Rothe Erde GWL GB Támcsapágy

8/9/2019 Rothe Erde GWL GB Támcsapágy

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T K

Rothe Erde ®Slewing Bearings.

A ThyssenKruppTechnologies

companyRothe Erde


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With slewing bearings and quality ringsto global success.

Rothe Erde is the worldwideleading manufacturer ofslewing bearings (includingball and roller bearing slewingrings) and of seamless-rolledsteel and non-ferrous metalrings. In addition Rothe Erdeis a well known manufacturerof turntables.

Rothe Erde slewing bearings

are for decades state of theart technology and practice-proven all over the world, in awide variety of applications.

Rothe Erde manufactures slew-ing bearings up to 8,000 mmdiameter as monobloc systemsand segmental bearings inlarger dimensions.

Rothe Erde slewing bearingsare manufactured in Germanyand by Rothe Erde subsidiariesin Great Britain, Italy, Spain,

the United States, Brazil, Japanand China. The market pres-ence of Rothe Erde in all majorindustrialised countries is main-

tained by own distributors orsales agencies.

Total commitment to quality iscommon to Both, our domesticand foreign production facili-ties.All service and areas from

applications consulting todesign and manufacturing,including comprehensivecustomer service, are basedon the international DIN/ISO9001/2000 quality standardseries.

Examples for applications:• Antennas and Radar• Equipment• Areal Hydraulic Platforms• Aviation and Aerospace Units• Bogie Bearings for Vehicles• Communication Systems• Excavators• Harbour and Shipyard Cranes• Machine Tools• Mechanical Engineering

• Mobile Cranes• Offshore Technology

• Packaging and FillingMachines

• Rail Vehicles• Ship Deck Cranes• Stackers and Reclaimers• Steelmill Equipment• Telescopes• Tower Cranes• Tunnelling Machines• Water Treatment Equipment• Wind and Solar Energy Plants


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Plant Dortmund

Plant Lippstadt


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Rothe ErdeSlewing Bearings

Index.

Bearing design typesBasic information Pages 6 – 41

Standard series KD 210Single-row ball bearing slewing rings

Profile bearings Pages 43 – 55

Standard series KD 320Double-row ball bearing slewing rings

Double axial ball bearings Pages 57 – 83

Standard series KD 600Single-row ball bearing slewing ringsFour-point contact bearings Pages 85 – 121

Standard series RD 700Double-row slewing rings

Roller/ball combination bearings Pages 123 – 133

Standard series RD 800Single-row roller bearing slewing rings

Cross-roller bearings Pages 135 – 153

Standard series RD 900Three-row roller bearing slewing rings

Axial-roller bearings Pages 155 – 167


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Rothe Erde GmbHD-44137 Dortmund© 2004All rights reserved.


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Bearing design typesBasic information Pages 4 – 41

Bearing design types 6 – 7

Load transmission 8

Bearing selection 9 –10

Load factors for bearing selection 11

Example of a bearing selection 12 – 14

Service life 15

Example of a service lifecalculation 16 – 17

Fastening bolts 18 – 23

Loctite-586improvement in the frictional bond 24

Gearing 25

Pinion tip relief 26

Turning torque calculation 27

Raceway hardening 28

Quality assurance 29

Finite Elemente calculations 30-31

Companion structures 32

Measurement and machine handlingof the area surface,

admissible plan deviations andbending of the mounting structure 33 – 34

Operating conditionsand special requirements 35

Wear measurement 36 – 37

Installation, lubrification, maintenance 38 – 40

Drawing number composition 41


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Bearing design types.

Standard series KD 210

Single-row ball bearing slewing ringsProfile bearings


Double-row ball bearing slewing ringsDouble-axial ball bearings


Single-row ball bearing slewing ringsFour-point contact bearings

KD 210 standard bearing types 21 and 110are available

without gearwith external gearwith internal gear

Type 13 is suppliedwithout gear

Applications:e.g. vehicle construction,general mechanical engineering.

For bearings with similar dimensions astype 21, but with higher load capacities:see standard series KD 600, Pages 90 and 91.

KD 320 standard bearings are available

without gearwith external gearwith internal geardrawing position = mounting position

Applications:e.g. mechanical handling, mining and materialshandling.

KD 600 standard bearings are available


Applications:e.g. hoisting and mechanical handling, generalmechanical engineering.


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Bearing design types.

Standard series RD 700

Double-row slewing ringsRoller/ball combination bearings


Single-row roller bearing slewing ringsCross-roller bearings


Three-row roller bearing slewing ringsAxial-roller bearings

RD 700 standard bearings are available


Applications:e.g. mining and materials handling.



Applications:e.g. hoisting and mechanical handling, generalmechanical engineering.



Applications:e.g. hoisting, mechanical handling, miningand materials-handling, offshore technology,general mechanical engineering.


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8


Rothe Erde large-diameter antifriction bearingsare ready for installation, transmitting axial andradial forces simultaneously as well as theresulting tilting moments.

Fig. 1:Large antifriction bearings are generallyinstalled supported on the lower companionstructure.

Fig. 1

Fig. 2

Fig. 2:Suspended installations require an increased

number of fastening bolts. The bolt curvesshown in the diagrams do not apply in such acase. Calculation to be carried out by RE.

Load transmission.


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Rothe Erde Large-Diameter Bearings KD 100 QuestionnaireCompany: Department:

Name: Phone:

Address: Fax:

e-mail:

Country: Phone/Visit on:

Customer project:

Rothe Erde Inquiry-No.: Rothe Erde Order-No.:Application: Axis of rotation:

Horizontal vertical mutualBearing under:compression tension*

*Bolts under tension by axial loadsGear:

external free choice internal

as per annex B without

Movement:

Positioning only Intermittent rotation Continuous rotation

No. of revolutions [rpm]:

norm.: max.:

B e a r i n g l o a d sA B C

Magnitude and directionof loads and their distance(related to axis of rotation)

max. working load max. test loade.g. 25% overload

condition

Extreme loade.g. shocks or

out of operationAxial loadsparallel to axis of rotation [kN]

Radial loadsat right angle to axis of rotation(without gear loads)

[kN]

Resulting moment [kNm]

Tangential force per drive [kN]:

norm.: max.:

No. of drives:

Position: apart

Existing or chosen bearing per drawing No.:For continuous rotation, variable and B10 life requirements, please complete annex A.Annex A is enclosed:Remarks: (e.g. special working conditions / temperatures, required accuracies, bearing dimensions, inspection- or

certification requirements, material tests etc.)

Please fully complete this form. Incomplete information will delay our proposal.

Individual consultation required. Please call for appointment

Date Si nature07.05.2003 TA / Habener


Bearing selection.

The final and binding selection of a large-diameter antifriction bearing is principally madeby us.

Selection determines the correct dimensioningof bearing races, gearing and bolt connections.

We, therefore require that you complete ourKD 100 applications questionaire to provide uswith all necessary data to help in selection of the appropriate bearing.

The most important data for choosing the right

bearing are:

1. Applied loads2. Collective loads with respective time per-

centages3. Speed or number of movements and angle

per time unit together with the relatingcollective loads

4. Circumferential forces to be transmitted bythe gearing

5. Bearing diameter6. Other operating conditions.

Full completion of the KD 100 form will enableus to largely respect your requests and preparea technically adequate and economical bearingproposal.

Whenever possible, the completed KD 100form should be submitted to us during theplanning stage, but no later than the order pla-cement to allow for confirmation of the bearing.

Bearing selection by catalogueThis catalogue permits you to make an approxi-mate bearing selection to be used in your pro-

ject work.

The Rothe Erde bearings listed in this catalogueare allocated critical load curves for their staticload capacity as well as service life curves.

For defining the required bearing load capacity,the determined loads must be multiplied by the”load factors” indicated in Table 1 for thevarious application cases, except for types 13and 21 of the KD 210 type series. If no applica-tions are indicated, comparable factors have tobe used, depending on the mode of operation.


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Static load capacityThe determined loads must be multiplied witha factor fstat allocated to the application. Theproduct Fa’ or Mk’ must be below the staticcritical load curve of the selected bearing.

With regard to radial loads in load combinations

Fa = axial loadFr = radial loadMk = tilting moment

the reference loads for the “static” bearing

selection from the KD 210 and KD 600 typeseries are computed as follows according toI or II:

Load combination IFa’ = (Fa + 5,046 · F r) · f statMk’ = Mk · f stat

Load combination IIFa’ = (1,225 · F a + 2,676 · F r) · f statMk’ = 1,225 · Mk · f stat

I and II apply analogously to types 13 and 21,but without the factor fstat..

A bearing is statically suitable if one of the twoload combinations (I or II) is below the static

critical load curve.

The reference load for the RD 800 type seriesis:

Fa’ = (Fa + 2,05 · F r) · f statMk’ = Mk · f stat

The bearing is statically suitable if one of thetwo load combinations (I or II) is below thestatic critical load curve.

For the KD 320 and RD 700 type series, radialloads Fr ≤ 10 % of the axial load can be neglec-ted in selecting bearings by critical load curves.

If the radial load is Fr > 10 % of the axial load,the supporting angle must be taken intoaccount. The respective calculation will then bedone by us.

In the RD 900 type series, radial loads have noinfluence on the critical load curve.

Service lifeThe operating load multiplied by factor f L isanalogously transferred to the service lifecurve.

If the expected service live deviates from thevalue allocated to the factor, or if the service lifeis to be determined by the collective loads andtime units, see “Service life”, Pages 15–17.


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Load Factors forbearing selection.Except for Standard series KD 210, types 13 and 21

Table 1

Service Timein Full Load

Application f stat. f L Revolutions

Floating Crane (Cargo)Mobile Crane (Cargo)

Ship Deck Crane (Grab) 1.10 1.0 30,000

Welding PositionerTurntable (Permanent Rotation)

Mkrü ≤ 0.5 Mk 1.0 30,000

0.5 Mk ≤ Mkrü ≤ 0.8 Mk 1.15 45,000

Mkrü ≥ 0.8 Mk 1.25 60,000Bearing at base 1.25 1.0 30,000

Slewing Crane (Cargo)

Shipyard CraneRotatable Trolley (Cargo) 1.15 45,000

Shiploader/Ship Unloader

Steel Mill Crane 1.5 100,000Mobile Crane(Grab or heavyhandling service)

1.7 150,000Slewing Crane (Grab/Magnet) 1.45**

Rotatable Trolley(Grab/Magnet)

Bridge Crane(Grab/Magnet)

Floating Crane(Grab/Magnet)Main slewing gear of Bucket Wheel ExcavatorReclaimer 2.15 300,000

Stacker

Boom ConveyorOffshore Crane subject to special criteria

Railway Crane 1.10

Deck Crane (Cargo) 1.00Stacker

Boom Conveyor

Conveyor Waggon 1.10Cable Excavator/Dragline

Swing Shovel1.25

Hydraulic ExcavatorBearing from KD 320 series 1.25Other bearing typesHydraulic Excavator up to 1.5 m 3 1.45

exceeding 1.5 m 3 subject to special criteria

Ladle Car 1.75

Static rating principally requires taking into account the maximum occurring loads whichmust include additional loads and test loads.

Static safety factors (f stat. e.g. for erection loads, higher test loads etc.) must not be redu-ced without prior written approval from us for exceptional cases.

The “f L” values shown refer to a rating for max. operating load and have been obtainedfrom operating experience and tests. If a load spectrum with an assumed average load isused to obtain the required full load revolutions, the service time values must be increasedaccordingly.

For applications not listed in the chart, guidance values for similar operating conditionsand comparable applications may be used.

*) Tower Cranes with bearing at top:

Mkrü = restoring moment without loadMk = Moment at max. radius with load

**) For applications requiring a rating of f stat.= 1.45, multi-row designs should be givenpreference because of the high average loads and arduous operating conditions.

Note:In these applications, the operating conditions, particularly the operating time and theloads during the slewing process, vary considerably. Infrequent slewing motions, e.g.occasional positioning for certain jobs, may permit a rating on static criteria alone. On theother hand, continuous rotation or oscillating motions will require a rating on the basis of

service time criteria. Selections based on service time may also be required if the bearingcarries out relative movements, which is often the case with the discharge boom con-veyors in bucket wheel units.

F o r

t h e s e a p p l i c a t i o n s p

l e a s e

m i n d t h e a c c o m p a n y i n

g n o t e

.

T o w e r C r a n e

B e a r i n g

a t t o p *


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O


Example of a bearing selection.

The maximum load must be determined using the formulae listedopposite.

The loads thus determined must be multiplied by the load factors(see Table 1, Page 11) before the bearing can beselected.

The following factors will apply to the examples given:

Cargo duty: Load factor f stat. = 1.25

Grab duty: Load factor f stat. = 1.45

1 Lifting load at maximum radius

1.1) Max. working load including wind:Axial load Fa = Q1 + A + O + GRes. moment Mk = Q1· lmax +A·amax +W·r–O·o–G·g

1.2) Load incl. 25% test load without wind:Axiallast Fa = 1,25· Q 1 + A + O + GRes. Moment Mk = 1,25· Q 1· lmax +A·amax– O · o – G · g

2 Lifting load at minimum radius

2.1) Max. working load including wind:Axial load Fa = Q2 +A+O+GRes. moment Mk = Q2 · Imin+A·amin+ W · r – O · o – G · g

2.2) Load incl. 25% test load without wind:Axial load Fa = 1,25· Q 2 + A + O + GRes. moment Mk = 1,25·Q 2 – Imin + A · amin– O · o – G · g

Fig. 3

Portal crane


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Crane for cargo dutyat maximum radiusQ = 220 kN Imax = 23 mA = 75 kN amax = 11 mO = 450 kN o = 0.75 mG = 900 kN g = 3 mW = 27 kN r = 6.5 m

1) Maximum operating load including wind

Fa = Q +A + O+G= 220+75+450+ 900

Fa = 1645 kN–––––––––––––

Mk = Q · lmax + A · amax +W·r–O·o–G·g= 220· 23+ 75·11+27·6.5– 450·0.75–900·3Mk = 3023.0 kNm–––––––––––––––––

2) Load case incl. 25 % test load without wind

Fa = Q · 1.25 + A + O + G= 275+ 75+ 450+ 900

Fa = 1700 kN–––––––––––––

Mk = Q · 1.25 · I max + A · amax – O · o – G · g= 275· 23+ 75· 11–450· 0.75– 900· 3

Mk = 4112,5 kNm––––––––––––––––

3) Maximum operating load without windFa = 1645 kN–––––––––––––

Mk = Q · Imax + A · amax –O·o–G·g= 220· 23+ 75·11–450· 0,75– 900· 3

Mk = 2847,5 kNm––––––––––––––––

When selecting the bearing, load case 2) should be used for staticevaluation, and load case 3) for service life.

The static load capacity of the bearing, taking into account load factorf stat. = 1.25, is checked against the “static limiting load curve”, reference

load:

Load case 2) F a’ = 1700 kN · 1.25 = 2125 kNMk’ = 4112,5 kNm · 1.25 = 5140.6 kNm

A load factor of f L = 1.15 is used for a service life of 45 000 revolutionsunder full load,reference load:

Load case 3) F a’ = 1645 kN · 1.15 = 1891.7 kNMk’ = 2847.5 kNm · 1.15 = 3274.6 kNm

The number of bolts and strength class will be determined for the max.load without a factor:

Load case 2) F a = 1700 kNMk = 4112.5 kNm

Crane for grab dutyat maximum radiusQ = 180 kN Imax = 19 mA = 110 kN amax = 9 mO = 450 kN o = 0.75 mG = 900 kN g = 3 mW = 27 kN r = 6.5 m

1) Maximum operating load including wind

Fa = Q +A +O+G= 180+110+450+ 900

Fa = 1640 kN–––––––––––––

Mk = Q · lmax + A · amax +W·r–O·o–G·g= 180· 19+ 110·9+ 27·6.5– 450·0.75–900·3Mk = 1548 kNm––––––––– ––––––

2) Load case incl. 25 % test load without wind

Fa = Q · 1.25 + A + O + G= 225+ 110+ 450+ 900

Fa = 1685 kN–––––––––––––––

Mk = Q · 1.25 · I max + A · amax – O · o – G · g= 255· 19+ 110· 9– 450· 0.75– 900· 3

Mk = 2227.5 kNm––––––––––––––––

3) Maximum operating load without windFa = 1640 kN–––––––––––––

Mk = Q · Imax + A · amax –O·o–G·g= 180· 19+ 110·9– 450· 0.75–900· 3

Mk = 1372.5 kNm–––––––––––––––––

When selecting the bearing, load case 2) should be used for staticevaluation, and load case 3) for service life.

The static load capacity of the bearing, taking into account load factorf stat. = 1.45, is checked against the “static limiting load curve”, reference

load:

Load case 2) F a’ = 1685 kN · 1.45 = 2443.3 kNMk’ = 2227.5 kNm · 1.45 = 3230.0 kNm

A load factor of f L = 1.7 is used for an overall service life of 150 000revolutions under full load,reference load:

Load case 3) F a’ = 1640 kN · 1.7 = 2788 kNMk’ = 1372.5 kNm · 1.7 = 2333.3 kNm

Number of bolts and strength class will be determined for maximum loadwithout a factor:

Load case 2) F a = 1685 kNMk = 2227.5 kNm


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Reference loads for cargo service ( black), grab service (red)For the above-mentioned load cases, the following bearings may be selected:e.g. bearings acc. to drawing No. 011.35.2620 with external gear see Page 64, curve 14 ; grab operation requires service life evaluation

Static limiting load curves Service life curves · 30 ,000 revolutions

Static limiting load curves Service life curves · 30 ,000 revolutions

e.g. Bearings acc. to drawing No. 012.35.2690 with internal gear see Page 76, curve 40 ; for cargo servicee.g. Bearings acc. to drawing No. 012.35.2500 with internal gear see Page 76, curve 39 ; for grab service

0

500

1000

1500

2000

2500

3000

3500

4000

4500

5000

5500

6000

6500

7000

0 500 1000 1500 2000 2500 3000 3500 4000 4500 5000 5500

R e s . m o m e n t ( k N m )

Axial load (kN)1700 21251685 2443

2227

3230

4112

5141

reference load(bolts)

+ reference load

+

reference load (bolts)

+ reference load

+

14

13

12

0

400

800

1200

1600

2000

2400

2800

3200

4000

4400

4800

0 200 400 600 800 1000 1200 1400 1600 1800 2000 2200

Axial load (kN) 1892 2788

2333

3275

+reference load

+reference load

2400 2600 2800

3600

14

13

12

0

500

1000

1500

2000

2500

3000

3500

4000

4500

5000

5500

6000

6500

7000

0 400 800 1200 1600 2000 2400 2800 3200 3600 4000 4400


Axial load (kN)1700 212516852443

2227

3230

4112

5141


+ reference load

+


+ reference load

+

40

39

38

0

400

800

1200

1600

2000

2400

2800

3200

4000

4400

4800

0 400 800 1200 1600 2000


Axial load (kN) 1891 2788

2333

3275

+reference load

+reference load

2400 2800

3600

40

39

38


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Service life.

In antifriction bearing technology, theoreticallife is a well-known term. Due to a multitude of influential factors, nominal life acc. to DIN/ISO281 cannot in practice be taken as an absolutevalue but as a reference value and designguide. Not all bearings will reach their theoreti-cal life, although most will generally exceed it,often by several times.

Theoretical life criteria cannot be applieddirectly to large-diameter bearings, particularywith bearings performing intermittant slewingmotions or slow rotations.

In most applications the speed of rotation inthe race will be relatively low. Therefore, thesmooth operation and precise running of thebearing are not adversely influenced by wear orby the sporadic occurrence of pittings. It is,therefore, not customary to design large-dia-meter bearings destined for slewing or slowrotating motion on the basis of their theoreticallife. For better definition, the term “service life”was introduced. A bearing has reached its ser-vice life when torque resistance progressivelyincreases, or when wear phenomena haveprogressed so far that the function of the bea-ring is jeopardized (see Page 36).

Large diameter antifriction bearings are used inhighly diverse operating conditions. The modesof operation can be entirely different such asslewing over different angles, different opera-ting cycles, oscillating motions, or continuousrotation. Therefore, apart from static aspects,these dynamic influences have to be taken intoaccount.

The service life determined with the aid of thecurves shown is only valid for bearings carryingout oscillating motions or slow rotations. Thismethod is not applicable to:

– bearings for high radial forces,– bearings rotating at high speed,– bearings having to meet stringent precision

requirements.

In such cases Rothe Erde will carry out the cal-culations based on the load spectra includingthe speed of rotation and the percentage of

operating time.

We must clearly distinguish between the opera-ting hours of the equipment and the actualrotating or slewing time. The various loads

must be taken into account in the form of loadspectra and percentages of time. For servicelife considerations another influential factor notto be neglected is the slewing angle under loadand without load.

For an approximate determination of theservice live of a bearing, service life curves areshown next to the static limiting load diagrams.This does not apply to profile bearings types 13and 21 .

These service life curves are based on 30,000

revolutions under full load. They can also beemployed to determine the service life with dif-ferent load spectra or to select a bearing with aspecified service life.

Symbols used Unit

G U Service life expressed in revolutionsG1; G2; ...G i U Service life for load spectra 1; 2; ...iFa kN Axial loadMk kNm Tilting momentFao kN Axial load on the curveurveMko kNm Resulting tilting moment on the curveFa’ kN ”Reference load” determined with f LMk’ kNm ”Reference load” determined with f LFam kN Mean axial loadMkm kNm Mean tilting momentED1; ED2; ...EDi % Percentage of operating t imep Exponent

Ball bearings p = 3Roller bearings p = 10/3

f L =Fao = Mko Loads/curve ratio [1]––– ––––Fa Mk (Load factor)

G = (f L)p · 30 000 [2]


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The known load case F a and Mk is plotted on the respective diagram.The line from the zero point of the diagram through the given load caseintersects the curve of the bearing, in this example 011.35.2220..., atpoint (Fao; Mko).Using formulae [1] and [2] this will give

f L =Fao = Mko [1]–––– ––––Fa Mk

f L =1750 = 1.4; f L =

2800 = 1,4–––––– –––––––1250 2000

G = (f L)p · 30 000 [2]

G = 1,43 · 30000 = 82320 revolutions


Example of a service life calculation.

A bearing according to drawing No. 011.35.2220 issubjected to the following loadsFa = 1250 kNMk = 2000 kNmWhat is its expected service life?Bearing and diagram, see Page 64 and curve 13

Example 1

Conversion into time can be obtained via slewing or rotation angle pertime unit.

If several different load combinations can be defined, example 2 shouldbe followed to determine the operating life.

0

400

800

1200

1600

2000

2400

2800

3200

4000

4400

4800

0 200 400 600 800 1000

R e s

. m o m e n t ( k N m )

Axial load (kN)

3600

1200 1400 1600 1800 2000 2200 2400 2600 48002800 3000 3200 3400 3600 3800 4000 4200 4400 4600

Fa = 1250Mk = 2000

+

+

14

13

12

Fao = 1750Mko = 2800


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First the service life G1;2;...i is determined for each load case according tothe above diagram.

Then these values and the operating percentages given for the individualload cases are compiled into an overall service life using formula [3].

Gges = 100 [3]–––––––––––––––––––––––––––––––––––ED1 + ED2 + ...... + EDi–––– –––– ––––G1 G2 Gi

1) f L =2990 = 1.07 f L =

1480 = 1.06–––––– ––––––2800 1400

used in calculation f L = 1.06

2) f L =2800 = 1.40 f L =

1750 = 1.40–––––– –––––––2000 1250



The following load spectra are assumed for the bearing in example No. 1:

Example 2

3) f L =2660 = 1.77 f L =

1960 = 1.78–––––––– –––––––1500 1100


4) f L

= 2450 = 0.91 f L

= 2280 = 0.91–––––––– –––––––2700 2500


Summarization:G1 = 1.06 3 · 30000 = 35730 U; ED 1 = 10%G2 = 1.40 3 · 30000 = 82320 U; ED 2 = 25%G3 = 1.77 3 · 30000 = 166360 U; ED 3 = 60%G4 = 0.91 3 · 30000 = 22607 U; ED 4 = 5 %

Gges =100 = 85807revolutions––––––––––––––––––––––––––––––––––

10+

25+

60+

5––––– ––––– – ––––– –––––––35730 82320 166360 22607

0

400

800

1200

1600

2000

2400

2800

3200

4000

4400

4800

0 200 400 600 800 1000


Axial load (kN)

3600

14

13

12

1200 1400 1600 1800 2000 2200 2400 2600 48002800 3000 3200 3400 3600 3800 4000 4200 4400 4600

loadspectrum

1)2)3)4)

oper.time %

102560 5

Fa[kN]

1400125011002500

given loads loads on curveMk[kNm]

2800200015002700

Fao[kN]

1480175019602280

Mko[kNm]

2990280026602450

13

3)

4)

+

++

+

+

+

2)+

1)+


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M k

Fa

M k Fa

Bolt strenght class 8.8 8.8 10.9 10.9 12.9 12.9Rate of thread d/P < 9 ≥ 9 < 9 ≥ 9 < 9 ≥ 9St 37 1.0 · d 1.25 · dSt 50, C 45 N,46 Cr 2 N, 46 Cr 4 N 0.9 · d 1.0 · d 1.2 · d 1.4 · d

C 45 V, 46 Cr 4 V,42 CrMo 4 V 0.8 · d 0.9 · d 1.0 · d 1.1 · d


Fastening bolts.

Bolts

The critical load curves shown in the staticdiagrams relate to strength class 10.9 boltswith a clamping length of 5 · d and prestressedto 70 % of the yield point.

For bearings without indicated bolt curves, theentire load capacity range below the criticalload curves is covered by strength class 10.9bolts.

Analysis of the bolt curves must be based on

the maximum load without factors.

Our technical quotation will show the number of bolts, strength class and required prestress forthe bearing concerned and the loads indicated.Unless mentioned otherwise, thefollowing shall be assumed:

a) The axial load Fa is supported, i.e. the axialoperating force F A from the axial load doesnot exert any tensile stress on the bolts,see figures 4 and 5.

b) The bolts are equispaced around the holecircles.

c) The mating structures are meeting ourtechnical requirements, see Page 32.

d) Bearing and mating structures consists of steel.

e) No resin grouting provided.

f) The clamping length Ik is at least 5 · d forbearings with a fully annular cross sectionand at least 3 · d for profiled bearings, e.g.KD 210 type series.

g) There are at least six free threads availablein the loaded bolt section.

Where deviations in these conditions occur,prior consultation with us will be required.

In order to avoid prestress losses due tocreeping, the surface pressures shown inTable 3 (see Page 19) in the contact areasbetween bolt head and nut/material of theclamped parts should not be exeeded.

The selected product and strength classes of bolts and nuts must be guaranteed by themanufacturer to DIN/ISO standards.

Table 2: Minimum engagement for blind hole threads. Applies to medium tolerance class (6 H)Deviating tolerance classes require specific insertion lengths

d – Thread O.D. [mm]Bolts with metric ISO-thread(standard thread)

P – Pitch [mm]up to M 30 = d/P < 9

> M 30 = d/P ≥ 9

Fig. 4: Axial load “compressive”

Fig 5: Axial load “suspended”

The angularity between support and bolt/nutthread axles must be checked.

Pitch errors which will falsify the tighteningtorque and lead to lower bolt prestress forces,especially if the reach is > 1 · d, must be avoi-ded.


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Table 4 does not show any tightening torquesfor bolts > M 30, as experience has shown thattheir friction coefficients vary too much. Thesebolts should preferentially be tightened using ahydraulic tension cylinder, see Page 20.

The increased space requirement for bolt head,nut and tightening tool must be taken intoaccount as early as possible during the designphase. The thickness of the washer must beadapted to the bolt diameter. Observe plane-parallelism.

Approximate determination of surfacepressure underneath the bolt head or nutcontact area.

Conditions:

p = FM/0.9 ≤ pG–––––––Ap

FM – Mounting prestressing forcefor selected bolt [N]

Ap – Contact area under bolt heador nut [mm 2]

pG – Limiting surface pressure [N/mm2]

for the pressed parts

With hexagon head bolts, the reduced contactarea due to hole chamfer and seating platemust be taken into consideration.

Ap =π (d2w – d2h)––4for dh > da

dh – Bore diameterda – I.D. of head contact areadw – O.D. of head contact area

Tightening torque

The tightening torque is dependent on manyfactors, in particular however on the frictionvalue in the thread, as well as on the head res-pectively the nut contact area.

For a medium friction value of µG ≈ µK = 0.14(threads and contact surface is lightly oiled) thetightening torque MA to pre-load F M for thehydraulic torque wrench is indicated.Considering a divergence of ± 10 % theassembly torque M A’ has been determined for

the torque spanner.

Material pG Limiting surface pressureSt 37 260 N/mm 2

St 50, C 45 N, 46 Cr 2 N, 46 Cr 4 N 420 N/mm 2

C 45, profile rolled (KD 210) 700 N/mm2

C 45 V, 46 Cr 4 V, 42 CrMo 4 V 700 N/mm2

GG 25 800 N/mm2

If these surface pressures are exceeded, washers of respective sizes and strengths

must be provided.

Table 3: p G - Limiting surface pressure [N/mm 2] for the pressed parts

Table 4: Clamping forces and tightening torques for bolts with metric regulation threads DIN 13, for µ G ≈ µK = 0.14

Strength class to DlN/lSO 898 8.8 10.9 12.9

Yield limit Rp 0 ,2 N/mm2 640 for ≤ M 16 940 1100660 for > M 16

Metric Cross section Cross Clamping For hydr. M a’ =0.9 MD* Clamping For hydr. Ma’ =0.9 MD* Clamping For hydr. Ma’ =0.9 MD*ISO- of area section force and electric for spanner force and electric for spanner force and electric for spannerthread under stress of thread torque wrench torque wrench torque wrenchDIN 13 AS A3 FM MA MA’ FM MA MA’ FM MA MA’

mm2

mm2

N Nm Nm N Nm Nm N Nm NmM 12 84.3 76.2 38500 87 78 56000 130 117 66000 150 135M 14 115 105 53000 140 126 77000 205 184 90000 240 216M 16 157 144 72000 215 193 106000 310 279 124000 370 333M 18 193 175 91000 300 270 129000 430 387 151000 510 459M 20 245 225 117000 430 387 166000 620 558 194000 720 648M 22 303 282 146000 580 522 208000 830 747 243000 970 873M 24 353 324 168000 740 666 239000 1060 954 280000 1240 1116M 27 459 427 221000 1100 990 315000 1550 1395 370000 1850 1665M 30 561 519 270000 1500 1350 385000 2100 1890 450000 2500 2250M 33 694 647 335000 determined bolt through 480000 determined bolt through 560000 determined bolt throughM 36 817 759 395000 yield measurement 560000 yield measurement 660000 yield measurementM 39 976 913 475000 670000 790000M 42 1120 1045 542000 772000 904000M 45 1300 1224 635000 905000 1059000M 48 1470 1377 714000 1018000 1191000M 52 1760 1652 857000 1221000 1429000M 56 2030 1905 989000 1408000 1648000M 60 2360 2227 1156000 1647000 1927000

* = MA will change with deviating µG or µK


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Prestressing of fastening bolts by hydraulictension cylinder (Stretch method)

Tests and practical experience have shown timeand again that the calculated torques for bolts> M 30 or 11/4“ are not coinciding with theactual values with adequate precision.

The main influential factor for these differencesis thread friction in the bolt and nut contactarea, for which to a large extent only empiricalor estimated values are available. The effectivefriction force is determined by the friction

coefficient. In addition, a bolted connection willundergo settling which is predominantly causedby the smoothing out of surface irregularities.

As these factors are of considerable importancein calculating the tightening torque, they canlead to substantial bolt stress variations.

The following lists of factors influencing frictioncoefficient variations are toillustrate this uncertainty:

1) Thread friction is a function of:

• the roughness of the thread surface i.e. theway how the thread is produced, whethercut or rolled

• surface roughness, i.e. bright,phosphated or blackened;

• type of lubrication: dry, lightly oiled,heavily oiled;

• surface treatment of the mother thread;

• inserted thread length;

• possibly repeated tightening andloosening of the bolts.

2) Friction variations between head or nutcontact area are a function of:

• roughness of the contact surfaces;

• surface condition (dry, lubricated,painted);

• hardness differences between thecontact surfaces or material pairing;

• dimensional and angular deviationsbetween contact surfaces.

Hydraulic tension cylinders often requiremore space than torque spanners,because the entire device must bepositioned in the bolt axis.

We recommend to use bolt tension cylindersby GmbH, Auf’m Brinke 18,D-59872 Meschede, Germany. The followingtables show the tension forces and dimen-sions for single and multistage bolt tensioncylinders.

Torque spanners for bolts requiring torque-

type prestressing can also be obtained from.

Information available upon request.

The factors influencing the bolt stress canmost effectively be reduced by using hydrau-lic tension cylinders, especially in the case of larger-diameter bolts. Compared with theconventional torque method, the tensioncylinder offers the advantage of eliminatingthe additional torsional and bending stressesover the bolt cross section. Even more deci-sive is the lack of any type of friction whichallows to precisely determine the remainingbolt prestress by previous tests, taking intoaccount respective design parameters.

It is possible to calculate with a tighteningfactor of aA of 1.2 to 1.6, depending on thediameter/length ratio, and to use the yieldpoint of the bolt up to 90%. The prestress of the bolt tightened first is influenced by thetightening of the other bolts so that a mini-mum of two passes is required.

This will at the same time compensate forthe settling produced by the smoothing outof the unloaded mating surface during pre-stressing (thread and nut contact area).

Table 7 shows the theoretical tension forcesfor a selected bolt series.

Due to the non-parallelism between nut andcontact area and the thread tolerance,settling phenomena after the nut has beenthightened cannot be included by thismethod either. (It is recommended to requestthe bolts and nuts manufacturer to observestrict squareness tolerances.)

As the tension force applied in this methodwill not only cause elongation in the shaft butalso in the thread, it is important to choosethe correct thread series or thread tolerancesacc. to DIN 2510. An inadequate threadclearance may cause jamming of the nut,when the bolt is elongated. Taking into aac-count the nut height consultation with thebolts manufacturer is absolutely necessary.

The bolts should be long enough to leave atleast 1 · d above the nuts free for positioningthe tension cylinder.

The exact minimum lenght will depend onthe strength class of the bolts and the tensio-ning tool used. Washers should be largeenough to be pressed onto contact surfaceby the tension cylinder during bolt thigh-tening. Enlarged washers should be prefer-red over standardised washers. Consultationwith the tension cylinder supplier is neces-sary.


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Table 5: Single-stage bolt tension cylindersCat.- Tension Thread dia.

Type No. force in kN D1 D2 D3 D4 H1 H2 H3ES20 33.10040 200 M 20 x 2.5 42 52 65 6 19 94

ES24 33.10041 290 M 24 x 3 49 60 78 8 22 102

ES27 33.10042 380 M 27 x 3 55 67 86 10 25 108

ES30 33.10043 460 M 30 x 3.5 61 74 97 12 27 107

ES33 33.10044 570 M 33 x 3.5 66 80 105 14 29 115

ES36 33.10045 670 M 36 x 4 71 86 117 16 32 118

ES39 33.10046 800 M 39 x 4 77 94 124 15 34 128

ES42 33.10047 920 M 42 x 4.5 83 102 137 20 37 134

ES45 33.10048 1080 M 45 x 4.5 89 110 148 22 39 135

ES48 33.10049 1220 M 48 x 5 94 116 158 24 42 140ES52 33.10050 1450 M 52 x 5 102 126 166 28 46 151

ES56 33.10051 1680 M 56 x 5.5 106 135 181 31 49 158

ES60 33.10052 2010 M 60 x 5.5 114 142 199 34 52 167

ES64 33.10053 2210 M 64 x 6 120 150 206 37 55 172

ES68 33.10054 2600 M 68 x 6 124 155 228 40 58 180

ES72 33.10055 2880 M 72 x 6 130 168 238 44 62 186

ES80 33.10056 3610 M 80 x 6 142 188 267 50 68 202

ES90 33.10057 4650 M 90 x 6 160 210 300 58 77 220

ES100 33.10058 5830 M 100 x 6 178 237 340 66 85 240

Table 6: Multi-stage bolt tension cylinders

Cat.- Tension Thread dia.Type No. force in kN D1 D2 D3 H1 H2 H3

MS 20 33.10090 200 M 20 x 2.5 43.3 51 6 19 156

MS 24 33.10091 290 M 24 x 3 50 59 8 24 192

MS 27 33.10092 380 M 27 x 3 55 65 10 25 188

MS 30 33.10093 460 M 30 x 3.5 61 73 12 27 182

MS 33 33.10094 570 M 33 x 3.5 66 80 14 29 198

MS 36 33.10095 670 M 36 x 4 71 84 16 32 246

MS 39 33.10096 800 M 39 x 4 77 90 18 34 260MS 42 33.10097 920 M 42 x 4.5 83 98 20 37 253

MS 45 33.10098 1080 M 45 x 4.5 89 107 22 39 256

MS 48 33.10099 1220 M 48 x 5 94 112 24 42 265

MS 52 33.10100 1450 M 52 x 5 102 123 28 46 278

MS 56 33.10101 1680 M 56 x 5.5 106 129 31 49 288

MS 60 33.10102 2010 M 60 x 5.5 114 136 34 52 328

MS 64 33.10103 2210 M 64 x 6 120 150 37 55 330

MS 68 33.10104 2600 M 68 x 6 126 155 40 58 346

MS 72 33.10105 2880 M 72 x 6 130 164 44 62 358

MS 80 33.10106 3610 M 80 x 6 142 183 50 68 385

MS 90 33.10107 4650 M 90 x 6 160 203 58 77 418

MS 100 33.10108 5830 M 100 x 6 178 232 66 85 446


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Table 7: Bolt tension forces including tolerances for “large-clearance metric thread” – DIN 2510 – Sheet 2 – using hydraulic tension cylinders

Strength class to DlN/lSO 898 8.8 10.9Yield point Rp 0 ,2 N/mm2 660 940

Tolerances to DIN 2510Metric Tension Core Tension Theoretical Tension TheoreticalISO-Thread cross-section cross-section force at use of tension force at use of tensionDIN 13 yield point force yield point forceNominal dia. Pitch AS A3 F0,2 FM= 0,9 · F 0,2 F0,2 FM= 0,9 · F 0,2mm mm mm 2 mm2 N N N N16 2 148 133 94700 85200 139100 12520020 2.5 232 211 153000 137000 218000 19600024 3 335 305 221000 199000 315000 28300027 3 440 404 290000 261000 413000 37200030 3.5 537 492 354000 319000 504000 45400033 3.5 668 617 440000 396000 627000 56400036 4 786 723 518000 466000 738000 66400039 4 943 873 622000 559000 886000 79700042 4.5 1083 999 714000 642000 1018000 91600045 4.5 1265 1174 834000 750000 1189000 107000048 5 1426 1320 941000 846000 1340000 120600052 5 1707 1590 1126000 1013000 1604000 144300056 5.5 1971 1833 1300000 1170000 1852000 166600064 6 2599 2426 1715000 1543000 2443000 219800072 6 3372 3174 2225000 2002000 3169000 285200080 6 4245 4023 2801000 2520000 3990000 359100090 6 5479 5226 3616000 3254000 5150000 4635000

100 6 6858 6575 4526000 4073000 6446000 5801000

Determination of tightening torques forfastening bolts > M 30 or 1 1/4“

Tightening torque variations can beconsiderably reduced if the tightening torquefor bolts > M 30 or 1 1/4“ is not theoreticallydetermined but by the longitudinal elongationof the bolt.

This procedure can be easily performed if bothbolt ends are accessible in the bolted condition.Structures not allowing this will requie ed amodel test (Fig. 7, Page 24).

The equivalent clamping length must be simu-lated by identically dimensioned steel blocks.The condition of the surface underneath theturned part (bolt head or nut) should also beidentical with the object itself. Generally harde-ned and tempered washers are used, so thatthese conditions can be easily complied with.The influence of a different number of joints ishardly measurable and can therefore beneglected.

The expected standard variation must be takeninto account in the calculation of the tighteningtorque. The test is to assure that the minimumclamping force of these larger bolts is withinthe values assumed in the calculation.

For the bolt to be used, the elastic longitudinalelongation at 70 % prestress of the yield pointis determined theoretically via the elastic resili-ence of the bolt with respect to its clampinglength.

The bolt is prestressed until the previouslydetermined bolt elongation I is displayed onthe dial gauge. This torque is then read off thetorque spanner. To account for any variations,an average value from several measurementsshould be determined.

When using a torque spanner with wrenchsocket, the measuring caliper must be removedduring tightening of the nut, and the test boltsshould be provided with center bores at bothends in order to avoid errors due to incorrectpositioning of the measuring caliper, (Fig. 6,Page 23).

All fastening bolts on the bearing are thenprestressed to this tightening torque using thesame torque spanner as in the test. It must beassured that the actual bolts used and the testbolts come from the same production batch.


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After a certain operating time the bolt connec-tion must be rechecked for prestress andregtightened, if necessary. This is required tocompensate for any settling phenomena whichmight reduce the bolt prestress.

The required longitudinal elongation is theo-retically determined by the elastic resilienceof the bolt.

Symbols used

AN Nominal bolt cross section ......................................................................... mm 2

A3 Thread core cross section .......................................................................... mm 2

AS Bolt thread tension cross section ................................................................. mm 2

ES Young‘s modulus of the bolt .......................................................... 205 000 N/mm 2

FM Mounting tension force .............................................................................. NF0.2 Bolt force at minimum yield point ................................................................ NI1 Elastic bolt length..................................................................................... mmI2 Elastic thread length ................................................................................. mm

I Linear deformation at bolt tightening ............................................................ mmS Elastic resilience of the bolt ........................................................................ mm/N

Rp 0.2 Tension at yield point of bolt material ........................................................... N/mm 2

Ik Clamping length of the bolt ........................................................................ mmIGM Thread length I G and nut displacement I M · IGM= IG + IM used incalculating the resilience of the inserted thread portion ................................ mm

Fig. 6

Fig. 7

This gives

= I–––––E · A

S = K + 1 + 2 + GMinserted thread portion

head shaft not inserted thread portion

with IG = 0,5 d and I M = 0,4 d

for nuts ac. to DIN 934

S =0.4 d + I1 + I2 +––––––– ––––––– –––––––ES · AN ES · AN ES · A30.5·d + 0.4·d––––––– –––––––

ES · A3 ES · AN

The force allocated to the elastic longitudinalelongation is:

FM=1 · I [N]––

S

Determination of the prestressing forceusing 70% of yield limit relative to thetension cross section:

FM = 0.7 · Rp 0,2 · AS [N]

F0,2

= Rp 0,2

· AS

[N]

Rp 0,2 for strength class 8.8= 640 N/mm 2 for d ≤ 16;= 660 N/mm 2 for d > 16.

Rp 0,2 for strength class 10.9= 940 N/mm 2

Rp 0,2 for strength class 12.9= 1100 N/mm 2

Therefore:

I = FM · S [mm]


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Loctite-586Improvement in the frictional bond.

Bearing installation using Loctite-586

The roughness of the surface to be joinedshould not exceed a value of Rt 65(peak-to-valley height) since shear strength willdecrease at greater roughness values.

Theoretically, the quantity required for a layer of 0.1 mm is 100 ml/m 2. However, if the layer is tobe applied by hand, it is advisable to usedouble or triple this quantity, since dosage byhand cannot always be absolutely accurate.

The following points must be observed duringinstallation:

1) Cleaning of contact surfaces with acommercially available cleaning agent toremove any oil or grease

2) Inactive surfaces must be pretreated withthe T 747 activator. Loctite-586 must onlybe applied to the nonactivated surface. If both sides are active, or if Loctite is appliedonto the activator, premature curing mayresult(drying within a few minutes).

3) Loctite must be applied with a stiff brushonto one surface.

4) Spigot locations must not come intocontact with Loctite since this would renderlater dismantling difficult. They must becoated with a separating agent, e.g. wax orgrease.

5) Tightening of fastening bolts. Loctite willstart curing as soon as 2 hours after posi-tioning of the bearing. If it is not possible tofully tighten the bolts during this period,manual tightening will suffice as a prelimi-nary solution.

6) Through holes and tapped holes have to beprotected against Loctite.

Dismantling

As already mentioned, the Loctite joint willresist compressive and shear forces, but nottension. Therefore, separating the bearing fromits companion structure does not present anydifficulties.

When using Loctite, the best solution is toincorporate tapped holes for jacking screwsright at the design stage of the companionstructure. For large and heavy bearings and/ora horizontal axis of rotation, the use of jacking

screws is imperative, especially when themounting space is restricted.

To lift the bearing off, the jacking screws aretightened consecutively until the bearing worksitself free.

With smaller bearings and easily accessiblemounting space, it may suffice to carefully liftthe bearing at one side, e.g. by applying apinch bar at several points around the circum-ference.

Under no circumstances should the bearing besuspended from eye bolts and lifted off beforethe joint has been released in the mannerdescribed above.

Before reassembly, the surfaces are best clea-ned by means of a wire brush.


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Gearing.

Rothe Erde large-diameter bearings are in themajority of cases supplied with spur gears. Agear cut into one of the bearing rings offers theadvantage that an additional driving gear wheelis not required, which helps to reduce designwork and costs.

In the case of highly stressed gears, a tipradius should be provided on the pinion, and inthe case of tip relief, an additional radius at thetip edge will be necessary (see Page 27).

Mainly provided are bearings with corrected

gearing, addendum modification coefficientx = 0.5 see DIN 3994, 3995.

For gears subjected to high tooth flank stress,hardened gears have proven very satisfactory.Depending on module and ring diameter, thegear rings are subjected to spin hardening orindividual tooth induction-hardening, the latterpredominantly in the form of tooth contour har-dening. Both methods provide improved flankload carrying capacity as well as higher toothroot strength. Flank hardening with hardnessphase-out in the region of the root radii leavingthe root radius unhardened will reduce the loadcapacity at the root. Hardened gears willrequire an individual calculation.

We need to know the pinion data in order to beable to check the meshing geometry.

During the installation of the large-diameterbearing and the drive pinion, adequate back-lash must be assured.

The backlash is adjusted at 3 teeth marked ingreen and is to be at least 0.03 x module.

After final assembly of the equipment and aftertightening all of the fastening bolts, the back-lash must be checked using a feeler gauge or alead wire.

Fig. 8: Backlash


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Pinion tip relief.

Despite geometrically correct profiles and theo-retically adequate gears, meshing problemsmay still occur in highly stressed gears, e.g.“scuffing” or “chipping” at the dedendum flankof the wheel, as shown in Fig. 9.

LubricationThe three influential factors mentioned willresult in high peak loads acting on the tip edgeof the pinion, which can cause the lubricant filmto break.

The direct metallic contact will increase thechipping effect.

Occasional damage which has occurred in thepast, can now be prevented by providing a tiprelief at the pinion and a radius at the tip edgeof the pinion.

Tip relief has become a means of reducing theeffects of vibration (noises) in high-speed gearmechanisms.

Investigations have led us to specify pinionswith a tip edge radius of 0.1 – 0.15 timesmodule for applications with extreme loadconditions.

The radius must blend into the addendum flankwithout forming an edge.

Fig. 10

This phenomenon occurs primarily in gearswith hardened pinions where the tip edges of the pinion act as scrapers.

Various causes may be responsible.

BendingDynamic load peaks under high force applicati-ons, accelerations, braking actions or vibrationswill cause elastic deformations in the meshingteeth.

Pitch errorsManufacturing tolerances in gears cannot beprevented, especially pitch errors, which incombination with the bending effect can pro-duce negative influences.

Drive unitMost slewing drive units are mounted in anoverhung arrangement, and deflections of thepinion shaft are unavoidable. The high forceswill simultaneously produce elastic deformati-ons at the interface of the slewing drive andmounting structure. Such deformations mayalso lead to meshing problems.

Fig. 9

Ca = 0.01 · mh = 0.4 – 0.6 · mCa:h = 1: 40 – 1: 60

(based on fulltooth width)

an ca. 0.1 – 0.15 · m


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Turning torque calculation.

The calculation of the torque Mr, detailed belowis based on theoretical and empirical know-ledge. The torque is affected by the rollingfriction coefficient, the rolling elements,spacers, seals, load distribution and the load.

Some other factors affecting the torque are:– The out-of-flatness including the slope of

the upper and lower companion structure.– The grease filling and the type of grease.– The lubrication of the lip seals and the seal

preload.– The variation in the bearing‘s clearance

resulting from installation.

The torque calculated is, of course, subject tocertain fluctuations, which can be estimatedwith approx. ± 25 %.

1. Starting Torque M r

Ball bearing slewing rings

Mr = µ (4.4 · Mk + Fa · DL +––2

2.2 · F r · DL · 1.73) [kNm]

Roller bearing slewing rings

Mr = µ (4.1 · Mk + Fa · DL +––2

2.05 · F r · DL) [kNm]

2. Power of Inertia

PBeh. = Mr · · –1 [kNm · s–1]

=Mr · n [kW]––––––––––9.55 ·

Specially designed bearings with reducedtorque can be supplied. Please contact usregarding the applications for such bearings.

In order to assess the total moment necessaryfor rotating the bearing, the acceleration powerof all the individual masses must still be confi-gured as a product using the squared distanceof their centres of gravity from the axis of rota-tion. The strength of the wind, which maypossibly act upon the bearing, and anycomponent parts under slope must also betaken into account.

Symbols used

Fa = axial load [kN]Fr = radial load [kN]Mk = resulting tilting moment [kNm]DL = bearing race diameter [m]µ = friction coefficient

= angular velocity= π · n [s –1]––––

30n = number of bearing revolutions

per minute [min –1]= drive efficiency

Various friction coefficients

µ = 0.008 for Type KD 210 (Type 13and 21, normal bearings)

0.006 for Type KD 210 (Type 110)

0.004 for Type KD 320

0.006 for Type KD 600

0.003 for Type RD 700



For precision bearings, bearings withoutclearance and preloaded bearings, the turningtorque calculation has to be performed byRothe Erde.


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Raceway hardening.

The bearing types described here are providedwith induction-hardened raceways. This ensu-res good reproducibility of hardening specifica-tions and, therefore, consistent quality. Thehardening coils used have been adapted to thevarious raceway designs. They are configuredso as to guarantee the load capacities specifiedfor the respective rolling element sizes.

Our patented coil shape ensures a goodhardness pattern in the raceways and in thetransition radii in three-row roller bearings.

Fig. 11: Raceway of a supporting ring in adouble-row ball bearing slewing ring.

Fig. 14: Raceways in a single-row roller bearingslewing ring.

Fig. 12: Raceways of a nose ring in a double-rowball bearing slewing ring.

Fig. 13: Raceways in a single-row ball bearingslewing ring.

Fig. 15: Raceways of a nose ring in a three-rowroller bearing slewing ring.


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Quality assurance.

The Rothe Erde quality assurance system hasbeen approved by internationally accreditedagencies and surveying authorities in accor-dance with the latest DIN EN ISO 9001:2000quality requirements, Environmental protectionto DIN/ISO 14001 and Occupational safety toOHSAS 18001.

We must first determine whether the custo-mer‘s requirements or ideas can be translatedinto a product that will not only meet the designcriteria but will also provide a good service life.

When the requirements have been clearly defi-ned, the desired quality level is established incollaboration with the relevant departments andincorporated in the drawings, production plans,testing instructions etc.; this also includespacking, delivery and after sales service.

An effective quality control procedure monitorsand ensures the quality of the product. Basedon drawings, testing plans, etc. a basis checkis carried out on the parts by the operators of the machinery within their responsibility.Moreover, members of the Quality AssuranceDepartment carry out systematic randomchecks.

There is 100% inspection of all of theproduction processes affecting the functioningof the product.

Material testing, i.e. the determination of mechanical properties, full analyses, structuraltests, ultrasonic tests and crack tests, guaran-tees uniform material quality.

Should any deviations be found during qualitychecks, the quality assurance system preventssuch defective parts from remaining in themanufacturing process.

Upon completion, every large-diameter bearingis subjected to a functional and dimensionalcheck.

Regular computerised monitoring of the mea-suring equipment ensures that during the entiremanufacturing and quality assurance proces-ses, only inspected or calibrated measuringunits are used.

We expect our supplier to attach the sameimportance to the quality of their products aswe do to ours. The “goods inward” inspectionis supplemented by regular system audits of allof our suppliers. This ensures that subcontractsare granted only to those companies who haveproven the quality of their products and theirability to deliver to schedule.

Internal audits in our company guarantee thequality of the manufacturing process and thefunctional safety of the quality assurancesystem. The information which these auditsobtain and the data on quality which aregenerally stored on computer make for effectivequality control.

The increasing requirements are met in ourcompany by regular in-house training of ourstaff. This also serves to make our employeesaware of the important contribution each indivi-dual has to make to the standard of quality.

Fig. 16


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Finite elements calculations.

The bearing rings used for slewing bearingshave only a relatively large diameter in relationto their cross-sectional area. Consequently,their inherent stiffness is limited. For this rea-son, the supporting stiffness provided by thecompanion structures has a major influence onthe load bearing characteristics of such a sle-wing bearing.

To be able to exploit the optimisation potentialconsistently, an integral calculation by aid of the finite elements analysis is imperative. The-refore, an optimum design is a joint task invol-

ving the slewing bearing manufacturer and themachine manufacturer.

The machine manufacturers calculate normallythe adjoining companion structures by aid of finite elements models.

Defined interfaces can feed the information of the companion structures into the finite ele-ments model of the slewing bearing enablingthe calculation to consider the stiffnesses of thesuperstructure and of the undercarriage. Fullycentralized development of an expensive overallmodel is no longer necessary.

This avoids the problems involved with havingto interpret unfamiliar design documentation.A know how transfer does not take place.

The following part models are linked for the

purpose of the analysis:

• the upper companion structure from thecustomer

• the slewing bearing including fasteningbolts

• the lower companion structure from thecustomer.

The information of the part models can beeasily exchanged by e-mail.

Mobile harbour crane; divided into three partmodels (Fig. 17)

The special software is able to directly importthe files with the stiffnesses of the companionstructures as generated by the customer and toadd them to the model of the bearing (Fig. 18).Thus results a complete overall model with onecalculation method which considers all majorinfluence quantities simultaneously.

Mobile harbour crane; divided into three part models(Fig. 17)

Finite elements model of the lower companion structure

Finite elements model of the slewing bearing

Finite elements model of the upper companion structure


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Figure 18

Schematic of the calculation method designed by Rothe Erde (Fig. 5)

customer Rothe Erde

Condensed stiffnessmatrix of the upper

companion structure

Condensed stiffnessmatrix of the lower

companion structure

Stresses,deformations...

Rolling elementforces, additional

bolt stresses

Stresses,deformations...

Create Finite-Element-Modelof the upper companion structure

and calculate its stiffness

Calculate the innerunknown magnitudes

of the uppercompanion structure

Displacements androtations of the inter-

facenodes of the upper

companion structure

Assess the results of the part model

“slewing bearing”

Solve non-linearintegrated systemusing an iterative

algorithm

Join the stiffnesses of all three part modelsinto one integrated

system

Create Finite-Element-Modelof the slewing bearing and

calculate its stiffness

Calculate the innerunknown magnitudes

of the lowercompanion structure

Displacements androtations of the inter-

facenodes of the lower

companion structure

Create Finite-Element-Modelof the lower companion structure

and calculate its stiffness

Per e-mail Per e-mail

Per e-mailPer e-mail

Once the bearing calculation has been comple-ted, the customer receives an e-mail with filescontaining data on the displacements and rotati-ons of the interface nodes on the flanges of thecompanion structures. These data can be directlyimported from his finite elements programme andused to calculate the corresponding internalstresses and deformations to which the compa-nion structures are subject (Fig. 2). This offers aneconomical means of optimising prototypes inwhich design weaknesses can be rapidly identi-fied.

The developed calculation method offers the cus-tomer and the manufacturer an opportunity toparticipate in a long-term development partner-ship. The new method enables both, a highlyeconomic and a thorough analysis from themechanical point of view.

Experiments acknowledge that the use of thissystem allows a very precise calculation.This in turn greatly reduces the expenditure requi-red for prototype development.


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Companion structures.

Due to their specific load carrying capacity,Rothe Erde bearings can transmit very highloads even at relatively small diameters. Thebolts provided for mounting the bearing to itscompanion structure must be rated accordingly.

For economic reasons, the cross sections of thebearings are kept relatively low in relation totheir diameters. The bearings therefore dependon a rigid and distortion-resistent structurewhich to a large extent will prevent deformati-ons in the bearings under the operating loads,provided a positive bolt connection is used.

The formation of peaks in smaller sectors hasto be avoided, i.e. the curve must progressgradually, rising and falling just once in therange from 0° to 90° to 180°. Otherwise tightspots may develop in the raceway which lead tolocal overload.

Fig. 20 illustrates that the vertical support in thecompanion structures must be in the vicinity of the track diameter. This is in order to keep anydeflection of the support surfaces under maxi-mum operating load within the permissiblelimits.

Rothe Erde offers seamless rolled rings for sup-port structures in a multitude of cross sectionsand profiles, unmachined or machined tocustomer‘s drawings which, for instance forflange ring supports (e.g. angular mountingring, Fig. 19). provide decisive advantages:

– Distortion-resistant fastening of the large-diameter bearing,

– Optimum load transfer between antifrictionbearing and companion structure.

The contact surfaces for the bearing mustalways be flat to prevent the bearing frombecoming distorted when it is bolted down.Careful machining of the contact surfaces is,therefore, absolutely essential.

As a rule, bearings and their companion struc-tures should be connected by means of through-bolts.

Fig. 19

Fig. 20


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Measurement and machining of contact areas,admissible flatness deviations includingslope of the companion structure.

Generally it applies that the companion struc-tures for large antifriction bearings must notonly be distortion-free but the contact surfacesfor the mounting of the bearings must be as flatas possible.

Measurements of the contact areas

Before the installation of a large antifrictionbearing Rothe Erde recommend that the con-tact areas be measured by means of an optical

Fig. 21 Fig. 22

machine or a laser measuring system. If themeasured values are outside of the Rothe Erdetolerances (Table 8) Rothe Erde would recom-mend mechanical re-working. In some casesthe re-working of spacious mating structuresproduces difficulties. However as a remedy weoffer the use of portable processing machines(picture 22 + 23) (also for upper constructionsand overhead machining).Reputable companies can execute this workaccording to Rothe Erde tolerances as a localservice (a reference list of these companiescan be requested from Rothe Erde). The ideal

installation conditions for large antifrictionbearings are steel/steel contact.

Fig. 23


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Track diameter Out-of-flatness including slope per supportin mm surface “P” in mm

DLDouble row ball Single row ball Roller bearing

bearing slew ring bearing slewing slewing ringaxial ball bearing ring Combination bearing

4 point contact bearing*double 4 point contact baring

to 500 0.15 0.10 0.07to 1000 0.20 0.15 0.10to 1500 0.25 0.19 0.12to 2000 0.30 0.22 0.15to 2500 0.35 0.25 0.17to 4000 0.40 0.30 0.20to 6000 0.50 0.40 0.30to 8000 0.60 0.50 0.40

Table 8:Tolerated out-of-flatness including slope “P” of the machined support surface.

For special applications such as precision bearings with a high running accuracy and low bearing play,the values in Table 8 may not be used.If the admissible values are exceeded, consultation should take place with Rothe Erde.*) For normal bearing type 13 and normal bearing type 21 double values are certified.

Tolerated out-of-flatness including slope “P”of the machined contact surfaces for RotheErde Antifriction bearings.

The maximum admissible flatness deviationsare listed in Table 8 as reference values.

Fig. 24

d · π

Regarding the slope of the machined contactarea the table values refer to 100 mm contactwidth.


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Operating conditions and special requirements.

The data contained in this catalogue refer tooscillating motions or slow rotating movements.

It is, of course, possible to use large-diameterbearings for higher circumferential speed. Forsuch requirements it is necessary to carry outspecial checks on the raceways and gears andto adapt these to the operating conditionsif need be. Enquiries concerning such applicati-ons should include a description of the opera-ting conditions as well as the customer‘s requi-rements.

If the bearing is to be installed with its axis inthe horizontal position, we must be consultedbeforehand.

Operating temperatureStandard design bearings are suitable foroperating temperatures ranging from 248 K(– 25° C) to 333 K (+60° C). The various opera-ting temperatures require suitable lubricants,see information on Page 40.

For higher or lower operating temperaturesand/or temperature differences between theouter and inner rings we must be advised befo-rehand so that checks can be carried out.Requirements regarding the mechanical pro-perties of the ring material are of particularimportance. In many cases, for instance, aminimum notch impact strength will be requiredfor applications at sub-zero temperatures.

Classification/special conditionsQuite a number of applications as in offshoreinstallations and ship deck cranes require clas-sification. For this purpose, the respective clas-sification agencies have produced a catalogueof requirements and specify acceptance of thebearing in accordance with that document.

In order to be able to take such specificationsinto account when preparing our offer, we needto kow the specifications in detail beforehand.

SealsThe seals provided in the bearing gaps preventdust and small particles from directly enteringthe raceway and retain fresh lubricant in thebearing gaps. In this function, they have provensatisfactory under normal operating conditionsfor many years. With adequate relubrication,i.e. until a uniform collar of grease appearsaround the circumference of the bearing, theircorrect functioning will be assured.

In case of considerable dirt sediments appro-priate covers should be provided at the com-

panion structure.

As sealing materials are subject to ageing whenexposed to a number of environmental conditi-ons, seals require maintenance and, dependingon their condition, may have to be replaced.Controlling: every 6 months.

Applications in a heavily dust-laden atmos-phere, such as mechanical handling equipmentfor coal and ore, will require special seals. TheRD 700 type series is, for instance, equippedwith additional steel labyrinths at the upperbearing gaps, which have proven very satisfac-tory in open cast-mining. The steel labyrinthprotects the seal against mechanical damageand it can be bolted in segments so that thespace containing the grease can be cleaned, if necessary.

Bearings in ship deck and floating cranes areoften exposed to splash and surge water. Insuch cases we use a special seal as shown inFig. 25.

Installing this type of seal may increase theheight of the bearing.

For the above applications it is preferable touse bearings with internal gears where the gearis protected by the surrounding structure.

RacewaysPlastic spacers are inserted between the rollingelements in the raceways. The bearings aresupplied already greased. Penetration of aggressive materials into the raceways must beprevented on all accounts. Aggressive materi-als will alter the lubricating properties which willlead to corrosion in the raceways and damagethe plastic spacers.

Special designsApart from the standard bearing series shown,we offer bearings tailored to specific operating

conditions with regard to dimensions, runningaccuracies, bearing clearances and materials.

We also manufacture wire-race bearings. Thisbearing permits the use of non-ferrous metalrings and thus meets any special requirementsregarding minimum weight, resistance to corro-sion, etc.

PackingGenerally, large-diameter bearings will be wrap-ped in foil or a similar material for transport.The external bearing surfaces are protectedagainst corrosion by means of Tectyl 502 C(oily) and by filling the raceways with lithium-based grease.

The method of transport will determine the typeof packing used (e.g. pallets, crates).

Standard packing will provide adequate protec-tion for storage times of up to one year in enc-losed, temperature-controlled areas.

Upon request, other preservation and packingmethods can be provided for longer storagetimes (e.g. long-term packing for 5 yearsstorage).

Fig. 25


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Messsung

Measuringpoint 1below

Base measurement Test measurement Test measurement

boomcounterweight

Measurement


Wear measurement.

For assessing the condition of a bearing, werecommend that its normal wear rate is deter-mined. The wear present in the raceway systemshows itself by a change in the axial motion of the bearing. Depending on the individualconditions, wear can be determined either bymeasuring the tilting clearance or by depres-sion measurements.

Tilting clearance measurementFor equipment allowing both positive and nega-tive application of moment loads, a respectiveloading principle is shown in Fig. 31.

Mark the respective measuring positions on thecircumference while keeping the boom in aspecified position.

The measurement is performed between thelower mating structure and the bearing boltedto the superstructure (Fig. 35).

Record the base values obtained in tabular formand allocate them to the respective base mea-surements (Fig. 36).

The axial reduction measurement should berepeated every twelve months as a minimum,under identical conditions.

In case of heavy wear the time intervals bet-ween measurements should be shortened.

Axial reduction measurementIn cases where the combination of both positiveand negative loads are not possible, the follo-wing procedure should be applied. The loadingprinciple is shown in Fig. 33.

The first measurement should be performedwhen the equipment is put into operation inorder to obtain a base value for subsequentrepeat measurements.

Fig. 32: Three-row roller bearing slewing ring –

basic test setup for tilting clearance measurement

Fig. 33: Loading prinziple for axial reductionmeasurement

Fig. 34: Value recording for tilting clearance measurement

Fig. 31: Loading principle for tilting clearanc

measurement

The first measurement should be performedwhen the equipment is put into operation inorder to obtain a base value for subsequentrepeat measurements.

The measuring points should be markedaround the circumference while the boom iskept in a specified position.

The measurements are then taken between thelower mating structure and the bearing boltedto the superstructure (Fig. 32).

The measurements should be taken as close tothe bearing as possible in order to minimize theeffect of elastic deformations in the system.

The dial gauges should have an accuracy of 0.01 mm. Start with applying the maximumbackward moment and set the dial gauges tozero. Then apply a forward tilting moment, withload uptake, if necessary.

Turn the superstructure to the next position andrepeat the measurement procedure.

When all positions have been measured, recordthe base values obtained in tabular form(Fig. 34).

The measurements should be repeated everytwelve months as a minimum and under identi-cal conditions as the base measurement.

The difference between the values measuredand the base values represents the wear thathas occurred.

If the wear is found to have heavily increased,the time intervals between measurementsshould be shortened.

If the acceptable wear values (Tables 9, 10 and11) are exceeded, please consult Rothe Erde.


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Table 9: Double-row ball bearing slewing rings (standard series KD 320)

Track diameterup to mm

ball diameter mm

10001250150017502000225025002750300032503500

3750400045005000550060006500700075008000

1.81.9

1.81.92.0

1.92.02.12.2

1.92.02.12.22.3

2.02.12.22.32.42.5

2.12.22.32.42.52.62.72.8

2.52.62.72.82.93.03.13.23.33.43.5

3.6

2.82.93.03.13.23.33.43.53.63.73.8

3.94.04.2

3.43.53.63.73.83.94.04.14.24.3

4.44.54.74.95.15.3

3.63.73.83.94.04.14.24.34.44.5

4.64.74.95.15.35.55.75.9

4.04.14.24.34.44.54.64.7

4.85.05.25.45.55.75.96.16.36.5

18 20 22 25 30 35 40 45 50 60 70

Table 10: Single-row ball bearings (4-point bearings) double four-point contact bearings and standard series KD 210


ball diameter mm

10001250150017502000225025002750300032503500375040004500500055006000650070007500

8000

1.4 1.41.5

1.41.51.6

1.51.61.71.71.8

1.71.71.71.81.92.02.0

1.92.02.02.12.22.32.32.42.52.6

2.12.22.32.32.42.52.62.62.72.82.93.0

2.52.62.62.72.82.92.93.03.13.23.23.33.33.53.73.94.1

2.72.82.92.93.03.13.23.23.33.43.53.63.84.04.24.54.64.8

3.03.13.23.23.33.43.53.53.63.73.94.14.34.64.74.95.1

5.3

20 22 25 30 35 40 45 50 60 70

Table 11: Roller bearing slewing rings


roller diameter mm

permissible increase in bearing clearance mm

permissible increase in bearing clearance mm

400500630800

10001250150020002500315040005000600070008000

permissible increase in bearing clearance mm20

0.220.220.270.270.320.420.520.62

0.200.200.250.250.300.400.50

16 25

0.240.240.290.290.340.440.540.640.74

28

0.260.310.310.360.460.560.660.760.86

32

0.280.330.330.380.480.580.680.780.880.98

36

0.310.360.360.410.510.610.710.810.911.011.111.21

40

0.380.380.430.530.630.730.830.931.031.131.23

50

0.600.700.800.901.001.101.201.301.40

60

0.901.001.101.201.301.401.501.60

70

1.101.101.201.301.401.501.601.70

80

1.101.211.321.431.541.651.761.87

90

1.331.451.571.691.811.932.05

100

1.601.751.902.052.202.35

45

0,460,560,660,760,860,961,061,161,261,36

Maximum permissible increase of bearing clearances

Fig. 36: Value recording in axial reductionmeasurement

Fig. 35: Three-row roller slewing bearing ring – basictest setup for axial reduction measurement

Messsung Basemeasurement

Testmeasurement

Measuringpoint 1

Measuringpoint 2

Measuringpoint 3

Measuringpoint 4

TesterSignatureDate

Measurement

If the deviation from the base measurementexceeds the maximum values shown in Tables9, 10 and 11, please consult Rothe Erde.

Such increases of the bearing clearance are not permissible for special applications (consultation withRothe Erde), e.g. 50% of the above values for large-diameter antifriction bearings for roundabouts.


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Installation, Lubrication, Maintenance.

Transport and storageLarge-diameter antifriction bearings, like anyother machine part, require careful handling.They should always be transported and storedin the horizontal position; if they must be trans-ported vertically, they will require internal crossbracing. Impact loads, particulary in a radialdirection, must be avoided.

State of delivery(unless otherwise agreed):

Running system

lubricated with one of the quality greasesspecified.

Gear ringnot lubricated; treatment as for externalsurfaces.

External surfacesTectyl 502-C-EH. Washing with ageneral-purpose cleaner such asShell Callina 2306.

StorageApprox. 6 months in roofed storage areas.Approx. 12 months in enclosed, temperature-controlled areas. Longer storage periods willrequire special preservation.

After a long storage period the large antifrictionbearing could incur high rotational resistanceduring start up and running through suction of the seal rim. Careful lifting with a blunt articlearound the entire circumference and repeatedrotation of the large antifriction bearing over360° right and left this reduces the torque tonormal values.

InstallationA flat, grease- and oil-free rest is essential forthe upper and lower ring to seat solidly. RotheErde recommend examination of the bearingsurfaces with a levelling instrument or lasermachine. Only in exceptional cases for bea-rings Ø 2,5 m (with corresponding large crosssections) a feeler gauge should be used.

With the feeler gauge measuring method, it isrecommended that after the first measurementthe bearing is offset by 90° and the measure-ment repeated.

Machining of the contact area, as a final manu-facturing step is necessary (after welding).Welding beads, burrs and strong paint residuesand other uneven surfaces are to be removed.

Dependent upon the design type and the trackdiameter “DL” of the bearing, the maximumout-of-flatness including the slope “P” (mm) of the upper or lower contact surface should notexceed the values shown in Table 8.

Regarding the slope of the machined surfaces,the figures shown in the table refer to a supportwidth of 100 mm.

To avoid larger deviations and the occurrenceof peaks in smaller sectors, any deviation in therange of 0° – 90° – 180° may only rise or fallgradually. Prior to installation, the bearing

Thread/bolt Drilling Tightening torque Nm with bolts with a strength classdiameter diameter µ G ≈ µK = 0.14

for hydr. for Md for hydr. for Mdmm Md torque wrench key Md torque wrench key

DIN/ISO 8.8 8.8 10.9 10.9

273

M 12 14 87 78 130 117

M 14 16 140 126 205 184M 16 17.5 215 193 310 279

M 18 20 300 270 430 387M 20 22 430 387 620 558

M 24 26 740 666 1060 954

M 27 30 1100 990 1550 1395M 30 33 1500 1350 2100 1890

Grade 5 Grade 5 Grade 8 Grade 8

UNC 5/8“-11 18 200 180 286 260UNC 3/4“-10 21 352 320 506 460UNC 7/8“- 9 25 572 520 803 730UNC 1“- 8 27.5 855 770 1210 1100UNC11/8“- 7 32 1068 970 1716 1560UNC11/4“- 7 35 1507 1370 2410 2190

Grade 5 Grade 5 Grade 8 Grade 8

UNF 5/8“-18 18 230 210 320 290UNF 3/4“-16 21 396 360 560 510UNF 7/8“-14 25 638 580 902 820UNF 1“-12 27.5 946 860 1330 1210UNF11/8“-12 32 1210 1100 1936 1760UNF11/4“-12 35 1672 1520 2685 2440

Table 12

should be checked for smooth running by rota-ting the unbolted bearing around its axis, twice.

Should the permissible out-of-flatness, includ-ing the slope, be exceeded, we recommendthat the contact surfaces for the bearing bemachined.

For bearings of standard series KD 320, RD700 and RD 900 the bearing has to be installedas shown on the drawing.

Remove the protective coating from the bea-ring’s upper and lower support surfaces as wellas from the gear. No solvent should be allowedto come in contact with the seals and race-ways. Do not clean the gears if these aregreased.


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Diameter tolerancesMachined diameters with untoleranced drawingdimensions have the following tolerances:

≤ 315mm ± 1,6mm≤ 1000mm ± 2,5mm≤ 2000mm ± 3,5mm≤ 4000 mm ± 5,0 mm≤ 6300mm ± 7,0mm≤ 10000 mm ± 10,0 mm

Allgrease nipples must be easily accessible. If necessary, grease pipes should be provided toallow relubrication through all grease connec-

tions. We recommend automatic lubricatingsystems.

Hardness gapThe unhardened zone between the beginningand end of the hardened region of the racewayis identified by a punched-in letter “S” near thetype plate or filler plug at the inner or outer dia-meter of each bearing ring. On the gear ring,the hardness

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