Rolling bearings in industrial gearboxes

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Publication 4560 E

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1 Industrial gearboxes – overview

2 Bearing types for industrial gearboxes

3 Design of bearing arrangements

4 Dimensioning the bearing arrangement

5 Lubrication and maintenance

6 Recommended fits

7 Mounting and dismounting bearings

8 Application examples

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Rolling bearings inindustrial gearboxes

Handbook for the gearbox designer

Foreword

This Handbook is intended to provide the gearbox designer

with the knowledge required to select bearings for gearboxes

and to correctly design gearbox bearing arrangements. Recom-

mendations are given based on experience gained by SKF

during decades of cooperation with gearbox manufacturers the

world over.

General information regarding the selection, calculation,

mounting and maintenance of ball and roller bearings is given

in the SKF General Catalogue. The questions arising from the

use of rolling bearings in industrial gearboxes are dealt with

here. Data from the General Catalogue are only repeated here

when it has been thought necessary for the sake of clarity.

The application examples described comprise proven

gearbox designs from major manufacturers which are worthy

of note.

Grateful thanks are extended to the companies concerned

for the provision of the detailed information about their prod-

ucts and the permission to publish.

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1 Industrial gearboxes – overview............................... 9

Types of gearbox............................................................ 9

Geared transmissions.................................................... 10

Demands made on gearboxes ...................................... 14

Selecting the gears ........................................................ 14

Designing the casing ..................................................... 15

2 Bearing types for industrial gearboxes.................. 17

Deep groove ball bearings ............................................ 18

Angular contact ball bearings....................................... 20

Cylindrical roller bearings ............................................. 22

CARB™ roller bearings ................................................. 24

Spherical roller bearings ............................................... 26

Taper roller bearings ...................................................... 28

Spherical roller thrust bearings .................................... 30

3 Design of bearing arrangements............................. 33

Shafts and gear wheels in spur gearboxes ................. 33

Shafts in bevel gearboxes ............................................. 44

Shafts in worm gearboxes............................................. 50

Shafts and gear wheels for planetary gearboxes........ 56

Made by SKF® stands for excellence. It symbolises our consistent endeavour to achieve total quality in everything we do. For those who use our products,“Made by SKF” implies three main benefits.

Reliability – thanks to modern, efficient products, basedon our worldwide application know-how, optimised materials, forward-looking designs and the most advanced production techniques.

Cost effectiveness – resulting from the favourable ratiobetween our product quality plus service facilities, andthe purchase price of the product.

Market lead – which you can achieve by taking advantage of our products and services. Increased operating time and reduced down-time, as well as improved output and product quality are the key to a successful partnership.

Contents

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4 Calculation of bearing arrangement ....................... 65

Bearing loads ................................................................. 65

Determination of external forces .................................. 66

Calculation of bearing loads ......................................... 74

Dimensioning the bearing arrangement ...................... 76

5 Lubrication and maintenance.................................. 91

Grease lubrication.......................................................... 92

Oil lubrication ................................................................. 95

Maintenance ................................................................... 98

6 Recommended fits..................................................103

7 Mounting and dismounting bearings....................109

Adjustment of angular contact bearings....................109

8 Application examples ............................................. 115


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SKF is an international industrial Groupoperating in some 130 countries and isworld leader in bearings.

The company was founded in 1907following the invention of the self-align-ing ball bearing by Sven Wingquist and,after only a few years, SKF began toexpand all over the world.

Today, SKF has some 43 000 em-ployees and more than 80 manufactur-ing facilities spread throughout theworld. An international sales network includes a large number of sales com-panies and some 20 000 distributorsand retailers. Worldwide availability ofSKF products is supported by a com-prehensive technical advisory service.

The key to success has been a con-sistent emphasis on maintaining the

highest quality of its products and services. Continuous investment inresearch and development has alsoplayed a vital role, resulting in manyexamples of epoch-making innovations.

The business of the Group consistsof bearings, seals, special steel and acomprehensive range of other high-tech industrial components. The ex-perience gained in these various fieldsprovides SKF with the essential know-ledge and expertise required in orderto provide the customers with the mostadvanced engineering products andefficient service.

The SKF Group – a worldwide corporation

SKF manufacturesball bearings, roller bearings and plainbearings. The smal-lest are just a fewmillimetres (a frac-tion of an inch) indiameter, the largestseveral metres. Inorder to protect thebearings effectivelyagainst the ingressof contaminationand the escape oflubricant, SKF alsomanufactures oiland bearing seals.SKF's subsidiariesCR and RFT S.p.A.are among the world's largest pro-ducers of seals.

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The SKF house colours are blue and red,but the thinking is green. The latest exampleis the new factory in Malaysia, where thebearing component cleaning process con-forms to the strictest ecological standards.Instead of trichloroethylene, a water-basedcleaning fluid is used in a closed system.The cleaning fluid is recycled in the factory'sown treatment plant.

SKF has developed the Channel concept infactories all over the world. This drasticallyreduces the lead time from raw material toend product as well as work in progress and finished goods in stock. The conceptenables faster and smoother informationflow, eliminates bottlenecks and bypassesunnecessary steps in production. TheChannel team members have the know-ledge and commitment needed to share theresponsibility for fulfilling objectives in areassuch as quality, delivery time, productionflow etc.

The SKF Engineering & Research Centre is situated just outside Utrecht in TheNetherlands. In an area of 17 000 squaremetres (185 000 sq.ft) some 150 scientists,engineers and support staff are engaged inthe further improvement of bearing perform-ance. They are developing technologiesaimed at achieving better materials, betterdesigns, better lubricants and better seals – together leading to an even better unders-tanding of the operation of a bearing in itsapplication. This is also where the SKF New Life Theory was evolved, enabling thedesign of bearings which are even morecompact and offer even longer operationallife.

1 Industrial gearboxes – overview

Types of gearbox . . . . . . . . . . 9

Geared transmission . . . . . . 10

Demands on gearboxes . . . 14

Selecting the gears . . . . . . . 14

Designing the casing . . . . . . 15

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Industrial gearboxes –overview

Gearboxes are devices for the transmission ortranslation of movement. In industry gearboxesare used to transform the speeds and torquesproduced by the prime mover in order thatthey are appropriate to the machine which is tobe driven. The speeds and torques required bythe machine are dictated by its use. Primemovers can generally only meet these require-ments when combined with gears.

Types of gearboxGearboxes are characterised by havingat least three members: the power in-put, power take-off and the casing. Thecasing transmits the support momentto the base.

In contrast, a coupling has only twomembers: the power input and power

1 Industrial gearboxes – overviewTypes of gearbox

P1

M1

PV

P2

M2

n2

n1

P2P1M2

Pv

M1 n1 n2

take-off. The coupling housing has nopart in the flow of force.

The symbols used for powertransmission by gearboxes and coup-lings are shown in figs and .21

Fig 1 Fig 2GearTorqueM1

< M2>

Rotational speedn1

≤ n2>

PowerP1 = P2 + Pv

CouplingTorqueM1 = M2

Rotational speedn1 ≥ n2

PowerP1 = P2 + Pv

(with slip)

Geared transmissionsGeared transmissions are the mostcommonly used. They transmit powerwithout slip, have high operational re-liability and long life, require little main-tenance and are characterised by theability to accept overloading, small sizeand high efficiency.

Spur gearsThe spur gear is the most well-knownand commonly used design of gearedtransmission. The dimensioning andmanufacture of the gear wheels arethe easiest to control. Their kinematicbehaviour also forms the basis of plan-etary gears. Spur gears are in rollingcontact and, irrespective of tooth type,have parallel axes.

1 Industrial gearboxes – overviewGeared transmissions

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The main types ofpower transmis-sion equipmentare shown in thefollowing.In addition, thereare many com-binations, for ex-ample bevel/spurgears, spur gearswith belt drive input,or variable tractiondrives combinedwith a planetarygear.

Types of gearbox

Fixed ratio transmissions, shift transmission

Geared transmissions• Spur gears• Planetary gears• Bevel gears• Worm gears• Hypoid gears• Helical gears

Eccentric drives• Cyclo drives• Harmonic drives

Traction drives• Belt drives• Chain drives

Infinitely variable transmissions

Mechanical transmissions• Belt drives• Roller drives• Ratchet gears

Hydraulic transmissions• Hydrostatic transmissions• Hydrodyanmic transmissions

● Gear wheels with straight cut teeth(➔ fig a) are simple in designand can be accurately produced.The axial forces generated by in-accuracies and deformations (twisting) are negligible.

● Gear wheels with helical teeth (➔ fig b) run more smoothly andcan carry heavier loads than thosewith straight cut teeth. A more elab-orate bearing arrangement is re-quired because of the axial forces.

● The double helix or herringbone (➔ fig c) allows for large toothwidths and can carry particularlyheavy loads. The axial forces canceleach other out. Deviations in thehelix angle cause axial vibrations.

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● Internal gearing (➔ fig d) hasgreater load carrying capacity thanexternal because of the favourableosculation, but is more difficult toproduce. The bearing arrangementis more complicated. The most fre-quent use is in planetary gears.

Bevel gearsThe common characteristic of this type of rolling contact gearing is thatthe axes of the wheels intersect eachother. There are three basic designscategorised by the form of the flank.

● With straight cut teeth (➔ fig a),the mesh begins and ends acrossthe total tooth width. The noise pro-duced considerably limits the use-fulness of straight cut bevel gears.

● Bevel gears with helical teeth (➔ fig b) have straight flanks.The teeth are usually ground andthe mesh is gradual. The total over-lap is bigger and the noise behav-iour better than with straight cutteeth.

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3 ● Bevel gears having spirally cut teeth(➔ fig c) with curved flanks haveclear advantages in respect of loadcarrying capacity. Particularly thosewith ground teeth are quieter thanthe types described above. Forbevel gears which have to transmithigh power, the spiral bevel gearsare the most frequently used.

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Spur gear unita) straight cut teethb) helical teethc) double helixd) internal gearing

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Fig 3

Fig 4

a b c d

Bevel gear unita) straight cut teethb) helical teethc) spirally cut teeth

a b c


Worm gearsThe worm and wheel axes cross eachother at a considerable distance andusually at an angle < 90° (➔ fig ).Worm gears are suitable for large single stage speed reduction. Theiroperation is quiet and vibration damp-ing. The efficiency is lower than that ofcompeting bevel/spur and planetarygears, because of the higher propor-tion of sliding motion. To reduce thefriction, the use of synthetic lubricantsis favoured.

The most commonly used design isthe cylindrical worm paired with a glob-oid wheel (➔ fig a). The cylindricalworm can be hardened and groundwhich improves load carrying capacity;it is also freely adjustable in the axialdirection so that bearing arrangementand mounting can be simplified. Twoother designs – globoid worm with spurwheel (➔ fig b) and globoid wormwith globoid wheel (➔ fig c) – arealso used.

Depending on the flank form, theworm types are classified as follows:

● ZA worm: trapezoidal worm threadin the axial cross section;

● ZN worm: trapezoidal worm threadin the normal cross section;

● ZK worm; trapezoidal tool (in normalcross section);

● ZI worm; evolvent thread in end facecross section;

● ZC worm: concave worm flanks

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Hypoid gearsThe pinion axis is displaced so that theaxes of this type of bevel gear do notintersect but are crossed (➔ fig ).

The wheels of hypoid gears are usu-ally spirally cut. The advantages of thistype of gear derive from the larger pin-ion and thus the smaller circumferentialforce for the same torque, as well asfrom the axis displacement which oftenallows the pinion to be supported atboth sides so that the bearing arrange-ment is stiffer. The noise behaviour isalso improved by the sliding motion inthe longitudinal direction of the teeth.However, the additional sliding motionincreases the friction, wear and risk of smearing and requires the use ofhypoid oils with high additive content.

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Fig 5Hypoid gear unit

Fig 6

a b c

Worm gear unita) cylindrical worm

with globoidwheel

b) globoid wormwith spur wheel

c) globoid wormwith globoidwheel

H

SP

Z

The ZI and ZC designs are the mostpopular. The ZI worm can be very ac-curately ground whilst the favourableosculation conditions of the ZC worm(concave worm, convex wheel) bringload carrying advantages.

Planetary gearsFrom the point of view of the toothflanks, planetary gears are mostly spurgears. In contrast to the spur gear unitsso far described, the shafts of whichare supported in stationary casings,the planetary gear unit has gear wheelswhich circulate. They are also referredto as epicyclic gears.

In the simplest design (➔ fig ),which is that most commonly used inindustry, the sun wheel drives the plan-etary wheels (when acting as a speedreducer). These are supported in thehollow wheel and drive the planetarycarrier from which the power is takenoff.

Planetary gears have the followingimportant advantages compared withconventional spur gear units:

● the volume, weight and centrifugalmass are smaller;

● the rolling and sliding velocities inthe mesh are lower, so that noise isreduced;

● some of the power is transmitted ascoupling power, so that efficiency ishigher.

These advantages have led to a continuous increase in the economic importance of planetary gear units inspite of their disadvantages whichinclude more difficult inspection, main-tenance and repairs.

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Simple planetarygear unit (prin-ciple)Z sun wheelP planetary wheelH hollow wheelS planetary carrier

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Fig 7

Demands made on gearboxesThe most important demands whichmust be fulfilled are:

● there must be a sufficient safetymargin in respect of fatigue and/orrequisite life for all components sothat the torques and speeds can bereliably transmitted;

● there must be sufficient cooling evenunder maximum power transmissionconditions;

● noise emission should not exceedthe permitted limits.

In addition to these demands, specialrequirements in respect of operationand design are dictated by the variousapplications. Some examples:

● radial and/or axial forces on the in-put and output shafts, e.g. for ex-truders;

● external forces on the casing, e.g. inmining;

● heavy impacts, torque peaks, e.g.when driven by single cylinder com-bustion engines or when drivingbucket excavators;

● vibrations, e.g. in wire drawing;● extreme environmental influences

in respect of temperature, dirt, dust,water, e.g. in arctic or tropical opencast mining and in continuous cast-ing plant;

● seals subjected to pressure, e.g. insubmerged gearboxes of dredgersor in mixing equipment in the chemi-cal industry;

● reversing operation, e.g. for rollingmills;

● return stop, e.g. for conveyors;● operation with little or no clearance

and torsional stiffness, e.g. for posi-tioning antennae and for robots;

● precision, e.g. for printing presses;● lubrication with non-flammable lub-

ricants, e.g. in mining;● minimum maintenance, e.g. in wind

power plant;● arrangement, e.g. slip-on gears for

converters;● accessibility of measuring points to

monitor lubrication, temperature,vibrations or torque, e.g. for largeplastic extruders.

Selecting the gearsTo avoid either under or over-dimen-sioning a gear unit the load and theload carrying capacity of the gear mustbe able to be determined as accuratelyand reliably as possible. The size iscorrectly chosen when a comparisonof the load spectrum and the load carrying capacity gives the desiredservice life. The determination of theload spectrum is a time-consumingand costly exercise calling for con-siderable measurements. Therefore,dimensioning is usually based on therated torque of the driven machine, i.e. the operating torque for the mostarduous work conditions. For a rollingmill, for example, this is the maximumcontinuous rolling torque (not the initialentry). The actual loads are higherbecause of additional external forces,produced by accelerations and vibra-tions, for example. When calculatingthe load carrying capacity of the gearwheels, these additional loads are considered by an application factor KA according to DIN 3990.

One standard work on the subjectlists the following criteria for evaluatingthe load carrying capacity of gear whe-els:

● resistance to pitting (tooth flank fatigue),

● root strength (tooth fracture fromfatigue),

● resistance to scuffing (hot toothflank welding),

● wear strength (slow wear of toothflanks),

● “grey spot” resistance (fatigue frommicro pores on the tooth flanks, and

● lubricant film formation.

The load carrying capacity which isused as the basis for dimensioninggear wheels is determined in rig testsunder standard conditions (partly stand-ardised: FZG test to DIN 51 354).

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1 Industrial gearboxes – overviewDemands made on gearboxes/Selecting the gears

Designing the casingThe following functions have a de-cisive influence on the design of thecasing:

● forces and supporting momentsmust be taken up and transmiitted atthe same time as the position of thegear wheels and the form of thebearing seatings must be accuratelymaintained;

● there must be adequate heat removal;

● noise radiation must be at a min-imum;

● gear wheels and bearings must beprotected against contamination byforeign matter;

● lubricant loss must be prevented.

The increase in load carrying capacityof gear wheels and rolling bearingsresulting from design improvements,improved materials and enhancedquality has enabled gearboxes to bedownsized or uprated. The higher specific loads, frictional losses and in-creased noise resulting from this trendmean that the casings must be morestable so as to keep deformations to aminimum, but also that they shouldhave a sufficiently large surface to pre-vent inadmissible heating and prema-ture lubricant ageing, and should beproperly designed with respect to mini-mising noise so as not to exceed thenoise emission limits.

1 Industrial gearboxes – overviewDesigning tha casing

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2 Bearing types for industrialgearboxes

Deep groove ball bearings . 18

Angular contact ball bearings . . . . . . . . . . . . . . . . 20

Cylindrical roller bearings . . 22

CARB™ roller bearings . . . . 24

Spherical roller bearings . . . 26

Taper roller bearings . . . . . . 28

Spherical roller thrust bearings . . . . . . . . . . . . . . . . 30

For the support of the shafts and gear wheels ofindustrial gearboxes, rolling bearings are usedalmost exclusively. The exceptions are in somespecialised areas, such as turbo drives, wherehydrodynamic plain bearings are used.

There are many good reasons for thisdominance of rolling bearings:

● good location with minimum radialand axial play enables optimummeshing to be achieved;

● high specific load carrying capacitywith low friction;

● wide range of internationally stand-ardised products produced in highvolumes at reasonable prices andhaving good availability;

● can be calculated using reliable loadcarrying capacity values;

● little design work for the user;● simple arrangement;● axially compact so that short and

stiff shafts can be used;● normal tolerances and surface fin-

ishes for shaft and housing seatings;● less sensitive to misalignment than

plain bearings;● ability of radial bearings to accept

axial loads;● not influenced by direction of load or

rotation;● low starting torque;● no starting problems in intermittent

operation;● relatively easy to lubricate;● favourable behaviour under emer-

gency conditions;● economic maintenance.

Almost all bearing types are used inindustrial gearboxes and almost all theavailable sizes. In the majority of appli-cations, standard “catalogue” bearingscan be used; any variants with respectto clearance or cage design are alsogenerally common, so that the com-prehensive range of SKF “catalogue”bearings for general engineering appli-cations covers the needs of gearboxesvery well and enables the designer tomake an optimum selection. The mostimportant bearing types for gearboxesare described in more detail in the following.

Bearing types forindustrial gearboxes

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2 Bearing types for industrial gearboxesDeep groove ball bearings

2 Bearing types for industrial gearboxesDeep groove ball bearings

Deep groove ball bearingsDeep groove ball bearings are themost popular of all bearing types andthis also applies for gearboxes. Themost important characteristics whichmake them so popular are

● they are able to carry radial loads as well as axial loads acting in bothdirections;

● they are suitable for high and veryhigh speed operation as their frictionis low;

● they have practically no tendency tosmear, i.e. cold welding when theballs are accelerated;

● they run quietly, particularly if theyare lightly preloaded by axial force;

● they are robust in operation andrequire little maintenance;

● they are favourably priced.

The dominant role for deep groove ballbearings is where shafts have to belocated axially and loads are relativelylight. This is the case in

● spur gear units (drive shaft and hol-low take-off shaft),

● multi-ratio gear units (switching spurgear wheels),

● geared motors● worm gear units (worm wheels),● planetary gears (drive shaft, planet-

ary carrier) and● coupling shafts.

These improvements also bring ad-vantages when the bearings are usedin gearboxes. In particular the reducedsensitivity to misalignment means thatthere is no reduction in bearing lifeunder the slight misalignments of up toapproximately 3 minutes of arc whichare normally encountered. The im-proved surfaces reduce friction lead-ing to lower running temperatures sothat lubrication conditions are im-proved and bearing life extended.

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Benefits offeredby SKF In recent years SKF has made a

number of improvements to deepgroove ball bearings which haveresulted in further performanceenhancements. The more import-ant include

● optimised raceway geometryand finish, reducing friction, run-ning noise and sensitivity tomisalignment;

● improved cages which are morestable, thus increasing reliabilityat high speeds;

● improved seals, thus enhancingthe sealing efficiency of sealedbearings.

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2 Bearing types for industrial gearboxesAngular contact ball bearings

Angular contact ballbearingsThe raceways of these bearings arearranged at an angle to the bearingaxis (contact angle), so that they areable to carry heavier axial loads thandeep groove ball bearings. Slidingmovements of the balls are superim-posed on their rolling motion, so thatthe single row bearings require ac-curate adjustment or a minimum axial load to function properly.

Angular contact ball bearings areavailable in the following designs:

● single row, single direction angularcontact ball bearings,

● double row, double direction andpaired single row angular contactball bearings and

● four-point contact ball bearings, i.e.single row, double direction ball bearings.

Single direction implies that axial loads acting in one direction only canbe accommodated, whereas doubledirection bearings (and paired singledirection bearings, depending on thearrangement) can take axial loadsacting in both directions.

The single and double row angularcontact ball bearings are preferred as locating bearings for worm shafts.Four-point contact ball bearings areused primarily as thrust bearings inhigh speed spur gear units, where theouter ring is radially free.

The improvements made by SKFto single and double row angularcontact ball bearings, e.g. reinfor-cement of the ball set (single row –BE design, double row – A and Edesigns) to give higher load carry-ing capacity means that worm gearunits can transmit more powerand, at the same time, the reduc-tion in friction means that bearingtemperature can be lowered. Thereduced tolerances for axial clear-ance and for dimensional and run-ning accuracy which are standardfor SKF single row angular contactball bearings for paired mountingof the CB design, because of theimproved location and reducedrunning noise, are advantageousin low-noise worm gear units suchas those required for lifts andescalators.

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2 Bearing types for industrial gearboxesAngular contact ball bearings

Benefits offeredby SKF

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2 Bearing types for industrial gearboxesCylindrical roller bearings

2 Bearing types for industrial gearboxesCylindrical roller bearings

Cylindrical roller bearingsThe special properties of cylindricalroller bearings make them a popularchoice for gearboxes and include:

● high radial load carrying capacity;● low friction – the lowest of any roller

bearing under purely radial load;● suitable for a wide range of operating

speeds, including very high speeds,as the cage has the correct combina-tion of roller guidance, strength andsliding friction properties;

● ability to accommodate moderateaxial loads, when they are simulta-neously under radial load, via theslid-ing surfaces of the rollerend/flange contact, although the inc-reasedfriction means that lubrication andcooling must be adapted to the conditions;

● the ease with which lateral displace-ment can take place within the bear-ing makes them ideal as non-locat-ing bearings;

● proven good performance underexternal radial accelerations;

● most designs are separable so thatmounting and dismounting are simple.

These characteristics make cylindricalroller bearings ideal for the followingapplications:

● as the non-locating bearings of allhigh-performance units; the NUdesign with its flangeless inner ringis perhaps the most used, but alsothe NJ, NJG and NCF find applica-tion; the rings of these bearingsneed only be axially located at oneside, and by mounting the rings withrelative axial displacement the bear-ings can accommodate lateraldisplacement in both directions.

● in spur gear units, even where com-bined radial and axial loads are pro-duced by the helical teeth; the mostpopular positions are those on theintermediate shaft, as the axial forcesfrom the driven and driving wheelsgenerally act in opposite directionsso that the resultant axial load islight.

Practically all improvements madeto cylindrical roller bearings by SKFcould be considered as tailored togearbox needs, so that they makean appreciable contribution to in-creased performance. The maincharacteristics are

● the reinforced roller complementsand “opened” flanges of the ECdesign give increased radial andaxial load carrying capacity;

● the logarithmic roller profile en-sures an optimum stress distribu-tion over the whole roller lengthso that edge stresses are avoid-ed even under heavy loads andthe permissible misalignments;

● the refined raceway micro-geo-metry reduces friction and im-proves lubricant film formation;

● newly developed cages ensureproper bearing function over theincreased performance range; the standard polyamide cages(designation suffix P) of smallbearings have low friction, areelastic and have good slidingproperties;the steel window-type cages(designation suffix J) which arestandard for medium-sized bear-ings and can also be fitted to thesmaller sizes (to special order)withstand high temperaturesand also medium to strong vib-rations;the machined brass cages (forgearbox bearings preferably out-er ring centred and in two parts,designation suffix MA, or in onepiece, suffix MP or ML) are stan-dard for large bearings and canbe fitted to other sizes to specialorder; they can tolerate highspeeds and are resistant to vib-rations and accelerations.

The range of cylindrical roller bearingsis large compared with other bearingtypes. The various flange configurations(NU, NJ, NUP, N and NCF designs)make the bearings suitable for a multi-tude of applications and the differentcage designs extend the usefulness ofthese bearings.

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2 Bearing types for industrial gearboxesCARB™ roller bearings

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CARB™ roller bearingsCARB is a completely new type ofbearing: a Compact Aligning RollerBearing. This single row roller bearing,developed by SKF, is characterised bya combination of properties whichmake it interesting for a multitude ofapplications:

● the ability to compensate for angularmisalignments or initial errors ofalignment typical of spherical rollerbearings;

● the ability to take up axial displace-ments in the bearing itself typical ofcylindrical roller bearings;

● the low cross section typical ofneedle roller bearings;

● the high radial load carrying capacityimparted by long sphered rollers;

● the low friction obtained from optim-ally matched raceway profiles;

● the quietness of operation.

Because of its many advantages, theCARB makes an ideal non-locatingbearing. The points in favour of its usein industrial gearboxes include, in addi-tion its compact design and high radialload carrying capacity even whenmisaligned, the potential for downsiz-ing or increasing operational reliabilityor the power rating. The CARB is par-ticularly suitable for the bearingarrange-ments of

● heavily loaded shafts in spur gearboxes,

● pinion shafts in bevel gearboxes,and

● planetary gears.

Two versions of CARB are available: a bearing with cage and a full comple-ment bearing.

SKF has introduced a completelynew roller bearing, the CARB. It isthe only bearing available whichcombines the advantages of threedifferent bearing types without, atthe same time, incorporating theirdisadvantages. For gearbox ap-plications, these advantages trans-late into the following opportunit-ies for enhanced performance.

● Up to 30 % higher load carryingcapacity at the bearing positioncombined with small radialspace requirements

● The low cross section allowsdownsizing or increased per-formance

● Compensation for errors of po-sition and also form of bearingseatings in housings thus allow-ing machining costs to be reduced

● Both bearing rings can bemounted with an interference fitso that there will be no wear inthe bore and no additional axialloads under conditions of axialdisplacement

● Quiet running and little vibration


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2 Bearing types for industrial gearboxesCARB™ roller bearings

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2 Bearing types for industrial gearboxesSpherical roller bearings

Spherical roller bearingsThe self-aligning capability (also inoperation) of spherical roller bearingsmakes their use advantageous whereshaft bending occurs or where thereare errors of alignment between shaftand housing (casing). They are there-fore used in all cases where misalign-ment of the bearing rings would pro-duce inadmissible edge stresses if rigidbearings were used. Additional import-ant characteristics make the sphericalroller bearing a reliable “all-rounder” forgearbox applications. These include

● the high radial load carrying capacityand the ability to accommodate axialloads acting in both directions;

● the wide range of dimension seriesand very wide range of sizes

● even very large sizes.

The many successful developmentrefinements and the improved charac-teristics resulting from them explainthe popularity of spherical roller bear-ings for gearboxes (particularly in spur,bevel and planetary gear units).

The design and functional charac-teristics substantiate the leadingposition of SKF spherical rollerbearings:

● long, symmetrical rollers givevery high load carrying capacity;

● the “floating” guide ring betweenthe rows of rollers ensures thatthe rollers are properly guided(without “wobble”) into the load-ed zone and, in cases whereaxial loads predominate, that theload is correctly carried by therollers and symmetrically distrib-uted over the roller length;

● the special form and optimumsurface finish of the racewaysminimise friction and operatingtemperature enabling highspeed operation;

● the latest development – the Edesign – has even higher loadcarrying capacity as the bearingsection is more efficiently ex-ploited;

● the position of the guide ringabove the pitch diameter in theE design favours lubricant filmformation between the rollersand guide ring;

● all SKF spherical roller bearingsare fitted with robust metalliccages which perform well evenunder arduous conditions.

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2 Bearing types for industrial gearboxesSpherical roller bearings


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2 Bearing types for industrial gearboxesTaper roller bearings

Taper roller bearingsThe tapered form of the racewaysmakes these bearings eminently suit-able for combined radial and axialloads. There is a choice of contactangles so that the appropriate bearingfor the particular combination of radialand axial loads can be found. Thenecessity for functional reasons to usetwo bearings adjusted against eachother enables the force distribution onthe rollers to be controlled so that maxi-mum life can be obtained at the sametime as the stiffness and guidance ofgear shafts can be optimised. Themain gearbox applications are

● spur gear units with helical teeth,● bevel and bevel/spur units and● worm gear units.

As taper roller bearings can supportvery heavy loads, they are alwaysused when the load carrying capacityof other bearings for combined loadconditions (deep groove and angularcontact ball bearings) is inadequate.

Because the raceways are at anangle to the bearing axis, an internalaxial force is produced when the bear-ing is radially loaded, which acts on the housing via the outer ring and can deform it. With larger units (fromapproximately 90 mm shaft diameter)and specifically high performancerequirements, the casing walls areoften not sufficiently stiff, so that theuse of double row or paired single rowtaper roller bearings (or spherical rollerbearings) is recommended, becausethe internal axial forces cancel outeach other and the casing walls willnot be deformed.

Paired single row taper roller bear-ings in a face-to-face arrangement(designation suffix DF) are always usedwhen the preset axial play can be ex-ploited and when adjustment duringmounting is to be avoided.

SKF taper roller bearings have anumber of advantages which makethem suitable for industrial gear-boxes. These include

● the ideal form and optimumfinish of the roller end/guideflange contact enable hydrody-namic lubrication to be achievedand mixed lubrication conditionsavoided, so that the critical run-ning-in process normally re-quired when commissioning a gearbox is not needed;

● the logarithmic raceway profilesguarantee optimum stress dis-tribution over the whole rollerlength and prevent edge stresses;

● the improved surface topographyof the raceways enhances lubric-ant film formation and reducesbearing noise.

29

2

2 Bearing types for industrial gearboxesTaper roller bearings


machine, e.g. in extruder gearing andwater turbine gearboxes. The bearingsare used successfully as thrust bear-ings for the pinion and worm shafts oflarge and very heavily loaded beveland worm gear units.

Spherical roller thrustbearingsThe special feature of these bearingsis their self-aligning capability. Thismeans that their full load carryingcapacity can be utilised, in contrast tothe very stiff cylindrical roller thrustbearings, even when the bearing washers are slightly out of alignmentwith each other. The even distributionof load is still maintained when thereare small angular misalignments of theseating surfaces. Such misalignmentswould considerably shorten the life ofcylindrical roller thrust bearings.

Spherical roller thrust bearings areused in gearboxes, particularly whereaxial forces are produced by the driven

30

2 Bearing types for industrial gearboxesSpherical roller thrust bearings

SKF spherical roller thrust bearingshave particularly low friction thanksto the special roller end/flangecontact geometry.


Marine gearboxwith spherical roller bearings,cylindrical rollerbearings, four-point contact ballbearings andspherical rollerthrust bearings

31

2


3 Design of bearing arrangements

Shafts and gear wheels in spur gearboxes . . . . . . . . 33

Shafts in bevel gearboxes . . 44

Shafts in worm gearboxes . 50

Shafts and gear wheels for planetary gearboxes . . . . . . 56

It is quite possible that several different bearingtypes are used in one gearbox, and where com-bined gear units are concerned, there are sev-eral types of gearing. A stepwise approach is therefore appropriate when selecting bearings,taking each shaft in turn so that the differentconditions for the individual shafts and gearwheels can be fully considered. The bearingarrangements described in the following arewell proven and the conditions specific to a certain shaft are covered. A presentation of themost commonly used bearing series facilitatesthe initial selection.

Design of bearingarrangements

Shafts and gear wheelsin spur gearboxesSpur gearboxes are generally used toreduce speed. There are three maintypes which differ in the way they aremounted: stationary units (mounted onthe machine base), cartridge units(mounted on the drive shaft of the driven machine) and flanged units(flanged to the casing of the primemover and/or driven machine).

The drive from the prime mover isvia a coupling or a belt. The drive istransmitted to the driven machine via acoupling, a quill shaft connection or viaa pinion.

Input shaftsThe input (drive) shafts have the high-est speeds and lightest loads providedno additional external loads have to beconsidered, e.g. belt tension forces.Vibrations and imbalance forces maybe produced by the prime mover. It isalso necessary to consider the prob-

33

3

3 Design of bearing arrangementsShafts and gear wheels in spur gearboxes

lems of high angular accelerationswhen starting without load as well asoperation without load at maximumspeed in order to prevent bearing dam-age caused by the rolling elements sliding on the raceways. There is adanger of this occurring when loadsare suddenly applied. The temperaturedifferences and the associated thermalexpansions in the radial and axialdirections are high for input shafts, asthe speed related large power loss andrelatively small masses as well as therelatively small surface of the pinionshaft mean that there is insufficientheat removal. The distance betweenbearings is dictated by the casing andthe low torque often means that slimshafts are used. This means that shaft

bending and bearing misalignment mustbe taken into account, particularly if abelt drive is used.

Two deep groove ball bearingsarranged for cross location (➔ fig )provide a cost-favourable bearing ar-rangement for moderate power require-ments. Deep groove ball bearings aresuitable for high-speed operation. Be-cause of the low friction, small quanti-ties of oil are adequate for lubricationand cooling so that the collected oilsplashed by the gear wheels dippinginto the oil bath is generally sufficient.

In order to prevent axial clamping ofthe bearings being caused by thermalexpansion of the shaft there should besufficient axial clearance between theouter ring and the cover.

For shaft diameters of up to approx-imately 90 mm, two taper roller bear-ings arranged face-to-face (➔ fig )are advantageous both from technicaland cost considerations. The taper roller bearings are adjusted againsteach other via the cover so that theywill have zero clearance when at theoperating temperature or, for reasonsof stiffness, they may have a slight pre-load. When determining the initial axialclearance it is necessary not only toconsider the temperature differentialbetween shaft and casing but also thedeformation of the shaft and, above all,the casing. The casings of larger unitsare often not stiff enough with respectto the axial forces (tooth force + in-ternal axial forces in the bearings). Insuch cases bearing adjustment is dif-

2

1

Bearing arrange-ment for an inputshaft with twocross-locateddeep groove ballbearings

Bearing arrange-ment for an inputshaft with two cylindrical rollerbearings

Bearing arrange-ment for an inputshaft with twotaper roller bear-ings arrangedface-to-face

34


ficult and shaft guidance is not suffici-ent-ly accurate. The taper roller bear-ing arrangement shown is, therefore,not always suitable.

Cylindrical roller bearings (➔ fig )have a high radial stiffness and guidethe shaft very accurately without havingto be adjusted as taper roller bearings.Axial forces are transmitted via theflanges and roller ends. Because thiscauses more frictional heat, lubricationand cooling must be particularly good.

In order to prevent axial clamping ofthe bearings when thermal expansionof the shaft takes place, there shouldbe adequate axial play between theflanges.

The classical locating/non-locatingarrangement (➔ fig ) is more com-plicated from a design point of viewthan the cross-located arrangementsdescribed above, as the inner andouter rings must be axially located atboth sides. However, it has advant-ages with regard to dimensioning asthe axial force is always taken up by agiven bearing – in this case the spher-ical roller bearing – irrespective of thedirection of the load. Additionally,displacement of the non-locating bear-ing is always assured so that there isno risk of axial clamping occurringwhen the shaft expands.

Two NU-design cylindrical roller bear-ings as radial bearings together with afour-point contact ball bearing as thethrust bearing (➔ fig ) have provedsuitable for very high-speed operation(up to n × dm ≈ 1 000 000). For such

5

4

3

high-speed operation the bearingsmust have

● machined brass cages, centred inthe outer ring,

● increased internal clearance: C3 forthe cylindrical roller bearings and C4 for the four-point contact ballbearing, and

● seatings having increased accuracyof form and position (IT4/2).

At high circumferential speeds thebearings will reject normal oil supplies.Therefore, it is necessary to inject oilat high speed (v ≈ 15 m/s) into the gap between cage and inner ring. Oildrainage facilities should be providedat the injection side of the bearings.

35

3


Bearing arrange-ment for an inputshaft with two cylindrical rollerbearings as theradial bearingsand a four-pointcontact ball bear-ing as the thrustbearing

Classiclocating/non-locating bearingarrangement witha spherical rollerbearing and a cylindrical rollerbearing

other bearing types or arrangementswhich are less unfavourable in respectof casing deformation.

In comparison with input shafts, the axial loading of cylindrical rollerbearings used to support intermediateshafts (➔ fig ) is less critical. Theaxial forces at the gears act in oppos-ite directions and cancel each otherout, at least partially, so that the axialload on the bearings is light. Also thespeeds are lower so that frictional losses deriving from the axial loadremain small.

The high radial load carrying capa-city of the cylindrical roller bearings isan advantage as the intermediate shaftbearings are heavily loaded. The choicebetween caged or full complement cylindrical roller bearings is determinedprimarily by the factors load, speed,lubrication conditions, friction and cost.

Compared with the input shaft, thereis only a small temperature gradientbetween the intermediate shaft and thecasing. This makes it possible to usespherical roller bearings in a cross-located arrangement as shown in fig which is simple in design andtherefore cost-favourable.

There is a wide range of sphericalroller bearings available, particularlyfor medium and large shaft diameters,and there is also a choice of severalcross sections for each diameter. It isthus possible to easily find bearingswhich can support the heavy loadsacting on the intermediate shaft but

8

7

Intermediate shaftsIntermediate shafts are the most heav-ily loaded as they are subjected to theforces from two gear meshes. Thespeeds are moderate. The axial forceson pinion and wheel oppose eachother when the direction of the teeth isthe same so that they partially balanceeach other. There are no additionalexternal forces but vibrations may betransmitted from the input or outputshafts. As there is no torque acting atthe shaft ends, reasonably small dia-meters can be used enabling a rela-tively large bearing section to be util-ised for the accommodation of the high radial forces. Design limits for thebearing outside diameter are set bythe distance between input and outputshafts.

When using taper roller bearings (➔ fig ) it should be rememberedthat axial forces are produced eventhough the load is purely radial. Thismay lead to axial deformation of thecasing. These deformations occur in the central, less stiff region of thecasing because of the position of theintermediate shaft, and are larger than for the input shaft. They lead to achange in position of the shaft and cantherefore cause inadmissibly high mis-alignment of the bearings and themesh.

Experience shows that the casingdeformations occurring in smaller unitswith shaft diameters up to 90 mm aregenerally within acceptable limits. Forlarger units it is necessary to resort to

6

36

Bearing arrange-ment for an inter-mediate shaft withtwo taper rollerbearings arrangedface-to-face

Bearing arrange-ment for an inter-mediate shaft withtwo cylindrical roller bearings


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which have outside diameters withinthe limits set by the distance betweenthe shafts.

A locating/non-locating bearingarrangement as per fig with a spherical roller bearing at the locatingside and a CARB as the non-locatingbearing offers the possibility of reduc-ing the cross section of the non-locat-ing bearing arrangement, because ofthe high load carrying capacity of theCARB, so that the available space canbe better exploited. In many applica-tions there is a risk that the bearingseating in the housing will be ”ham-mered out” so that an intermediatesleeve must be incorporated. By usinga CARB bearing this is no longer a problem as the outer ring is mountedwith an interference fit in the housing,so that a sleeve is not needed.

9

37

3

Bearing arrange-ment for an inter-mediate shaft withtwo spherical roller bearings

Bearing arrange-ment for an inter-mediate shaft with one spherical roller bearing(locating) and one CARB (non-locating bearing)


Fig 9

(EHD) lubrication, i.e. the formation of a separating lubricant film between rolling elements and raceways, cannotbe achieved. Operating bearings under conditions of mixed friction orboundary lubrication will result in wearand shorter bearing life. Besides rota-tional speed, operating temperatureand lubricant viscosity are the mostimportant factors determining EHDlubrication.

There is a limit to how high the viscosity of the oil can be becauseconsideration must be paid to the high-speed gears and bearings in theunit. Therefore, a cooling of the gear-box in the region where the low-speedbearings of the drive shaft are situatedis often the most effective means ofincreasing bearing life. Suitable ad-ditives in the oil can also contribute to a reduction in wear.

Other factors influencing drive shaftbearings depend on the gearbox design:

● In stationary, base-mounted gear-boxes, depending on the type ofpower take-off, it is necessary toconsider the forces of the coupling,the propeller shaft, a pinion or of thedirectly coupled driven machine(e.g. extruders).

Bearing arrange-ment for an outputshaft with twospherical rollerbearings

Locating/non-locating bearingarrangement foran intermediateshaft with twomatched singlerow taper rollerbearings and onecylindrical rollerbearing

The locating/non-locating arrange-ment shown in fig can carry veryheavy radial as well as axial loads.Two matched single row taper rollerbearings (DF execution) are used forthe locating arrangement. In contrastto the cross-located bearing arrange-ments shown in figs 2 and 6, the inter-nal axial forces of the taper roller bear-ings compensate each other within thebearing pair and do not deform thecasing. The intermediate ring suppliedwith the bearing pair ensures that thereis a minimum axial clearance within thebearings. This is adequate for temper-ature differentials between shaft andcasing of up to 20 °C. To avoid deform-ation of the thin-walled inner ring asthe cover screws are tightened, thelength of the centring surface (spigot)of the cover should be chosen to givea preload of approximately 0,01 mm.

Drive (output) shaftsThe conditions for the drive shafts are characterised by high torques and low speeds. The torque calls for a large shaft diameter so that therequisite load carrying capacity can beobtained even when using bearingswith low cross sections. There arepotential problems with lubrication ofthe rolling contacts if, because of thelow speeds, elastohydrodynamic

10

38


● The bearings in cartridge-type gear-boxes are subjected to the reac-tionary forces of the torque support.Additional forces may also be pro-duced as a result of casing deforma-tion.

● The casings of flanged gearboxesare bolted to the driven machine.The shafts are generally rigidlycoupled so that the double supportof the output shaft becomes a mul-tiple support in practice. Centringerrors of the coupled componentsproduce additional forces in the bearings so that narrower tolerancesfor the centring should ensure theaccuracy of alignment of the bearingarrangement.

The arrangement with spherical rollerbearings (➔ fig ) is especially suit-able for applications where rough operation, external additional forces,misalignments and shock loads placeheightened demands on the bearings.Axial shock loads are taken up by theless sensitive raceways in the absenceof flanges on the rings.

For cartidge-type gearboxes, therelatively large diameters of the hollowshaft mean that bearings having lowcross section are suitable. Figshows a well-proven bearing arrange-ment incorporating full complement

12

11

cylindrical roller bearings of seriesNCF 29 V. For lighter loads but withsimilar diameters, deep groove ballbearings of series 619 can be used in the same arrangement. For heavierloads as well as larger deformations,but still with the same diameters andarrangement, spherical roller bearingsof series 239 are appropriate. Deepgroove ball and spherical roller bear-ings have cages and are thus less susceptible to wear when inade-quately lubricated than full comple-ment bearings.

Intermediate gear wheelsAn internal bearing arrangement ismost suitable for intermediate gears as it takes up the least space. Internalbearing arrangements are character-ised by rotating outer rings. Therefore,there is rotating outer ring load andstationary inner ring load. This meansthat the outer rings should have inter-ference fits and the seatings should bevery accurately machined in order tokeep the rotating inaccuracies – whichcause increased friction and additionalforces on the bearing cage – to a mini-mum.

With opposing meshes the circum-ferential forces are added, so that highradial load carrying capacity is re-quired. The axial forces from the

39

3

Bearing arrange-ment for an outputshaft of a cart-ridge-type unitwith full comple-ment cylindricalroller bearings ofseries NCF 29 V

Bearing arrange-ment for an inter-mediate gearwheel with twocylindrical rollerbearings of the NJ design


the inner ring) oil should be supplied atthe side. To prevent the supplied oilfrom being rejected by the bearing, theseal gap at the supply side should notexceed 1 mm.

Shifting gear wheelsFor reasons of space these gear whe-els are supported internally in a similarmanner to the intermediate gears. Thetorque is transmitted in the engagedcondition so that the bearings are sub-jected to the tooth forces. The innerand outer rings rotate but the relativespeed is zero. Both rings have rotatingload but the rolling elements do notroll. The continuous changes in loadunder these stationary conditionscause micro-sliding to take place at therolling element/raceway contacts. Asthere is no relative rotation of the rings,a ”washboarding” type of wear will beproduced in the raceways. This wearcan be reduced by using highly viscouslubricating oil containing anti-wearadditives.

Where the wheels have helicalteeth, the axial force produces a tiltingmoment and consequently a rotatingtilting motion which leads to axial move-ment in the rolling element/racewaycontacts. This increases wear. Ballbearings, adjusted to zero clearance,behave favourably as the balls can

helical teeth oppose each other andpartially cancel each other producing atilting moment on the bearing whichcan cause misalignment.

Two cylindrical roller bearings of theNJ design provide the requisite highradial load carrying capacity in a re-stricted space as shown in fig . Thedesign of the associated components ofthe arrangement is simple. The bearingarrangement of helical intermediategear wheels must be checked for angu-lar misalignment. An unfavourable com-bination of wheel diameter, pitch anddistance between bearings can produceinadmissible values of misalignment.An extended support width (distancebetween bearing pressure centres)can be achieved using, for example,angular contact ball bearings.

Taper roller bearings in a back-to-back arrangement (➔ fig ) also inc-rease the support width as well asreducing the influence of the tilitng moment on the misalignment if theyare adjusted to zero clearance, or alight preload.

Straight cut gear wheels may besupported by a single spherical rollerbearing (➔ fig ). The intermediategear wheels are thus free to align sothat a good mesh is achieved.

In order to be able to use standardbearings (without lubrication holes in

15

14

13

40


Bearing arrange-ment for an inter-mediate gearwheel with twotaper roller bear-ings arrangedback-to-back

Bearing arrange-ment for an inter-mediate gearwheel with a single sphericalroller bearing

also roll in the axial direction and be-cause the movement is reduced by theclearance-free adjustment. Wear isalways load-dependent so that bear-ings under low specific loads wearless. The washboarding effect is alsoless prominent as engagement alwaystakes place at new positions so thatthe wear is evenly spread over theraceway.

For the support of shifting wheels,deep groove ball bearings have provedto give good performance (➔ fig ).Bearings with increased radial internalclearance (C3) are used. The clear-ance-free adjustment via the inner ringsproduces a contact angle in the bear-ings of approximately 15°, so that thesupport width of the bearings is ex-tended. This reduces movement in therelatively stationary bearings underrotating load and thus reduces wear. Inaddition, the clearance-free back-to-back arrangement improves guidanceof the wheel.

Lubrication of the bearings from theoutside is difficult as all components ofthe arrangement – shaft, bearings andwheel – rotate and because the bear-ings are partly covered e.g. by thecoupling. The most reliable method isto supply oil internally through theshaft.

16

41

3


Bearing arrange-ment for shiftinggear wheel withtwo deep grooveball bearings

each particular bearing position mustbe considered. To make the situationclearer in Tables to , the text hasbeen kept as short as possible.

42

Demands on the bearingsModern spur gears generally have hardened gear wheels with groundteeth. It is then possible to obtain highperformance with relatively little frictionand low noise. A prerequisite for this isthe use of high-performance bearings,which should have the properties listedin Table .

In addition to these general require-ments with respect to ball and rollerbearings for high-performance gear-boxes, other demands deriving fromthe specific operating conditions at

1

42


Demand Required bearing design feature

High load carrying capacity Optimised rolling element size and number.Logarithmic roller/raceway contact.Good lubricant film formation through low friction andlow raceway surface roughness.

High stiffness Optimised rolling element size and number.Logarithmic roller/raceway contact.

High dimensional and running accuracy Particularly the inner ring running accuracy should preferably be to tolerance class P6 or better.

Low friction Low friction in roller end/flange contact for taper and cylindrical roller bearings.Low friction in roller/raceway contact.Lightweight precision cage.Low raceway surface roughness.

Low running noise High precision of all bearing components.

Specific operating conditions Requirements of bearings/steps to guarantee performance

High speed and thus high friction Use low-friction bearings. and high operating temperature Avoid over-dimensioning.

Ensure lubricant supply when starting up cold. Provide good cooling.

Large temperature differential when Check required bearing internal clearance; if necessary starting up (slim input shaft heats up select bearings with C3 clearance.more quickly than the better cooled Ensure axial displacement at non-locating bearing position.solid casing)

Vibration from drive; imbalance Use bearings with stable cages, e.g. cylindrical roller bearings forces with steel window-type cages or outer ring centred machined

cages, or spherical roller bearings with steel window-type cages.

Idling under light load Check minimum load. Avoid over-dimensioning.Use bearings with small roller masses where possible.Do not use full complement cylindrical roller bearings.Choose bearing types less susceptible to smearing,e.g. spherical and taper roller bearings.

Demands on rolling bearingsfor spur gears

Demands on inputshaft bearings

Table 1

Table 2

Specific Requirements of bearings/steps operating conditions to guarantee performance

Heavy radial loads Use bearings with high load carrying capacity.

Low to Check lubricant film formation. If necessary increase viscosity or moderate speeds improve cooling. Use lubricants with wear-reducing additives.

Bearing selectionThe following check list will be founduseful when selecting bearings in ordernot to forget any important factors.

● Adjusted basic rating life● Axial load carrying capacity when

the flanges of cylindrical roller bear-ings are under load

● Friction● Stiffness● Misalignment

● Sufficient play to prevent inadmiss-ible clamping when temperature differentials are large

● Minimum load● Static safety under peak loads

A preliminary choice can be made fromthe bearing series shown in Table .5

Demands on inter-mediate shaft bearings

Demands on output shaft bearings

Bearing selection

43

3


Table 3

Specific Requirements of bearings/steps operating conditions to guarantee performance

Very low speeds When lubricant film formation inadequate, i.e. a viscosity ratio (actual to required viscosity) κ < 1, use lubricants with suitable EP additives. When κ < 0,5 bearings with cages (not full complement bearings) must be used.When κ < 0,1 reduce the specific bearing load; aim for s0 > 10.

Shock loads from power Use robust, self-aligning, spherical roller bearings.take-off;deformations

Operating conditions Bearing series normally usedInput shaft Intermediate Output Intermediate Shifting

shaft shaft gears gears

Light loads 62 63 619 60 618/C363 NJ 2 EC 160 62 619/C3

60

Moderate loads NJ 2 EC NJ 22 EC NCF 29 V NJ 2 EC 160/C3320 X 322 239 CC 320 X 60/C3222 E(CC) 222 E(CC)

Heavy loads 322 NJ 23 EC 230 CC NJ 3 EC 62/C3232 CC NJG 23 VH 303223 E(CC) 223 E(CC) 232 CC

322/DF 223 E(CC)

High speeds NU 2 ECMA/C3QJ 2 N2MA/C4

In addition to the bearing series listed above, a CARB can be used as the non-locating bearing for locating/non-locating bearing arrangements

Table 4

Table 5

Shafts in bevel gearboxesBevel gears are generally speedreduction gears. The high-speed driveshaft is termed the pinion shaft and theslow-speed driven shaft carries the larger bevel gear wheel.

The pinion shaft is driven by themotor via a coupling, a spur gear or abelt drive. The power take-off is eithervia a coupling or with bevel/spur gearsvia a pinion.

Pinion shaftsThe pinion is generally supported in anoverhung arrangement. In a few casesthe pinion is supported between thebearings but it is difficult to design in abearing with sufficiently high load car-rying capacity at the head. The over-hung arrangement offers more space.

Two taper roller bearings in a back-to-back arrangement as shown in fig offer a cost-favourable and axi-ally as well as radially stiff arrangementfor small to medium diameter shafts (d < 90 mm). The bearings are adjustedusing a shim between the shaft shoul-der and the inner ring of the bearing atthe input side. The adjustment is deter-mined to give zero clearance when thebearings are in operation and warm or, if required for stiffness reasons, aslight axial prelod. When determiningthe initial axial clearance the tempera-ture differential between shaft andcasing must be considered as well asthe deformations of shaft and casing.

17

Oil should be supplied between thetwo bearings. A baffle plate ensuresthat both bearings are reliably suppliedwith lubricant. The oil drain at the coverside reduces the amount of lubricantreaching the seal.

Bearing arrange-ment for a bevelpinion shaft withtwo taper rollerbearings arrangedback-to-back

44

3 Design of bearing arrangementsShafts in bevel gearboxes

For larger shafts, the requisite loadcarrying capacity can be achieved usinga locating/non-locating bearing arrange-ment as shown in fig . The locatingarrangement is at the drive side andconsists of two matched single rowtaper roller bearings (DF execution).The intermediate ring which is suppliedwith the bearing pair ensures that aminimum axial clearance remains whenthe bearings are mounted which cancope with temperature differentials be-tween shaft and casing of up to 20 °C.For greater temperature differentialssuch as may occur, for example, inoperation when ambient temperaturesare very low, paired bearings with larger axial clearance are required(special execution). In order not todeform the thin-walled intermediatering when tightening the cover screws,the length of the centring flange (spigot)on the cover should be such that a pre-load corresponding to approximately 0,01 mm is obtained.

The matched taper roller bearingsoperate as a double row bearing. Asthe axial load from the pinion domin-ates, one of the two bearings – de-pending on the direction of the load –is completely unloaded. Experienceshows that this is not a disadvantagewhen there is little vibration.

The non-locating bearing adjacent tothe bevel pinion may be either a spher-ical roller bearing, a cylindrical rollerbearing or a CARB.

18

For one-piece casings, spherical rol-ler bearings offer mounting advantagesand they are also relatively insensitiveto smearing when loads vary consider-ably and there are long periods ofidling. If cylindrical roller bearings areused, the requisite axial displacementcan always take place in the bearingitself so that the outer ring can have an interference fit in the housing, andradial guidance is enhanced. The sameis true of CARB (➔ fig ). At thisposition the bearing will not only enablethe axial displacements to be easilyaccommodated, it will also accept theangular misalignments caused by theoff-centre point of action of the toothforces with no reduction in life.

Oil should be supplied to the twotaper roller bearings between the outerrings. Experience shows that for smalland medium-sized gears (up to approx-imately d = 150 mm) the non-locatingbearing can be adequately lubricatedby the oil returning from the locatingbearings. For larger gears, however, itis necessary to arrange for a separateoil supply to the non-locating bearing.For spherical roller bearings, the oilshould be supplied via the lubricationgroove and holes in the outer ring forthe best results.

19

Bearing arrange-ment for a bevelpinion shaft withtwo matched single row taperroller bearingsarranged face-to-face (locatingposition) and one spherical roller bearing(non-locatingposition)

Bearing arrange-ment for a bevelpinion shaft withtwo single rowtaper roller bear-ings arrangedback-to-back(locating) and one CARB (non-locating bearing)

45

3


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�Fig 19

Although the bearing arrangementshown in fig is similar to that in fig

, there are considerable functionaldifferences. All roller rows are alwaysunder load irrespective of the directionof the axial load. If the direction of axialload is from the pinion tip to the driveinput, the taper roller bearing at thecover side with its radially free outerring will be axially loaded, and the op-posing bearing will be radially loaded.If the load direction is reversed thenthe smaller axial load will act on theinboard bearing which is also underradial load. The taper roller bearing atthe cover side will then only be sub-jected to a minimum axial load by thesprings. Because all roller rows arealways under load, this arrangement isless sensitive to vibrations than thatshown in fig .

Mounting is more complex becausethere is no intermediate ring betweenthe taper roller bearings which have tobe adjusted on mounting. The radiallyfree outer ring of the taper roller bear-ing at the cover side is prevented fromturning by an O-ring.

A variant of this bearing arrange-ment incorporates a spherical rollerthrust bearing which has a higher loadcarrying capacity. It replaces the taperroller bearing which only carries axialloads.

With respect to lubrication, the samerecommendations apply as for thearrangement shown in fig .18

18

1820

Output shaftsThe gear wheels are generally ar-ranged between the bearings fordesign reasons. This is also true forthe bevel/spur gearboxes.

For shaft diameters up to approx-imately 90 mm, two taper roller bear-ings mounted back-to-back (➔ fig )provide a technically advantageousand cost-favourable arrangement. Withlarger dimensions, the casings areoften inadequately stiff with regard tothe axial forces (tooth force + internalaxial force of the bearings). This makesadjustment of the bearings difficult andshaft guidance is generally not suffi-ciently accurate. The bearing arrange-ment with cross location is then notaltogether suitable.

The axial force from the gear wheelalways acts in one direction. As theaxial force from the pinion dominates,it is possible that the direction of theresultant axial force will change. Thismust be taken into consideration whenadjusting the mesh.

When adjusting the taper roller bear-ings, the shim at the gear wheel sidedetermines the position of the wheel inthe gearbox. The shim at the pinionside is used to set the axial clearanceof the taper roller bearings.

Oil from the collecting pockets abovethe bearings runs down at the coverside of each bearing. From there theoil must pass through the bearing andthus lubricate it. Oil retaining plates en-sure that there is an adequate supplyof oil available even when starting up.

21

Bearing arrange-ment for a bevelwheel shaft withtwo taper rollerbearings arrangedface-to-face

Bearing arrange-ment for a bevelpinion shaft withone taper rollerbearing as a thrustbearing and onetaper roller bear-ing as a radialbearing (locatingposition) and onecylindrical rollerbearing (non-locating position)

46


The locating/non-locating bearingarrangement shown in fig has theadvantage, compared with that shownin fig , that no bearing adjustmentis required. The bearings are also in-sensitive to axial deformation of thecasing. This will only be subjected tothe tooth forces and not to the internalbearing forces, so that there will beless deformation.

A double row angular contact ballbearing is used as the locating bearing.Alternatively, single row angular con-tact ball bearings in matched setshaving the same diameters as the double row bearing and being margin-ally wider can be used for higher loadcarrying capacity.

To determine the position of the gear wheel in the gearbox and to adjustthe mesh, a split washer is insertedbetween the bearing outer ring and the retaining ring. When doing this thebearing can remain on the shaft. Acylindrical roller bearing of the NUdesign is used as the non-locatingbearing at the other side where theradial load is heavier.

The locating/non-locating bearingarrangement shown in fig is similarin design and function to that shown infig . At the locating side, two singlerow taper roller bearings are arrangedface-to-face. Compared with the doublerow angular contact ball bearing, thetaper roller bearings provide higherload carrying capacity and greater stiffness.

22

23

21

22Adjustment of the bevel gear wheel

is simplified using a special (hook-shaped) sleeve. In order to prevent thethin-walled intermediate ring of the paired taper roller bearings from beingdeformed as the cover screws are tightened, the length of the spigot inthe cover should be chosen to give apreload corresponding to approxim-ately 0,01 mm.

Oil should be supplied to the taperroller bearings via the lubrication groove and holes in the intermediatering. To allow an even distribution overthe two bearings, an oil drain shouldbe provided at the cover side.

47

3

Bearing arrange-ment for a bevelwheel shaft withtwo matched single row taperroller bearings(locating position)and one cylindricalroller bearing(non-locatingposition)

Bearing arrange-ment for a bevelwheel shaft with a double rowangular contactball bearing (locating position)and a cylindricalroller bearing(non-locatingposition)


Demands on the bearingsModern bevel gearboxes usually havehardened gear wheels with groundhelical teeth. This enables high powertransmission to be achieved with littlefriction and little noise generation. Aprerequisite for this good performanceis the use of high-performance ball androller bearings which should have theproperties listed in Table .

In addition to these general require-ments for bearings for high-perform-ance gearboxes, there are additionalrequirements which are specific to the actual bearing position.

6

Bearings for the pinion shaftHigh radial and axial forces act simul-taneously on the pinion shaft. There-fore high radial load carrying capacityis required of the non-locating bearingand high axial load carrying capacity of the locating bearing. Because of thehigh speed, bearings having low frictionshould be used. These two require-ments are in part contradictory.

Experience shows that pinion bear-ings do not fail from fatigue but are en-dangered by other influences. Fromthis it is possible to derive the require-ments and actions listed in Table .7

Demands on rolling bearingsfor bevel gears

Demands on bevelpinion shaft bearings

48


Demand Required bearing design feature

High load carrying capacity Optimised rolling element size and number.Logarithmic roller/raceway contact.Good lubricant film formation through low friction andlow raceway surface roughness.

High stiffness Optimised rolling element size and number.Logarithmic roller/raceway contact.

High dimensional and running accuracy Particularly the inner ring running accuracy should preferably be to tolerance class P6 or better.

Low friction Low friction in roller end/flange contact for taper roller bearings.Low friction in roller/raceway contact.Low raceway surface roughness.

Low running noise High precision of all bearing components.

Most frequent reason for How to alleviate problem/demandspinion bearing damage on bearings

Lubrication breakdown Guarantee lubrication when starting up in the cold state.

Overloading because of too When selecting bearing size, check the temperature differential heavy a preload between shaft and casing. C3 internal clearance often required.

Inadequate lubricant film generation be- Use low friction bearings.cause of too high operating temperatures Avoid over-dimensioning.

Improve cooling.

Smearing on rollers and Avoid over-dimensioning.raceways caused by roller Spherical roller bearings are more favourable than cylindrical slip or sliding roller bearings in larger size range (d > 150 mm).

When using cylindrical roller bearings aim for small roller diameters; use a full complement bearing.

Wear caused by contaminants Avoid contaminating the gearbox during production, assembly and in operation.

Table 6

Table 7

To obtain good meshing it is neces-sary among other things to have abearing arrangement with high radialand axial stiffness. The locating bear-ing should therefore have a large contact angle and as small an axialclearance as possible.

Bearings for the output shaftThese bearings are predominantlyradially loaded so that high radial loadcarrying capacity is also required ofthe locating bearing. Because of theslow speeds the risks in respect ofthermal behaviour and over-dimension-ing compared with the pinion shaft arenegligible. The requirements for axialand radial stiffness, for minimum axialclearance and for bearing accuracycorrespond to those for the pinionshaft bearings.

Bearing selectionWhen selecting the bearings it is use-ful to refer to the cheklist given below.

● Adjusted basic rating life● Permissible speed● Axial and radial stiffness● Sufficient bearing clearance in the

mounted but cold state to avoid inad-missible preload under conditions ofmaximum temperature differentials

● Minimum load

A preliminary selection can be madeusing the overview of the bearing series commonly used; see Table .8

Bearing selection

49

3


Bearing Bearing series normally usedarrangement

Bevel pinion shaft Bevel gear wheel/Bevel/spur gear wheelInput side Pinion side Gear wheel side Opposite side or

spur pinion side

Cross location 72 BE 72 BE 72 BE 72 BE73 BE 73 BE 73 BE 73 BE313 323 B 322 322323 B 323 B 332 332

303 303323 323

Locating bearing(s) (2×) 72 BECB 33(2×) 73 BECB (2×) 72 BECB313/DF (2×) 73 BECB322 + 293 E 320 X/DF

322/DF303 + 294

Non-locating NU 22 EC(/C3) NU 2 ECbearing NU 23 EC(/C3) NU 22 EC

232 CC(/C3) NU 3 EC223 CC(/C3) NU 23 EC

223 EC

In addition to the bearing series listed above, a CARB can be used as the non-locating bearing for locating/non-locating bearing arrangements

Table 8

Shafts in worm gearboxesGenerally worm gearboxes are used to reduce speed. There are two maintypes: one for mounting on the machinebase and a cartridge type for mount-ing on the input (drive) shaft of themachine.

The drive from the prime mover iseither via a coupling or a belt drive.The power take-off is via a coupling ora quill (hollow) shaft connection.

Worm shaftsThe heaviest axial loads act on theworm shaft at the same time as speedsare high. Where there is a belt drive,the radial loads will also be heavy.

ment (➔ fig ) offer a cost-favour-able, low friction bearing arrangementwith low noise for moderate perform-ance and where diameters are small(bearing bore diameter d ≈ 50 mm).The angular contact ball bearings aresuitable for high speeds and becauseof the large contact angle they are alsoappropriate for predominantly axialloads. The two bearings are adjustedagainst each other via the cover so thatthey will have a slight preload whenrunning at the operating temperature.When determining the degree of adjust-ment it is necessary to consider thetemperature differential between shaftand casing, but also casing deforma-tion.

24

The temperature differences and theassociated thermal expansion in theradial and axial directions are also largein worm gearboxes. Only small massesand surfaces of the worm shafts areavailable to remove heat. Therefore,there are large temperature gradientsfrom the shaft to the casing and thesemust be considered when adjustingthe bearings.

The distance between bearings isdictated by the casing and togetherwith the small torques this often leadsto the use of slim shafts. If there is abelt drive, then shaft bending shouldbe calculated so that inadmissiblebearing misalignment can be avoided.

Two single row angular contact ballbearings in a cross-located arrange-

The same type of arrangement butusing two steep-angled taper rollerbearings (➔ fig ) can carry heavierloads than that with the angular con-tact ball bearings for the same shaftdiameter. Therefore, taper roller bear-ings are preferred for higher perform-ance gearboxes and for medium tolarge diameters. When determining the degree of adjustment, it must beremembered that taper roller bearingsare axially stiffer than angular contactball bearings and are therefore moresensitive to excessive preload. It isthus advisable to aim at zero clear-ance when the bearings are running atthe operating temperature. When start-ing up (worm already warm, casing still cold) a slight preload will be pre-

25

Bearing arrange-ment for a wormshaft with twotaper roller bear-ings arrangedface-to-face

Bearing arrange-ment for a wormshaft with twoangular contactball bearings in a cross-locatedarrangement

50

3 Design of bearing arrangementsShafts in worm gearboxes

sent which experience shows can betolerated when lubrication is good.

The locating/non-locating bearingarrangement (➔ fig ) is more costlyfrom a design point of view and be-cause a third bearing is involved but ithas the following advantages:

● higher load carrying capacity (e.g.for belt tension forces);

● if paired angular contact ball bear-ings are used, no individual adjust-ment is required;

● axial displacement at the non-locat-ing bearing position is guaranteed.

A further performance increase can beobtained by replacing the pair of angu-lar contact ball bearings by two match-

26

ed single row taper roller bearings (DF execution) as shown in fig .The intermediate ring which is sup-plied with the bearing pair ensures thatthere is a minimum axial clearance inthe mounted condition, which is suffi-cient for temperature differentials be-tween shaft and casing of up to 20 °C.In order not to deform the thin-walledintermediate ring when the cover scre-ws are tightened, the spigot (centringshoulder) in the cover should have alength such that a preload correspon-ding to approximately 0,01 mm can beobtained.

27

Bearing arrange-ment for a wormshaft with twomatched angularcontact ball bear-ings (locatingposition) and onecylindrical rollerbearing (non-locating position)

Bearing arrange-ment for a wormshaft with twomatched taper roller bearings(locating position)and one cylindricalroller bearing(non-locatingposition)

51

3


The bearing arrangement shown infig is particularly suitable when theaxial load in one direction predomin-ates, as for example in lifting gear. Thespherical roller thrust bearing takes thedominant axial load as well as the axialforce produced in the taper roller bear-ing, which in this case is only subjectedto radial load. If the axial load changesdirection, then the taper roller bearingtakes the radial as well as the axialload, while the spherical roller bearingis spring loaded to give the minimumaxial load required for the correct mo-tion of the rollers. Both bearings areadjusted via the cover. When determin-ing the axial clearance it is necessaryto consider the temperature differentialbetween shaft and casing.

28The advantages of this bearing

arrangement are the very high loadcarrying capacity in the one directionand also that all three bearings arealways under load. Bearing noise isthen particularly low and the bearingsare less sensitive to vibration.

Fig shows a bearing arrange-ment for maximum loads and shock-type operation as encountered, forexample, in rolling mills when the rollsare set. The radial forces are taken upby two radial spherical roller bearingsmounted as non-locating bearings,whilst the axial forces act on the spher-ical roller thrust bearings which haveradial freedom in the casing. The axialclearance of the spherical roller thrustbearings is obtained by adjusting the

29

Bearing arrange-ment for a wormshaft with a taperroller bearing asthe radial bearingand a sphericalroller thrust bear-ing as the thrustbearing (locatingposition) and acylindrical rollerbearing (non-locating position)

Bearing arrange-ment for a wormshaft for maximumloads with tworadial sphericalroller bearingsand two sphericalroller thrust bearings

52


Fig 30: Bearingarrangement for aworm wheel shaftwith two deepgroove ball bear-ings in a cross-located arrange-ment

Fig 32: Bearingarrangement for a worm wheelshaft with two cylindrical rollerbearings

Fig 31: Bearingarrangement for aworm wheel shaftwith two taper roller bearingsarranged face-to-face

width of the spacer sleeve. The springsensure that the requisite minimum loadis applied to the bearing which is re-lieved of axial load.

Worm wheel shaftsThe high torques on the worm wheelshafts require large shaft diameters.As speeds are slow, the load carryingcapacity of low cross section bearings(light series) is adequate.

Because of the low speeds, lubrica-tion of the worm wheel bearings by oilspray is usually not sufficient and spe-cial arrangements must be made forlubricant supply. An oil wiper on theworm wheel or separate grease lubri-cation of the bearings have been foundto give good results.

In most cases the worm wheel has agloboid form and requires accurateaxial guidance, but it must also be pos-sible for the axial position of the meshto be changed.

Two deep groove ball bearings inthe cross-located arrangement shownin fig generally have adequateload carrying capacity and are verycost-favourable. The adjustment of themesh and the bearings is made via the covers. The mesh should prefer-ably be adjusted via the one cover first and then the bearing clearance via the other cover. The temperature ofthe slowly rotating worm wheel shaft is usually low, so the bearings can beadjusted to almost zero clearance. Tokeep the oil level down and still pro-

30

vide adequate lubrication for the bear-ings they are greased and a gap-typeseal is provided on the inboard side.

The arrangement shown in figwith two taper roller bearings is in-tended for heavier loads than thatshown in fig but is otherwise similar. It should be remembered whenusing taper roller bearings that – incontrast to deep groove ball bearings – the axial adjustment of the bearingswill influence the radial guidance of the worm wheel. Therefore, the casingmust be sufficiently stiff so that it willnot be deformed (beaten out) underload. This would otherwise lead to too large a bearing clearance and inadmissible alterations to the mesh.

30

31

When two cylindrical roller bearings(➔ fig ) of the NJ or NCF (full com-plement) design are used, the radialguidance of the worm wheel is notinfluenced by any axial adjustment, sothat setting the mesh is simplified.However, axially loaded cylindrical rol-ler bearings are particularly susceptibleto wear, so that it is important that theyare adequately supplied with lubricantof sufficient viscosity (κ > 0,5) and thatthe specific bearing load is not too high (s0 > 10).

32

53

3


The design of the locating/non-locat-ing bearing arrangement shown in fig

is more complex since all bearingsrings have to be axially located at bothsides. The double row angular contactball bearing (alternatively two matchedsingle row angular contact ball bearings)guides the worm wheel axially with prac-tically no clearance, so that adjustmentis not required. The inner ring withoutflanges of the cylindrical roller bearing(NU design) allows free axial displace-ment at the non-locating side.

Demands on the rolling bearingsThe demands on the rolling bearingsare derived from the specific operatingconditions at each bearing position.They are briefly summarised in Tables and .109

33

Bearing selectionThe following checklist may be helpfulwhen selecting the bearings.

● Adjusted basic rating life● Permissible speed● Maximum preload when starting

up for the maximum temperature differential

● Zero clearance or slight preload atthe operating temperature

● Misalignment● Static safety under shock loads

The commonly used bearing series arelisted in Table to facilitate a prelim-inary choice.

11

Bearing arrange-ment for a wormwheel shaft with a double rowangular contactball bearing (locat-ing position) and a cylindrical rollerbearing (non-locating position)

54



Specifically heavy axial loads Use bearings with large contact angle to take up the axial load.

Need for clearance-free operation and Aim for slight preload when running at the operating quiet running temperature. When adjusting or selecting initial bearing

clearance remember the expected temperature differential.

Large temperature differentials during When adjusting remember the expected temperature start-up (slim worm shaft heats differential to avoid inadmissible preloads.up faster than cooled casing)

High operating temperatures; the use For gearboxes which operate constantly, or mainly of lubricants with large proportion (high frequency of use) at high operating temperatures of additives which are chemically (> 80 °C) and which must also have a long service life aggressive to plasticwhen aged (> 20 000 hours) bearings fitted with metal cages should

be used.


Accurate axial guidance of Adjust bearings to zero axial clearance.worm wheel

Very slow speeds With inadequate lubricant film formation corresponding to a viscosity ratio (actual to required viscosity) of κ < 1 use lubricants with suitable EP additives. When κ < 0,5 only use bearings with cages (not full complement bearings). When κ < 0,1 reduce the specific bearing load; aim for s0 > 10.

Operating conditions Bearing series normally usedWorm shaft Worm wheel shaft

Light loads 72 BEP 61873 BEP 619

160

Moderate loads 313 60(2×) 73 BECBM + NJ 2 ECJ 62313/DF + NJ 3 ECJ 32 + NU 10

(2×) 72 BECBM + NU 2 ECJ

Heavy loads 293 E + 303 + NU 3 ECJ 320 X294 E + 313 + NU 3 ECJ NCF 29 V(2×) 293 E + (2×) 230 CC NJ 2 ECJ

In addition to the bearing series listed above, a CARB can be used as the non-locating bearing for locating/non-locating bearings arrangements

55

3


Table 9

Table 10

Table 11

Demands on wormshaft bearings

Demands on worm wheel shaftbearings

Bearing selection

Shafts and gear wheelsfor planetary gearboxesPlanetary gearboxes usually come ascartridge-type or flanged units andmore seldom for mounting on the base.Power input and also output from thesun wheel is almost exclusively viacouplings so that the sun wheel cancentre itself in the planetary wheels.Power take-off via the planetary car-riers is either via a coupling or, for thecartridge-type units via a hollow shaftconnection.

Sun wheelsThe sun wheel meshes with severalplanetary wheels, so splitting the power.The arrangement is always symmetri-cal so that with straight-cut teeth, thereaction forces on the sun wheel bear-ings should cancel each other out the-oretically. In practice, however, this isnot the case. The even distribution ofload over all the planetary wheels isinfluenced by many factors. The mostimportant are the design (radial align-ment of the sun wheel), the accuracyof manufacture, and the specific load.When the load is heavy the relativedeviation in the load distribution will besmaller because of deformation. Be-cause of the ability of the sun wheel toalign radially and/or the high manufac-turing precision common today, thebearing forces resulting from the un-even load distribution are so small thatthey can be neglected when selectingthe sun wheel bearings. At high

speeds, it is even sensible to subjectdeep groove ball bearings to a mini-mum axial load by springs, in order toprevent them from rotating withoutload, and to achieve smooth running.

Fig shows the bearing arrange-ment of the input shaft which incorpor-ates two deep groove ball bearings ina cross-located bearing arrangement.Transmission of the torque from theinput shaft to the sun wheel is via atoothed coupling. This allows the sunwheel to adjust easily and the load dis-tribution will be good as a result.

The sun wheel shaft shown in fig is also supported by two deepgroove ball bearings, but these are in alocating/non-locating bearing arrange-ment. The springs acting on the outerring of the non-locating bearing subjectboth bearings to axial load. This in-creases the smooth running of thebearings, particularly at high speedsand under vibrating conditions.

35

34

Bearing arrange-ment for an inputshaft with twodeep groove ballbearings

56

3 Design of bearing arrangementsShafts and gear wheels for planetary gearboxes

Bearing arrange-ment for a sunwheel with twodeep groove ballbearings

Bearing arrange-ment for a plane-tary wheel with aneedle roller andcage assembly

Planetary wheelsThe conditions for the planetary wheelsare characterised by heavy radial loadfrom the forces of two meshes as wellas by the infuence of radial accelera-tions and the mass inertia forces result-ing from these. Bearings having highradial load carrying capacity are need-ed, and their cages should be able toendure the mass forces.

An internal bearing arrangement issuitable for the planetary wheels as ittakes up the least space. This meansrotating load for the outer ring andpoint load on the inner ring. Thus, theouter rings must have interference fitsand the seatings must be accuratelymachined in order to keep the rotatinginaccuracy – which leads to increased

friction in the bearings and additionalforces on the cages – to be kept assmall as possible.

The specifically heavy radial loads,the rotating outer rings, and not least,the mass inertia forces cause high fric-tion. Therefore special demands areplaced on the lubrication and coolingof the planetary wheel bearings.

The least space is taken up by aneedle roller and cage assembly asshown in fig . This very cost-favour-able bearing arrangement is very suitable for small units (up to approx-imately 50 mm between shafts) as wellas for light loads or short periods ofoperation as, for example, with smalllifting gear.

The pins and bores of the planetarywheels serve as bearing raceways.Recommendations regarding racewayhardness and design are given in thesection ”Recommended fits” (➔ page106). The planetary wheel is axiallyguided by thrust washers. These aresecured on the planetary carrier sothat they cannot turn.

The bearing arrangement shown infig with two cylindrical roller bear-ings of the NJ design offers the advant-ages of very high radial load carryingcapacity and very high accuracy aswell as high rupture strength in respectof the cage forces if window-typecages are used.

The planetary wheel is guided axi-ally by the flanges of the cylindrical rol-ler bearings. To prevent the bearingsfrom being axially clamped, the inter-mediate ring on the pin should be atleast 1 mm wider than the retainingring in the bore of the wheel.

Even though the two cylindrical rollerbearings are virtually immediately ad-jacent to each other, it is not necessaryto resort to special bearings for pairedmounting (DR execution). Modernmanufacturing methods mean thatstandard bearings differ only slightly intheir cross section (bore and outsidediameters + internal clearance) fromeach other. When using two bearingsper wheel the deviation will, at themost, cause a slight angular misalign-ment which is largely compensated forby deformation, so that any effect onthe mesh or the load carrying capacityof the bearings is negligible.

37

36

57

3


To achieve the maximum load carry-ing capacity in the limited space, thebearing outer rings can be dispensedwith, as shown in fig . Cylindricalroller bearings of the RN design areused. The wheel is guided axially bythe flange rings and the inner ring flanges. The dimensions of the ringsare not standardised and should beagreed with the bearing manufacturer.Recommendations regarding design ofthe raceways in the wheel bore will befound in the section “Recommendedfits” (➔ page 106).

Another way to increase load carry-ing capacity is to use full complementcylindrical roller bearings as shown infig . In this case, a special doublerow bearing without outer ring is used.

39

38

The design provides very high loadcarrying capacity in a small space.However, full complement cylindricalroller bearings cause more friction andare susceptible to wear. They are notsuitable for high normal accelerations.Therefore, this bearing arrangement ismore appropriate for short-term opera-tion, also with heavy load shocks, rather than for constant operation. A typical application area is that of mobile gear units.

The use of a spherical roller bearingto support a planetary wheel, as shownin fig , allows the wheel to adjust tothe mesh. When the planetary carriersdeform, so that the overhung pins become misaligned, the mesh is im-proved by the use of a self-aligning

40

Bearing arrange-ment for a plan-etary wheel withtwo cylindrical roller bearings

Bearing arrange-ment for a plan-etary wheel withtwo cylindrical roller bearings without outer ring

Bearing arrange-ment for a plane-tary wheel withone spherical roller bearing

Bearing arrange-ment for a plan-etary wheel with adouble row fullcomplementcylindrical rollerbearing withoutouter ring

58


Planetary wheelbearing arrange-ment with oneCARB

bearing arrangement, when comparedto a rigid bearing arrangement incorp-orating more than one bearing. Theadvantage of this self-alignment canalso be exploited at high speeds andcorrespondingly small tooth forces, as there is not much deformation in the tooth contact and the mesh will begood. The easy adjustment of the meshis also an advantage when the wheelsare wide. The smaller theoretical loadcarrying capacity of the single spher-ical roller bearing as compared withrigid arrangements where two or morebearings are used is partly compens-ated for by the even distribution of load over the two rows of rollers.

Because of its exceptionally highload carrying capacity compared withother roller bearings and its low crosssection, the CARB is eminently suit-able for planetary gear bearing ar-rangements (➔ fig ). Its insensitivityto angular misalignment is particularlyimportant for correct meshing in thiscase. The planetary wheel can alignitself so that even meshing is obtainedacross the whole tooth width. Thefavourable distributon of the tooth forces thus obtained has a positiveinfluence on the life of the gearbox.

41

59

3


��

��

��

��

��

��

��

Fig 41

Planetary carriersTo achieve equal power splitting in planetary gear units, it is possible toavoid the need for an additional bearingsupport for the planetary carrier if the following conditions apply:

● the planetary carrier is not subjectedto load from the output shaft or thetorque support;

● the weight of the planetary carrier isnegligible.

The planetary carrier centres itselfunder load via the planetary wheelmeshes.

Fig shows a gearbox where theplanetary carrier of the high-speedstage is not supported by bearings. It

42

centres itself via the planetary wheelsin the hollow shaft (housing) and onthe supported sun wheel. This multiplecentring is only possible if manufactur-ing precision is adequate.

Planetary gearboxof cartridge typewith two deepgroove ball bear-ings supportingthe planetary carrier

60


The casing is supported by the plan-etary carrier of the slow-speed stage.The two deep groove ball bearings in across-located arrangement are underload from the restoring force of the torque support and from the weight ofthe gearbox. The resultant bearing forces are generally small and the rotational speed low so that the load carrying capacity of deep groove ballbearings is usually sufficient.

The planetary carrier with take-offshaft shown in fig is supported bytwo full complement cylindrical rollerbearings. This arrangement enablesadditional forces from the power take-off to be accommodated.

43

61

3


Bearing arrange-ment for planetarycarrier with twofull complementcylindrical rollerbearings

Demands on the rolling bearingsThe special requirements placed onthe bearings for planetary gearboxesare derived from the particular condi-tions pertaining at the various bearingpositions. A brief summary is given inTables to .1412

Bearing selectionThe following list may be found usefulto check that the chosen bearingssatisfy the demands.

● Adjusted rating life● Permissible radial acceleration● Permissible speed● Friction and cooling● Adequate bearing play to prevent

inadmissible preload at the maxi-mum operating temperature (sunwheel) or under interference fits (planetary wheel)

Demands on sunwheel bearings

Demands on planetary wheelbearings

62



Light loads; idling Use of deep groove ball bearings preferred to avoid over-dimensioning.

Requirements for clearance-free Adjust deep groove ball bearings axially by springs.operation and quiet running

Large temperature differentials Particularly where casings are solid and/or well cooled use when starting up (slim sun wheel shaft deep groove ball bearings with internal clearance to C3.heats up more quickly than the casing which is better cooled)

Table 12


Heavy radial loads Use roller bearings with high load carrying capacity. If lubricant film formation is also inadequate, corresponding to a viscosity ratio (actual to required) of κ < 1 use lubricants with suitable EP additives. When κ < 0,5 only use bearings with cages (not full complementbearings).When κ < 0,1 reduce the specific bearing load; aim for s0 > 10.

Radial accelerations resulting from move- Check cage stresses by calculating mass inertia forces. ment of the planetary wheels around the Pay consideration to mass inertia forces of planetary wheel axis of rotation of the sun wheel when calculating bearing life.

Increased friction caused by mass Ensure adequate lubricant supply and cooling.intertia forces and rotating outer Use heat-stable lubricants.rings (rotating inaccuracy) For gearboxes which continuously, or frequently (high

frequency of use) operate at high tempeatures (> 80 °C) and which should also have long service life (> 20 000 hours) bearings with metallic cages should be used.

Deformation of planetary wheel by For thin-walled planetary wheels (wall thickness < 3 × modulus)two meshes on opposite sides take into account the influence of the tension band load distribution

on the loaded zone of the bearing (FEM calculation).

Table 13

Operating Bearing series normally usedPlanetary wheels Sun wheels Planetary carriers

Low radial accelerations or NJ 23 ECP 60, 62, 63 618, 619short operation periods NCF 30 V NCF 18 V, NCF 29 V

NJG 23 VH 239 CC230 CC232 CC223 E(CC)

Moderate radial accelerations NJ 3 ECMA 60, 62, 63 618, 619and continuous operation NJ 23 ECMA NCF 18 V, NCF 29 V

230 CC 239 CC232 CC223 E(CC)

High radial accelerations NJ 2 ECML 60, 62, 63 618, 619NJ 3 ECML NCF 18 V, NCF 29 VNJ 23 ECML 239 CC223 CCJA/VA405

In addition to the bearing series listed above, a CARB can be used for planetary wheels


Very slow speeds with additional Use preferably bearings with small cross section.loads from the drive When κ < 1 use lubricants with suitable EP additives.

When κ < 0,5 only use bearings with cages (not full complement bearings). When κ < 0,1 reduce the specific bearing load; aim for s0 > 10.

● Deformation of planetary wheelwhen wall thickness small; influenceon the load distribution in the bearing

● Static load safety in respect of loadshocks

A preliminary bearing selection can bemade by referring to the most fre-quently used bearing series listed in Table .15

Bearing selection

Demands on planetary carrierbearings

63

3


Table 15

Table 14

4 Calculation of bearing arrangements

Bearing loads . . . . . . . . . . . . 65

Determination of external forces . . . . . . . . . . . . . . . . . . 66

Calculation of bearing loads . . . . . . . . . . . . . . . . . . . 74

Dimensioning the bearingarrangement . . . . . . . . . . . . . 76

Life calculation . . . . . . . . . 76Static safety factor . . . . . . . 79Axial load carrying capacity . . . . . . . . . . . . . . . 79Minimum load . . . . . . . . . . 80Normal acceleration and cage load carrying capacity 80Friction and cooling . . . . . . 81Permissible speeds . . . . . . 82Internal clearance and preload . . . . . . . . . . . . 83Adjustment values for single row angular contact bearings . . . . . . . . 85

Following the preliminary selection of bearingtype, it is necessary to determine all the exter-nal forces acting on a gear unit and from themto calculate the bearing loads. For the finalselection of bearing size (and execution) severalcriteria must be observed, the most importantof which is bearing life.

Calculation of bearingarrangements

Bearing loadsTo calculate the bearing loads it is firstnecessary to determine all the externalforces acting on the shaft/bearingsystem. The following are external for-ces:

● tooth forces,● mass inertia forces from radial ac-

celerations in planetary gears,● coupling and propeller shaft forces,● belt forces, and● weights of shafts and gear wheels.

An analysis of the force distributionover the bearings must then be made.There is a choice of method:

● conventional methods based on thebeam model, suitable for manualcalculations and the correspondingcomputer programs (included in theSKF CADalogue and ADAM soft-ware); these methods rely on simpli-

fying assumptions and models;● advanced methods where bearings,

shafts and in part also housings(casing) are considered as a non-rigid system; these methods involveextensive calculations and requirethe use of sophisticated computerprograms available in house at SKF.

Where experience is available from the same or similar designs it is still the custom to use the conventionalmethods for comparative calculations.Because of the greater informationobtained using the sophisticated methods it is recommended that theybe applied for new designs and alsowhen conducting damage analysis.Please contact SKF for assistance.

65

4

4 Calculation of bearing arrangements Bearing loads

Kp

KaKn

Kr

Determination of external forces

Tooth forcesThe magnitude of the tooth forces isdependent upon the torque which is tobe transmitted. As the torque is thefundamental criterium on which all cal-culations are based, and consequentlyalso the evaluation of the bearingarrangement, it should be determinedas accurately as possible, e.g. by measuring or based on experience.Additional forces caused by inaccur-acies in the mesh which come from the manufacturing process, or byshocks originating from the input oroutput drives, are taken into accountby selecting an application-relatedminimum life.

When calculating the forces for spur,bevel and planetary gears (➔ fig ),tooth friction is ignored. Friction is onlytaken into account for hypoid and wormgears where there is a larger propor-tion of sliding friction.

In the following equations the index1 is used for the driving wheel and theindex 2 for the driven wheel.

The peripheral force Kp dependson the torque or power and can be ob-tained from

M WKp = = 9,5517 × 106

r n r

For spur and bevel gears, the gearratio is

n1 r2 Z2= =n2 r1 Z1

1

SymbolsK tooth force acting at right angles

to the tooth flank, NKp tangential component of K

(= peripheral force), NKa component of K acting parallel to

the shaft axis (= axial force), NKn component of K acting at right

angles to the shaft axis (= normal force), N

M torque to be transmitted by gearwheel, Nmm

W power to be transmitted by gearwheel, kW

r pitch radius (mean radius forbevel gear wheels), mm

n rotational speed of gear wheel,r/min

α angle of engagement, degreesβ angle of inclination, degreesδ half cone angle of bevel gear

wheels, degreesγ pitch of worm, degreesZ number of teethµ coefficient of friction of tooth

flanks of hypoid and worm gearsη degree of efficiency for hypoid

and worm gears

Tooth forces

66

4 Calculation of bearing arrangements Determination of external forces

Fig 1

r

Kp

Kn

Kp

Knr

βKa

Spur gearFor straight cut spur gear units (➔ fig )

Ka = 0

Kn = Kp tanα

and for spiral cut spur gear units (➔ fig )

Ka = Kp tanβ

tanαKn = Kp cosβ

Bevel gear unitsFor straight cut bevel gear units (➔ fig )

Ka1 = Kp tanα sinδ1

Kn1 = Kp tanα cosδ1

Ka2 = Kp tanα sinδ2

Kn2 = Kp tanα cosδ2

4

3

2

For spiral cut or curved bevel gearunits (➔ fig ) the equations shownin Table should be used.

As the equations

Ka2 = Kn1

Kn2 = Ka1

also apply to bevel gear units wherethe shafts are at right angles to eachother, it is sufficient in this case to cal-culate the forces acting on the drivingwheel, as this will also determine theforces on the driven wheel.

Where the teeth are straight cut, theforces Kn and Ka always act in thedirections shown in fig 4. For spiral cutor curved teeth, the forces may act inthe opposite direction, depending onthe angles α, β and δ. In this case, thecalculated values for Kn and Ka arenegative.

15

67

4


Fig 2

r

Kp

Kn

δKa

Fig 4

Kp

Knr

Ka

β

Fig 5

Fig 3 Tooth forces ofhelical cut spurgears

Tooth forces ofstraight cut

Tooth forces ofhelical cut bevelgears

Tooth forces ofstraight cut

68

Tooth forces forhelical and curvetoothed bevelgears

Tooth forces forhypoid gears


Table 1

Table 2

Driving wheel

Driven wheel

Driving wheel

Driven wheel

Ka1 =Kp (− sinβ cosδ1 + tanα sinδ1)

cosβ

Kn1 =Kp (sinβ sinδ1 + tanα cosδ1)

cosβ

Ka1 =Kp (sinβ cosδ1 + tanα sinδ1)

cosβ

Kn1 =Kp (− sinβ sinδ1 + tanα cosδ1)

cosβ

Ka2 =Kp (sinβ cosδ2 + tanα sinδ2)

cosβ

Kn2 =Kp (− sinβ sinδ2 + tanα cosδ2)

cosβ

Ka2 =Kp (− sinβ cosδ2 + tanα sinδ2)

cosβ

Kn2 =Kp (sinβ sinδ2 + tanα cosδ2)

cosβ

Ka1 = K (− cosα sinβ1 cosδ1 + sinα sinδ1 + µ cosβ1 cosδ1)

Kn1 = K (cosα sinβ1 sinδ1 + sinα cosδ1 − µ cosβ1 sinδ1)

Ka1 = K (cosα sinβ1 cosδ1 + sinα sinδ1 − µ cosβ1 cosδ1)

Kn1 = K (− cosα sinβ1 sinδ1 + sinα cosδ1 + µ cosβ1 sinδ1)

Ka2 = K (cosα sinβ2 cosδ2 + sinα sinδ2 − µ cosβ2 cosδ2)

Kn2 = K (− cosα sinβ2 sinδ2 + sinα cosδ2 + µ cosβ2 sinδ2)

Ka2 = K (− cosα sinβ2 cosδ2 + sinα sinδ2 + µ cosβ2 cosδ2)

Kn2 = K (cosα sinβ2 sinδ2 + sinα cosδ2 − µ cosβ2 sinδ2)

Hypoid gear unitsAs can be seen from fig , the twoshafts of a hypoid gear unit do not liein the same plane. Therefore, theangle of inclination β1 of the drivingwheel is not the same as β2 of the driven wheel. The wheels are so chosen that β1 is larger than β2. Thedirections of the peripheral forces Kp1and Kp2 do not coincide, in contrast tospur and bevel gear units.

For hypoid gears, the ratio is

n1 Z2 r2 cosβ2= = ×n2 Z1 r1 cosβ1

As (cosβ2/cosβ1) > 1, the pitch radiusof the pinion is greater for a given ratioand a given size of the wheel whichthe pinion engages than is the case fora bevel gear unit.

The peripheral force Kp1 which actson the pinion is obtained from

M1Kp1 =r1

The tooth force which acts vertically onthe tooth flank is obtained from

Kp1K =cosα cosβ1 + µ sinβ1

and the peripheral force for the largewheel from

Kp2 = K (cosα cosβ2 + µ sinβ2)

6

Kn2

r1

Ka2

Kn1Kp1

Kp2

Ka1

r2

The forces Ka and Kn are obtainedusing the equations shown in Table ,taking into account the requirementsfor the direction of the spiral cut and ofrotation.

Worm gear unitsWhen calculating worm gears it iscommon practice to take the angle ofpitch γ instead of the angle of inclina-tion β. The following equations can beused

γ = 90 − β

htanγ =

2 π r1

where h in mm is the pass height ofthe worm on the partial cylinder and r1in mm the pitch radius of the worm.Generally, the worm drives the wormwheel and the following calculation isfor this case. Index 1 refers to theworm and index 2 to the worm wheel(➔ fig ).

The tooth forces are obtained from

M1Kp1 =r1

cosα cosγ − µ sinγKa1 = Kp1 cosα sinγ + µ cosγ

= Kp1 cotγ η

7

2

69

4

Hypoid gearsTooth forces ofworm gears


Fig 6 Fig 7

R R R R

S S S Sr

r r

r

pitch radius of the hollow wheel R andits planetary wheel to that of the sunwheel and its planetary wheel. Themagnitude of the radius s, which forcorrected toothing must not be (R + r) /2, does not matter here.

The advantage of this method is thatthe various types with double planetarywheels can be calculated in a simplemanner. The same equations can beused for all planetary gear units of ty-pes I to III (➔ fig a), which are equi-valent to the simple unit (➔ fig b).

It should be remembered that thevalues of R, s and r to be inserted inthe equation for u correspond to thelever of the three parts which act onthe assumed double lever in the plane-tary wheel (thus, R is not always theradius of the hollow wheel, s notalways the radius of the planetarywheel and r not always the radius ofthe sun wheel).

The basic ratio is obtained from

R (s − r)u =

(R − s) r

For normal toothing

(s − r) = (R − s)

is valid so that

Ru =

r

88

Construction ofplanetary gearbox(schematic)

sinαKn1 = Kp1 cosα sinγ + µ cosγ

tanα= Kp1 [sin2γ (1 − η) + η]

sinγ

As can be seen from fig the forcesacting on the worm wheel are deter-mined by calculating the forces on the worm as follows

Kp2 = Ka1; Ka2 = Kp1; Kn2 = Kn1

The reduction ratio for worm gear units is

n1 Z2=n2 Z1

where Z1 is the number of passes ofthe worm and Z2 the number of teethof the worm wheel.

Planetary gear unitsThe determination of the forces isshown for the most common type ofplanetary gear, i.e. with parallel shaftsand toothed pinion. Using the followingequations it is of no importance for thedetermintion of the speeds and torqueswhich of the three parts is connectedto the drive, the power take-off, or thestationary part (casing), or whether allthree parts are in motion and transmitpower.

The calculation starts with the basicratio u, which is the ratio of the rolling

7

70


Fig 8

a b

type I type II type III simple unit

71

If the symbols R, s and r are insertedfor the different equivalent planetaryunits according to fig a, the theupper equation for u is again valid.

Speeds:

nr = (u + 1) ns − u nR

(u + 1) ns − nrnR =u

nr + u nRns =u + 1

Speeds of the planetary wheels abouttheir own axes a) for simple planetary gear units

Rnpl = (nr − ns) R − s

r= (ns − nr) s − r

1= (nR − nr) s − r + R − s

r R

b) for double planetary unitstype I:

Rnpl = (nR − nr) R − r

s= (ns − nr) s − r

type II:

Rnpl = (nR − ns) R − s

r= (ns − nr) s − r

type III:

rnpl = (nR − nr) R − r

s= (nR − ns) R − s

8

Torques:

Ms = Mr + MR = (u + 1) Mr

1= ( + 1) MRu

1 1Mr = MR = Msu u + 1

uMR = u Mr = Msu + 1

Tooth forces:The peripheral force is obtained from

Mr MrKp = or Kp =r Zpl R Zpl

where Zpl = number of planetary whe-els.

For straight cut teeth

Ka = 0

Kn = Kp tanα

and for spiral cut teeth

Ka = Kp tanβ

tanαKn = Kp cosβ

4


72

ϕ

FG

a

FG

M

Kr

S2

S1

r

lowing radially acting pair of maximumforces should be used for calculations

MFG max = tanϕ

a

whereFG max = maximum, periodically

changing force, NM = torque to be transmitted, Nmma = distance between bearings,

mmϕ = bending angle of joint, degrees

As FG max is the maximum of the peri-odically changing force, an approxim-ate average force can be obtained from

Fm = 0,75 FG max

assuming that the bearings are onlysubjected to load caused by the jointforces. If the bearings are also sub-jected to other forces then the follow-ing approximation applies

1 2Fm = Fmin + Fmax3 3

Fmin = forces other than the joint forcewhich act on the bearings, N

Fmax = all forces acting on the bearing,including the joint force FG max, N

As the bending angle ϕ changes therewill be a compensation in length of thepropeller shaft which, because of fric-tion, will produce an axial force

Inertia forces from the radialaccelerationThe rotation of the planetary carrierabout its own axis causes inertia forceson the planetary wheels which must beconsidered when calculating the bear-ing load if speeds are high.

For the inertia (gyratory) force on aplanetary wheel

F = m rs ω2

whereF = inertia force, Nm = mass of planetary wheel, kgrs = radius of centre of gravity of

rotating planetary wheel, mω = angular velocity of the planetary

carrier

π ns(= ), s−130

ns = rotational speed of the planetarycarrier, r/min

Coupling and propeller shaft forcesWhen selecting and designing torque-transmitting couplings, it is desirablethat no reactionary forces act on theshaft/bearing system. Even though thisis not completely possible, because ofinaccuracies governed by manufactureor deviations when aligning thecoupled shafts, and not least becauseof deformations, it may still be assu-med that the coupling forces are negli-gible in comparison to the tooth forces.

With propeller shafts, forces are pro-duced when the torque is transmitted.These forces rotate with the rotation of the shaft and change periodically (➔ fig ). For two bearings, the fol-9


Cardan shaft forces

Belt forces

Fig 9 Fig 10

73

Type of Preload factor fbelt drive at peripheral speed (m/s)

< 5 5 to 20 > 20

Flat belts 3 to 4 2,5 to 3,5 2 to 3

V belts 1,5 to 2,5 1,5 to 2,5 1,5 to 2,5

Toothed belts 1,1 to 1,3 1,1 to 1,3 1,1 to 1,3

MFa = µ cosϕ

rm

whereM = torque, Nmmrm = mean radius of the sliding

profile, mmµ = coefficient of frictionϕ = angle of bending, degrees

As this axial force only acts during cer-tain periods – namely when the ben-ding angle changes – it should betaken into account for the time it actswhen calculating the life, or if thechange in angle occurs when the shaftis not rotating, it should be included inthe calculation of the static safety fac-tor s0.

Belt forcesThe gear unit may be driven by a beltand power take-off may also be via abelt. The radial force acting on the shaft

4


Preload factor

Torque supportforces

Table 3

G

K1

K2

a

l

Fig 11

Kr produced by the belt (➔ fig ) canbe calculated using

MKr = f Kp = f

r

whereKr = resultant belt force, NKp = peripheral force, NM = torque, Nmmf = tensioning factorr = radius of belt pulley, mm

Appropriate values of tensioning factor,depending on the peripheral speed,can be obtained from Table .

Forces from the torque supportIn cartridge-type gear units, the bear-ings on the output shaft are subjectednot only to the tooth forces, but also toforces from the reaction to the torqueand from the weight (➔ fig ).

The force K1 acting on the outputshaft bearings can be obtained from

M aK1 = + G

l l

whereK1 = force acting on the bearings, NM = reaction torque (for simplicity it

can be taken as being equal tothe torque of the output shaft),Nmm

G = weight of gear unit includingmotor and base plate, N

l = distance between torque supportand output shaft, mm

a = distance between torque supportand centre of gravity, mm

When calculating bearing load itshould be remembered that the forceK1 is introduced via the bearing outerrings from the casing.

Weights of shafts and gearsThe weights of shafts and gears aregenerally negligible compared with the tooth forces. They should not beignored, however, when dealing with vertical units as they act axially andmay constitute a considerable part ofthe total bearing load – particularly inlarge units.

11

3

10

74

0

0,5

0,4

0,3

0,2

0,1

0 21,81,61,41,210,80,60,40,2

Ka = Fa

Fa cotαFr

Kn Fr

ax

0b

can be made by taking into accountthe resilience of the shaft and bear-ings. This can be done with the in-house computer programs developedby SKF.

The external force acts on theshaft between the pressure centres of the bearingsThe bearings with pressure centres Iand II at a distance l corresponding tofig are loaded by a force K actingin any direction. The force is dividedinto the components Kp, Kn and Ka.

For the forces acting vertically at thebearing positions

l − a rF1 I = Kn − Kal l

a rF1 II = Kn + Kal l

and for the forces acting horizontally

l − aF2 I = Kpl

aF2 II = Kpl

12

Calculation of bearingloadsOnce the external forces have beendetermined it is possible to calculatethe bearing loads. It is sensible to divide the forces, as shown in fig ,into three vertically acting components.The forces act at the pressure centresof the bearings. For deep groove ballbearings, cylindrical roller bearingsand spherical roller bearings, the pres-sure centre is at the geometric centreof the bearing. For single row angularcontact ball bearings and taper rollerbearings, the distance between thepressure and geometric centres of thebearing will be found for each bearingin the SKF General Catalogue.

If a shaft is supported in a doublerow angular contact bearing, or in twosingle row angular contact bearingsarranged back-to-back, plus an addi-tional bearing, and if the distance be-tween the bearings is relatively small,under a load consisting of a radialforce component Kn and an axial com-ponent Ka, the position of the line ofaction of the radial force Fr acting onthe bearing pair or bearings will influ-ence the distribution of the externalload over the three rows of rolling ele-ments. The distance ax of the line ofaction can be determined approxim-ately from the diagram in fig inrelation to the contact angle of thebearing and the load ratio Fa/Fr.

A more realistic determination of theload distribution over the three rows

13

12

4 Calculation of bearing arrangements Calculation of bearing loads

Forces acting atthe bearing posi-tions when anexternal force isapplied at a pointbetween the pres-sure centres

Position of theforce produced bydouble row andpaired single rowangular contactbearings

r

Kp

Ka

Kn

F1I

FrI

F2I

I

F2II

II

F1II

Fa

l

a

K

FrII

Fig 12 Fig 13

Ball bearings

Roller bearings

75

The resultant radial load s for the bear-ings can then be determined using

Fr I = √ F1 I2 + F2 I

2

Fr II = √ F1 II2 + F2 II

2

The axial force Fa acts on one of thetwo bearings – the locating bearing –in addition to the radial forces. Whenthe bearing is not a single row angularcontact bearing, Fa = Ka. In single rowangular contact bearings under radialload, an axial force will be inducedwhich must be taken into accountwhen calculating the equivalent dy-namic bearing load. Details will befound in the SKF General Catalogue.

The external force acts on theshaft away from the pressurecentres of the bearingsThe force K is also divided into threecomponents: Kp, Kn and Ka. Accordingto fig , for the bearing forces actingvertically

a − l rF1 I = Kn − Kal l

a rF1 II = Kn − Kal l

14

and for the horizontally acting forces

a − lF2 I = Kpl

aF2 II = Kpl

The resultant radial loads acting on thebearing can then be obtained, as be-fore, from

Fr I = √ F1 I2 + F2 I

2

Fr II = √ F1 II2 + F2 II

2

Once the radial load Fr and the axialload Fa have been determined, theequivalent dynamic bearing load P andthen the bearing rating life L10h can bedetermined following the instructionsgiven in the SKF General Catalogue.

The conventional determination ofthe bearing load described here isbased on many simplifying assump-tions in order to permit manual calcula-tion. More realistic results are obtainedif the deformation of bearings, shaftsand possibly also of the casing can betaken into account. This can be doneusing the sophisticated SKF computerprograms available in house.

For shaft systems supported at three or more positions it is imperativethat deformations are considered, asthe conventional methods often lead to rather unrealistic results. Even for stat-ically determinate doubly supportedshafts, it is advisable to calculate usingthe more sophisticated methods whenthe application limits for a new designare being evaluated, or when addition-al information is required on bearingand gear displacements and misalign-ments, or on rolling element loads andstresses in the rolling contact, ratherthan the approximate life.

4

4 Calculation of bearing arrangements Calculation of bearing loads

Forces acting atthe bearing posi-tions when anexternal force isapplied at a pointoutside the pres-sure centres

r

a

l

Kp

Ka

Kn

F1II

FrII

F2II

II

Fa

KF2I

FrI

F1I

I

Fig 14

76

A more reliable selection can bemade by calculating the adjustedrating life L10ah which also takes intoaccount lubrication. The calculationrequires information regarding the vis-cosity of the lubricant to be used andthe bearing operating temperature inaddition to the load and speed. By cal-culating the adjusted rating life it isalso possible to determine whether thelubricant is suitable and whether cool-ing would give better results. A deter-mination of the adjusted rating life isalso helpful for the following reasons.

● Bearings operating at high speedsbut which are lightly loaded arenegatively influenced by high tem-peratures and large inertia forces.Lubricant film formation is promot-ed at high speeds and an adjustedrating life calculation can show thata smaller bearing can be used thanwould be suggested by a basicrating life calculation, so that frictionas well as inertia forces will be re-duced. The reliability of the bearingarrangement will be enhanced.

● Slowly rotating bearings operatingunder heavy loads are subject todeformations with correspondinglyhigh proportions of sliding in the rol-ling contact and are susceptible towear. The slow speeds mean thatlubricant film formation will be poor-er, and the adjusted rating life cal-culation will lead to the choice ofbearings having higher load carryingcapacity. This will mean that thespecific bearing loads will be lighter,deformations and wear will be re-duced, and reliability enhanced.

Contamination has a considerableeffect on the life of gearbox bearings.The influence of contamination can becalculated using the SKF New LifeTheory. The fatigue load limit is alsoconsidered when calculating the ad-justed rating life L10aah according to theNew Life Theory so that it is possibleto design an arrangement for infinitelife.

4 Calculation of bearing arrangements Dimensioning the bearing arrangement

Dimensioning the bearing arrangementThe bearing size and execution re-quired for a given bearing arrange-ment are determined based on the following criteria:

● life● static load carrying capacity● axial load carrying capacity● minimum load● normal acceleration and cage load

carrying capacity● friction and cooling● speed capability● internal clearance and preload● adjustment values for single row

angular contact bearings.

In many cases bearing size is simplyselected on the basis of the calculatedlife. The list above and the followingcomments serve to show that for reli-able performance of the bearing, anumber of other criteria should be con-sidered in addition to the calculatedbearing life.

Life calculationThe Lundberg and Palmgren theory ofbearing fatigue life forms the basis forbearing life calculations. The life equa-tions derived from the theory are to befound in the SKF General Catalogue.Their use for gearbox bearing calcula-tion will be discussed here.

Bearing life can be calculated withgreater accuracy and reliability, themore accurately the operating condi-tions are known or can be determined.To calculate the basic rating life L10haccording to ISO it is only necessary to know the basic dynamic load ratingof the bearing, the equivalent bearingload and the rotational speed. Import-ant influences such as lubricant filmformation in the bearing and lubricantcleanliness are not considered in theL10h calculation. In spite of this, if ex-perience of similar bearing arrange-ments is available and the other para-meters which affect bearing life, butwhich are not considered in the calcu-lation are reasonably constant, a basicrating life calculation may be sufficientto determine the appropriate bearingsize.

77

Gearbox application L10h(operating hours)

Machines and equipment infrequently used: 300 to 3 000Household appliancesAgricultural machineryMedical equipment

Machines used for brief periods or intermittently: 3 000 to 10 000CranesLifts and elevatorsConstruction machinery

Machines for daily (8 hour) use: 10 000 to 30 000Machine toolsWoodworking machinesFansConveyor drivesCentrifuges

Machines for 24-hour use: 30 000 to 50 000Rolling millsCompressorsPumpsBarges

Machines for 24-hour operation where high reliability is required: 50 000 to 100 000Cement millsRotary furnacesPower generating plantLarge-size open cast mining equipmentWind and water turbinesOcean-going ships

The following parameters are con-sidered when calculating L10aah:

● dynamic load rating of the bearing,● fatigue load limit of the bearing,● equivalent dynamic bearing load,● rotational speed,● lubricant viscosity,● operating temperature and cooling,

and● contamination and sealing.

Calculations according to the New LifeTheory are particularly suitable formaking parametric studies to deter-mine the influence of the different fac-tors. It should be noted that the variousfactors have a strong influence oneach other, and such calculations areonly meaningful when the operatingconditions are exactly known. Whenbearing life calculations for the selectionof bearing size are made, only thoseresults obtained using one and thesame method should be compared.

When determining a suitable life it isnecessary to consider how the gear-box is to be used. The requisite basicrating life is dependent on the type

and size of the driven machine, on the length of service and on demandsregarding operational reliability. If noexperience is available then the guide-line values for the requisite basic ratinglife L10h given in Table can beused.

In similar applications, the drives oflarge machines are generally subject-ed to more arduous conditions thanthe drives of smaller machines be-cause of stronger shock loads and larger defomations. This should betaken into consideration when choosingthe guide-line value from Table 4.

When bearing arrangements are in-tended for very slow rotational speedsand/or are to have a very short life, the requisite basic dynamic load ratingof the bearing is very small. This canlead to an unsuitable bearing beingchosen which will give inadequate static safety, or the formation of onlyan inadequate lubricant film, or to theoverloading and consequent deforma-tion of the associated components. If,in addition to the requisite life, a mini-mum requisite value of the static safetyfactor s0 is also to be considered, this

4

Guideline valuesfor the requisitebasic rating lifeL10h for gearboxesfor various appli-cations

4


Table 4

78

should be based on the κ value (ratio of actual to required viscosity). Thedecision not only depends on the oper-ating speed therefore, but also on theviscosity at the operating temperatureand on the mean bearing diameter.

Table contains recommendationsas whether the bearing selection shouldbe based on the requisite life or on thestatic safety, taking the value of κ intoaccount. Thus

● when κ > 0,5, the static safety factors0 should be checked in addition tothe requisite life;

● when κ ≤ 0,5 then the static safetyfactor s0 must be considered;

5

● when κ < 0,1 no life should be given;the material will fatigue under condi-tions of small κ, but the operationalreliability and service life will notdepend on fatigue but on other fac-tors which are indirectly accountedfor by the static safety factor s0.

Selection criteria

Guideline valuesfor the static safety factor s0


Bearing type Type of operationRotating, Rotating, Rotating at very slow Stationarystatically brief shock speeds under loadloaded loadsnrel = 0 nrel > 0 κ < 0,1 κ = 0,1 to 0,5

Ball bearings 2 2 10 5 0,5

Roller bearings 3,5 3 10 5 1

Full complement – 3 20 10 1cylindrical roller bearings

Viscosity Bearing selection based on ratio fatigue life static safety factorκover incl. L10h L10ah L10aah s0

0,1 − − − +

0,1 0,5 − o + +

0,5 1 + + + o

1 + + + o

Symbols+ recommended– not appropriateo can also be used

Table 6

Table 5

79

Static safety factorThe basic static load rating C0 is usedto select bearing size in the followingcases:

● when the bearing rotates at a rel-ative speed of 0 (bearing arrange-ments of shifting gears) under load(rotating static load);

● when the bearing rotates and must,in addition to the normal loads, takeup heavy shock loads for a fractionof a revolution (e.g. rolling mill drives);

● when the bearing rotates very slowlyunder constant load;

● when the bearing is stationary and isunder constant load or is subjectedto shock (short duration) loads, e.g.in mobile gearboxes.

The guideline values of the static safety factor s0 for different bearingtypes given in Table are valid whenthere is adequate lubrication using aCLP oil to DIN 51 517 which offersgood protection against wear.

Bearing selection based on the static safety factor s0 is described inthe SKF General Catalogue.

Axial load carrying capacityThe axial loads acting on rolling bear-ings are considered when calculatingthe equivalent dynamic and static bear-ing loads, see SKF General Catalogue.However, the axial load carrying capa-city of cylindrical roller bearings is primarily determined by the load carry-ing ability of the sliding surfaces of the roller ends and flanges and is verystrongly dependent on the lubricationand cooling. When calculating the permissible axial load according to theSKF General Catalogue, a viscosityratio κ ≥ 2 is presupposed. When κ issmaller friction and wear will increase.Based on experience these effects canbe kept at an acceptable level forslowly rotating gearbox bearings if thefol-lowing favourable conditions pertain

6

● light axial load,when 0,1 < κ ≤ 0,5: Fa max = 0,05 Fap,when 0,5 < κ ≤ 1: Fa max = 0,1 Fap,when 1 < κ ≤ 2: Fa max = 0,2 Fap,where Fap is the maximum permiss-ible axial load at κ ≥ 2

● there is an adequate supply of a CLPoil which offers good protectionagainst wear

● the arrangements for oil supply anddrainage are designed so that wearparticles will not collect in the bearing

4


80

Minimum loadIn order for bearings to perform corr-ectly they must always be subjected to a given minimum load. This will pre-vent the rolling elements from slidingon the raceways which would lead tosmearing and premature bearing fail-ure. This minimum bearing load can becalculated using the information givenin the SKF General Catalogue. Whenthis minimum load is constantly applied,there will be practically no sliding in thebearings. This load can be applied rather easily to thrust bearings, e.g. bysprings, even when they are idling, butmay be more difficult to arrange forradial bearings.

In cases where the weights of shaftand gears are insufficient for the mini-mum load requirements, the risk of sliding can at least be reduced if thefollowing recommendations arerespected.

● Use ball bearings, taper roller bear-ings or spherical roller bearingswhere possible (full complement cylindrical roller bearings are mostat risk).

● Use bearings with small rolling elements – in critical cases at theexpense of basic rating life.

● Keep bearing internal clearancesmall and – if at all possible – applya preload.

● Avoid metallic contact in the rollingelement/raceway contacts (ensureadequate supply of lubricant havingsufficient viscosity; if necessary usebearings with black oxidised rollingelements).

● Ensure high accuracy of positionand form of the associated com-ponents and use bearings of corre-spondingly high precision.

● Avoid vibrations wherever possible.● Limit periods of idling under insuffi-

cient load as far as possible.

Experience shows that idling under in-sufficient load in gearboxes cannot al-ways be avoided. The bearings whichare most susceptible to damage undersuch conditions are large cylindricalroller bearings (d > 150 mm) as well asfull complement cylindrical roller bear-ings. Often the bearings are damagedduring test running without load.

The development of smearing – thetypical damage caused during idling –and its prevention are being studied.SKF application engineers will gladlyprovide information on the latest re-search results.

Normal acceleration and cageload carrying capacityThe movement of a planetary gearbearing is made up of a guidance orlocating movement, resulting from therotation of the planetary carrier, and arelative movement resulting from thebearing turning in the planetary carrier.In comparison with bearings mountedin stationary housings, the guidanceand coriolis accelerations cause addi-tional inertia forces to act on the plan-etary gear bearings. The mass of theplanetary gear and the associatedbearing rings produces a force as aresult of the normal guidance accelera-tion which the bearing arrangementmust also accommodate. These accel-erations also mean that the masses ofthe rolling elements and cage will exertadditional forces as well as the bearingitself.

These additional inertia forces acton the rolling elements, bearing ringsand, to a high degree, also the bearingcage. It is thus possible that a bearingwill fail not from fatigue but because ofcage fracture.

The additional forces increase thesliding friction in the contacts whichguide the rolling elements and cage. In full complement cylindrical roller bearings, because of the normal accel-eration, the rollers are in contact witheach other, so that friction increasesand lubricant film formation is hin-dered. As a result the risk of scuffing or seizure is increased.

SKF has specially developed com-puter programs for the calculation ofthe cage carrying capacity and also forhow much the friction will be increasedby the additional forces as well as therisk of seizure for full complement cylindrical roller bearings.


81

An estimate of the permissible nor-mal acceleration for the bearing andcage designs most frequently used forplanetary gears can be made using the following equation and cataloguedata

dm0,8

an ≤ ka g × 103

C0

wherean = permissible normal accelerationka = a factor (➔Table )dm = mean bearing diameter

= 0,5 (d + D), mmC0 = basic static load rating, N

7

Friction and coolingBearing friction depends on the follow-ing factors:

● load,● speed,● bearing type,● bearing size,● lubricant properties (viscosity in

operation), and● lubricant quantity.

The total frictional resistance in a bearing is made up of

● rolling and sliding friction in the rolling element/raceway contacts,

● sliding friction in the rolling element/cage contacts (rolling element guidance),

● sliding friction in the cage/bearingring contact (cage guidance),

● friction in the lubricant, and● sliding friction of the rubbing seals

in sealed bearings.

Friction influences heat generation and consequently bearing operatingtemperature. In gearboxes, the gearsproduce more friction than the bear-ings. When making arrangements forcooling, therefore, it is necessary toconsider the total friction in the gear-box.

Factor ka

4


Bearing type Bearing design Factor kafor circulating for oil bath lub-oil lubrication rication withoutwith good cooling special cooling

Cylindrical roller ECP 120 40bearings ECJ 170 50

ECM 150 50ECMR 400 150ECMA 700 250ECMP 1 400 500ECML 1 800 600

Spherical roller E 250 100bearings CC 600 200

CC/VA405 1 400 500

Table 7

82

Operating temperatures in the gear-box (and thus the bearings) shouldpreferably not exceed 100 °C and de-finitely not be higher than 150 °C forthe following reasons:

● High lubricant viscosity enhanceslubricant film formation.

● The lubricant ages more slowly, thelower the temperature.

● The dimensional changes in thebearing rings and rolling elementsresulting from micro-structural changes in the material are smaller,the lower the temperature.

● The temperature differential acrossa bearing is smaller, the lower thetemperature, so that preset bearingclearance or preload will not changeas much.

The power loss resulting from the bear-ing friction can be calculated using infor-mation given in the SKF General Cata-logue. Heat is removed from a bearingby conduction, convection, radiationand by the lubricant. If circulating oillubrication is to be used, the requisitequantity of oil can be calculated from

NRQ = 0,039Ta − Te

whereQ = requisite quantity of oil (oil flow

rate), l/minNR = power loss, WTa = oil temperature at exit, °CTe = oil temperature at inlet, °C

By experience, approximately 1/3 ofthe power loss is dissipated by the oiland 2/3 through heat conduction, con-vection and radiation. A value of 10 °Ccan be assumed for the temperaturedifference (Ta – Te).

The guideline values obtained usingthe equation below have been found tobe good estimates of the oil flow rates.

Q = f D B

whereQ = oil quantity (oil flow rate), l/minf = factor depending on bearing type

and duty = 0,00003 for radial ball bearings,

and radial roller bearings formoderate duty

= 0,00005 for radial roller bearingsin general

= 0,00001 for thrust bearings,radial roller bearings with rotat-ing outer ring and planetarygear bearings

D = bearing outside diameter, mmB = bearing total width (radial bearings)

or height (thrust bearings), mm

The guideline values for the oil flowrate are generally on the safe side. Forsmall bearings only very small quantit-ies are required and it is difficult to ar-range for a correct supply, particularlywhen the temperature varies. Often,the oil from pockets which capture oilwill be sufficient. As there is a risk withforced oil circulation that the leads andnozzles become blocked it is recom-mended that either at least 0,25 l/minis supplied to each bearing, or supplypumps should be used which allow larger supply cross sections evenwhere oil quantities are small andpressures high.

Permissible speedsWhen considering the operatingspeed, the speed ratings quoted in theSKF General Catalogue should beused as a reference. Bearing speedswhich are higher than 70 to 80 % ofthe catalogue speed ratings are con-sidered high. In such cases the follow-ing influences must be specially takeninto consideration.

● The heat produced as a result of thefriction increases bearing tempera-ture; lubrication (viscosity, type oflubricant, lubricant supply) and cooling must be checked.

● As the heat loss via the casing isusually good, the temperature dif-ferential from inner to outer ring islarger, and a bearing having in-creased internal clearance (e.g. to C3) is required.


83

● To ensure proper performance of thebearing (no slip and proper rollingmotion of the rolling elements) a correspondingly higher minimumload is required.

The maximum permissible speeds aremuch higher than the speed ratings(see SKF General Catalogue, factor fn).This also applies to gearbox bearings,so that the maximum permissiblespeeds for deep groove ball bearings,cylindrical roller bearings (with cage),angular contact ball bearings and four-point contact ball bearings are twicethe speed ratings. If operating speedsare to exceed the speed ratings bymore than 50 %, however, it is not onlynecessary to consider the points out-lined above but also the followingpoints.

● Use oil jet lubrication with a jetspeed of approximately 15 m/s. Theoil should be directed at the innerring raceway or the gap betweencage and inner ring.

● Particularly stable cage designsshould be chosen, e.g. one-pieceouter ring centred machined brasscages (window-type), designationsuffix ML, for cylindrical roller bearings.

● Minimise the vibrations produced inthe complete drive system. Thismeans using bearings with increas-ed accuracy of dimensions and formand associated components withcorrespondingly high accuracy.

● Take into account the critical bend-ing and torsional vibrations whendesigning the gearbox shafts.

In cases where bearings fitted withspecial cages or with increased ac-curacy are required, it is advisable tocontact the SKF application engin-eer-ing service.

Internal clearance and preloadThe clearance in a bearing in opera-tion is important with regard to properperformance of the bearing and to proper load distribution on the rollingelements. The following conditionsshould be aimed for when the bearingshave reached their operating temper-ature.

● For radial roller bearings in gearbox-es (e.g. cylindrical, spherical anddouble row taper roller bearings) aslight radial internal clearance isfavourable as the bearings andassociated components (shaft,casing) usually have high radial stiff-ness. Radial preload combined withthe deviations from form normallytolerated in gearboxes, or combinedwith unexpected differences in tem-perature would increase the risk ofinadmissibly high additional stressesoccurring which would overload thebearing.

● For single row taper roller bearings,although they have high radial stiff-ness, an axial preload can alwaysbe allowed if it can be expected thatbearing overloading can be avoidedby the casing walls ‘‘giving” in theaxial direction.

● For ball bearings zero clearance isbest; a slight preload is less criticalfor ball bearings than for the muchstiffer radial roller bearings.

When calculating the clearance in op-eration it must be rememberd that theclearance range quoted in the GeneralCatalogue will be reduced when thebearing is mounted with interferencefits and by the temperature differentialfrom inner to outer ring. The Normalbearing clearance is sufficiently largeso that if the fits are as normallyrecommended and operating condi-tions are normal, a sensible operation-al clearance will be obtained. In gear-boxes, unusual operating conditions(e.g. in the cases below) often requirethe use of bearings with greater thanNormal internal clearance to C3 or C4.In such cases it is advisable to checkthe operational clearance.

● Bearings mounted inside gears forwhich an interference fit for the outerring is required. This will furtherreduce internal clearance.

● Bearings on high-speed slim shaftswhich will heat up much more rapid-ly than the casing. The temperaturedifferential across the bearing willthen be particularly large.

4


84

Bearing (designation): ..........................

Tolerances (shaft/housing bore): .......................... low high

Radial clearance (µm)1 Bearing bore (deviation ∆dmp) .......... ..........2 Shaft (deviation) .......... ..........3 Theoretical interference (+) or clearance (−): Zth = Point 2 − Point 1 .......... ..........4 Expected interference Z = Zth − smoothing1) .......... ..........5 Expansion of inner ring:

d d/F [1 − (di/d)2]el = Z (solid shaft) el = Z (hollow shaft) .......... ..........

F 1 − (d/F)2 (di/d)2

6 Bearing outside diameter (deviation ∆Dmp) .......... ..........7 Housing bore (deviation) .......... ..........8 Theoretical interference (+) or clearance (−): Zth = Point 6 − Point 7 .......... ..........9 Expected interference: Z = Zth – smoothing1) .......... ..........10 Compression of outer ring:

E/D [1 − (D/Da)2]eA = Z .......... ..........

1 − (D/Da)2 (E/D)2

11 Total radial clearance reduction (Point 5 + Point 10) .......... ..........12 Radial internal clearance before mounting (min/max) .......... ..........13 Radial internal clearance after mounting (Point 12 − Point 11) .......... ..........14 Thermal expansion:

∆tet = 1,1 dm (µm, with dm in mm) .......... ..........

100

15 Radial clearance in operation (Point 13 − Point 14) .......... ..........

Axial clearance (µm) for double row angular contact bearings11a Total axial clearance reduction (Point 11 × cotα) .......... ..........12a Axial internal clearance before mounting (min/max) .......... ..........13a Axial internal clearance after mounting (Point 12a − Point 11a) .......... ..........14a Thermal expansion:

∆teta = 1,1 dm cotα (µm, with dm in mm) .......... ..........

100

15a Axial clearance in operation (Point 13a − Point 14a) .......... ..........

1) For guideline values for smoothing see Table 9.

● Gearboxes where the casing is wellcooled. Again there will be a largetemperature differential across thebearings. Examples include gear-boxes operating out of doors whereambient temperatures are low andgearboxes having thick-walled orfan-cooled casings.

The operational clearance (mountedbearings which have reached the oper-ating temperature) can be calculatedby following the scheme shown inTable .8

Calculation ofoperationalclearance


85

Nominal Smoothingdiameterover incl

mm µm

– 50 4

50 100 6

100 – 8

Guideline valuesfor smoothing ofmating surfaces

Definition of distance betweenbearings

Adjustment values for singlerow angular contact bearingsSingle row angular contact bearings(angular contact ball bearings, taperroller bearings) are adjusted axially onmounting. The adjustment values(axial clearance or preload) are basedon the operating conditions when thebearing is under load and has reachedits operating temperature. Light pre-load is recommended for gearbox bearings and provides the followingadvantages compared with clearance:

● accurate shaft guidance,● increased stiffness,● extended calculated and service

lives,● quiet running, and● compensation for settling move-

ments in operation.

As the bearings have to be adjustedon mounting, i.e. in an unloaded con-dition at ambient temperature, thechanges produced when the bearingsare in operation must be consideredwhen determining the adjustment values. The main influences are thoseof temperature and deformations.

Influence of temperature on theadjustment of angular contact bearingsThe inner rings of bearings mountedon gearbox shafts are generally hotterthan the outer rings. This will reducethe set clearance or increase the setpreload. The influence of temperatureon the adjustment can be calculatedusing the following equation providedboth shaft and casing are of steel or a material with the same thermal behaviour

∆a = 11 × 10−6 [0,5 (dmA T∆A cotαA

+ dmB T∆B cotαB) ± T∆m L]

whereDa = reduction in axial internal

clearance caused by tem-perature differential, mm

dm = mean bearing diameter = 0,5 (d + D), mm

L = mean distance between bearings (➔ fig ), mm

α = contact angle of bearing,degrees (cotα = 1,5/e; for values of bearing-dependent factor e see SKF General Catalogue)

T∆A, T∆B = temperature differential frominner to outer ring acrossbearings A and B, °C

T∆m = temperature differentialfrom shaft to casing, °C

The plus sign is used for bearingsarranged face-to-face, the minus signfor bearings arranged back-to-back.

15

4


Table 9

Fig 15

Face-to-face arrangement Back-to-back arrangement

��

��

��

L

A B

��

��L

A B

��

86

a1δ

F01

Ka

1δ

If the value of the temperature dif-ferential T∆ is not known from experi-ence or measurements, the followingguideline values can be used:

T∆ = 5 to 10 °C for slowly rotating gear-box shafts

T∆ = 10 to 20 °C for intermediateshafts and moderate speeds

T∆ = 20 to 30 °C for slim high-speedshafts

T∆ = 30 to 40 °C for high-speed inputshafts and well-cooled gearboxes

Influence of deformations on theadjustment of angular contact bearingsWhen considering deformations itshould be remembered that the totalresilience is influenced not only by theresilience of the bearings but also bythe elasticity of the associated compo-nents, the fits and the elastic deforma-tions of all other components throughwhich the forces pass, including thegearbox support. The effects of the different stiffnesses of the associatedcomponents can be represented inpreload force/preload path diagrams.

The three preload force/preload pathdiagrams shown in Diagrams toshow the influence of casing stiffness

31

Preload force/pre-load “path” dia-grams for a bear-ing arrangement(Design 1)



Diagram 1

F02

Ka

a2δ

2δ 1δ=

Diagram 2

Preload force F0

Bearing ABearing B

Bearingposition B total

Bearing position Atotal

Axial displace-ment δa

Preload force F0

Bearing ABearing B

Bearingposition B total

Bearing position A total

Axial displace-ment δa

87

��

��

Ka

B A

δa

��

3δ

F01 =

F03

Ka

a3δ

4


Diagram 3

Fig 16


Pinion shaft bearing arrange-ment

Preload force F0 Bearing ABearing B

Bearing position Btotal

Bearing position Atotal

Axial dis-placement δa

on the axial displacement δa for thepinion shaft shown in fig as a re-sult of the external force Ka.

In all three cases, the bearing stiff-ness and the external force Ka are thesame. The casing in case 1 is very stiffwhereas the casings in cases 2 and 3are less stiff. Cases 2 and 3 differ onlyin the preload. Whereas in case 2 thepreload path d is kept constant withrespect to case 1, for case 3, the pre-load force F0 is the same as for case 1.Irrespectively of whether the preloadpath or the preload force is kept con-stant, the axial displacement δa willchange depending on the casing stiff-ness. Thus it is imperative that thetotal resilience at the bearing positionsis taken into account when determining

16the preload in order to limit the axialdisplacement.

Using the application exampleshown in fig (a bevel/spur gear)the choice of adjustment (axial clear-ance, zero clearance or preload) willbe discussed.

The locating bearings for the bevelpinion shaft have axial clearancebecause the temperature differentialfrom shaft to casing is relatively largeas the speed is high and the pinionshaft has a small mass. Also the bear-ings are arranged in the (hook-shaped)sleeve and this arrangement is relatve-ly stiff in the axial direction.

The intermediate shaft bearings andthose on the output (power take-off)shaft can be either clearance-free or –

17

88

depending on casing stiffness – evenbe adjusted to preload. The reason forthis is that the speeds are low (lessfrictional heat), the masses of the shaftsare relatively large, and the axial stiff-ness of the casing is lower. In fact, because of the axial forces generatedin the bearings, the casing tends todeform (bulge).

Influence of adjustment on bearing lifeThe adjustment has different effects onthe life of the two bearings shown in

fig . Whereas the life of bearing Awhich is subjected to the external forceKa immediately drops with increasingpreload, bearing B will achieve itsmaximum life when it has a slight preload.

Diagram shows qualitatively thedependence of bearing life on preloadand clearance. From this it will be seenthat the stiffness does not increasevery much with increasing preloadwhereas there is a risk that bearing lifewill be shortened and there will be inc-reased friction and heat. Thus it isadvisable to choose the adjustment sothat when under load and at the operat-ing temperature the bearing arrange-ment will have virtually zero clearance.An adjustment to give a distinct pre-load should only be chosen if the oper-ating conditions (loads, temperatures,deformations) are accurately known,so that the preload force can be deter-mined using sophisticated computerprograms.

4

16

Influence on bear-ing life of preloadand clearance

Bevel/spur gear-box bearing arrangements


Diagram 4

Life

Bearing A

Bearing B

Preload Clearance

89

When selecting the bearings there-fore, not only must the complete bear-ing designation (cage design, bearingclearance) be established, but informa-tion regarding adjustment values, oilflow rates and minimum load must alsobe given to production and assemblyas well as to the end user, so that proper bearing performance can beguaranteed.

4



Grease lubrication . . . . . . . . 92

Oil lubrication . . . . . . . . . . . 95

Maintenance . . . . . . . . . . . . . 98

5

Rolling bearings will only perform reliablywhen they are adequately lubricated. The lub-ricant prevents intermetallic contact betweenrolling elements, raceways and cage and alsoprotects the bearing surfaces against corrosion.The importance of lubrication can be seen fromthe fact that of all premature bearing failures,some 80 to 90 % are caused by faulty lubrica-tion and/or contamination. Long experienceindicates that the same estimate holds true forgearbox bearings.

Lubrication and maintenance

The task of the gearbox designer tochoose the most suitable method oflubrication as well as the most suitablelubricant is made more difficult be-cause of the different and varying de-mands on lubrication which exist forone and the same gearbox. Generally,the lubrication must not only be appro-priate for the bearings but also for thegears. Additionally, the operating con-ditions for the individual bearings in agearbox are often very different. Onetype of lubrication can be the optimumfor high-speed, lightly loaded bearings,but unsuitable for heavily loaded bear-ings which rotate slowly. The operatingtemperature, which has a significantinfluence on the quality of the lubrica-tion, is often not only dependent on the

load and speed but is also affected by changes in ambient temperature.Since, generally, only one method oflubrication and one lubricant are to be used for a gearbox, the optimum will never be achieved. To find the bestcompromise all the demands regardinglubrication and lubricant propertiesmust be weighed against each other.The explanations and recommenda-tions given in the following may behelpful.

91


Grease lubricationThe most important advantages of grease lubrication are:

● good protection against corrosion asthe grease adheres well to the bear-ing surfaces;

● the efficiency of seals against ex-ternal contaminants is reinforced;

● there is little risk of leakage;● reliable lubricant supply – particularly

when operation is intermittent – asthe grease is retained at the bearingposition;

● freedom from maintenance for lubricated-for-life bearings.

From this it is possible to define themain areas where grease lubricationcan be employed in gearboxes. It isused mostly for small units and particu-larly for geared motors, and the gearsare also grease lubricated. Small gear-boxes may often be used in varyingpositions (horizontal, vertical or in-clined at an angle). In such cases lub-ricant supply is more reliable if greaseis used rather than oil bath lubrication.Sealing arrangements can also besimpler if grease is used. The liferequirements are often very moderate for small units and if they are onlyused for short periods at a time, theywill require no maintenance, being literally lubricated for life.

For oil bath lubricated vertical gear-boxes it is sensible to grease the upperbearings as the amount of oil splashedup is generally inadequate. The greasecan be retained in position by baffleplates.

GreasesThe following properties must be con-sidered when selecting an appropriategrease.

Base oil viscosityGenerally speaking, the base oil vis-cosity of a grease can be used to calcu-late the adjusted rating life Lna, seeSKF General Catalogue. This viscosity,ν, should preferably be greater thanthe required viscosity ν1, both viscos-ities being at the bearing operatingtemperature.

ConsistencyGreases of consistency 2 and 3 aregenerally used for rolling bearing lub-rication. Greases with lower consist-ency are easier to pump; those with higher consistency are easier to retainat the bearing position. At low temperat-ures soft greases of consistency 0 or 1may be used, but special grease supplyarrangements must then be made(e.g. 100 % grease fill, or a centrallubrication unit and short relubricationintervals). For gearboxes subjected tovibrations or which are arranged verti-cally, a consistency 3 grease with highmechanical stability is preferable.

When ‘‘gearbox greases” are usedfor small gearboxes, lubrication is atype of ‘‘dip” lubrication. The greaseshave a consistency of 0 or 00.

Temperature rangeThe expected operating temperatureshould lie within the temperature rangepermitted for the grease. When thetemperature is too low, the grease willnot have sufficient lubricating propertiesand when it is too high, ageing will beaccelerated. An increase of 15 °C halves the original relubrication inter-val.

Load carrying ability and wear protectionFor heavily loaded bearings (C/P < 10,e.g. bearings on the intermediate andoutput shafts) or in cases where a fullyseparating lubricant film is not present(κ < 1), EP greases are used. As theeffect of some EP additives may bedetrimental to bearing life, it is advis-able to contact the lubricant supplierfor recommendations.

92

5 Lubrication and maintenanceGrease lubrication

5

Protection against corrosionUsually gearboxes are well protectedagainst the penetration of water. Never-theless the presence of water ormoisture cannot be completely preven-ted as differences in temperature allowcondensation to form. Since any waterin the rolling contacts of a bearing willquickly destroy the bearing surfaces,only greases having good rust inhibit-ing properties should be used.

Oil bleedA grease must bleed oil to allow theformation of a lubricant film in the rol-ling contact. At low temperatures con-siderable bleeding is advantageous toensure lubricant supply. At very slowspeeds grease will be pushed awayfrom the raceways and will no longerparticipate in bearing lubrication. Oilwill not bleed to the raceways so thatstarvation will occur in the rolling con-tact. Consequently, oil lubrication is tobe preferred for very slow speed oper-ation. A much more moderate oil bleed is preferred at higher temper-atures (> 80 °C) in order to give longrelubrication intervals.

MiscibilityIf, for some reason, it is necessary tochange to another grease it should bechecked whether the base oil and

thickener of the old and new greasesare compatible. When a combinationof oil and grease lubrication is used(e.g. grease lubricated bearings and oillubricated gears) the lubricants shouldalso be compatible with each other ifnegative results are to be avoided.This is particularly important when syn-thetic gear oils and mineral oil basedbearing greases are used.

93


SKF greases

SKF grease Use, propertiesDesignation

LGMT 2 Small bearings (outside diameter D up to approx. 62 mm)Light to moderate loadsModerate temperatures up to 80 °C (max 120 °C) Low friction, quiet, good protection against corrosion

LGMT 3 Medium-sized bearings (outside diameter > 62 mm up to approx. 240 mm)Moderate loadsModerate temperatures up to 100 °C (max 120 °C)Multi-purpose grease, good protection against corrosion

LGEP 2 Heavily loaded roller bearingsModerate temperatures up to 80 °C (max 110 °C)Good protection against corrosion

LGEM 2 Heavily loaded roller bearings at low speedsModerate temperatures up to 90 °C (max 120 °C)Water repellant

LGLT 2 Small, lightly loaded bearings at high speedsLow temperatures down to −20 °CLow friction, water repellant

LGHQ 3 High temperatures above 80 up to 150 °CModerate loadsModerate speedsWater repellant

94

1Suitable SKF lubri-cating greases forgearbox bearings


SKF greasesThe SKF range of lubricating greasescovers nearly all the requirements forgearbox bearing lubrication. Thesequality greases were specially devel-oped for bearing lubrication. The mostimportant technical data will be foundin the SKF General Catalogue. Table

gives recommendations regardingthe particular suitability of the variousgreases for different gearbox applica-tions.

Methods of grease lubricationThe selection of the lubrication methodis basically governed by the relubrica-tion interval which can be determinedusing the information given in the SKFGeneral Catalogue.

● In cases where the relubricationinterval is longer than the expectedservice life of the bearings a singlegrease fill will suffice. This presup-poses that the grease can be retain-ed in the bearings and that any oilbled from the grease cannot escapethrough openings below the bear-ings. Lubrication for life has onlybeen found suitable for small andmedium-sized bearings (bearingoutside diameter up to 240 mm).

● Manual relubrication using a grease gun is suitable when relubri-

1

cation intervals are in the range oneweek to six months and the quant-ities required are up to 500 g. Thismeans that manual relubrication can be used for bearings with out-side diameters up to 420 mm.

● For larger bearings (D > 420 mm),larger quantities of grease (G > 500 g), or shorter relubrication intervals than one week, a continu-ous supply of grease is more reliableand also more economic. This is alsotrue where the number of bearingsto be grease lubricated is large.

When designing the grease supply,care should be taken to ensure thatgrease cannot escape at the supplyside of the bearing, i.e. that it is com-pelled to pass through the bearing. Atthe opposite side of the bearing, theemerging used grease will preventcontaminants from entering the bear-ing. For double row bearings, the mostefficient method is to supply the greasevia the lubrication holes in the outerring or, for paired taper roller bearings,through the lubrication holes in theintermediate ring.

Table

5

Oil lubricationGearbox bearings are generally oillubricated when the gears are to be oil lubricated and it is simpler to use asingle lubricant. The use of oil lubrica-tion for bearings has the followingadvantages:

● oil can remove heat when bearingsoperate at high speeds and hightemperatures;

● at very slow speeds and underheavy loads, oil penetrates to thebearing surfaces more easily thangrease;

● less maintenance is required inrespect of supplying oil to the bear-ing position than for grease lubrica-tion, so that operational reliability isenhanced;

● the intervals between oil changesare longer than the grease relubri-cation intervals, particularly formedium and large-sized bearings;

● changing oil is simpler than chan-ging grease.

Lubricating oilsThe following lubricant propertiesshould be considered when selectingthe oil.

ViscosityPreferably the viscosity of the oil νshould be greater than the required oilviscosity ν1, both viscosities being atthe bearing operating temperature(see under adjusted rating life in theSKF General Catalogue). When deter-mining the appropriate viscosity for thedifferent bearing requirements (speeds,temperatures etc.) in a gearbox, aswell as for gear lubrication, it is advis-able, if κ values < 1 are found for someof the positions, to err on the side ofhigher viscosity for the compromisesolution.

The intention is to improve the lub-rication conditions for the heavily load-ed bearings rotating at slow speed atthe expense of generating more friction,because of the higher viscosity, in thehigh speed bearings. The operatingviscosity and lubricant film formationcan be influenced by selecting an oil of the appropriate viscosity class, butalso by cooling.

Load carrying ability, wear protectionEP oils (lubricating oils CLP to DIN 51 517) are preferred for the lubricationof spur, bevel and planetary gearboxes.As some EP additives have a detri-mental effect on bearing life and EPoils also have varying load carryingability and wear protection properties,it is advisable to contact the lubricantsupplier for recommendations regard-ing the particular application.

Protection against corrosion, behaviour in presence of waterThe rust inhibiting lubricating oils CLPto DIN 51 517 provide enhanced pro-tection against corrosion as they havegood surface wetting properties. Freewater in the rolling contact is extremelydamaging even when the actualamounts are very small. This is par-ticularly true of bearings where theproportion of sliding is high (e.g. hea-vily loaded spherical roller bearingsandall bearings subjected to centrifugalforce). It is thus desirable that the oilwill emulsify the small quantities ofwater which cannot be avoided.

Behaviour in presence of airAt moderate to high speeds there is adanger of air becoming mixed into the oil (foaming). Gear oils should becapable of expelling dispersed air andshould not be able to form a stablefoam.

AgeingLubricating oils oxidise as a result ofexternal influences, mainly high tem-peratures and exposure to air. Thisoxidation is catalysed (accelerated) inthe presence of some metals such ascopper or iron (wear particles). Anti-oxidant additives will slow down theprocess. Synthetic lubricating oils aremore resistant to oxidation than min-eral oils, but are not always as good in respect of lubricant film formation.Synthetic oils are used for worm gearsbecause of lower friction, and for gearswhich are to be used in a wide rangeof temperatures, e.g. wind turbinegears.

95

5 Lubrication and maintenanceOil lubrication

the casing walls is collected. The feedto the bearings should be designed to lead the oil through the bearingsbefore it flows back to the sump. If thefeed is on the seal or cover side, thenthe drainage should be laterally displaced and should be positionedsufficiently high so that the oil mustpass through the bearings but at thesame time, any surplus oil can run offwithout impinging on the seals. Thisalso supports oil circulation andexchange at the bearing position onthe cover side, thus improving cooling(➔ fig ).

If there is a risk that insufficient oilwill be caught by the oil pockets, the oilsupply can be improved by providingbaffle plates or wipers.

Bearings with asymmetrical crosssection which dip into oil have a pump-ing action by virtue of their design, and this can contribute to cooling.Appropriate feed and return ductsshould be provided.

1

Oil supply andreturn ducts for oilbath lubrication

Oil lubrication methodsWhen selecting the method of lubrica-tion the first aim should be to ensure areliable supply of lubricant to the bear-ings. The oil mist inside a gearbox isnot sufficient as bearings in moderngearboxes are heavily loaded andunder conditions of lubricant starvationwill wear and fatigue prematurely. Themost used methods are described inthe following.

Oil bath lubricationThis method is commonly used forgears operating at peripheral speedsof up to 15 m/s. The oil level shouldreach the centre of the lowest rollingelement. Greater depths mean lossesbecause of churning and higher friction.This is often accepted for small andmedium-sized vertical gears (for oscil-lation and agitation, and submergedunits) where the bearings may be fullysubmerged.

Bearings which are arranged abovethe surface of the oil must be suppliedwith oil which is captured by oil pocketsor grooves where the oil running down

96


5

Circulating oilCirculating oil lubrication should beconsidered above all when

● circulating oil is to be used for thegears,

● the oil is to be used for heat re-moval,

● speeds are high to prevent rapidageing of the oil,

● oil bath lubrication will not provideenough oil for the bearings, e.g. onvertical or inclined shafts,

● very large quantities of oil are re-quired for oil bath lubrication be-cause of the size of the gearbox, or

● the oil is to be continuously ‘‘fresh-ened” by filtration or centrifuging.

When designing for oil circulation thefollowing points should be remem-bered.

● To guarantee that the bearings arelubricated right at the start, the oilsupply leads must be dimensionedto provide oil even when the gear-box is first started up. There is other-wise a risk that oil will only arrive atpositions where the feed cross sec-tion is larger (e.g. for the gears).

● To prevent the oil nozzles from be-coming blocked they should have anopening diameter of at least 1,5 mm.Where oil pressures are high a suit-able throttle length can be used tolimit the oil flow. The throttle shouldbe positioned immediately in front ofeach bearing, so that larger andthus more reliable oil lead diameterscan be used with high oil pressures.

● Bearings operating at high speedsproduce turbulence which rejectsthe oil. Care must be taken to seethat the oil can actually enter thebearing at the feed side. Double rowbearings are usually best lubricatedvia the lubrication holes in the outerring (or paired single row taper rollerbearings through the holes in theintermediate ring). For single rowbearings the oil should preferably be supplied at the cover side.

Oil jet lubricationAt very high speeds (n × dm > 106) oiljet lubrication must be used. As shownin fig , the oil should be injected inthe gap between inner ring and cageat high speed (v ≈ 15 m/s). Rejected oil must be able to run off between thebearings so that heat can be removedwithout excessive losses.

2

97


Oil jet lubrication

Machine: ............................................................. Oil:........................................................................Type: ................................................................... Oil quantity in system: .......................................No.: ...................................................................... Sample taken, date:............................................Location: ............................................................ Sample taken by: ................................................

Property or Test method Unit Analysis Data forguideline value (Standard) result for new oil

used oil

Colour, appearance Visual inspection – .................. ..................

Smell – – .................. ..................

Density at 15 °C DIN 51 757 kg/m3 .................. ..................

Kinematic viscosity DIN 51 562 mm2/sat 40 °C .................. ..................at 80 °C .................. ..................at 100 °C .................. ..................

Acid number DIN 51 588, Part 1 mg KOH/g .................. ..................

Water content ISO 3733 % wt/wt .................. ..................

Solid contaminants e.g. IR % wt/wt .................. ..................> 3 µm analysis(quantity + type) DIN 51 451

Four ball test DIN 51 350, Part 4 N .................. ..................

Special test(s): ...............................................................................................................................................

Remarks: ...............................................................................................................................................

Characteristic Deviation from new oilAs new slight moderate large very large

Ageing

Contamination

Recommended action: .........................................................................................................................

............................................ ................................................ ..............................................................Test date Test carried out at Tested by (Signature)

Monitoring lubricationLubricant supply and lubricant qualityshould be checked. To check the lub-ricant supply, simple means are avail-able, e.g. a dip stick for oil bath lubrica-tion. For circulating oil lubrication, onthe other hand, complex systems arerequired to check the oil pressure, flowrate and temperature at each lubrica-tion position, and include an alarmsystem. When choosing the monitoringarrangements lubricant supply relibilityshould be weighed against the costswhich would occur in the event of a

MaintenanceGearbox bearing maintenance consistsbasically of monitoring the operatingconditions in the gearbox and of monit-oring the condition of the bearingsthemselves. This preventive mainten-ance should enable early identificationof any malfunction so that remedialaction can be taken. Such action shouldeither prevent premature ending of thebearing service life or, at least, enablebearing replacement to be planned sothat downtime costs can be minimised.

98

2

5 Lubrication and maintenanceMaintenance

For for analysis ofused oil

Table

5

blockage. Oil quality can be monitoredby measuring the temperature in theoil bath, in the return duct and in thebearings either continuously or atregular intervals. This allows the oper-ating viscosity to be evaluated. Addi-tionally, regular analysis of the used oil is recommended (according to thescheme shown in Table , forexample). The results should alwaysbe compared to a similar analysis ofthe fresh oil.

Monitoring loadThe power consumption of the drive issometimes used as a measure of theload, but this is not suitable for monit-oring bearing loads, as the peak loadsare very much smoothed in the record-ing. Better information is obtained bymeasuring torque and measuringstress at the root of the gear teeth. A reliable bearing load measurementcan only be obtained by using specialforce measuring bearings equippedwith strain gauges. As this method isvery expensive, it is generally onlyused for new developments or duringdamage analysis.

Monitoring temperatureAn indication of incipient bearingdamage will be given quite late by thetemperature, and at low speeds theremay be no indication at all. Therefore,measuring bearing temperature is onlyappropriate for condition monitoring ofbearings at high speeds, and then onlyas an indication of trends. To be of anyuse, the temperature should preferablybe measured directly on the bearingrings.

Temperature measurements of bear-ings, gearbox and oil are very suitablefor monitoring the operating viscosityof the oil. This allows important deduc-tions to be made with respect to theoperating conditions.

Monitoring wearUnder favourable operating conditions(adequate lubricant film thickness andclean lubricant) bearings will operatepractically without wear. Where thereis a clear indication that particles ofbearing steel are among the wear par-ticles the conclusion is that a bearinghas already become damaged. It is

2

99


then recommended that the gearboxbe inspected to determine the sourceof the wear and to take remedial actionto prevent further damage.

Wear particle analysis also enablesgear wear and seal efficiency to bemonitored.

Monitoring vibrationsBearings in operation generate slightnoise even when in perfect condition.This running noise could be listened toby holding a wooden stick to the hous-ing and to the ear. In the past this wasone of the most reliable monitoringmethods in spite of human failingssuch as limited frequency spectrum,subjective judgements and inability torelate frequencies heard to causes.With the methods and equipment avail-able today diagnoses can be madeand condition monitoring is effective.Suitable proven procedures are:

● comparative measurements on similar gearboxes under the sameoperating conditions, allowing dif-ferences to be observed, and/or

● trend measurements on one gear-box at given intervals, again allow-ing differences to be noticed.

SKF has developed special measuringtechniques as well as the requisiteequipment allowing a broad spectrumof vibrations to be monitored andmaking it possible to analyse the typeand magnitude of incipient damage ina bearing. The more important items of equipment and associated softwareare described in the following.

SKF VIB PenThis very handy vibration measuringprobe (dimensions 150 × 20 × 18 mm;mass 80 g) can measure vibrationvelocities of 0,1 to 99,9 mm/s in thefre-quency range 10 to1 000 Hz. It isposs-ible to determine whether themachine vibrations are in the rangeallowed according to ISO 3945.Bearing dam-age can only be identifi-ed when it is in an advanced stageusing this method. However, as inad-missible vibrations will considerablyshorten bearing life, the VIB Pen is asimple and reliable instrument formaintenance personnel to monitoroperating conditions.

SKF MultilogThis is a system for plant monitoringwith permanently installed sensors andis more powerful than the SKF Microlog.It can be used for the continuous monitoring of rolling bearings andmachines.

In practice, the SEE method indic-ates incipient bearing damage earlierand more clearly than other methods.This is particularly true when thedamage consists of micro cracksand/or cold welding (lubricant starva-tion) in the rolling contact. Because ofthe early warning, the user has time toplan bearing replacement.

SKF SEE PenThe SEE Pen measures differences invibration acceleration with time in thefrequency range 250 to 350 Hz. Thesignals in the high frequency bandwhich are measured, evaluated andrecorded using the SEE (SpectralEmitted Energy) method are only pro-duced by ‘‘damaged” bearings. Theindications may be for lubricant starva-tion, contamination or actual bearingdamage. Thus the SEE Pen is an idealcomplement to the VIB Pen (both havethe same dimensions) to give simpleand reliable bearing condition monit-oring. Here too, trend measurementsgive the optimum evaluation.

SKF PicologThis compact, breast-pocket sizeapparatus combines the measuringcapabilities of the VIB and SEE Pensand can also be used for ‘‘enveloping”.The peaks of the enveloped bearingnoise are evaluated. The distance be-tween peaks enables the bearing com-ponent which is damaged to be identi-fied. Up to 500 recorded measure-ments and alarm levels can be storedand downloaded on to a PC. Evalu-ations can be made using PRISM2 Jr.software. The Picolog is an excellenttool for bearing condition monitoring.

SKF MicrologThis portable equipment (mass 2 kg)can be used for frequency analysisand gives optimum evaluation in thelow and high-frequency range (SEE).The Microlog is a powerful data logwith a display panel. The PRISM2

software permits a variety of evalua-tion methods to be used, e.g.waterfalldiagrams, storage of critical frequen-cies, determination of alarm levels etc.The Microlog can be used with hand-held sensors or with permanentlyinstalled sensors. As it records electric-al signals, it can be used to measurenot only vibration velocities and accel-erations but also distances, pressuresand temperatures.

100


Photograph (fromleft to right)SKF Thermo Pen,SKF Picolog, SKFSEE Pen, SKF VIBPen (upper), SKFTachometer(lower), SKF OilCheck, SKFStethoskop, SKFMicrolog

e

-

-

5

101


6

The selection of fits is dealt with indetail in the SKF General Catalogue.The following recommendations complement the catalogue information,giving the usual, proven tolerances forhigh-performance gearboxes for themost common case, i.e. rotating innerring load and stationary outer ring load(➔ Table ).1

The recommendations given inTables and are for specialcases which differ from the above, butwhich are typical of certain types ofgear.

32

103

6 Recommended fits

Recommended fits

The rings of rolling bearings deform elasticallyunder load and adapt themselves to their seat-ings. To be able to fully exploit the load carry-ing capacity and accuracy of the bearings, thebearing rings must be supported with sufficientfirmness and accuracy by the associated com-ponents. Where the load rotates with respect tothe ring, the ring should have an interference fiton or in its seating (shaft, housing or gear). Thisprevents a loosening of the bearing fit and thering will not ‘‘wander” under load. Fretting cor-rosion will also be prevented. It is not possibleto provide a sufficiently tight fit for the ringsimply by clamping it axially.

Bearing type Shaft tolerances Housing tolerances(for solid steel shafts and rotating inner ring load) (for steel, spheroidal graphite or

grey cast iron and stationary outer ring load)

Shaft diameter (mm) Housing bore diameter (mm) Bearing≤18 (18) (40) (100) (140) (200) (280) >500 ≤300 (300) >500 arrangement

to to to to to to to40 100 140 200 280 500 500

Deep groove ball bearings j5 k5 k5 k6 k6 m6 m6 m6 J6 J6 H7 Locating(for light loads P ≤ 0,06 C) G6 G7 F7 Non-locating

Angular contact ball bearings

single row (adjusted j6 k6 k6 m6 m6 n6 p6 p6 J6 J6 H7 Cross locatedvia the outer ring)

double row, paired single row j5 k5 k5 m5 m5 m5 – – J6 J6 H7 Locating(series 32, 33, 70 BG, 72 BG, 73 BG)

double row (series 33 D) k5 k5 m5 m5 – – – – J6 J6 H7 Locating

Four-point contact ball bearings k5 k5 m5 m5 n6 – – – approx. 1 mm radial Thrust

clearance (locate to bearing prevent turning)

Cylindrical roller bearings k5 k5 m5 m5 n6 p6 p6 r6 J6 J6 H7 –(N, NU, NJ designs)

Spherical roller bearings k5 k5 m5 m5 n6 p6 p6 r6 J6 J6 H7 LocatingG6 G7 F7 Non-locating

Taper roller bearingssingle row (adjusted k6 k6 m6 m6 n6 p6 p6 – J6 J6 H7 Cross locatedvia the outer ring)

double row, paired single row k5 k5 m5 m5 n6 p6 p6 r6 J6 J6 H7 Locating

Thrust ball bearings h6 h6 h6 h6 h6 g6 g6 g6 G7 G7 F7 Thrust bearing

Spherical roller thrust bearings j6 (for all diameters) approx. 1 mm radial Thrust

clearance bearing

Form and position tolerances, surface roughness

Cylindricity IT5/2 (for all diameters)

Rectangularity IT5 (for all diameters)

Permissible surface 4 4 4 6,3 6,3 6,3 6,3 10 8 10 16roughness Rz (µm)

When shaft tolerances p6 and r6 are used, use of the oil injection method will ease dismounting

Recommendedfits, form and positiontolerances forgearbox bearings

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1

6 Recommended fits

Table

Case Housing toleranceHousing bore diameter (mm)< 300 (300) > 500

to500

Deep groove ball bearings and spherical roller bearings as non-locating bearings G7 F7 E8with rotating inner ring load and stationary outer ring load and a temperature differential > 10 °C from outer ring to housing (e.g. when heating via the shaft, high speed operation, very solid housings, low environmental temperatures)

Deep groove ball bearings and spherical roller bearings, cross located, withrotating inner ring load and stationary outer ring load

a) axial displacement of outer ring in housing required, e.g. with thermal G6 G7 F7expansion of shaft and axially stiff housing

b) axial displacement of outer ring not required, e.g. when thermal expansion J6 J6 H7 of shaft is compensated by elastic deformation of housing without overloading bearings

Cylindrical roller bearings of NUP design with rotating inner ring load and stationary outer ring load

a) locating bearing G6 G7 F7

b) non-locating bearing J6 J6 H7

Locating bearings and cylindrical roller bearings under oscillating outer ring JS6 JS6 JS7 load, e.g. when weight and tooth force act in different directions. Special steps have to be taken when mounting in one-piece (non-split) housings (e.g. heating the housing)

Shaft tolerances ➔ Table 1

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105

2

6 Recommended fits

Housingtolerancesfor special cases

Table

Bearing type Bearing arrangement Shaft tolerance Housing toleranceShaft diameter (mm) Housing bore diameter (mm)< 120 (120) (250) < 120 (120) > 250

to to to250 315 250

Deep groove ball Shifting gear j5 js6 k6 M61) M61) N61)

bearings (inner and outer rings rotate atsame speed)

Planetary gear, intermediate gear h5 h6 h6 M61) M61) M61)

(outer ring rotates, inner ring stationary)

Spherical roller bearings Planetary gear, intermediate gear h5 h6 h6 N6 P61) R61)

Cylindrical roller (outer ring rotates, inner ring stationary)bearings

Cylindrical roller Planetary gear, intermediate gear see Table 1 N61) P61) R61)

bearings (rotating inner and outer ring load)

Cylindrical roller bearings Planetary gear, intermediate gear h5 h6 h6 G62) F62) F62)

without outer ring (planetary gear rotates, inner ring stationary)

Cylindrical roller bearings Planetary gear, intermediate gear f62) e62) e62) N6 P6 R6without inner ring (outer ring rotates)

Needle roller and Planetary gear, intermediate gear g52) g52) – G62) G62) –cage assemblies

1) C3 internal clearance required

2) For raceways on the planetary pins and in gear hubs,

the deviation from circularity should be < 25 % of actual diameter tolerance;

the deviation from cylindricity should be < 50 % of actual diameter tolerance;

the surface roughness should be Ra ≤ 0,2 µm and Rz ≤ 1 µm;

hardness should be 58 to 64 HRC

and the case depth when finish machined should be Eht = 0,5 √Dw – 0,5 ≥ 0,3 mm, with Dw = rolling element diameter in mm

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3

6 Recommended fits

Tolerances forbearings mountedin gear hubs

Measuring adistance foradjustment oftaperroller bearings

Table

6

107

6 Recommended fits

7 Mounting and dismountingbearings

Adjustment of angular contact bearings . . . . . . . . .109

7

Rolling bearings are precision products whichmust be carefully handled when they are beingmounted if they are to perform properly. Equalcare must be taken when dismounting if thebearings are going to be re-used.

Mounting and dismountingbearings

Basically, there are three things toremember when mounting:

● cleanliness, to prevent damage tothe raceways by contamination andcorrosion;

● accuracy of all associated compon-ents, to avoid additional forces arising from deformations and toavoid imprecise running;

● the force used to mount and dis-mount should not be applied via therolling elements and cage; directblows should be avoided so thatindentations and initial damage tothe raceways are prevented.

The SKF General Catalogue containsmore detailed instructions regardingmounting and dismounting based onthe above requirements. A compre-hensive selection of SKF tools, equip-ment and maintenance products arepresented in publication 4100 “SKFBearing Maintenance Handbook”. SKFalso offers various training coursesand seminars for personnel involved inmounting and dismounting.

Adjustment of angularcontact bearingsWhen mounting angular contact bear-ings (angular contact ball bearings,taper roller bearings) in gearboxes,particular attention should be paid tothe adjustment of the bearings as thisdetermines not only the performanceof the bearings themselves but alsothe guidance of the shafts and conse-quently the load carrying ability of thegears. The calculation of the adjust-ment value is described in the section‘‘Dimensioning rolling bearings” (➔Section 4). The choice of adjustmentmethod depends on whether the bear-ings are to be adjusted to axial clear-ance or to preload.

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7 Mounting and dismounting bearingsAdjustment of angular contact bearings

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1B

1

1B

1A 1A

C

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a

x

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Adjustment of taper roller bearings arranged face-to-faceto axial clearanceFirst it is necessary to determine thezero clearance condition as accuratelyas possible. This is rather difficult fortaper roller bearings on horizontalshafts, as the weight of the shaft andgears displaces the outer rings axiallybecause of the taper angle, so that theclearance-free roller end/flange con-tact, which is decisive for the adjust-ment, is difficult to achieve. The pro-cedure described in the following iswell proven and is very much simplerand more reliable. A device is used toswing the gear shaft into the verticalposition for the adjustment (➔ fig ).

● Mount the inner rings on the shaft(take care that the rings abut theshoulders correctly).

● Push the outer rings over the rollerand cage assemblies.

● Place the shaft with bearings in thegearbox which should be horizont-ally positioned

● Mount the top of the casing.● Screw down the cover at one side of

the casing.● Tilt the casing so that the shaft is

supported via the bearing by thecover.

1

● Rotate the shaft by hand (if neces-sary by turning the input or outputshaft) and press the outer ring of theupper bearing downwards in its seat-ing until all the rollers in the bearingturn about their own axes. The bear-ing arrangement is now free of clear-ance.

● The requisite length of the spigot inthe cover is determined from a = x – s where s is the required axial clearance.

● Mount the finish machined coverwith shims (if necessary).

The weight of the shaft and gears actsas a measuring load on the lowerbearing. The upper bearing is free ofclearance as soon as all the rollersrotate about their own axes when theshaft is rotated.

A limited range of matched singlerow taper roller bearings (DF execu-tion) is available. The bearing pairs aresupplied with an appropriate interme-diate ring, so that adjustment is notrequired. The user can also matchsingle row taper roller bearings him-self; the requisite width of the interme-diate ring, taking into consideration thefit, is determined as follows.

● Mark the bearing components asshown in fig using an electric pen.

● Place bearing A on three gaugeblocks (➔ fig ).

● Apply the measuring load:300 N for bearings with outside dia-meter up to and including 240 mm500 N for bearings with outside dia-meter over 240 mm.

● Turn outer ring 1A by hand so thatthe rollers abut the flange of theinner ring 1A.

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2

110


Fig 1

Fig 2a = x – ss = requisite

axialclearance

Adjustment oftaper roller bearings arrangedface-to-face withaxial clearance

Marking of bearingcomponents

Bearing bore Shaft Maximum standard axial clearance ∆a before mountingdiameter toleranced Bearings of seriesover incl. 329 320 X 330 331 302,322 332 303,323 313 (X)

mm – mm

– 30 k5 – 0,120 – – 0,140 0,150 0,170 0,10030 40 k5 0,200 0,140 – 0,160 0,160 0,170 0,180 0,11040 50 m5 0,220 0,160 0,220 0,180 0,180 0,170 0,200 0,120

50 65 m5 0,250 0,180 0,240 0,200 0,200 0,190 0,220 0,14065 80 m5 0,270 0,200 0,290 0,240 0,220 0,220 0,260 0,17080 100 m5 0,310 0,230 0,390 0,270 0,270 0,260 0,300 0,170

100 120 m5 0,330 0,280 0,400 0,300 0,280 0,300 0,340 0,190120 140 m5 0,370 0,300 0,400 – 0,300 – 0,390 0,220140 160 n6 0,430 0,330 0,400 – 0,330 – 0,430 0,240

160 180 n6 0,430 0,370 – – 0,370 – 0,450 –180 190 n6 0,430 0,400 – – 0,400 – 0,500 –190 200 n6 0,450 0,400 – – 0,400 – 0,500 –

200 225 p6 0,500 0,450 – – 0,450 – 0,550 –225 250 p6 0,500 0,500 – – 0,500 – 0,600 –250 280 p6 0,600 0,550 – – 0,550 – – –

280 300 p6 0,700 0,600 – – 0,600 – – –300 340 p6 0,700 0,650 – – 0,650 – – –340 360 p6 0,750 0,750 – – – – – –

7

● Measure the standout FA at threepoints using the gauge blocks.

● Calculate the average value of FAfrom

FA1 + FA2 + FA3FA = (mm)3

● Repeat the above procedure forbearing B.

● Calculate the average value of FBfrom

FB1 + FB2 + FB3FB = (mm)3

● Determine the width of the interme-diate ring from

C = FA + FB + ∆a (mm)

where ∆a = maximum axial clearanceaccording to Table or for specialbearings, the maximum value of thespecial clearance.

The following tolerances apply to the width C of the intermediate ring:0/–0,04 mm for bearings with outsidediameter D ≤ 140 mm and 0/–0,06mm for berings with outside diameterD > 140 mm

The axial clearance values given takeinto account the clearance reductioncaused by the interference fit when theshaft tolerances (also given in Table

) are applied. These tolerances arerequired for rotating inner ring loadswhich are moderate to heavy. Theouter ring with its point load shouldhave a seating to tolerance J6 or H7.

1

1

111


Measuring the standout F

Maximum stand-ard axial clearanceof matched taperroller bearings

�F

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Table 1

Measuringload

Gaugeblock

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112

Principle offorce/pathmeasurement

Recorded force/path diagram forshaft/bearing/housing system


There will be virtually no reduction inclearance from any deformation of theouter ring. If looser fits are chosen,then the axial clearance value will beslightly larger when the bearing ismounted. If a tighter fit is used then itis advisable to check that the bearingwill not be axially preloaded.

Adjustment of taper roller bearings arranged face-to-faceto preload When adjusting bearings which are to have a preload it is necessary toachieve a certain preload force. If thepreload distance (path) method is tobe used, it is first necessary to meas-ure the force and displacement in themounted condition. This is the onlyway to be able to take housing resili-ence into account when determining theappropriate distance. Fig shows theprinciple of a force/distance measure-ment. Diagram shows the result ofsuch measurements. The characteris-tic curve has been extrapolated (brok-en line) for small loads because themeasurements are not sufficientlyaccurate under such light loads. Usingthe characteristic curve the desiredpreload can be set by fitting a shim orspacer ring.

The adjustment of taper roller bear-ings using the friction torque as a

1

4

basis or a collective method based ontolerances has not found acceptancein gearbox applications as there isexcessive scatter of the preload forcewhen these methods are used.However, the friction torque can beused indirectly for adjustment as willbe seen from the following.

● The friction torque of the two bear-ings which are to be adjustedagainst each other is measured in a rig for a given preload force and ata defined measuring speed andrecorded.

● After mounting the bearings in thegearbox, the preload force is appliedby inserting shims until the recordedfriction torque is obtained. Thespeed and lubrication conditionswhen the torque is measured mustbe the same as when the originalrecorded measurements were made.

This method is advantageous particu-larly when large numbers of bearingsare to be adjusted if it is easier to measure torque than force at theassembly position.

Adjustment of taper roller bearings arranged back-to-back To ensure sufficient accuracy, thebearing rings should always be mount-

Fig 4 Diagram 1

Axial preload path s

Axial load

Axial force

��Z

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7

ed against a fixed abutment face whichis at right angles to the shaft axis. Fortaper roller bearings arranged back-to-back, therefore, a shim (spacer ring) isinserted between one of the two bear-ings and a shaft shoulder. The follow-ing procedure allows the shim to be fitted without having to mount and dis-mount bearing V; this would be ratherdifficult because of the requisite inter-ference fit of the bearing on the shaft.

● Mount outer ring of bearing H in thecasing.

● Mount the inner ring with roller andcage assembly of bearing H on thepinion shaft and introduce the shaftinto the casing.

● Rest the shaft with bearing H andcasing in a vertical position on the face of the pinion (➔ fig ).

● Turn the casing by hand so that therollers of bearing H abut the innerring flange.

● Measure the standout X at threepoints (dial gauge).

● Calculate average standout X.

X1 + X2 + X3X = (mm)3

● Lay bearing V on measuring plate(➔ fig ).

● Apply measuring load:300 N for bearings with outside dia-meter D up to and including 240 mm 500 N for bearings with outside dia-meter D > 240 mm.

● Rotate the outer ring of bearing V byhand so that the rollers abut theinner ring flange.

● Measure the standout Z at threepoints (dial gauge).

● Calculate average standout Z.

Z1 + Z2 + Z3Z = (mm)3

● Determine reduction in axial clear-ance ∆p taking into account the shaft fit from

∆p = ∆r 0,4 cotα

= ∆r 0,4 1,5 (mm)e

where∆p = reduction in axial clearance, mm∆r = radial interference, mme = bearing-related calculation fac-

tor, see bearing tables in SKFGeneral Catalogue

● Determine width B of shim for agiven adjustment of the pinion bear-ing arrangement using

B = X + Z + ∆p ± ∆a (mm)

where+ ∆a is the desired axial clearanceand − ∆a is the desired preload.

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Fig 5

Fig 6

Bearing H

Measuring load

Bearing V

Measuringplate

Spacer ring(shim)

Bearing V

Adjustment oftaper roller bearings arrangedback-to-back onpinion shaft

Measurement ofstandout Z

Rolling bearings in industrial gearboxes · 2020-06-16 · Industrial gearboxes – overview...

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Transcript of Rolling bearings in industrial gearboxes · 2020-06-16 · Industrial gearboxes – overview...