneous response to control is vital and/or where air is thecooling medium, the cooler should be provided with a bypass line and a control valve todivert 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 athermostatic 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 regulateany flow variation due to temperature fluctuations, and to accommodate any changes indemand 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

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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 varythe system pressure and the diaphragm, sensing this, will reposition the valve to adjust thespill-off rate This valve will maintain the pressure within acceptable limits provided theviscosity 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 ifthe 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 offthe main pump supply, the tank is fitted with an overflow connection This method ensuresthe continual change of tank contents to maintain the oil at system temperature As an addedadvantage, 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 tothe main reservoir

Providing different pressures within the same system involves the use of pressure-reducingvalves 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 Thiscan be done by use of flow or pressure-operated switches, or both Control panel lights

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dished end is included in the tank sizing calculations As some air will be absorbed in theoil, it will be necessary from time to time to add air A check valve should be fitted in theair supply line to prevent oil from entering the air main and an air regulator installed toavoid accidental overpressurizing.

SYSTEM PIPINGSizes 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 pumpsuction lines, to avoid undue pressure drop in pump supply lines (minimizing pump drivepower), and to avoid backup in drain lines

To determine the friction loss in pipework, multiply the length by 0.1 m head Frictionloss caused by fittings and valves can be determined by converting them to equivalent pipelengths in Table 2 These, together with pressure losses through the filter and cooler, willdetermine the system losses

Drain

Drain pipes should be sized to run not more than half full so as to encourage escape ofentrained 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 A minimum 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

FIGURE 6 Supply line sizing.

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Design Principles

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JOURNAL AND THRUST BEARINGS

A.A Raimondi and A.Z Szeri

INTRODUCTIONThis 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 emphasizelubrication theory and solution techniques Most of the material has been summarized indesign charts and tables for estimating performance of a variety of applications

It is not within the scope of this chapter to recommend bearing proportions, allowabletemperature rise, etc These are left to the designer to decide on the basis of experience andtest The charts provided here will serve for performance calculations on many representativebearings; similar information is available in the literature for a variety of designs Computerprograms are also available for studying design parameters for specific applications

LUBRICANT PROPERTIES IN BEARING DESIGNThe 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 usuallyexpressed in lbF · sec/in.2 (reyn) The ratio of absolute viscosity to density (ρ) is termedthe kinematic viscosity,  = μ/ρ It is measured in m2/sec in SI units and commonly in

in.2/sec in English units Table 1 contains conversion factors for commonly used viscosityunits

Increasing temperature lowers the viscosity of lubricating oils as shown in Figure 1 fortypical 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 bearingtemperature rise.1.19This 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

Ts = (Ti + ΔT/2), where ΔT = To – Ti is the mean temperature rise across thebearing

This mean temperature rise, ΔT, can be calculated from a simple energy balance whichgives:

Flow Transition

Two basic modes of flow occur in nature: laminar and turbulent Flow transition fromlaminar 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 instabilitycharacterized by propagating waves in the boundary layer

Instability between concentric cylinders was studied by Taylor.3He found that when theTaylor number (Ta = Re2C/R = Rω2C3/) 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 spacedalong the axis

As the Taylor number is increased above its critical value, the axisymmetric Taylor vorticesbecome unstable to produce nonaxisymmetric disturbances, and turbulence eventually makesits appearance.4 If the Reynolds number reaches 2000 before the Taylor number achievesits critical value, turbulence is introduced rapidly5without appearance of a secondary laminarflow

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, isalso destabilizing,6 as is heat generation by viscous dissipation.7 These statements drawsupport from experimental journal bearing data.8The local critical Reynolds number Rh =Rωh/ seems to be in the 400 to 900 range.2Accepting Rh = 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 v-iscosity in the film to its value at theleading edge Thus, for a fourfold decrease in viscosity and an 0.8 eccentricity ratio, thecritical global Reynolds number is Re = 1125 A global value of 1000 has been used inlater examples as a criterion for onset of turbulence

In thrust bearings, it was found9turbulent transition takes place within the range 580 <

Re < 800, where Re = Uaha/ is calculated on average conditions Reference 10 reportsagreement, 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 pressureshow random variation with time and with position Once a relationship is established betweenthe 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-:

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 axialdirection, thermohydrodynamic (THD) journal bearing lubrication is represented by thefollowing equations of pressure and temperature:12

(4)

Here, −the mean dissipation and the turbulent functions kx, k2, and F, as well as the velocitytemperature correlation V’t’—, depend on the turbulence model used A minimum list ofnondimensional parameters that characterize bearing performance according to THD theorymust include

Inertia Effects

Lubricant inertia can have a significant effect on bearing performance if Re (C/R) 1.2While the equations of motion are nonlinear when convective inertia is retained, the problembecomes tractable as pointwise lubricant inertia is replaced by its average value, obtainedvia integration across the film.14,15The averaged equations of motion and continuity combine

in a single equation in lubricant pressure For journal bearings:

(7)

Table 2 PARAMETERS FOR THD SAMPLE SOLUTION

Equation 7 was made nondimensional through:

The entries of Table 3 were calculated from Equation 7 for a 160° partial arc journal bearing(L/D = 1.0)

Dynamic Properties of Lubricant Films

Figure 5 represents an idealized configuration where rotor weight (2W) is supported ontwo bearings Under steady load W (Figure 5b), the journal center Ojsis displaced from thebearing center to the steady operating position shown

Rotor response to a small excitation, say imbalance, assuming the bearings to be rigidsupports, will be as shown in Figure 6 (The same curve applies with roiling contact bearings.)Such rotors cannot be operated at the critical speed and can become “hung” on the criticalwhen attempting to drive through When hydrodynamic bearings arc used, the lubricant filmadds another spring (in addition to the shaft spring in bending) and, importantly, considerabledamping Two effects can be noticed in the rotor response curve: (1) critical speed is loweredbelow that calculated for rigid supports, and (2) vibration amplitude is reduced

In this example, the excitation is imbalance and occurs at running speed In practice,exciting frequencies can be different from the shaft speed: magnetic pulls, gear impacts,out-of-round shaft, steam, or aerodynamic forces16 on turbine or compressor blades, etc.The latter has been known to cause large self-excited vibrations In addition, lubricant filmsthemselves can originate destructive self-excited vibrations A classical case is oil whip atslightly less than one-half running speed.17Another oil film phenomenon is a self-excitedvibration at exactly one-half (or other exact submultiple) of the running speed known assubharmonic resonance.18

The rotor-shaft configuration of Figure 5 is reduced to a simple dynamical system ofsprings 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

FIGURE 5 Dynamical elements of rotor-shaft configuration.