Lubricant selection: Function and composition - STLE · grade selection based on oil temperature, rpm and moderate shaft loads (150-200 psi). For higher loads (~300 psi), raise the

18 A P R I L 2 0 0 8 T R I B O L O G Y & L U B R I C A T I O N T E C H N O L O G Y18 A P R I L 2 0 0 8 T R I B O L O G Y & L U B R I C A T I O N T E C H N O L O G Y18 A P R I L 2 0 0 8 T R I B O L O G Y & L U B R I C A T I O N T E C H N O L O G Y18 A P R I L 2 0 0 8 T R I B O L O G Y & L U B R I C A T I O N T E C H N O L O G Y

ubricant selection is a pivotal starting point in thepursuit of precision lubrication practices. All theeffort applied to clean delivery and handling, filtra-

tion, dehydration, alignment, balancing, etc., is lost if thelubricant selected for the application cannot support themachine’s demands. Many criteria must be consideredwhen selecting a lubricant for a set of machines ormachine components.

Lubricant chemistry has an influence on the finaldecision. The basestock (synthetic or mineral) and theadditive systems in use (EP, AW, R&O) exert tremen-dous influence on the performance qualities of thelubricant. It is common to have multiple choices for agiven grade and type of lubricant within a productline (i.e., ISO 220EP/SAE 90EP gear oil), with eachexample designed to perform effectively for a set ofgiven conditions.

The machines themselves have an influence onfinal lubricant selections. OEMs design and buildmachines for general types of service. A gear drive

manufacturer could not possibly consider each andevery type of application for a given make and modelwhen it begins the design process. Designers can build

BEST PRACTICES NOTEBOOK

Lubricant selection:Function and composition

Lubricant selection:Function and composition

Make sure you’re specifying the right product by using

a system based on objective, repeatable and widely

recognized engineering practices and principles.

By Mike Johnson, CLS, CMRP

Contributing Editor

LArticle highlights

:■ Lubricant selec

tion is the

starting point for precision

lubrication practices.

■ Guidelines for selecting lubri-

cants for plain and element

bearings, gears and hydraulic

systems.

■ Analyzing the additive mix for

each component.

■ Making technically accurate

baseline lubricant selection

decisions without exacting

mathematical expressions.

sufficient durability into the basic design thatallows customers to use those products for similaruses within widely varying production environ-ments. The nature of the production environment(wet, dry, hot, cold, abrasive dirt, harsh chemicalexposure, steady state or intermittent operation,etc.) influences the degree of effectiveness for agiven lubricant type and grade.

Plant maintenance strategy has an influence onlubricant selection as well. Where management isparticularly forward thinking and willing to investin modifications that improve lubricant manage-ment effectiveness (filter connections, continuousfiltration, embedded sample ports, bearing isola-tors, etc.), the company is positioned to maximizethe superior value that can be achieved throughthe use of high-performance lubricants, both min-eral- and synthetic oil-based.

With the variety of factors that can impact lubri-cant film formation and effectiveness, it is to thepractitioner’s benefit to follow a lubricant selectionprocess that is objective, repeatable and based onwidely recognized engineering practices and princi-ples. This article addresses a formal lubricant selec-tion process that could be used to make technical-ly accurate baseline lubricant selection decisionswithout using exacting mathematical expressions.

Parsing the machineIn the January TLT I proposed cataloging all lubri-cated components within a machine for evaluationand specification of a lubricant. The purpose of theprocess is to consolidate the collection of machinecomponents into a concise set of types and catalogwhich components rotate, slide, pivot or haveother dynamically interacting surfaces. There arerelatively few unique types of components, butthere are many permutations of each of the fewunique options. Let’s address these componentsby general type that represent most industrialmachinery, including:

■ Plain bearings■ Element bearings■ Gears■ Hydraulic systems

Plain bearing lubricant selectionProbably the most common manifestation of aplain bearing is a round steel journal riding on aconforming one or two-part brass or babbit bush-ing. There are plenty other machine componenttypes that could be categorized similarly, includingmachine tool gibbs (slideways), brass bushing,pivot pins, ball screws, worm gears, etc. These aresimilar to plain bearings in composition, form(shape), and film characteristics.

Regardless of shape and form, all plain bearingcomponents have a common requirement for afull-fluid hydrodynamic film to sustain componentlife cycles. Two types of decisions must be made:

1. Viscosity grade according to the machine’soperating profile.

2. Lubricant chemistry/additive type (R&O, EP,AW, Compounded, Tackified).

The viscosity grade considerations for slidingsurface interaction are rooted in a common set ofphysical realities. Total available surface area, lin-ear surface speed, surface unit loading, lubricantviscosity grade and lubricant replacement rate allhave an influence on the formation of a hydrody-namic oil film.

STLE member Bob Scott, manager of Lube-Works, Ltd., in Calgary, offers a method of lubricantselection for plain bearing applications. Accordingto Scott, “The minimum information required forthe determination of the proper ISO grade for jour-nal bearings includes:

1. Shaft rpm.

2. Temperature of the oil in the bearing.

3. Approximate unit loading pressure (PSI orNewtons /m2).”

Scott adds that it would be best to also haveadditional information, including:

■ Bearing length, number of bearings androtor weight (to calculate the bearing pres-sure to verify the load).

■ Shaft diameter to calculate the shaft surfacespeed.

■ Driving horsepower.

■ Knowing whether shock loading is present.

■ Knowing if the unit is in a warm or cold envi-ronment.

■ Knowing if cooling water is being applied.

Scott explains there are several different chartsthat recommend ISO viscosity grades for journalbearings, and most are based on oil temperatureand shaft speed. Figure 1 provides an ISO viscosity

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CONTINUED ON PAGE 20

Figure 1. Viscosity selection chart for ‘mediumload’ applications (150–200 psi) Shaft speed Operating oil temperature

20 C–50 C 60 C 75 C 90 C

800 rpm ISO 68 ISO 100 ISO 1501,200 rpm ISO 46 ISO 68 ISO 100-1501,800 rpm ISO 32 ISO 46 ISO 68-100 ISO 1503,600 rpm ISO 32 ISO 32 ISO 46-68 ISO 68-100 10,000 rpm ISO 32 ISO 32 ISO 32 ISO 32-46

T R I B O L O G Y & L U B R I C A T I O N T E C H N O L O G Y A P R I L 2 0 0 8 1 9

grade selection based on oil temperature, rpm andmoderate shaft loads (150-200 psi).

For higher loads (~300 psi), raise the viscosityby 2 ISO grades. For lower loads (~100 psi), lowerthe viscosity by 1 or 2 ISO grades.

According to Scott, temperature estimation iscritical. “While some tables only account forapproximate load and surface shaft speed, the tem-perature of the oil in the bearing itself must be con-sidered. With this approach, the lubrication engi-neer must account for how and where the oil tem-perature was taken. A temperature from the bear-ing housing external surface is likely to be 5 C-10 Cless than the bulk oil temperature and may be asmuch as 25 C below the actual bearing surface tem-perature (at the dynamic interface).”

Scott continues: “Additionally, after settling ona target ISO viscosity grade, it is essential to checkthe product selection against the minimum and opti-mum recommended viscosity for the actual(machine dynamic interface) operating tempera-ture. Most journal bearings require a minimum of10 to 13 cSt at operating temperature. Some tur-bines may only require 7 or 8 cSt as a minimum,but the 13 cSt minimum is a good rule of thumbthat provides some margin for error. If low temper-ature startups are involved, the pour point andBrookfield viscosity data at low temperatures alsoneed to be investigated. Figure 2 provides a set ofgenerally accepted minimum and optimum vis-cosities.”

The Tribology Handbook1 provides a useful chart (seeFigure 3) that can be used to plot speed againstload to arrive at a more narrowly selected lubricantgrade.

Figure 4 shows the formulas that would be usedto calculate both bearing linear speed (in meters/second) and unit load (in kilo-Newtons/m2). Onceknown, these values can be placed on a viscosityselection chart similar to Figure 1 to arrive at a tar-get viscosity grade in centipoise.

Two steps remain. First, for both approaches thepractitioner should plot the selected product on anASTM viscosity-temperature graph to find the actu-al viscosity in centistokes for the expected operat-ing temperature. This step provides a useful profileof the viscosity of the oil across its operating tem-perature range.

Second, for the approach using Figure 1 (units incentipoise), the practitioner should multiply theselected oil viscosity in centistokes (at operatingtemperature) by the typical specific gravity for theproduct to arrive at the expected viscosity in cen-tipoise.

CONTINUED FROM PAGE 19

Figure 4. Calculations for plain bearing surfacespeed and unit load

Suggested Minimum Allowable Viscosities

■ 7 - 8 cSt. Westinghouse, Elliott, other process turbines ■ 13 cSt. General guidance

Generally Accepted Optimum Viscosities

■ 20 cSt. (for 3,600 RPM, No Shock Loading) ■ 35 cSt. (for 1,800 RPM, No Shock Loading) ■ 50 cSt. (for 1,800 RPM, Heavy Load & Shock Loads) ■ 72 cSt (for 500 RPM, No Shock Loading) ■ 95 cSt (for 500 RPM, Heavy Load & Shock Loads)

All Viscosity units are centistokes at operating temperature

Figure 2. Generally accepted minimum andoptimum viscosity grades

(Courtesy of Lubeworks, Inc.)

Figure 3. Viscosity estimation chart based onactual speed and unit loads

20 A P R I L 2 0 0 8 T R I B O L O G Y & L U B R I C A T I O N T E C H N O L O G Y

(Courtesy of The Tribology Handbook, Second Edition, Michael J. Neale)


Plain bearing additivesPlain bearings traditionally have been lubricatedby either compounded type products (fortified withfatty-acid compounds to improve lubricity) for low-speed and heavily loaded sliding surfaces, or rustand oxidation inhibited type products for high-speed sliding surfaces. A strong argument can bemade for sticking with these general product cate-gories, particularly in defense of avoiding EP typelubricants.

Given the thermal energy state required to acti-vate EP (sulfur and phosphorous) agents and thelow likelihood that the steel-on-soft metals interac-tion (bronze or babbit) can deliver the heat loadsufficient to cause these agents to perform theirlocalized function, there is likely more risk from theuse of EP products (accelerated lubricant degrada-tion and increase of acidic oil conditions) thanthere is a benefit from their use. If high load capa-bility is required, and there is concern that the vis-cosity cannot deliver the required results, then thereliability engineer should consider the use of solidfilm (molybdenum disulfide, graphite) fortifiedlubricant that can provide the added protection.

Element bearing lubricant selectionThere are many types of element bearings, includ-ing ball, roller, spherical roller, thrust roller, needleroller and single- and double-row types. Fortunate-ly the selection process for the various types issimilar. The following discussion pertains to thecommon ball, roller, spherical roller and thrustroller type element bearings.

Element bearings operate under a wide array oftemperature, speed, load and environmental con-ditions. The first step is selecting an appropriateviscosity grade followed by selection of additivetype (R&O, AW, EP).

Bearing manufacturers provided fairly usefulcharts that assist us with this process. The fivesteps to this process are:

1. Determine the bearing physical size (width,outer diameter, inner diameter), elementtype (ball, cylindrical roller, spherical roller,thrust ball or roller) and shaft speed. Thesedetails often can be found in the machineconstruction diagrams and drawings. Exam-ple: A roller bearing designated as a 6320bearing (by DIN Standard 623) has a boredimension of 100 mm, an outer diameterdimension of 215 mm and a race width of 47mm.

2. Determine the bearing’s limiting speed fac-tor. This value is referred to as the “pitch line

velocity” or the “nDm.” Figure 5 shows thespeed limit calculation for element bearings.Bearing speed limits differ for each bearing,but general rules exist around bearings bytype and function.

Figure 6 provides practical speed limit val-ues for four common bearing types. Anygiven bearing may have a unique value for itsgiven quality, but these table values ade-quately represent range limits for decisionsfor selecting either oil or grease lubrication.2

Example: The 6320 roller bearing turningat 1,200 rpm has an nDm (speed limiting)value of 120,000. This information is particu-larly useful to assess whether the bearingshould be lubricated by oil or grease. It iscommon for equipment operators to convertoil-lubricated bearings to grease, eventhough the operating speed of the bearing isbeyond that which is considered safe forgrease lubrication.

3. Calculate the Pitch Diameter (PD) of thebearing ([inner diameter + outer diameter] ÷2). Example: The 6320 bearing pitch diameteris ([100 mm + 215 mm] / 2 = 157.5 mm). This

T R I B O L O G Y & L U B R I C A T I O N T E C H N O L O G Y A P R I L 2 0 0 8 2 1T R I B O L O G Y & L U B R I C A T I O N T E C H N O L O G Y A P R I L 2 0 0 8 2 1T R I B O L O G Y & L U B R I C A T I O N T E C H N O L O G Y A P R I L 2 0 0 8 2 1


Figure 5. Bearing speed limit parameters arecalculated using the following equation:

Figure 6. General accepted speed limiting valuesfor industrial bearings2

parameter will be used to identify the mini-mum viscosity for the bearing at its operatingspeed.

4. Plot the bearing PD on a bearing specific vis-cosity limit reference chart. Locate and plotthe shaft rotational speed and triangulate tothe suggested viscosity for the given runningspeed. The result provides the lowest viscos-ity point at which the bearing would beexpected to achieve its projected minimumlife cycle. The optimum viscosity is three tofive times the minimum recommended vis-cosity.

Example: This is a multistep process.Begin by locating the pitch diameter valuefrom Step Four at its appropriate location onthe X (bottom–horizontal) axis of Figure 7.Next, draw a vertical line from the X axistoward the top of the chart until it intersectswith the diagonal line that represents theshaft rotational speed. Next, at that inter-secting point, draw a line to the Y axis fromthat point. Make sure that the final line ishorizontal to the X axis. The Y axis representsthe bearing minimum suggested viscosityvalue for a given size and shaft speed.

In this instance, the PD value (157.4) islabeled Point A. The intersection with thediagonal line for shaft speed is labeled PointB, which is the initiating point for the line thatterminates at Point C, this bearing’s requiredviscosity in centistokes. This element bearingappears to need at least 13 cSt viscosity gradeoil when operating at 1,200 rpm.

The objective is to apply a lubricant that isoptimum, not minimum. Consequently, theexact minimum is less important then itmight seem. Ideally, the operator shouldmaintain three to five times the minimum vis-cosity in the element at its normal operatingtemperature. Therefore, for this bearing, thelow-end optimum is (13 cSt * 3 = 39 centis-tokes), and high-end optimum is (13 cSt * 5 =65 centistokes). If viscosity error must be tol-



There are many crite-

ria to consider when

selecting a lubricant

for a set of machines

or machine components.

Figure 7. Element bearing viscosity selectiongraph

Figure 8. Operating viscosity for mineral oils(with a VI of 100)

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Bea

rin

g C

o.)

(Co

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of F

AG

Bea

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g C

o.)


erated it would be best to err on the side ofincreased oil thickness as long as the overagedoes not induce heat through fluid friction.

5. Plot the viscosity of the target lubricant onan ASTM viscosity-temperature graph anddetermine the actual viscosity at the expect-ed operating temperature. Repeat theprocess until the correct lubricant has beenselected.

Example: Figure 8 is a temperature viscos-ity chart provided by FAG Bearing Co. Thebearing operating temperature is located onthe Y axis, and is labeled Point A. Draw astraight line from the point that is horizontalto the X axis until it intersects with a red line,representing a standard viscosity grade (32,46, 68, 100, etc.). I believe an ISO 68 shouldprovide the desired viscosity at the normaloperating temperature of the application (50C). The intersection point with this line forISO 68 is labeled Point B. Lastly, from thisintersection point (B), draw a straight line tothe X axis to locate the actual operating vis-cosity, 38 cSt, which is labeled Point C.

It is evident from this exercise that thisbearing, operating at a running speed of1,200 rpm at 50 C temperature requires anISO 68 viscosity grade (with a VI of 100).

Element bearing additivesBearing manufacturers suggest that as long as theoperating viscosity is above the minimum it wouldbe best to lubricate without “doped” lubricants.Doped pertains to lubricants with AW or EP agentsthat form physi-chemical barrier films followingsurface contacts. AW and EP agents may over timeinfluence the microcrystalline structure of the bear-ing metals in a negative way, even if the additiveagent is not fully engaged in surface protection.3

If shock loading is expected, then plan to useEP/AW-fortified lubricants to provide a margin oferror. If not, assuming that the operating viscosityis in an ideal range, use an R&O lubricant. If theoperating conditions include momentary high-temperature excursions that bring the operatingviscosity below the allowable minimum, thenEP/AW type lubricants are in order.

Industrial gearing lubricant selectionGears are used to increase or decrease transmittedspeed and load and change the direction of rota-tion from one shaft to another. As shown in Figure9, gears come in a variety of configurations andtooth shapes.

These types of gears can be identified by observ-ing the shaft characteristics (see Figure 10).

The majority of industrial gears are spur, helical,worm and bevel types.

Gears are designed to optimize energy inputwith speed or load output. Large loads requirelarge gear teeth and heavy construction. Converse-ly, small loads can be accomplished with smallteeth and light construction.

Gear drives can be lubricated with either greaseor oil. Gear type, size and function are the leading

Figure 9. Types of common industrial gears

Figure 10. The shaft characteristics of different gear design types

Gear design type Shaft characteristic

Spur, helical, herringbone Non-intersecting and non-crossing shafts

Bevel (straight and spiral tooth) Intersecting shafts (any angle)

Worm and hypoid gears Non-intersecting shafts



factors in this decision. While the great majority ofindustrial gear applications operate in an oil bath,some gears operate without one. These gears,called open gears, are lubricated with specializedgreases that are sprayed onto the gear set on anintermittent basis. Open gear lubrication will beaddressed later in this series.

Once again, viscosity is the central feature ofthe lubricant that must be accurately identifiedand selected in order for components to achievetheir expected life cycles. As is the case with thepreviously evaluated components, the oil mustpossess viscometric properties that cover cold-flow requirements at low starting temperaturesand provide sufficient oil thickness at normal oper-ating temperatures to resist the squeezing effect ofthe meshing gears.

The low temperature flow requirement is depend-ent on a variety of application-specific factors,including weather, degree of shelter (indoor, out-doors), presence of thermal controls and likelihoodthat the unit will be started at extreme (cold) condi-tions. A safe pour-point limit for gear oils is 10 F/5.2C below the coldest likely starting temperature.4

Most modern gear drives provide lubricantselection guidelines that conform to the currentAGMA standard for industrial gear lubricant, Stan-dard 9005-EO2. In most instances gear drive hous-

ings include selection advice on the gear caseitself. Figure 11 shows how the information is dis-played on a common gear drive.

Over time reliability engineers may have torevisit the lubricant selection following a changefrom the original operating state. In theseinstances, there needs to be a relatively uncompli-cated and accurate way to estimate the targetlubricant type and grade. AGMA standard 9005-EO2 for lubricant selection is a helpful documentto support effective lubricant selection and will bereferenced for the balance of this selectionprocess.

AGMA promotes viscosity selection based onthe pitch line velocities of each gear in a set. Pitchline velocity (PLV) is calculated based on the fol-lowing formula:

PLV, meters/second = (Gear pitch diameter in millimeters X Gear rpm) ÷ 19098

The gear pitch circle dimensions and shaft speedsmust be known in order to make these calcula-tions. These values can be provided by the drivemanufacturer or derived from other tooth form andratio data.

Once known, the PLV can be correlated to tablesfrom 9005-EO2, Appendix B, to reference the re-quired viscosity by ISO grade for a given sump tem-perature and for the referenced gear tooth forms.

Consider the following example for a drive on awire-drawing machine, as seen in Figure 12:

Step 1..Determine the dimensions of the bevel gear set

■ pinion pitch diameter of .032 meters

■ bull (driven) gear pitch diameter of .093meters.

Step 2..Measure or calculate the gear shaft speed (pinion= 3750, bull= 350 rpm).

Step 3..Measure the internal oil sump temperature (50 C).

Step 4..Calculate the PLV of the each gear.

Step 5a..Using the appropriate table from the 9005.EO2standard, locate the column for PLV and the row foroil sump temperature. The appropriate oil grade isfound where the column and row intersects.

Step 5b..If the lubricant under consideration is a synthetic,


Figure 11. Gear drive faceplate label showingviscosity parameters

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T R I B O L O G Y & L U B R I C A T I O N T E C H N O L O G Y A P R I L 2 0 0 8 2 5T R I B O L O G Y & L U B R I C A T I O N T E C H N O L O G Y A P R I L 2 0 0 8 2 5

use the table that corresponds to the high viscosityindex (VI) property. Figure 13 corresponds to 9005-EO2, Table B-3 for synthetics oils with VI = 160.

Where dimensions and speeds produce gradesthat are at least one or more standard ISO viscos-ity grades apart, the owner must rationalize theselection based on lubricant application (forcefeed vs. splash feed), risk imposed by likely load,risk imposed by temperature spikes and risk ofcold start-up to arrive at a final selection. In thisexercise, where there is a high pinion speed, lowheat loads and little risk of load spikes, the pre-ferred mineral oil selection would be an ISO 150,and the preferred synthetic oil selection would bean ISO 100. A synthetic VG 150 also would beacceptable. Bearing viscosity selection should beevaluated for all gears with high operating speedsor temperatures.

The 9005 standard provides additional chartsthat are useful for selecting lubricant viscositygrades for worm gears and large open gears withcaveats and explanations for each type of appli-cation.

Industrial gear drive additivesThere are three classifications of oils that apply togeared drives: R&O, EP/Antiscuff and Compoundedtypes. R&O oils are intended where pitch linevelocities are high and shock loading is low.EP/Antiscuff oils are intended where gear materialsare steel (no brass or bronze) and where PLV valuesare low and the risk of shock loading is high. Mostindustrial applications require EP oils. (EP oilsmust be avoided for gears with internal backstops.)

Lastly, compounded oils are used on gears withhigh sliding contact surface pressures and softmetal gears–such as with worm gear type applica-tions. Although some lubricant suppliers advocatethe use of antiscuff oils for worm gear applications,the 9005 standard does not recommend this prac-tice, nor do I.

Hydraulic system lubricant selectionLubricant selection for hydraulic systems is a topicfor an entire book. Let’s look at a few centralthemes at the heart of all hydraulic oil selectionsthat can be addressed:

1. Viscosity grade at operating temperature.

2. Base fluid stock type (mineral oil, synthetic,water-based, biodegradable).

3. Additive type.

The hydraulic system, in a simplistic view, ac-complishes work by squeezing a small quantity of

Figure 12. Gear oil viscosity calculation and selection

Figure 13. AGMA Standard 9005-EO2 (Appendix B, Table 3.)


(Courtesy of Advanced Machine Reliability Resources)

Bearing manufacturerssuggest that as long asthe operating viscosityis above the minimumit would be best tolubricate without‘doped’ lubricants.


oil into a pipe very rapidly and then directing thepressurized fluid through a set of valves to be re-leased where specific mechanical force is needed.

There are several common components in ahydraulic system, including:

■ Reservoir. Provides a supply of oil to thepump.

■ Pump. Adds flow/pressure to the oil.

■ Piping system. Contains the pressurized oil.

■ Valve/control system. Channels energy inthe form of pressurized oil.

■ Work components of some type. Performsthe required work.

■ Filters. Protects critical components fromcatastrophic damage, cleans the oil.

These various components represent a mixtureof frictional properties, including rolling friction(hydraulic motor and input motor bearings), slid-ing friction (pump vanes and pistons, valve spools,hydraulic rods and cylinders) and a combination ofthe two (gear pump tooth contacts, hydraulicmotors). All these working components are highlysensitive to the physical, chemical and cleanlinessproperties of the fluid.

Flow control systems tend to dictate fluid clean-liness requirements. Cleanliness will be addressedin a future article. Pump type (mechanical inter-face) and efficiency parameters typically dictatephysical viscosity parameters. Fluid type is dictat-ed by plant environmental factors: basestock typeand additive performance selection. It is notunusual for a given product selection to have con-flicting priorities.

Despite the many differences between systems,they are all intended to operate within an efficien-cy range. The allowable viscosity for a given system

is established by pump type and the specificpump’s efficiency parameters. If the fluid is toothick the pump may cavitate, overheat due to fluidfriction, increase mechanical stress on bearing ele-ments and consume excessive energy to fulfill itsfunction. If the fluid is too thin the pump contactsurfaces will wear and lose efficiency, leak acrossinternal and external seals and clearances andoverheat. For that reason the viscosity selectionmust meet the recommended physical propertiesat the system operating temperature.

System manufacturers offer a window for appro-priate viscosity performance. The equipmentowner must fit the product into the viscometricwindow. The following seven steps represent a sys-tematic approach that may be used to arrive at thebest overall viscosity selection.

1. Plot the operating temperature range for thesystem on a standard temperature-viscosityselection chart. The green lines representhigh and low temperature operating ranges.If the system hasn’t been commissioned, usethe OEM’s suggested temperature profile.

2. Hydraulic pump designers provide viscositylimits that include absolute maximum andminimum limits (shown as blue boxes) andoptimum maximum and minimum limits.Plot these provided values on the chart.

Example: a pump manufacturer providesan absolute maximum limit of 1,000 cSt and aminimum of 10 cSt. These lines establishfinite limits for operating temperatures (coldstart, high operating temperature limits) forthe final fluid selection. Additionally, theOEM offers the ideal minimum (16 cSt) andideal maximum viscosity (36 cSt) values. Thegray lines on the chart represent viscositiesthat support optimum volumetric efficiency.

3. Find the intersecting lines between the oper-ating temperature limits (green lines) andoptimum viscosity limits (grey lines). Thiscross-section represents the target spacethrough which we would like our best prod-uct choice to pass.

4. Review and plot the viscosity parameters foreach of the available fluid choices on the vis-cosity temperature chart. Example: There arethree available options. Option A is an ISO 68mineral oil with a VI of 100. Its default vis-cosity measurements are 64.6 cSt at 40 C,and 8.4 cSt at 100 C. These points are plottedand a line is drawn from the far left of thechart through these points to the far right ofthe chart. The extreme ends of this line rep-


Gear drives can be

lubricated with either

grease or oil. Gear

type, size and function

are the leading factors

in this decision.


resent temperature extremes that will beconsidered later. This process is repeated forthe other available products (options B and Care provided).

5. Evaluate the cold flow temperature limit foreach of the plotted lines and eliminate anythat do not meet the limits for cold weatheroperation.

Example: Identify the product that has thelongest line segment passing through theoptimum area (noted in Step 3). In this caseit appears to be option A. Observe the leftside of the chart where the ‘A’ line passesthrough the OEM’s absolute low viscositylimit. Line A crosses this (1,000 cSt) line atabout 0 C. This temperature is the lowestpossible temperature that the system can beallowed to run with this fluid. If it is likelythat the system will be expected to start attemperatures lower than 0 C, then anotheroption must be considered. Repeat the stepsfor each additional product option.

6. Make the final selection based on the best fitfor optimum viscosity within the operatingtemperature range and the least risk of coldstart failure.

Example: If the fluid is being selected for ahydraulic system that is bolted to the floor, itis unlikely that the cold start temperatureswill ever fall below 0 C. To control this risk,the OEM might install a thermostatically-controlled heater to assure safe start condi-tions. If the system happens to be installed inoutdoor construction machinery, then it is

likely that eventually this machine will haveto function below the cold start limit sug-gested for fluid A, making it undesirable forthe application.

7. Repeat this evaluation process for each prod-uct type until the best option is identified.

Industrial hydraulic applicationsAs shown in Figure 14, several categories ofhydraulic fluids may be considered. Water-basedfluids are preferred for applications where fire is athreat, and oil-based fluids are preferred for appli-cations where either there is no risk of fire or wherethere is a risk of fire and a high degree of compo-


Compounded oils areused on gears with highsliding contact surfacepressures and softmetal gears–such aswith worm gear typeapplications.

Figure 14. Cross-section of types of industrial hydraulic fluids

(Courtesy of Advanced Machine Reliability Resources)


nent dependability is necessary for other reasons(i.e., safety factors such as the fire- resistance syn-thetic fluids used in aircraft hydraulics).

In each instance, regardless of the basestocktype, wear-resistance is critical. The primarymechanical component subject to wear is thepump. Pumps operate with significant but lowincremental-force mechanical interaction (bound-ary and mixed film state). AW fluids are appropri-ate for this type of service. These additives gener-ate a physi-chemical organo-metallic barrier layeron contacting surfaces that can be repeatedlywiped off and replaced many times during normaloperation. These metallic soap films provide ongo-ing wear resistance without excessive chemicalreactivity.

Lubricant manufacturers have many differentlubricant formulas that accomplish these results,with varying degrees of longevity, additive stabilityand overall system friendliness. Standardized wear-resistance performance testing should be includedin the final lubricant selection process.

Mike Johnson, CLS, CMRP, MLT, is the principal consult-ant for Advanced Machine Reliability Resources, headquar-tered in Franklin, Tenn. You can reach him at [email protected]


References1. Heale, M.J. (1985), The Tribology Hand-book, Second Edition, Elsevier.

2. Pirro, D.M. and Wessol, A.A. (2001),Lubrication Fundamentals, Second Edition,Marcel Dekker.

3. SKF Bearing Installation and Mainte-nance Guide, 140-710.

4. AGMA 9005-E02, American NationalStandard for Industrial Gear Lubrication.

Documents

Lubricant selection: Function and composition - STLE · grade selection based on oil temperature, rpm and moderate shaft loads (150-200 psi). For higher loads (~300 psi), raise the