Selection of Lubricant

557

Paper 36

SELECTION OF LUBRICANTS

By I?. Beuerlein* and W. H. Karat

INTRODUCTION To DEFINE THE OBJECTIVE of this paper the title needs some comment. The fact that a separate paper deals with the selection of lubricants does not mean that this is an isolated problem. In all the papers previously presented at this Conference lubricants or at least certain properties of lubricants were mentioned. From this one can conclude that the lubricant has to be regarded as an integral part of a machine design. In the same manner in which an engi- neer designs a structural element, for instance a journal using the structural material steel, the structural element lubricating film has to be designed using the structural material lubricant. This way of looking at things has gained more and more supporters in the last few years (I) (2)$-though it must be admitted that even nowadays many machines are still manufactured, the designers of which start thinking of the lubricant when the first test-run is about to commence.

Engineers have certain difficulties with this more modern concept to regard the lubricating film as a structural element and the lubricant as a structural material. The reason may be that in this case we have to deal with a structural element which has, because of the material it is made of, no rigid form but changes its dimensions with the operating conditions and geometrical configuration at the point of contact. However, this very property of the film forces one to look at the film not in isolation but as part of the whole machine.

These thoughts have been expressed explicitly and implicitly by many other authors at this Conference. How- ever, the separate treatment of this topic is justified for a number of reasons. First, it is necessary for economic reasons to limit the number of lubricants necessary for a machine to a minimum. This asks for compromises. Secondly, certain recommendations based on field experience can be given to help the designer in his task to select lubricants. Moreover separate treatment will demonstrate The MS. of this paper was received a t the Institution on 18th July

* F A M Fachausschus, Mineraloel und Brermstoffnormung, Ham-

+ Deutsche Shell A G , Hamburg-Wilhemsburg, Germany. $ References aregiven in Appendix 36.1. Proc Instn Mech Engrs 1967-68

1967.

burg, Germany.

certain difficulties the designer faces, as some of the results of research are not available in a form easily applicable to practical problems.

ANALYSIS OF POINT OF CONTACT When assessing a lubricant in regard to its suitability as structural material for the lubricating film, one has to distinguish between properties which influence its function and those which influence economy. Among the functional properties the most important ones are viscosity and viscosity-temperature relationship. For certain applications extreme pressure (e.p.) properties come into the picture as well. The necessary functional properties can only be considered in connection with a defined point of contact, for instance, a bearing or a toothflank. There- fore, an analysis of the contact conditions is the first step towards selecting a lubricant.

Among the properties which have to be taken into account from the economic point of view are oxidation stability, thermal stability, separation from water and corrosion protection. The level of performance in respect of these properties is dictated to a lesser extent by the conditions at the point of contact than by environmental influences and by the lubricating system. A sharp distinction between functional and economic properties is, of course, not possible in all cases. In modern internal combustion engines, for instance, dispersancy has to be regarded as a functional property, as lack of dispersancy very quickly endangers the function of the engine.

Environmental influences discussed in this paper will not include highly specialized cases like high vacuum (space) or radiation as it is intended to deal preferably with problems encountered in normal engineering practice.

Required functional properties of the lubricant To specify certain functional properties of a lubricant one has to consider two aspects:

(1) Damage to the mating surfaces must be avoided and

(2) Frictional heat must be kept below a safe limit. Normally the first aspect will have priority. In certain

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558 P. BEUERLEINAND W. H. KARA

cases (for instance, at very high speeds) the second criterion might be of overriding importance.

When selecting lubricants for a range of machinery (and more important even for a single machine) it is necessary to keep the number of lubricants as low as possible. For this reason one has to check which point of contact is the most critical.

Journal bearings With journal bearings the Sommerfeld number (So) offers a simple way of comparing different bearings in regard to film thickness. The higher this value the smaller the minimum film thickness will be. The results of hydrodynamic research are available in well-arranged graphs which permit the calculation of the critical dimension of the lubricating film, namely minimum thickness (2)-(6). Although this calculation is simple, a number of uncertainties remain. The first one is the estimation of minimum permissible film thickness. It is usual practice to sum the height of asperities of journal and bearing plus the deformation of the journal over the bearing length.

Further uncertainties in these calculations are the difference in clearance as designed and under operating conditions and the assumption that the viscosity of the lubricant is constant throughout the film. Sufficiently great safety factors are therefore required.

A very simple method to check a journal bearing in respect of safety against contact of the mating surfaces is the Volume Formula derived by G. Vogelpohl (5). This formula is based on the analysis published by Sassenfeld and Walther (3) and allows the calculation of the speed at which transition from mixed friction to fluid friction occurs.

~ x I O - ~ W Nt = -- 7r Q D 2 L

iVi (revjmin), W (kP), C = constant = 1, 7 (cP), L and D (cm). (The assumptions made in this formula are: not too small clearance, good surface finish and LID ratio between 0.5 and 1.5.)

Clearance and film thickness are contained in the constant. Each factor for itself can vary in practice far more than the product of film thickness and clearance, so that in this formula some of the difficulties mentioned before are diminished (7).

The transition speed Nt should correspond to a circumferential speed of 1 m/s with the second condition that the ratio of operating speed to transition speed N/Nt > 3. Under these conditions overheating is in general avoided.

As mentioned before, frictional heat is the second entity which has to be computed. P. Ramsden showed in his paper about the design of large bearings for steam turbines that bearing temperature may be the overriding criterion of bearing design. This depends mainly on the type of operation and bearing metal used. White metals conform more easily to deflecting shafts and are able to embed foreign matter particles. Owing to their comparatively low melting point they are sensitive to bearing temperature. Proc Instn Mech Engrs 196748

Only very rarely this calculation (and possibly optimization) is necessary from the energy point of view. An example of these exceptions is textile spinning machines, which contain a great number of spindles. Here unnecessary frictional losses are reflected in the energy cost. How- ever, in general the frictional heat developed is of interest in relation to the bearing temperature. In comparing different bearings of a machine for friction losses, one can assume that under otherwise equal conditions the heat developed is proportional to &'I2. u3I2. More detailed information regarding the frictional heat developed and the operating temperature of the bearing can be gained by a heat balance.

The computation of the frictional heat developed can be performed with a fair degree of accuracy using the well- established relationship between coefficient of friction and Sommerfeld number (2) (4) (5). The dissipated heat depends, however-apart from the heat carried away by the lubricant-to a great extent on the heat transfer coefficient. The numerical value of this coefficient is influenced by the air flow-rate around the heat dissipating surface. The values quoted in the literature should be applied carefully until practical experience has been gained. This calculation shows, moreover, whether a journal bearing can run safely under mixed friction conditions. Usually this will only be the case at very low speeds, as otherwise the frictional heat developed cannot be dissipated. An example for this is bearings of cement kilns, which have circumferential speeds below 0.1 m/s. Cases are known where, under these running conditions, e.p. oils became necessary.

If it is unavoidable that a journal bearing has to operate in the mixed friction region, one should examine whether hydrostatic lubrication is feasible.

In his paper H. Opitz supplied the necessary information for the design of hydrostatic bearings. Apart from examples of successful designs, the graphs and formulas enable optimization in respect of layout and the choice of viscosity.

Especially with high-speed bearings it can happen that not enough heat can be dissipated over the bearing surface to control the bearing temperature at a reasonable level. In these cases it should be investigated whether an increase in oil throughput would solve the problem. Under these circumstances additional grooves in the unloaded area of the bearing are advisable to enable flow of larger oil volumes through the bearing. As P. Ramsden pointed out, oil inlets and channels in the bearing shell need very careful design to ensure flow of the necessary amount of oil through the bearing.

Water cooling may be advantageous. Because of the intensive cooling action one has to be careful to avoid condensation of water, which could collect in the luboil.

High-speed bearings which are prone to overheating may also pose stability problems. The design charts presented by B. Sternlicht offer a first insight into the likelihood of whirl. The critical Sommerfeld number for a certain system can to a certain extent be varied by altering the viscosity of the lubricant. Again, this cannot be

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SELECTION OF LUBRICANTS 559

regarded as an isolated problem as lowering the viscosity of the lubricant could lead to unsatisfactory load-carrying capacity in the start and stop phase of operation.

Anti-friction bearings In contrast to journal bearings, the possibilities of calculating the necessary lubricant viscosity for anti-friction bearings are rather more limited. As was shown by T. E. Tallian the results of elastohydrodynamic research and flash temperature theory are applicable to the lubrication of anti-friction bearings. Besides the rather straightfor- ward influences of film thickness and contact temperatures on different forms of wear, even plastic deformation and fatigue phenomena are affected by lubrication. With the information available today rolling contacts can be designed in a comparable manner to conformal contacts. Owing to the high contact stresses statistical considerations are introduced and a number of stress parameters have to be taken into account. However, from the point of view of the engineers in the field it would not be advisable to rely solely on these theoretical results when selecting a lubricant. To a far greater extent than with journal bearings secondary influences, as, for instance, friction between rolling elements and cage, have to be taken into consideration. Based on a wealth of test results and field experiences the manufacturers of these bearings have developed empirical formulas and graphs, which give very good indications regarding the necessary viscosity of the lubricant in relation to specific operating conditions and bearing type and size. This is in any case true for normal conditions. Extremely high loads, speeds and temperatures require detailed analysis on the lines described by Tallian and empirical tests developed by the bearing manufacturers.

The empirical data for the required viscosity are, of course, related to the operating temperature of the anti- friction bearing. Because of the influence of sliding seals and cage friction the calculation of bearing temperature suffers from great uncertainties.

In most cases one will try to lubricate an anti-friction bearing with an oil that is needed at another part of the machine, for instance, for the gears. With anti-friction bearings this is more practicable than with journal bearings as they are less sensitive to viscosity from the wear and load-carrying point of view. On the other hand one has to be careful that the oil supply to these bearings (for instance, with splash lubrication) is not too generous as the operating temperature of anti-friction bearings is much more sensitive to the amount of oil in the bearing due to churning than to viscosity.

Gears The lubrication of gears is a considerably more complex field than the lubrication of, for instance, journal bearings. Even if one deals only with spur gears, it is difficult to decide under which conditions of load and speed damage to a pair of toothflanks has occurred. Proc Instn Mech Engrs 1967-68

Usually the requirements and operating conditions dic- tate which materials and heat treatments have to be used.

If one succeeds in establishing an uninterrupted lubricant film between the toothflanks practically no wear will occur. The first considerations in the design stage should therefore include the application of elastohydrodynamic theory to calculate the necessary viscosity. At this stage it should be checked whether the film thickness between the flanks can be increased by modifications in tooth geometry, taking other considerations like strength of the root of the tooth into account.

It holds generally, of course, that the possibilities of changing tooth geometry are more limited when fewer design parameters are available. This is especially the case with gear ratio and centre distance. According to elastohydrodynamic theory, the minimum film thickness changes only slightly with load. With given pinion speed and pressure angle the viscosity of the lubricant and pitch circle diameter have the greatest influence. However, film thickness varies proportionally to pitch circle diameter and with the square root of viscosity. Moreover it must be taken into account that under otherwise equal conditions a smaller modulus results in less variation in film thickness over the path of contact. These remarks only emphasize the importance of considering the lubrication of the tooth flanks at the very first design stages.

The calculation of film thickness suffers from two par- ticular uncertainties. First, the effective viscosity is the one at the bulk temperature of the flanks. For this temperature there are indications in the literature and recently experimental work has been published by G. Lechner (9). However, these values should only be regarded as guidance until further results are available. Secondly, as with all calculations of lubricating films, estimation of permissible values for minimum film thickness is difficult. As with bearings one uses the sum of the height of asperities as the limiting value until further experimental evidence makes a revision of this recommendation advisable.

If film thickness is insufficient, continuous wear will be encountered. At higher specific loads and at the same pitch line speed the ‘pitting barrier’, as it was called by H. Blok (IO), will eventually be reached. Though research in this field has not come up with final results it can be said that pitting formation is influenced by a number of factors which all have to be considered in the design stage.

Some papers at this Conference refer to Dawson’s work (11)-(14)~ in which the influence of initial surface roughness in relation to oil film thickness on pitting formation is demonstrated. Qualitatively similar conclusions were reached by G. Niemann and co-workers in their empirical investigation with actual gears (15). Further- more G. Niemann looked into the relative influence of other design parameters. His results can be summarized as follows.

With small tooth errors an increase in peripheral speed leads to extended pitting life due to greater film thickness. However, if tooth errors are great an increase in speed has a negative effect due to an increase of dynamic tooth loads.

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560 P. BEUERLEIN AND W. H. KARA

A greater pressure angle results in greater radii of curvature which reduces surface stresses and leads to greater film thickness. The overall effect is therefore a greater than linear increase of pitting life with pressure angle.

With smaller moduli bending stresses are increased. Therefore tangential tensile stresses at the toothflanks are higher which favours pitting formation.

Pitting resistance increases with viscosity of the lubricant proportional to to yo,*, Furthermore, it was found that certain synthetic lubricants increased pitting life considerably. Whether this is due to their different viscosity- pressure relationship or to their chemical structure which may influence frictional properties under boundary conditions is not known yet.

Fatigue strength of the gear materials has an overriding influence. Soft gears behaved relatively better when running against hardened ground flanks. This may be due to the smaller average tooth error resulting in thicker lubricating film and smaller dynamic loads.

Continuous smooth wear is observed at insufficient film thickness and low pitch line speeds. With increasing speed the elastohydrodynamically formed film on the flanks increases in thickness quite rapidly. On the other hand, at higher speeds and high loads temperatures at the points of contact become so high that local welding occurs. This limit to safe operation was called the ‘scuffing barrier’ by H. Blok (10).

The calculation of safety against scuffing of lubricated toothflanks is still rather difficult. Two procedures have been developed in the course of time. One is based on Blok’s flash-temperature hypothesis and starts from theoretical considerations. The other is an empirical method developed by G. Niemann. The paper by B. W. Kelley describes the calculation based on Blok’s hypothesis. Here some comment from the point of view of a designer seems necessary to show some of the difficulties arising in practical applications.

In the formula for flash temperatures the coefficient of friction appears. Since Ohlendorff’s thesis (16)~ certain data based on experiments with actual gears are available. B. W. Kelley and A. J. Lemanski present in their paper a formula for the instantaneous coefficient of friction. How- ever, as this formula is based on experiments with disc machines it should be used with a certain reserve, which means some factor of safety. The situation in respect of limiting values for flash temperatures (or better for contact temperatures which are the sum of bulk temperature of the teeth and flash temperatures) is similar. Based on tests with geometrically simpler configurations like discs, limiting values for flash temperatures using straight mineral oils have been published. T. I. Fowle (17) analysed test results with the I.A.E. and F.Z.G. machines in respect of critical flash temperatures. Dudley in his well-known book (18) presents values for e.p. oils as well. As these latter values seem to contain sufficient safety the use of the data quoted should not contain too great a risk. Proc Instn Mech Engrs 1967-68

It may be permitted to discuss here the procedure developed by G. Niemann in some more detail, as it was mentioned in this meeting only in passing (19). All calculations in this procedure are based on test results with the F.Z.G. rig. This rig has been described in the literature (20)-(22). G. Niemann postulates that a lubricant has to lubricate satisfactorily in the F.Z.G. rig at three to five times higher contact stresses than in a practical case in the field. When calculating effective contact stresses tooth geometry, operating conditions, empirically determined influences of dynamic loads, shocks and uneven load distribution are taken into consideration. Based on the empirically determined dependence of the permissible contact stress of speed, a quotient ‘contact stress of the test over actual contact stress in the gear’ can be formed, which should be three to five for safety reasons. The transposi- tion of permissible contact stresses ascertained with the test spur gears to other types of gears by means of ‘substi- tute spur gears’ was proved correct in a number of cases by experiment. This is even true to a certain extent for hypoid gears. However, one should not overlook that the comparative tests were run on a test bench under more favourable conditions than those encountered in an actual axle on the road.

The cases mentioned so far permit a comparatively safe calculation of the most important functional properties of the lubricant (which should, of course, be supported by field experience). Now we shall discuss some examples, which do not allow (or at least only rather incompletely) the necessary functional properties to be found by calculation. Here one uses empirical methods when selecting the lubricant.

(a) Most important in this respect are reciprocating engines. A relatively great amount of empiricism is encountered in this field. This handling appears reasonable for different reasons. We have the case here where one oil must be selected for a great number of contacts, which are quite differently stressed mechanically and thermally. This is true at least for the most widely represented trunk piston engines. Especially for high-speed piston engines the number of viscosity grades available in a market is limited. Thus the designer is forced to accept one lubricant with a certain nominal viscosity for all contacts and to design these contacts in regard to dimensions and material in such a way that trouble-free operation is possible. All bearings and other points of contact can be cal- culated. Based on research in the last decade even journal bearings with loads varying in magnitude and direction can be treated mathematically with very good correlation with experimental results, as Campbell and co-authors have shown.

Nevertheless the calculation of these bearings is difficult and the effort considerable. In contrast to bearings with constant stationary load, a squeeze film is generated, which is in general of advantage and improves load-carrying capacity.

Gudgeon-pin bearings need very careful development, as in this case a number of unfavourable influences come

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together (high temperature, small angular movement, difficult oil supply).

The background of the lubrication of valve trains has been considerably enlarged in the last years. The applications of elastohydrodynamic theory and flash-temperature hypothesis offer new insights and enable a more uniform stressing of the sliding surfaces to be reached, as H. Naylor has shown. Nevertheless, because of the very complex interaction of static and dynamic forces, materials and geometry, final conclusion about the lubrication of this contact can only be drawn after experiments.

Though the lubrication of the cylinder bore is hydrodynamic over a considerable part of the stroke the greatest difficulty exists here to keep the ring area of the piston clean to ensure free mobility of the piston rings. The influence of nominal viscosity on cylinder wear can be shown, especially at lower viscosities and high speeds and temperatures. However, this increase in wear is considerably less than the mechanical wear encountered with sticking rings or the chemical wear which occurs with diesel fuels of high sulphur content and operation at low jacket temperatures.

The wear caused by chemical mechanism and the necessary counter-measures are to a much lesser extent mathematically accessible than pure mechanical wear. The amount of sulphur in a diesel fuel gives certain indications regarding the necessary alkalinity of the engine oil. The required thermal and oxidation stability for certain engines and operating conditions, as well as the required ability of the oil to disperse foreign matter particles, can only be established by experiment.

In the course of the years a number of test methods have been developed to check these requirements. In most cases these are full-scale engine tests. It is very difficult to simulate the complex conditions at the different points of contact in glassware type of tests and this problem has not so far been solved satisfactorily. T. I. Fowle quoted in his paper some of the engine tests in use for this purpose.

(b) Furthermore, from the great number of cases where theoretical treatment is not yet possible, we should like to mention those which suffer from frictional vibrations, which phenomenon is known as ‘stick-slip’. Usually this vibration occurs at very low sliding speeds. Here, too, the actual point of contact should not be treated in isolation. The inertia and elastic properties of the sliding parts, stiffness of the driving member, material combinations, operating conditions and lubricant have likewise influence on the occurrence of frictional vibrations. In this context D. Tabor referred in his paper to the work of Rabinowicz who assumes an influence of contact time upon the static friction.

Stick-slip is most often observed on tableways of machine tools with low feed. It could be shown empirically that certain additives can influence the level of static friction in relation to kinetic friction and thus prevent frictional vibrations. In these cases empirical methods have also to be applied when selecting a lubricant.

(c> For trouble-free function of a lubricant, low tem- Proc Instn Mech Engn 1967-68

perature behaviour can be of paramount importance in some applications. This is especially critical when only starting occurs at low temperatures at the contacts, while in operation these contacts become hot. The most common cases are lubrication of aero and automotive engines.

Startability of automotive engines at low ambient temperatures is governed partly by the viscometric properties of the lubricant. As engine lubricants behave at low temperatures like non-Newtonian fluids, i.e. their viscosity depends not only on temperature and pressure but also on shear, viscosity measurements have to be performed at realistic shear rates. This can be done by using the engine itself as a viscometer and motoring it or by employing so- called rotational viscometers. For practical purposes bearing simulators have been developed. The permissible maximum viscosities of the lubricant at given temperatures and shear rates should be properly established in tests employing realistic temperatures. It seems that some manufacturers try to be so far on the safe side that un necessarily low viscosities (or unnecessarily low test tem peratures-which lead to the same effect) are specified. This forces the lubricant manufacturer to incorporate rather low viscous base stock in the engine oil which may have other adverse effects, like higher oil consumption.

Apart from good flow properties at low temperatures, excellent viscosity-temperature characteristics are important to ensure sufficient load-carrying capacity at high operating temperatures. In extreme cases these requirements cannot be fulfilled by lubricants based on mineral oil, so that synthetic fluids might be necessary.

Concluding the remarks on functional properties, it can be said that in all cases in which hydrodynamic lubrication is feasible, the thickness of the lubricating film can be computed.

If the resulting film under certain conditions is too thin or if the heat generated at the points of contact cannot be removed fast enough, theoretical treatment is only possible to a limited extent. Here an empirical approach must be used.

Required economic properties of the lubricant In connection with the properties so far discussed, it is the task of the designer, for given operating conditions, to select the geometry of the contact and the materials (which includes the lubricant) in such a way that wear and energy losses are minimized. When selecting a lubricant in respect of economic properties the main objective is to safeguard trouble-free operation over a long span of time. From the functional aspect, after analysing the operating conditions, one should not confine oneself solely to selecting a suitable lubricant, but one should examine critically whether changes in design can avoid exposing a contact to conditions which put unnecessarily great stresses on the lubricant.

Oxidation Among the most important economic properties are

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oxidation stability and stability against thermal decomposition. As T. I. Fowle has mentioned, all lubricants tend to oxidize in the presence of oxygen. This process is influenced, as are all chemical processes, by temperature and catalysts. By modern refining processes and adding anti-oxidants, it is possible to lengthen the service life of an oil considerably. Nevertheless, a designer should always try to avoid unnecessary heating of the contact to be lubricated.

The same is true in respect of the catalytic influences on oil oxidation. It has long been known that copper and copper-containing materials in the oil circuit influence the rate of oxidation catalytically. Nevertheless there still exist many designs in which this well-known fact is not appreciated.

Oxidation stability can only be tested in comparative experiments under defined conditions. T. I. Fowle men- tions in Table 19.4 of his paper a number of test methods which are in use today. When judging test results and applying them to practical cases, great care must be taken as oxidation processes are strongly influenced by the test conditions.

In general access of oxygen to lubricating systems cannot be avoided and the rate of oxidation is high at high temperatures. Fowle showed that the upper temperature limit for mineral-based oils lies between 150 and 200°C in the presence of oxygen, depending on type of crude, extent of refining and inhibition. Stability against thermal decomposition plays a comparatively small role in lubrication.

Corrosion protection In general mineral oils with oxidation inhibitors will contain anti-corrosion additives as well. The reason for this

additive combination is that polar substances are formed in non-inhibited oils in the course of oxidation which will improve the corrosion protection of a mineral oil. If this oxidation process is interrupted due to the presence of anti-oxidants, these polar substances (or other substances giving a similar effect) must be added.

Results of extensive test series have been published (21), in which straight mineral oil was tested parallel to an inhibited oil containing anti-oxidants and anti-corrosion additives as lubricants for a large tool machine complex. Figures 36.1 and 36.2 show the reduced maintenance and the extension of oil-change periods when using the inhibited oil. The saving in maintenance cost and oil con- sumed is far greater than the increase in cost due to the use of the higher quality oil.

The anti-corrosion additives mentioned prevent (or at least reduce) moisture corrosion. Moreover corrosive attack of sensitive materials (mainly yellow metal) in machine parts wetted by oil may occur. This can be observed especially with certain e.p. additives and high oil temperatures. If for certain reasons the use of yellow metal cannot be avoided (for instance, for thrust washers in a gear) care must be taken in this respect when selecting the oil as otherwise these parts may corrode and wear away very rapidly.

Separation from water Occasionally good demulsibility of the oil must be specified, as in some systems even with careful design and selection and maintenance of seals, entrance of water or steam into the luboil system cannot be avoided. Another reason for the presence of water may be condensation in cooler parts of the system. If one does not succeed, by

Averaqe Amount Of Time Saved Per Machine Per Year

HOURS Per Machine (Five Yeor Study I -

Average Maintenance Hours Per Yeor Per Machine Based On A Ten Yeor

Machines Lubricated Machines Lubricated With With Medium Ouolity A Goad Ouol~ty Rust And Straight Mineral Oi l Oridotion Inhibited Oil

F ~ . 36.1. InfZuence of lubricating oil quality on maintenance cost of tool machines Proc Instn Mech Engrs 1967-68 Vo1182 Pt 3A

SELECTION OF LUBRICANTS 5 63

Average Frequency O f Lubrication Average Frequency Of Lubricotion Of Machines Lubricoted With A Of Mochines Lubricoted With A Medium Ouolity Straight Mineral Oil Good Quality Rust And Onidolion

Inhibited Oi l

Fig. 36.2. Influence of oil quality on oil change periods in tool machines

proper design of the system, in settling the water and draining it from the bottom of the container, among other things emulsification, foaming and faster aging will occur. Apart from the fact that the container may overflow, foaming oil will be compressible and unsuitable for hydraulic use. Furthermore oils containing water favour corrosion.

There are different test methods which allow determina- tion of the demulsibility of lubricating oils. These methods usually involve stirring in water or blowing the sample with steam. The reliability of all these methods is not undisputed.

Sealing properties If a machine works in a dusty atmosphere, one must try all possible means to prevent access of dust to the contacts to be lubricated. This is mainly a problem of proper seal design. Sealing can be greatly improved if a grease is used instead of oil. However, in this case grease must be supplied at regular intervals as only the grease collar flowing out of the bearing will seal. The use of grease in journal bearings is limited to lower speeds, as the removal of heat by the lubricant is insufficient. According to experience the maximum permissible sliding speed is about 2 m/s.

Such a low limit of speed does not exist with anti-friction bearings. On the other hand these bearings are more sensitive to overfilling. However, grease lubrication has here the same advantage of acting as an additional seal.

Necessary alterations of design As was mentioned before, geometry of the mating sur-

faces, lubricant, material of the contacting parts and operating and environmental conditions influence the opera- tional performance of the point of contact.

In addition to these factors the designer has the possibility of reducing mechanical and thermal stresses imposed on the point of contact.

An example of reduction of mechanical stresses in a bearing housing is shown in Fig. 36.3. This work was published by A. Buske (24). To increase the load-carrying capacity of main bearings of a rotary piston engine he utilized the elastic deformation of the bearing housing to ensure uniform film thickness over the length of the bearing. Conventionally designed highly loaded journal bearings tend to suffer from high edge pressure due to deflection of the journal which results in very small film thickness at the bearing edge. By altering the wall thickness of the bearing housing Buske succeeded in designing bearings which performed satisfactorily up to specific loads of 1200 kP/cm2. While the bearings did not fail, the journals sheared off a t these high loads due to high frictional torque.

Another example for sound design is concerned with the reduction of thermal stresses. Drying kilns are very often equipped with ventilators to ensure good circulation of hot air. If ventilator and electric motor are arranged in the kiln itself extremely high bearing temperatures may be reached which necessitate the use of very expensive synthetic products or lead to very short overhaul periods. If on the other hand only the ventilator itself remains in the high temperature zone and the electric motor is placed outside the kiln, lubrication of the bearings no longer presents any problem. The same is, of course, true for lubrication in low temperature environments. D. H. Tantam quoted in his paper examples of designs which avoid

Proc Instn Mech Engrs 1967-68 Vol182 Pt 3A


X X

I

X

at; 7000

m

m

2w

0 X

P’ P’

a f i 7000

eoo

#

400

ZW

X

P” P’

X

I a- Yergekhsmafl f i die mgsie Schmierspo fldicbe I < I

a - fe@khsmoR fir die eogste Schm;ersbaltdicke

p = oil pressure in lubricating film. a = relative measure for minimum film thickness. C)lfilnidruck = oil-film pressure.

Fig. 36.3. Influence of bearing design on jilm thickness

operation of bearings and seals under extreme low temperature conditions.

A third example deals with the ingress of water into large lubricating oil systems which are common, for instance, with steam turbines. The oil temperature in the tank is usually around 50°C. If the tank is situated at a position where a draught of air passes oTer the lid, then condensation of water underneath the lid is very likely with the consequent accumulation of water in the oil charge.

Foaming of the oil is in very many cases caused by faulty design of the return line into the tank. T o avoid splashing, either slightly inclined baffle plates or sieves which allow the entrained air to escape or return lines below the lowest oil level usually cure excessive foaming. In this way the use of special anti-foam additives, which may have other disadvantages, can be avoided.

LUBRICANTS Liquids Although in the hydrodynamic sense gases can be regarded as liquids in the wider sense, we shall confine ourselves in this chapter to those fluids which are liquid under normal operating conditions.

Liquids are the most commonly used lubricants. The reasons are that due to their (compared with gases) high viscosity, hydrodynamic lubrication is possible over a very wide range of operating conditions. Application to the point of contact is easy and the liquid draining out of the contact carries away a considerable amount of frictional heat. Proc Instn Mech Engrs 1967-68

Each lubricant has a number of limitations which must be taken into account in the selection.

Mineral oils Among the liquids, mineral oils or lubricants based on mineral oil are by far the most widely used. The reason is that mineral oils are available in a wide variety of viscosities. This enables the designer to be more flexible in the choice of contact dimensions and specific loading. The temperature range in which mineral oils can be used satisfactorily covers all temperatures encountered in normal engineering practice. The engineering materials in common use are not attacked by mineral oils and a certain degree of rust protection is offered. Vapour pressure is low so that evaporation losses at normal operating temperatures are negligible.

However, there are certain exceptional cases where the physical and chemical properties of mineral oils prohibit their use. Among these limitations the most important ones are:

Viscosity-temperature behaviour and low temperature properties. In extreme cases, as mentioned before, fluidity at very low temperatures and sufficiently high viscosity at elevated temperatures are required. This asks for extremely good low temperature behaviour and better viscosity-temperature relationship than can be offered by mineral oils.

The service life of mineral oils in the presence of oxygen becomes unsatisfactory above about 130°C bulk temperature.

Mineral oils, as do all liquids, drain away from the VoZ 182 Pt 3A


point of contact. In certain applications this can lead to sealing difficulties, especially when contamination of other products in the neighbourhood of the point of contact with the lubricant must be avoided.

Mineral oils are not fire-resistant (although water-in- oil emulsions offer a great improvement in this respect).

Synthetic JIuids As T. I. Fowle has pointed out a great variety of synthetic fluids are available. The targets for the development of these fluids were mainly low temperature fluidity, good viscosity-temperature relationship and high thermal stability. Furthermore, certain synthetic fluids haye better fire-resistance than mineral oils.

In respect of the limitations, only general remarks are possible due to the wide variety of chemical compositions. Apart from the cost, the following technical and physiological aspects need attention :

Conventional sealing materials deteriorate in the presence of many synthetic fluids.

Owing to hydrolysis, corrosive attack on engineering materials may be encountered with certain fluids. Others attack some of the metals commonly present in engineering designs.

Physiological and toxicological hazards with some fluids need special consideration.

Owing to the high density of some synthetics, water separation offers greater problems,

Process JIuids In this context we shall consider fluids which are basically not especially suitable as lubricants but which are present or used in a process. Depending on the kind of process, other types of lubricants might introduce sealing difficulties or might be, due to ambient conditions, completely unsatisfactory.

In underwater pumps, for instance, bearings can be designed to use water as lubricant. In spite of the defici- encies of water, namely low viscosity, poor wettability, poor boundary lubrication, corrosive attack on iron alloys and high vapour pressure, very reliable water-lubricated bearings can be designed if the disadvantages quoted are taken into consideration. Examples from the nuclear engineering field were quoted by H. H. Heath.

Other examples for the use of process fluids as lubricants have been given by D. H. Tantam in his paper about ‘Cold environments’. Special emphasis is placed in this paper on the fact that certain process fluids have poor lubricating properties and act mainly as media to remove frictional heat. In these cases appropriate material combinations for the sliding parts have to be used to ensure trouble-free operation.

Plastic solids This family of lubricants comprises mainly the so-called ‘greases’, which consist of a liquid phase that has been gelled with a thickener. The vast majority of greases use mineral oil as the liquid and metallic soaps as thickeners. Proc Insrn Mech Engrs 1967-68

As N. A. Scarlett has pointed out, greases do not flow unless a minimum shear stress is applied. This rheologic behaviour ensures that the grease stays at or near the point of contact to be lubricated. This property enables very simple designs, especially when lubricating rolling bearings. Moreover, contamination of foodstuff, etc., in the neighbourhood of bearings can be avoided.

The second advantage of greases is that the grease collar being pressed out of a journal bearing acts as a seal. For this reason grease lubrication can be very useful in dusty atmospheres.

The high viscosity of greases at low rates of shear, caused by non-Newtonian behaviour, enables hydrodynamic lubrication under conditions where contacts lubricated with fluids would run in the thin film region.

Depending on the type of thickener and the type of liquid phase, greases of different thermal stability and behaviour against water are manufactured. Greases based on mineral oil have an upper operating temperature limit of about 140”C, depending on severity of operation and re-lubrication periods. However, higher operating temperatures are possible as long as the bearings are lubricated more often.

The mentioned advantage of high viscosity at low shear rates can be a drawback when greases have to be fed through long supply lines to the points of lubrication as in centralized systems. In these systems the requirements in respect to pumpability may indicate the selection of greases to a far greater extent than the requirements of the point of lubrication.

The selection of greases depends much more on empirically gained experience and testing in simulators than is the case with fluids. This is caused mainly by the unusual flow behaviour of these so-called Bingham bodies, which makes mathematical treatment very involved.

The greatest disadvantage of greases is, of course, that they cannot act as heat transfer media to remove frictional heat from the point of contact. In addition to the recommendations given by N. A, Scarlett in his paper on the use of greases in rolling bearings, it should be mentioned that, because of lack of heat transfer, greases can be used in journal bearings only up to peripheral speeds of about 2 m/s.

Gases W. A. Gross has mentioned the advantages gases offer as lubricants. Especially in high-speed applications, their low viscosities (compared with fluids) are favourable (Sommerfeld numbers with liquid lubricants considerably below 1). This very property on the other hand necessi- tates a high degree of precision in manufacture to avoid contact of the mating surfaces. Because of the relatively low load-carrying capacity of gas films, only conformal surfaces can be lubricated.

Besides the high speed and precision machinery mentioned by Gross, bearings in nuclear engineering lubricated with coolant gases were described by H. H. Heath in

Vo1182 Pt 3A


Small High

Moderate Small

his paper. This application makes use of the great resistance against radiation and thermal decomposition of certain gases.

Only for very low requirements Necessary oil throughput through bearings

Only for moderate circumferential speeds Onlv for moderate circumferential speeds

must be ensured

Solids The most widely used inorganic solid lubricants are graphite and molybdenum disulphide. Although the mechanism of lubrication with solids is still a matter of discussion, practical and research experience shows cases where lubrication with solids offers advantages.

The higher thermal stability of graphite and molybdenum disulphide compared with mineral oils is bene- ficial at temperatures above 200°C. Journal bearings of trucks in ceramic kilns are very often lubricated with graphite suspended in an easily vaporizing hydrocarbon. Rolling bearings for this application are available with molybdenum disulphide bonded in a surface layer.

Lubrication of slow moving flat surfaces with molybdenum disulphide has proved advantageous due to the low coefficient of friction. The same is true for the lubrication of highly stressed threads.

The application of molybdenum disulphide in high vacuum is, as was shown by Anderson and Glenn, not so far solved to complete satisfaction.

The success of lubrication with solids depends to a large extent on the proper application to the surfaces to be

Journal bearings

Rolling bearings

Gears

Lubricant

Oil

Grease

Oil

Grease

Oil

Grease

Lubricating system

By hand Circulating system

Ring lubrication Sinter bearings

By hand Centralized system

Oil mist

Bath

Splash

Packed

Centralized system

Bath

Circulating system

By hand Housing filled

Table 36.2

'Maintenance cost

High Low

Low Low

High Low

Low

Low

L O W

Low

Low

Low

Low

High Low

Table 36.1. Approximate cost reiative to mineral oil

Synthetic Oils Organic acid esters 5-1 5 Triarylphosphate esters 5 Chlorinated diphenyls 5

Polyglycol 5 Fluorocarbon 300

Silicones 30-70 Polyp heny lether 250

Process Fluids -

Plastic Solids Greases based on mineral oil Greases based on di-esters 10-30 Greases based on silicones 60-250

1-2

- Gases

Solids Graphite 2-5 Molybdenum disulphide 10-15

Solids Graphite 2-5 Molybdenum disulphide 10-15

lubricated. In the simplest form rubbing on to the surface to establish orientation of the lamellae seems essential. In more sophisticated applications bonding to the surface by means of metals or resins is recommended.

The best known organic solid lubricant is polytetra- fluorethylene (p.t.f.e.). Strictly speaking p.t.f.e. cannot be

Investment cost

Low High

LOW Low

L O W High

High

Low

Low

Low

High

Low

High

High Low

Rate of ieat removal by lubricant

Remarks

Nil Nil

and low specific pressures Only for very low requirements Good pumpability of grease required if

long lines to bearings and operation at low ambient temperature

Small If compressed air in necessary quantity and cleanliness available, investment

Small

Moderate

Nil

Nil

Moderate

High

Nil Nil

costs are moderate Careful design and filling required to avoid

excessive churning Careful design of housing (e.g. gears)

necessary to ensure adequate oil supply Overfilling must be avoided. Maintenance

costs are only low if re-lubrication period not too short

Possibility for used grease to escape must be provided. Shield delivery lines from heat

Careful design of housing required to ensure adequate oil supply to all gears and to avoid excessive churning

Jets have to be properly designed and placed to ensure even oil distribution and heat removal

Only for very low requirements Only for small, low-speed gears

Proc Instn Mech Engrs 1967-68 Vol182 Pt 3A


regarded as a lubricant but as a structural material with unique physico-chemical properties-low friction, little adhesion to other substances, chemical resistance against attack of almost all substances up to about 250”C, and fire resistance. Owing to the low mechanical strength and bad thermal conductivity p.t.f.e. is used normally either in thin foils or filled with metal powders.

Cost of lubricants It is not the intention to quote here prices for different types of lubricants, These prices vary due to individual quality, additive-content, offtake, etc. Table 36.1 gives only a general idea of relative costs of different lubricants.

For process fluids no relative cost is quoted as it is assumed for this purpose that the fluid in question is available anyway in the required amount and purity. The same is true for gases.

Cost of maintenance and investment Generally it can be said that when less maintenance effort is intended, the higher is the necessary investment cost for the lubricating system. Table 36.2 gives a very general survey over different lubricating systems and some indications of maintenance and investment costs. More detailed information can only be supplied for specific cases.

In this context development costs for special types of bearings, for instance, those which are lubricated with gases or certain process fluids, have to be taken into account as this special design effort may influence the cost of lubrication considerably.

APPENDIX 36.1

REFERENCES

(I) BLOK, H. ‘The conceptual integration of tribology and tribo- design into machine design’, Address to the American Society of Mechanical Engineers (New York), 30th November 1966.

(2) THEYSE, F. H. ‘Fundamentals of hydrodynamic lubrication and their consequences in design engineering’, Wear 1964 7 Part I: 419; Part 11: 477.

(3) SASSENFELD, H. and WALTHER, A. ‘Gleirlagerberech- nungen’, VDI-Forschhfr 1954, 441 (Diisseldorf, VDI- Verlag).

(4) BARWELL, F. T. Lubrication of bearings 1956 (Butter- worths Scientific Publications, London).

(5) VOGELPOHL, G.

(6) PINKUS, 0. and STERNLIMT, B.

Betriebssichere Gleitlager 1958 (Springer- Verlag, Berlin-Gtrttingen-Heidelberg).

Theory of hydrodynamic lubrication 1961 (McGraw-Hill, New York-Toronto- London).

Einfiihrung in die Schmiertechnik, Part 1, 1961 : Part 2. 1966 (Karl Marklein-Verlaa, Diisseldorf).

(7) GBTTNER, G. H.

(8) VOGELFOHL, G. - ‘Geringste zul;issige Sch&rschichtdicke und tfbergangsdrehzahl’, Komtruktion 1962 14 (Heft 12), 461.

‘Die Erwarmung der Zahnrader im Betrieb’, Schmiertechnik 1967 14 (No. l), 13.

(10) BLOK, H. ‘Lubrication as a gear design factor’, Inter- national Conference on Gearing (London) 1958.

(11) DAWSON, P. H. ‘Pitting of lubricated gear teeth and rollers’, Power Transmission 1961 30,208.

(12) DAWSON, P. H. ‘The effect of metallic contact on the pitting of lubricated rolling surfaces’,J. mech. Engng Sci. 1962 4, 16.

(13) DAWSON, P. H. ‘The effect of metallic contact and sliding on the shape of the S-N curve for pitting fatigue’, Proc. Symposium on Fatigue in Rolling Contact 1964 (Instn Mech. Engrs, London).

‘Experiments on the effect of metallic contact on the pitting of lubricated rolling surfaces’, Proc. Symposium on Elastohydrodynamic Lubrication 1965 (Instn Mech. Engrs, London).

(IS) NIEMANN, G. and RICHTER, W. ‘Versuchsergebnisse zur Zahnflankentragfahigkeit’, Konstruktion 1960 12: Parts I, 11: (No. 5), 185; Part 111, IV: (No. 6),236; PartsV,VI: (No. 7), 269; Part VII : (No. S), 319; Part VIII : (No. 9), 360; Part IX: (No. lo), 397.

(16) NIEMANN, G. and OHLENDORFF, H. ‘Verlustleistung und Erwarmung von Stirnradgetrieben’, Z. Ver. dt. Ing. 1960 (No. 6), 216.

(17) FOWLE, T. I. ‘Correlation of the I.A.E. and F.Z.G. gear rigs by the critical scuffing temperature theory’, Proc. Sym- posium on Gear Lubrication 1966 (Instn Mech. Engrs, London).

(IS) DUDLEY, D. W. PracticaE gear design 1954 (McGraw-Hill, New York-Toronto-London).

(19) NIEMANN, G. Maschinenelemente, Vol. 2: Getriebe 1961 (Springer-Verlag, Berlin-Gottingen-Heidelberg).

(20) NIEMANN, G., RETTIG, H. and LECHNER, G. ‘Scuffing tests on gear oils in the F.Z.G. apparatus’, A.S.L.E. Trans. 1961 4,71.

(21) NIEMANN, G. and ASSMANN, H. ‘Experience with the F.Z.G. apparatus for testing gear oils’, J. Insf. Petrol. 1966 52,88.

‘Der F.Z.G.-Zahnradkurztest zur Prufung von Getriebetrlen’, Erd6l Kohle 1954 7 , 640.

‘Advantages of a planned lubrication pro- gramme’, Mech. Engng 1955,23.

‘Der Einfiuss:der Lagergestaltung auf die Belast- barkeit und die Betriebssicherheit’, Srahl Eisen 1951 71 (No. 26), 1420.

(9) LECHNER, G.

(14) DAWSON, P. H.

(22) NIEMANN, G. and RETTIG, H.

(23) HARTLEY, A. L.

(24) BUSKE, A.

Proc Instn Mech Engrs 1967-68 37

Vol182 Pt 3A

Selection of Lubricant

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