Bolt Tightening Hadbook-SKF.pdf

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Bolt-tightening Hand

Linear Motion & Precis



The SKF Group

The SKF Group is an internationalindustrial corporation owned bySKF Sweden AB. Founded in 1907,it operates in 130 countries and hassome 40000 employees.

The company has over 80 manufactu-ring units throughout the world and anetwork of nearly 20000 distributorsand retailers. SKF is the world leaderin the rolling bearing business.

SKF Linear Motion &Precision Technologies

SKF Linear Motion & PrecisionTechnologies is an organization withinSKF which, as the name suggests,is dedicated to the manufacture, salesand service of linear motion products,high precision bearings and spindles.

It serves the market through itsorganization of 15 specialized salescompanies located in Europe,North America and Japan.

In addition to the services providedby these sales companies, productand application support is availableworldwide through the SKF internationalnetwork.

Specialized Linear Motion& Precison TechnologiesSales Companies

SKF Bearing salescompanies with LinearMotion sales staff

Production facilities



Introd

Traditional tightening me

Tightening with torqueTightening with he

Tightening by mechanical elo

Tightening with hydraulic bolt tens

PreseFeatures and

Measurement devices for hydraulic tig

Technical analysis of bolt-tigh

Comparison between torque w

and hydraulic bolt-tight

Tightening of an existing bolted as

Design of a new bolted as

Simultaneous hyd



Without a doubt, bolted assemblies are the most commonly

used joints in mechanics.These types of assemblies employ two basic elements:

• on the one hand, some kind of threaded component:- screws and nuts,- studs with nuts on one end,

- studs with nuts on both ends.These components are sometimes used with differentkinds of washers (Fig.1a below).

• on the other hand, some means for tightening.

These types of tightening means are the subjectof this Handbook.

In this document the word “bolt” will be used in a genericsense to cover all three of the types of screwingcomponents mentioned above.

Although bolted assemblies at first appear very simple,they cause several problems for design engineers,assemblers, and maintenance departments.

Rough-dimensioning methods are too often used atthe design stage, leading to substantial oversizing

of all the components of the asse

ensure assembly safety, quite theIn reality, the design of a bolted aa methodical and rigorous approacan lead to failures with often cosdisastrous consequences.

Many surveys show that failures mainly due to the fact that they w(analysis, drawing, calculation, chor implemented (tightening metho

The surveys also show that amonof assembly failure (overloading, manufacturing defects etc.) the massembly. Undertightening, overttightening alone cause 30% of al

Furthermore, in addition, 45% of

due to poor assembly (see Fig.1b

Introduction

Screw and nut Stud with nut on one end Stud with



Correct tightening of a bolt means making the best use ofthe bolt’s elastic properties.

To work well, a bolt must behave just like a spring.

In operation, the tightening process exerts an axial pre-loadtension on the bolt. This tension load is of courseequal and opposite to the compression force appliedon the assembled components. It can be referredto as the “tightening load” or “tension load”.

Depending on the application, the purpose of the tighteningload is multiple:

- ensure the rigidity of the whole assembly and make itcapable of supporting external loads due to traction,compression, bending moments and shear;

- prevent leakage at seals;

- avoid shear stresses on the bolts;

- resist spontaneous loosening effects;

- reduce the influence of dynamic loads on the fatigue lifeof the bolts (see Fig. 2 above).

Furthermore, all components (bolts and assembly parts)must perform these tasks while remaining below the yieldpoint of their respective materials.

Bolt-tightening is optimal when the bolt is properly

tightened: not too much, not too little! A bolt can fail just asoften - and even more so - when it is not tightened enough,as when it is over-tightened.

Controlling bolted assembl

It is fundamental to control the level of well as the accuracy of the tightening vrequired performance of the bolted ass

Complete control over the tightening outset of the design stage - ensures bolt’s mechanical properties of bolts,below, and page 6).

SealingTractionCompression

Shear stress Spontaneousloosening

Uncontr

tighteninfor overs

Controlle



Bolts are most often made of steel. Like mostmetals, steel is elastic, at least as long as the strain(elongation) does not exceed the “elastic limit”

beyond which permanent deformation occurs.Within the “elastic limit”, a metal part such as a boltfollows Hooke’s law, that is to say that the strain(elongation) is proportional to the stress (load),as shown on the graph opposite.

Any tightening method must ensure thatthe stress in the bolt never exceeds point “A”(the elastic limit or “yield point”), both duringthe tightening operation and when the assemblyis later exposed to efforts during operation.

Mechanical properties of bolts

When discussing structural mechanics, the following properties of materials will be considered:

E: Traction elastic modulus or Young’s modulus:

with F = traction force, S = cross-section, L = length, ∆t. = elongation

ν: Poisson’s ratio or lateral strain index:

for steel: 0.27/0.30for aluminium: 0.33/0.36for rubber: 0.49(the least compressible of all solids)for liquids: 0.5 (almost incompressible)for cork: 0.0x (very compressible)

K: Compressibility coefficient (by analogy with liquids):

for liquids: k g 0

E =F ∆L

=F.L

=σ.L

S L S.∆L ∆L

(∆L =σ

= F )L E S.E

for steel E: 200 000/210 000 MPa

ν =∆d ∆L

d L

K =dV

=3(1-2 ν)

dP E



There are several methods of tightening bolts.

The respective principles are quite different,as are the quality and accuracy levels achieved.

The following is a summary of the most commonly usedmethods.

The torque wrenchThis is probably the most common tightening method.

Its main advantage, especially when the bolt diameterdoes not exceed 30 mm, is that it is very simple and quickto use. But in spite of theoretical developments and muchexperimentation, this method suffers from the followingmajor intrinsic drawbacks:

Characteristics of torque tightening

High amount of uncertainty as to the final

bolt tension loadThe final tightening load depends on the friction coefficientsin the threads of the nut and the bolt, and on the bearing-contact surfaces between the nut and the flange.

In practical terms, it is impossible to know the value ofthese coefficients accurately and reliably.

For a given nominal torque value, the deviation in thefinal tightening load of the bolt can vary between +/-20%

when conditions are good, and +/-60% when conditions arebad (see Fig. 5 below).

This wide range is due to the combin

three phenomena:- the tolerance in the applied torque,

from +/-5% to +/-50%, depending o(see Fig. 6 p.8);

- geometric defects and surface roug

and the bearing surfaces of the fast

- degree of lubrication of bearing surf

Incorporation of additional torsion stress

In addition to the desired axial tensiotightening introduces a “parasite” torswhich can reach over 30% of the ten

The resulting equivalent stress in theTresca criteria) is greatly increased a

yield point of the material, whereas th

remains within admissible limits (seeFurthermore, the residual torsion streof spontaneous loosening at a later s

Furthermore, since the torque is mosin a non-symmetrical manner, there istress, but because its value is compit is often ignored. However, in casesconditions are near the limit, this bentaken into account.

Traditional tightening methods

Tightening method

• Calibrated torque wrenches

• Power tightening tools with regular calibrationon application (measurement of elongationof the bolt or measurement of torque valueusing a calibrated torque wrench)

• Impact wrenches with stiffness adjustmentand periodic calibration on application

(measurement of torque value using a calibrated

Accuracy

on pre-load

± 20 %

± 40 %



Accuracyrange

of torquetightening

method

D± 20% to ± 50%

C± 10% to ± 20%

B± 5% to ± 10%

Manual hand

tool

Calibrated wrenches

with simple releasedevice

Calibrated wrencheswith release deviceand automaticresetting

Calibrated wrenches

with dial gauge

Wrenches withangle drive andrelease device

Power tightening toolswith pneumatic adjust-ment

Power tightening toolswith electric adjustment

Impact wrencheswith stored energy (tor-sion bar or othermeans)

Adjustable wrenches

with angle drive

Simple shock wrenches

Power tightening toolswith positive clutch

Portable

power tool

Equipment type

Non-portable

power tool

Simple air-driven tools

Hydraulic screwing tools

Air-driven toolswith controlled torque

Air pulsed tools

Electric power tighteningtools



Damage to bearing surfaces

Friction between parts under very heavy loads leads togalling and damage to the friction surfaces, namelythe threads between nuts and bolts, and the bearingsurfaces between nuts and flanges.

At the next tightening operation, such damage will increasethe friction forces, and the error in the final tightening loadwill increase accordingly (Fig. 8a, p10).

Difficulties in untightening

It is often much more difficult to unscrew a torquedbolt than it was to screw it on in the first place. Damageto the contact surfaces and corrosion problems impose

Problematic tightening of la

When the required tightening torquevarious torque equipment must be uwrenches, torque multipliers or hydra(Figs. 9a and 9b, p10).

This equipment provides the requireHowever, with the impact wrench in is unreliable. Only the hydraulic torqcondition that top-quality equipment operators following correct procedurimprovement in accuracy.

Simultaneous tightening is

eqdd

2d

3+

2=

Tth load moment in the bolt/nut threads=

F0 tightening load in the bolt=

Torsion stress

Traction stress

Equivalent stress

with :

σ

A B X

X

A B X

eq

M

N

t

F0

π eqd2

d

σF0

AS=

τ =π eq

16 Tth

3

SA4

=

A B

σ eq =

Tth



Methods and devices for measuringtightening torqueIt is possible to reduce the deviation on the final tighteningload by using an instrument to measure either the torque orthe resulting bolt elongation. But whatever the means of

Checking by the angle of

There are two steps to this methotightened to a torque value whichrequired final torque. Then, a furtrotation is apllied.

Thi li htl d th d i ti

Fig. 9a: Torque multiplier Fig. 9b: Hydraulic torque wrenc

Fig. 8a: Tightening with a wrench causes damageto the surfaces of the assembly components.Successive assembly and disassembly increasethis phenomenon.

Fig. 8b: Tightening with a hydrpreserves the condition of the no matter how many successivuntightening operations occur.



Rod and knurled-wheel method

A rod topped with a knurled wheel is screwed into a holebored in the middle of the bolt.When the bolt elongation equals the initial clearance leftbetween the top of the rod and the knurled wheel, thewheel’s rotation is blocked, thereby informing the operator

that the bolt is tightened.This method has certain disadvantages:

- the extra cost of the additional parts and the drilling ;

- the bolt is weakened;

- need for preliminary calibration;

- uncertain degree of accuracy, in particular becausethe operator must turn the knurled wheel a little to checkthe remaining clearance.

Measurement by dial gauge or LVDT

The entire length of the bolt is drilled through to house ameasuring rod. The variation in the distance between thetop of the rod and the top of the bolt is measured witha dial gauge or an electronic sensor (LVDT).

This method is more accurate than the previous one,but it has similar disadvantages:

- the extra cost of drilling the bolt, and the extra parts;- the bolt is weakened;

- preliminary calibration is required.

Ultrasonic measuring (US) method

This consists in measuring the time it takes an ultrasoundwave to travel down and back the longitudinal axis of the bolt.The bolts are not drilled but they must be top-quality, andcareful calibration is required. The method requires qualified,

skilled operators. Constant improvements are making thismethod increasingly attractive, in particular for small-sizescrews (diameters under 20 mm).

Strain-gauge method

Strain gauges are attached to the bolt and connected

to a Weston bridge; the variation in the signal - whichcorresponds to the strain variation in the bolt - ismeasured. Preliminary calibration is required.

This is strictly a laboratory method and can not be used inindustrial applications.

In conclusion, the methods described above requirespecialised technicians and can be long to implement.Furthermore their accuracy is directly proportionalto their cost In addition they do not directly measure

The sensor washer(Fig. 10 above)

The sensor washer offers a major awith other methods, since it measurdirectly.

The “sensor” washer is an instrumenunder the nut on which the torque is

It is highly recommended to install a between the sensor washer and the friction on the sensor washer during untightening

This washer acts as a load-cell sens

The accuracy is good and the metho

When a torque wrench is used, the fgreatly from one bolt to another for tIf good accuracy is required, each asfitled with a sensor washer.

Furthermore, this method provides eor permanent measurement and rectensile stress when the assembly is

Changing from torquehydraulic bolt tension

Fig. 10: The “sensor



the residual tension load when a torque value is applied toa bolt. This formula was obtained by taking into account

the friction of the threads and the friction of the nut faceagainst the flange.

Fo =

T: tightening torque

p: thread pitchµ1: friction coefficient of the bolt/nut threadsµ2: friction coefficient at the nut face/flanged2: equivalent diameter of the boltrm: average radius of the nut face.

The paragraph “Comparison between the torque wrenchand hydraulic bolt-tightening” describes a real application ofthis formula.

Tighteningwith heater rodThis method consists of elongating the bolt by heatingit with a heater rod inserted down the bolt centre.It then suffices to turn the nut under low torque force until it

is in contact with the flange.Upon cooling, the bolt will contract lengthwise, therebytightening the nut. Simultaneous tightening of several boltsis theoretically possible. The method is theoreticallyaccurate but in fact has several disadvantages:

- A hole must be drilled down the centre of the bolt

to receive the heating rod.

- Heating systems, electrical connections, temperature-controldevices and handling means are required, especially in

the event of simultaneous tightening.

- The method is exceedingly slow, due to the time required

to heat the bolts, and the final tightening load can onlybe checked after the bolts have cooled down, which takeseven longer.

The process cycle includes: heating the bolt, advancingthe nut, cooling down the parts, and measurements.This cycle must be repeated several times in order toadjust the tightening.

The temperature required to reach suitable elongation isoften so high that it could modify the mechanical propertiesof the equipment. As a result, when thermal elongation is

insufficient, additional torque tightening must be performedand verified by measuring the nut angle.This thermal elongation technique is fairly rarely used,and is generally only applied to large sized bolts

deformation due to torque tightenof the nut. The following theoretic

estimating the residual tension loathis technique is applied:

Fo =

θ: angle of rotation of nut

∆t: temperature increase

α: expansion coefficient of the bo

p: thread pitch

S: section of bolt

l: tightened length of boltE: elastic modulus (Young's modu

Tightening by melongationWith this method, the tension loabolt (Fig. 11 below).

In general, the body of the nut is prthrust screws located symmetrically hole. These screws apply - either d

a bearing pressure on the contacThey are turned one by one and st

torque load until a suitable tension

The bolt elongation is most often the previously mentioned methodthis method eliminates torsion strseveral drawbacks:

- Simultaneous tightening is not eastep-by-step tightening process is

one bolt to the next. This is both teand the result is pseudo-simultane

- To precisely determine whether out correctly, an additional measprovided, such as the elongationload-measuring washers.

- The nuts are generally expensivand require several small thrust

several threaded holes

- When professionally applied, thito achieve the quality criteria of described in the introduction.

- The process is very slow becau

to be hand-tightened several tim

For all of these reasons the mec

θ p S E

l

α S E ∆t +T

0.16 p + µ1 0.583 d2 + µ2 rm



DescriptionThis tightening method uses SKF® HYDROCAM® bolt

For optimum accuracy, it is recommetion of the bolt and turning-down of th

In effect the first turning down opera

Tightening with hydraulic bolt tensi

Fig. 12a: SKF HYDROCAM bolt tensioner



Advantages of hydraulicbolt-tensioning

No torsion stress

The method eliminates all “parasite” torsion or bendingstress in the bolt (Fig. 7b below).

Good accuracy

It is accurate because the most important parameter, namelythe traction load, is perfectly controlled throughthe hydraulic pressure in the tensioner. The load does not

depend on the various friction coefficients in the assembly.The only uncertainty in the method arises from the degree

of torque applied when driving down the nut. However, thistorque’s influence is, by definition, of secondary importance.Through simple precautions of good workmanship, good

uniformity can be obtained in the driving-down operation.In addition, the Fh/Fo ratio must be taken into account(hydraulic load/residual load after pressure release). Thisratio is described in detail on page 25. Proper understandingof this ratio is important, for there are means available toobtain an accurate ratio for each assembly.

Easy implementationThis method is easy to perform and requires no physicaleffort, even for very large bolts.

Occupational hazards and physical exertionare significantly reduced.

Material Variety

Many different bolt materials, such as stainless steel,

titanium, composite materials and others, can very easily betightened with the hydraulic tensioning.

Suits a wide range of bolt diameters

The method can be used for a very wide range of bolt

diameters, from 5 to 500 mm!

No damage to components

Internal stresses are controlled, and no friction is generated

under heavy bearing pressure. Therefore this methodprotects the individual components of the assembly(Fig. 8b p10).

Easy untightening

The untightening operation is extremely easy: generallythe req ired h dra lic effort is appro imatel one percent

Process automation is po

The tightening and untigthtening automated, thereby providing:

- optimisation on the simulta

- better tightening accuracy,

- even distribution of tighten- reduced tightening time,

- better working conditions foof difficult access,

- remote control.

The remote control feature allowsall phases of the tightening or unt

from a safe area. This significantlto poor or dangerous working conhazardous media, high temperaturisk of component failure.



1 - The turndown socket is placed over the nut and thehydraulic tensioner grasps the bolt.

3 - After the hydraulic connections, the tensioner is pressu-rised and applies the required tractive force on the bolt.

2 - The brace/retraction unit is screwedend of the bolt.

4 - While the pressure is maintained, thwithout loading, using the socket an



Measurement devicesfor hydraulicbolt tensioningSeveral different control methods can be used to checkthe hydraulic bolt tensioning, depending on the requiredaccuracy for the assembly:

Measuring the hydraulic pressure

Once you have determined the Fh/Fo ratio, the precisemeasurement of the pressure applied to the hydrauliccylinder provides an 8 to 10% accuracy level for the finaltightening load, which is often quite acceptable.

The “double pressurisation” method

Once the bolt has been tightened by the initial cycle of

pressurisation on the tensioner, turning down the nut andreleasing of pressure, a second pressurisation is applied totake a more accurate measurement of the bolt load. Duringthis second pressurisation, the rising pressure value isrecorded and plotted versus the displacement of the toppart of the bolt. The resulting tension load of the bolt can be

determined graphically using the change in the slope of the

The “sensor washer”

This method, which was previoustightening, is perfectly suited to hand provides an accurate measutension load in the bolt (Fig. 16, p

Since this sensor washer remains

it is particularly useful for periodicmonitoring of load variations over

Unlike torque tightening, hydraulicreproducible, does not require onbolt. One washer every two, threemay suffice, depending on the re

With sensor washers, the measurto the bolt load is about five percto two percent by careful machinicomponents.

Note that no additional classic wabetween the tightening nut and thsince hydraulic tensioning genera

Fig. 13: Measurement of thebolt tip displacement with a dialgauge.

Fig. 14: Measurement of the boltelongation with an internal rod andan LVDT sensor.

Fig. 15: Ultrthe bolt elon



The hydraulic bolt-tightening method provides good controlover the tension stress. If the bolt is long enough, the finaltension stress can be safely brought very close to the yieldpoint without any risk of exceeding the latter.

This point is very important for the quality of the joint.

Contrary to what has been commonly believed, casestudies show that the closer the bolt stress to the yieldpoint:

- the better the behaviour of the joint,

- the longer the life of the bolts under cyclic loads(Fig. 17 below).

Controlling the stress in the bolt enables material choice

and dimensioning to be optimised at the design stage.

Compared with traditional methotorque-tightening, hydraulic bolt-reduces the dimensions of the asnumber of bolts. This in turn offer

- a reduction in overall dimension

- a reduction in weight,

- a reduction in cost.



Whatever the method used to tighten a bolt - such as the

assembly shown in Fig. 18 opposite - the goal is in fact to

apply a traction load to the bolt, and a compression loadto the assembled components. In general, the bolt has arelatively low stiffness compared with that of the structuralparts on which the compression stress is applied.

Fig. 19 below shows that line D1, which corresponds tothe bolt, has only a slight slope, whereas the slope forline D2, which corresponds to the structure (Fig. 20 p. 20)is steep. Lines D1 and D2 can be drawn on the samegraph, and the final tightening load - Fo - is shown by the

intersection of the two lines (Fig. 21b, p. 20).

Naturally, the tension in the bolt has the same valuebut exactly the opposite sign as the compression of thestructure (Fig. 21a, p. 20).

The elongation of the bolt is δB and the compression

of the structure is δS.

Technical analysis of bolt-tensionin

A very ela

Fig. 18: Example of bolte



As soon as external traction load FE is applied to theassembly, the tension in the bolt is not increased by FE but

only by F1 because the compression of the structure isreduced by F2.

This gives: FE = F1 + F2 (Fig. 22a1 p. 21).

It can be seen on the graphs that not all of the external load

is applied to the bolt, but only a fraction of it(Figs. 22a2 and 23, pages 21 and 22).

For a compressive external load FE the diagram inFig. 22a2 p. 21 is replaced by the diagram in Fig. 22b p. 22.

The F1 part of the load which is taken up by the bolt

can be be calculated as follows and depends on the boltstiffness - RB and the RS stiffness structure. This gives:

F1= FE RB / (RB + RS)

If the tightening load is insufficient compared with theexternal load, tightening is lost

(Figs. 24 and 25, p. 23).

In the case of cyclic external loads, the diagrams inFigs. 26 and 27, p. 24 show that the actual alternating loadon the bolt is very low. This is very important since we know

that the alternating fraction of the load has a strong influenceon the fatigue of the material (figure 17, p. 18).

Fig. 21a: Bolted joint tightened to a pre-load Fo

Fig. 20: Compression load/deof the structure



Fig. 22a1: Bolted joint with tightened pre-load F0 and external traction load FE

F1 increase of tension load applied to the boltF2 reduction of compression load applied to the structure



Other diagrams showing the relationship between the total load introduced in the bolt and thapplied to the assembly:

Fig. 22b: A compressive external load FE on the assembly induces a reduction in load and an increase in the compression strain in the structure



Fig. 24: When the tightening load is insufficient relative to the external load, tightenin



Fig. 26: Load diagram in the event of a cyclic external load

L/d i Fh/F i f h d li



L/d ratio, Fh/Fo ratio for hydraulictightening

It is clear that the greater the difference instiffness between the bolt and the structure, the smallerthe fraction of the external load of the assembly whichaffects the bolt.

Therefore, it is always better to use bolts with a high lengthto diameter (L/d) ratio, for stiffness is lowest in this case.

Furthermore, this L/d ratio is also the main parameterdetermining the ratio between the hydraulic load Fh appliedduring tightening, and the final residual load Fo(Fig. 28, p. 24). It is clear that the higher the ratio L/d,the lower the relation Fh/Fo. The benefits in a high L/d ratioare therefore twofold.

When the L/d ratio is low (< 1.5), the to obtain the final tightening load can the yield point of the material. In thesea top-quality hydraulic torque wrench

solution.However, in this case, the choice betwtightening methods will also depend oof the application, in particular the req

and distribution of tightening, accessisimultaneous tightening etc.

C



This comparison will be made by taking two cases: tighteningan existing assembly, and design of a new assembly.

Tighteningan existing assemblyFirst consider two flanges with an outer diameter of 600 mm.The bolted joint comprises sixteen M20x2.5 bolts on a PCD

of 500 mm (Fig. 18, p. 19).

Each bolt has a clamped length of approximately 200 mm,which means a length/diameter ratio L/d = 10 - a frequent

choice in mechanical applications.

The pitch of the threads is 2.5, in compliance with the ISOstandard, and the class of material is 10-9.

This bolt size has been deliberatecomparison because these bolts ctightened with a calibrated manua(torque < 700 Nm).

Naturally, tightening with a hydraueasier.

Fig. 29, page 26 gives the dimenISO threads:

d = 20 mm d2 = 18.376 m

deq = 17.655 mm As = 244.5 m

For calculation purposes, the scre

a cylindrical bar with an equivalena cross-section As.

Comparison between torque-wrenctightening and hydraulic bolt-tensio



Tensioner reference

Bolt dimensions Metric(M Ø x pitch) system

Ø (IN) - ImperialThread/in system

Max. pressure (MPa)(psi)

Hydraulic section (cm2)(in2)

Max. hydraulic load (kN)(lbf)

Piston stroke (mm)

D (mm)

H (mm)

H1 (mm)

D1 (mm)

D2 (mm)

A (mm)

U (mm)

X (mm)

HTA 20

M 20 x 2.5M 22 x 2.5M 24 x 3M 27 x 3

3/4” - 107/8” - 91” - 8

15021756

203.10

30067443

8

86

100

30

74

56

26

38424652

138142146

HTA 35

M 27 x 3M 30 x 3.5M 33 x 3.5M 36 x 4

1” - 81 1/8” - 71 1/4” - 7

1 3/4” - 6

15021756

355.43

525118025

8

109

116

40

97

73

31

52576369

168173179

HTA 60

M 42 x 4.5M 45 x 4.5M 48 x 5M 52 x 5

1 1/2” - 61 3/4” - 52” - 4 1/2

15021756

609.30

900202328

8

137

140

54

133

102

40

80869299

220226232

HTA 50

M 36 x 4M 39 x 4M 42 x 4.5M 45 x 4.5

1 3/8” - 61 1/2” - 61 3/4” - 5

15021756

507.75

750168606

8

128

128

49

116

90

38

69748086

197202208

HTA 90

M 45 x 4.5M 48 x 5M 52 x 5M 56 x 5.5M 60 x 5.5

1 3/4” - 52” - 4 1/22 1/4” - 4 1/2

15021756

9013.95

1 350303492

8

166

154

65

154

114

42

869299

107114

240246253

HTA 130

M 60 x 5.5M 64 x 6M 68 x 6M 72 x 6M 76 x 6

2 1/2” - 42 3/4” - 43” - 4

15021756

13020.15

1 950438377

8

198

179

82

187

137

50

114122130137145

293301309

HTA 160

M 72 x 6M 76 x 6M 80 x 6

2 3/4” - 43” - 43 1/4” - 4

15021756

16024.80

2 400539541

10

215

190

86

203

145

50

137145152

327335342

Class 10 9 material has the following characteristics: The uncertainty attributable to the t



Class 10-9 material has the following characteristics:

• ultimate tensile strength

Rmx 1 000 MPa

• yield point at 0.2%:

Re (0.2%)x 900 MPa

It is decided not to exceed a stress of 90% of the yieldpoint (0.2%) of the material

thus: σmax = 0.9 Re = 810 MPa

Therefore the maximum admissible load on the bolt will be:

Fmax = 0.9 Re As

Fmax = 198 000 N

Tightening with a hydraulictensionerThe HTA 20 tensioner, which suits the characteristics of thebolts being used, is selected from the SKF HYDROCAMstandard product range (Fig. 30 p. 20).

Uncertainty factors in hydraulic tensioning

The various factors of uncertainty generally encountered inhydraulic bolt tensioning are listed and analysed below.

Deviation due to dimensions and tolerances of bolts

and parts of assembly

Fig. 28 p. 25 helps determine the hydraulic tension load thatmust be applied to the bolt in order to obtain the desiredfinal tightening load.

It is seen that for a ratio L/d = 10, the Fh/Fo ratiobetween the necessary hydraulic load Fh and final residual

tightening load Fo is approximately 1.10 and 1.20,viz: 1.15 +/- 4.5%.

However, the graph shows a general deviation(for L/d =10: +/- 4.5%) which takes into accountthe various types of assembly, and the various shapes andcharacteristics of components, threads in particular, mostoften met in mechanics.

However, we know from experience that for a givenassembly, the Fh/Fo ratio will vary by +/- 2% or less forsingle bolt-tightening, since the dimensional tolerances,

geometric faults and deviations in material characteristicsneed to be considered solely for a given part or a givenbatch.

Therefore, we take 1.15 as the average value and1 18/1 12 as maximum and minimum respectively

The uncertainty attributable to the tThe deviation on the pressure va

It depends on the accuracy of the(manometer or sensor) and on thoperator’s reading.

In the final analysis, the accuracy

load comes to +/- 3%.We retain our preliminary conditioa tension load on the bolt of Fmax

(to avoid exceeding a maximum s

The hydraulic tension load will theto the same value:

Fhmax = 198 000 N

Taking the “hydraulic” uncertaintythe minimum hydraulic tension lo

Fhmin = 198 000/1.06so Fhmin = 186 800 N

And the required average hydrau

Fhm = (Fhmaxi+ Fhmini)/2 = 192 400

Since the SKF HYDROCAM HTA

hydraulic area of 20 cm2, and allo(98%), the nominal pressure will

Uncertainty due to the clampin

When the nut is turned down mauncertainty as to the clamping mo

leads to a +/- 3% uncertainty as tThe General Catalogue describeslimit turn-down deviation.

In our example, the deviation factdown.



Consequences of uncertainties on thefinal load

The total uncertainty on the final tightening load can now

be calculated.

The final tightening load is maximal when the hydraulictension load is maximal, the clamping moment to turndown the nut is maximal, and ratio Fh/Fo is minimal.

The maximum final tightening load will therefore be:

Fomax = 182 000 N (198 000 x 1.03/1.12)

The final tightening load is minimal when hydraulic tensionis minimal, the clamping moment to turn down the nutis minimal and ratio Fh/Fo is maximal.

The minimum final tightening load will therefore be:

Fomin = 154 000 N (186 800 x 0.97/1.18)

Maximum admissible externassembly

Now that our assembly is correctly fa

hydraulic tensioner, let us see what mcan be applied on each bolted point set limit of:

Fmax = 198 000 N.

For this purpose, the stiffnesses of thof the assembly are required.

The stiffness of the bolts can be calc

dimensions:

RB = AS E / L = 244.5 x 210 000/200 =

We have not described a precise destherefore we refer to the general expengineers and arbitraril set the stiffn

Fig. 31: Load/elongation diagram of a M20x2.5 lg 200 mm bolt tightened with a bolt tensioner

It is therefore easy to calculate the maximum admissible Uncertainty as to friction coeff



t s t e e o e easy to ca cu ate t e a u ad ss b eexternal load which can be supported by the bolts without

exceeding the set limit of 810 MPa.

Thus, the total load on each bolt - even includingthe external load - must not exceed:

Fmax = 198 000 N

The value of this maximum load is:Ftmax = Fomax + F1 = Fomax + FE RB / (RB + RS)

Thus the maximum admissible FE can becalculated:

FE = (Ftmax - Fomax) (RB + RS) / RB

FE = 140 000 N (Fig. 31 p. 29)

The total external load which can be supported by theassembly described on page 26 is:

16 x 140 000 = 2 240 000 N

The distribution of the load on the bolts is assumed to beuniform.

We know that bolts will all “work” in almost the

same way with very similar stress levels.There is greathomogeneity. However, for higher levels of external load,there is the risk of exceeding the safety limit set for thebolts.

Tightening with a torque wrench

Let us now tighten the same set of bolts with a torquewrench.

It is decided to use a manual calibrated torque wrench.

Uncertainty factors with torque wrenchtightening

Let us examine these factors and the values they canreach.

Uncertainty as to the torque itself

A calibrated torque wrench generally has a deviationof: +/- 5%.

However, it should be pointed out that with the commonly

used torque tightening equipment this deviation is often

U ce a y as o c o coe

Two friction coefficients need to bthe case of the torque wrench:

µth = friction coefficient in the nut/

µfl = friction coefficient of the nut f

For steel parts originating from a generally lies between0.08 and 0.12 (0.10 +/-20%)and the second is 0.10 to 0.15 (0

Analysing the componenof the tightening torque

The tightening torque TT delivered

to overcome two load moments:

- load moment in the Tth bolt/nut twhich generates bolt torsion)

- load moment in the nut/structure

Therefore, we can calculate TT =

Equations commonly used for the

Tth = T’th + T’’th = F (p/2π) + F (µ1 rTfl = F (µ2 rm)

where p = thread pitch,(2.5 for M20)1/2 π = 0.16

µth = friction coefficient in tµfl = friction coefficient in bd2 = average diameter of trr = 0.583 d2 = average rad

rm = average radius of bea(13 for nut M20)F = tightening load

This can be expressed as:

Tth = T’th + T’’th = F (0.16 p) + F (µTfl = F (µfl rm)

In fact only the part T’th of torque the bolt: the other torque compon

“parasite” torques.

The aim is to obtain the same finas with the hydraulic bolt tensioneFom = 168 000 N

The load moment on the bearing surfaces is Fmax = TT max /(0.16 x 2.5 + µth min 0.583



g

Tfl m = 168 000 x (0.125 x 13)

Tfl m = 273 000 Nmm

The required average torque will be

TT m = Tth m + Tfl m = 520 200 Nmm

520.2 Nm52.02 daNm

Only 67 200 Nm (just 13%) of this is actually used fortightening !

Given the accuracy of the tool itself, the real torque will bewithin the following two values:

TT min = 520 200 x 0.95 = 494 190 Nmm

TT max = 520 200 x 1.05 = 546 210 Nmm

Consequence of uncertaintyas to the final load

The corresponding minimum and maximum tension loadsin the bolt will be:

Fmin = TT min /(0.16 x 2.5 + µthmax 0.583 d2 + µflmax 13)

Fmin = 494 190/(0.16 x 2.5 + 0.12 x 0.583 x 18.376 + 0.15 x 13)

Fmin = 135 930 N

( µ

Fmax = 546 210/(0.16 x 2.5 + 0.08 x 0.58

Fmax = 213 610 N

It is already clear that the real averag

the targeted average load:Fomr = (135 930 + 213 610)/2 = 174 7Which is: + 4%

Expressing the tolerance as a functioaverage tightening load, the tension expressed as follows:

Fo = 168 000 N (+ 27%(or 174 770 N +/- 22%)

(Fig. 32 below)

The stress analysis is detailed in the but it can already be seen that, beca

torsion stress, to induce a tensioningin the bolt by the torque-wrench methequivalent stress of 795 MPa and the810 MPa leaves only a maximum tenthe external load FE) of 172 kN.

It is also seen that even with an accuthe tolerance range on the final tightein our example, almost three times gwrench than with a hydraulic bolt-ten

Effect on stresses in the bolt We can clearly see that the exter



Case where stresses are maximum

Let us now see what happens in the bolt in the caseof “stacking” of tolerances, that is, when the actual valuesof the various parameters simultaneously contributeto maximising the resulting loads.

The loads induce the following stresses in the bolt:- traction stress:σmax = Fmax /As = 213 610/244.5

σmax = 873.6 MPa

- torsion stress:τmax = 16 Tth max / (π d3

eq)

we know thatTth max = 21 610 x (0.16 x 2.5 + 0.08 x 0.583 x 18.376)

Tth max = 268 520 Nmm

therefore τmax = 248.5 MPa

- The equivalent Von Mises stress is calculated as:σeq max = σmax

2+ 3 τmax2

σeq max = 974 MPa

It is seen that in this case, where loads are maximal,the torsion stress - which attains 28 percent of the traction

stress - leads to an increase in the equivalent stress ofmore than 11%.

In this case, the yield point of the material will probably beexceeded as of the tightening operation itself, let alonethe 810 MPa limit which we had set ourselves.

External load on the assembly when stresses aremaximum

Now let us assume that an external load of 140 000 N isapplied to one bolt in the assembly after torque-tighteninghas resulted in maximum loads.

We have already seen that the load taken up by the bolt is:

F1 = FE RB / (RB + RS)

F1 = 15 900 N

The maximum tension load on the bolt increases toa level of:

Fmax = 213 610 + 15 900 = 229 510 N

worse.

In our example, there is a very hiwill occur during the tightening opwill break under working conditio

Case where stresses are minim

We can also examine what happwhen the actual values of the varsimultaneously contribute to mini

These minimum loads generate tthe bolt:

- traction stress:σmin = Fmini / As = 135 930/244.5

σmini = 556 MPa

- torsion stress:τmini = 16 Tth mini / (π d3

eq)

we know that

Tth mini = 135 930 x (0.16 x 2.5 + 0

Tth mini = 229 120 Nmm

so τmini = 212 MPa

- And the equivalent Von Mises sσeq mini = σmini

2 + 3τ mini2

σeq mini = 666 MPa

It can be seen that in “negative sthe resulting loads are minimal, th38% of the traction stress and lea

almost 20% in the equivalent stre

Fortunately, in this case, the loadare far below the yield point of th

However, it will be noticed that thobtained by the torque method istraction stress foreseeable in ave(666 MPa versus 687 MPa).

This serves to confirm that when h i i d d b hi i h

External load on the assembly when stressesi i l

Risks when tightening with



are minimal

Let us now apply the external load of 140 000 Nto the torque wrench tightened assembly under (chance)minimal load conditions.

It is clear that from the stress viewpoint, the consequencesare not very severe.

We know that the additional load on the bolt is:

F1= 15 900 N

The tension load is expressed by:

F = 135 930 + 15 900 = 151 830 N

- The traction stress is:

σ = 621 MPa

- The torsion stress is:τ = 212 MPa

- The equivalent Von Mises stress will be:

σeq = 721 MPa

We approach, but do not reach the yield point

of the material.

In fact in this case, the problem lies not in terms of thestresses, but more with the tightening itself as will be shown:

In our example, the external load on the bolt is:FE = 140 000 N

For “minimal” tightening conditions, the resultingtightening load is 135 930 N whereas the external load

which leads to looseness is:FE = Fo RS / (RB + RS)

which is 153 345 N !

In the case of a slightly higher external load (+10%),there is a real risk of looseness, unscrewing and leakage

(if it is a sealing application).

Furthermore, we also know that, under cyclic loads,fatigue failure is more likely to occur when the tighteningis not optimal

With torque tightening, the two followoccur on a given application:

- excessive stress, resulting in bolt faover-stress,

- insufficient tightening, with the risk oleakage and even fatigue failure.

To avoid these risks, the tendency is the bolts.

Design engineers often think that hig

can thus be applied.

However, this approach both worsenof fatigue resistance, and increases t

assembly, which in turn increases cobigger or heavier bolts, larger flangesimpact service and maintenance cosfor heavier and larger parts.

Conclusion of the comThe comparison between the torquinmethods was based on tightening of

The benefits of using the tensioner arAs described in the introduction, hydeven more advantageous when sevetightened, which is, of course, the sitoften encountered.

In this case, ease of use, safety, reliaand the possibility of performing simunot to mention better accuracy are juthe advantages.

In addition, accuracy can be further i

measurement system such as the prsensor washer.

Design of a new Increasing the number of



Design of a newassemblyLet us now examine the case involving the design of a newassembly.

Based on the preceding work, it is clear that designwill differ greatly depending on whether the bolt-tighteningmethod involves torquing or the use of hydraulic tensioners.

Let’s look again at the example of two flanges with an outerdiameter of 600 mm, assembled using sixteen M20x2.5bolts on a PCD of 500 mm (Fig. 18, page 19).

Tightening with hydraulic

tensionersIn the previous example, we optimised the size of thebolts and the value of the loads was optimised for hydraulictensioning.

Nothing needs to be added in this case.

Tightening with the torque methodLet us now examine the design considerations to allow use

of the torque-tightening method while ensuring the samesafety margin as with the bolt tensioner.

Three solutions exist at the initial design stage:

- selecting a higher class of bolt material

- using a greater number of bolts

- increasing the size of the bolts.

Let us analyse each of these three solutions in turn.

Selecting a higher class of bolt material

Given that the maximum stress isσeq maxi = 1 033 MPa, we conclude that the class of material

for the bolt must be 12-9.In this case, the ultimate tensile stress is Rm > 1 200 MPaand the yield point is Re (0.2%) > 1 080 MPa.

Yet, even with this solution, the safety margin will not bethe same as with the bolt tensioner because 90% of theyield point means a maximum stress of 972 MPa.

Therefore, in this example, changing the class

This solution can only be retainedspace available around the full peassembly.

The space between two successi

that the socket of the torque wreninterference with the adjacent bo

In our example, the overall diameapproximately 36 mm and the ouis 50 mm.

This means that the straight dista

two successive bolts must be gre

We assume a value of 50 mm to With sixteen bolts over a diametebetween two bolts equates to 97.sufficient.

It would even be possible to use

the distance between adjacent bo52.26 mm.

For 16 bolts the maximum stressσeq maxi = 1 033 MPa

whereas we intended not to excetherefore, we need to reduce the than 20%.

If we simply apply this coefficient

we find that 20 bolts are needed

Let us examine the situation of 20a torque wrench.

As we have already seen, the av

168 000 x 16 = 2 688 000 N

Therefore with 20 bolts:Fom = 134 400 N per bolt.

Average torque required to reach the averagetightening load:

Applying the external effort:

A h i l h f



tightening load:

average load moment in the threads:

Tth m = 134 400 x (0.16 x 2.5) + 134 400 x (0.10 x 0.583 x 18.376)

Tth m = 53 760 + 144 000 = 218 400 Nmm

average load moment on bearing surfaces of the nutsTfl m = 134 400 x (0.125 x 13)

Tfl m = 218 400 Nmm

average torque required


416.16 Nm41.616 daNm

Allowing for the accuracy of the tool, the actual torque will

lie somewhere between:TT min = 416 160 x 0.95 = 395 350 Nmm

TT max = 416 160 x 1.05 = 436 970 Nmm

Corresponding minimum and maximumtension loads in the bolt

Fmin = TT min /(0.16 x 2.5 + µth max 0.583 d2 + µfl max 13)

Fmin = 395 350/(0.16 x 2.5 + 0.12 x 0.583 x 18.376 + 0.15 x 13)

Fmin = 108 740 N

Fmax = TT max /(0.16 x 2.5 + µth min 0.583 d2 + µfl min 13)

Fmax = 436 970/(0.16 x 2.5 + 0.08 x 0.583 x 18.376 + 0.10 x 13)

Fmax = 170 890 N

Actual average tension load:

Fomr = (108 740 + 170 890)/2 = 139 815 N

(4% more than the expected value)

Uncertainty in the tension load in the bolt:

Fo = 134 400 N (+ 27% at - 19%)

Stresses in the bolt in the caseof maximum loads

- traction stressσmaxi = Fmaxi / As = 170 890/244.5

σmaxi = 699 MPa

As we have previously seen, the fracload which is applied in every bolt is:F1 = 12 720 N

Maximum tensioning load in the bolt:

Fmaxi = 170 890 + 12 720 = 183 610 N

Stress when external load is appl

- traction stress:

σmaxi = 751 MPa

- torsion stress:τmaxi = 126 MPa

- Equivalent stress (Von Mises):σeq maxi = 782 MPa

The objective is achieved. Therefore

instead of 16 can be a solution.

However, the cost of the bolts increamust be added the extra machining cthere is still a risk that problems will a

tightening.

(It should to be pointed out that a ca18 bolts yields a maximum stress of:

which does not meet our safety requ

Inversely, however, an assembly desibolts, tightened by the torque methoduse 16 M20x2.5 bolts tightened by themethod.

Increasing the bolt diamete

As with the increase in the number o

apply the 20% coefficient to the squadiameter. Given the standard dimens

a 24 mm bolt is needed instead of a

The dimensions of M24x3 ISO bolts

d 24 mm d 22 0508 mm

Average torque required to obtain the average load:

load moment in the threads:

With the external load, the stre

traction stress:



load moment in the threads:

Tth m = 168 000 x (0.16 x 3) + 168 000 x (0.10 x 0.583 x 22.0508)

Tth m = 80 640 + 215 970 = 296 610 Nmm

load moment on the bearing surfaces:

Tfl m = 168 000 x (0.125 x 16)Tfl m = 336 000 Nmm

average torque required:


632.61 Nm

63.261 daNm

Because of the uncertainty induced by the tool, the real

torque value will randomly lie between the following:

TT mini = 632 610 x 0.95 = 600 980 Nmm

TT maxi = 632 610 x 1.05 = 664 240 Nmm

Corresponding minimum and maximum tension loads

in the bolt:

Fmini = 600 980/(0.16 x 3 + 0.12 x 0.583 x 22.0508 + 0.15 x 16)

Fmini = 135 890 N

Fmaxi = 664 240/(0.16 x 3 + 0.08 x 0.583 x 22.0508 + 0.10 x 16)

Fmaxi = 213 690 N

Stresses in the bolt in the maximum case:

- traction stressσmaxi = Fmaxi / As = 213 690/352.5

σmaxi = 606.5 MPa

- torsion stressτmaxi =16 Tth maxi / (π d3

eq)

(Tth maxi = 213 690 x (0.16 x 3 + 0.08 x 0.583 x 22.0508) = 322 340 Nmm)

τmaxi = 173 MPa

- Equivalent stress (Von Mises criterion):σeq maxi = σmaxi

2 + 3 τmaxi2

σeq maxi = 676 MPa

Application of the external load:

- traction stress:σmaxi = 668 MPa

- torsion stress:τmaxi = 173 MPa

- equivalent stress (Von Mises):σeq maxi = 732 MPa

Once again, we have achieved owe can see that using M24x3 bolbolts is an acceptable solution.

However, the increased cost of thas 40%, and the risks of their comcase still remain.

It is seen that the calculation with(which is an unusual dimension), stress of σeq maxi = 839 MPa.

Therefore, if this solution is chose

to accept a lower safety margin.

Inversely, however, an assembly bolts tightened by the torque methuse 16 M20 x 2.5 bolts tightened

method.

Simultaneous tightening with bolt ten



Simultaneous tightening with bolt tenSimultaneous tightening consists in tensioning severalbolts, or even all of the bolts, of an assembly at the same

time.Some simultaneous tightening examples will be foundat the end of the chapter.

We have already seen that this tightening process presentsimportant, even decisive, advantages:

- great homogeneity in tightening all the bolts of theassembly,

- straightforward implementation,

- short process times.

Hydraulic bolt tensioning is the most convenient methodfor obtaining simultaneous tightening.The following information is of a general nature, and providesa series of examples. Results on actual applicationmay vary depending on the dimensions and tolerancesof the components, the quality of the tensioning tooling,and the implementation procedure. A particular application

can, if necessary, be examined in greater detail.

Simultaneous tig

of 100% of the boThis is the most accurate and quickeare tightened at the same time (Fig. However, it requires as many tensionTherefore in our previous example, 1required, including the necessary hosconnections.

The tightening operation is quite simpby a tensioner. All the tensioner oil-feeto a common source of high-pressuresimultaneously stretched (preferably irequired load. In our example: Fo resifor a hydraulic load Fh = 192 400, usinHTAtensioners pressurised to 98 MP

This method provides excellent tighteover all the bolts. We know from expthe previously described uncertaintiereduced:

- the uncertainty on the Fh/Fo ratio is

- the uncertainty on the hydraulic loa

on the tensioners only, and is there

- the uncertainty in the nut clamping

The resulting dispersion in the tightenof the different bolts will be as low as

Simultaneous tightening



of 50% of the boltsIn this case, every other bolt is tightened in one step,(Fig. 34 below), so that the number of tensioners requiredis halved. In our example, only eight tensioners are required

with the necessary hoses and hydraulic connections.

The tightening process requires two rounds of tightening onthe first half of the bolts (which we call “even”) and at leastone round on the second half (which we call “odd”) which

means three and possibly four tightening rounds.

Assuming that during the first round we apply a hydraulic

load of 192 400 N to the even bolts, and settle to a residualload of 168 000 N.

A first round on the second half of the bolts - the oddbolts - obtains the same residual load of 168 000 N.

However, tightening the odd bolts releases some of theload taken by the even ones, and reduces the load by10%, to an average tightening value of 151 000 N.

A second tightening round is thusback to 168 000 N.

However, this second step will onsecond half to loosen by approximtightening load to 158 000 N.

If this level is acceptable, the tighcan stop at this point. However, srequires a second round to bring to 168 000 N. This round will affeapproximately 3%, for an averageof 3%, which is often perfectly ac

The load homogeneity is not as g100% simultaneous tightening, buwithin an acceptable range of +5%

In many applications this deviatioacceptable.

Naturally, the intervention time is preceding case.

Simultaneous tighteningf 25% f th b lt



of 25% of the boltsIn this case, only one bolt in four is tightened at each step(Fig. 35 below).

The number of tensioners required is one-fourth the totalnumber of bolts, namely four in our example.

Each set of four bolts will need at least four tighteningsteps, which means that sixteen steps are requiredto guarantee acceptable load homogeneity.

The first set of bolts (say numbers 1, 5, 9 and 13) is firsttightened with a residual load of 168 000 N, but then willlose 20%, 30% and finally 35% of the load (109 000 N)while the neighbouring sets are successively tightened.

Four rounds are required before the fir5% of the nominal load after adjustmeThe load is then 159 600 N. If this vari

the application, we may stop the opera

However, the load homogeneity is obas in the case of 100% simultaneous

The level of variation will lie within th

One more round on the first set of bothe variation within the range 5%/-8%

Clearly, the use of control means succan reduce the uncertainties describe

Since there are more rounds, the inteproportionally.



Fig. 36a Fig. 36b

Fig. 37

Fig 36a and 36b: Multiple Stud TeMachine for nuclear reactor vesse

Fig 37: Simultaneous tightening wtensioners connected to a single o



Conclusion



The quality of a bolted assembly depends on twointerdependent parameters, namely:

- the design of the assembly,- the method used to tighten the bolts.

It is therefore necessary to choose the suitable tighteningmethod at the beginning of the design phase.

Tightening bolts by the torque method is simple, especiallyfor bolts of reasonable dimensions.

However, the torquing process generates “parasite” torsion

stress in the bolt, which weakens the assembly. More importantly,uncertainty regarding the final load remains very high.

Although the use of top-quality hydraulic torque wrencheshas slightly improved this situation, the inherent drawbacksof the method are incompatible with the requirements inmany applications.

Tightening by heater rods or by mechanical elongationoffers some improvement compared with torque-tightening,but these two methods are rarely used due to their cost

and the time-consuming processes involved.

Hydraulic tensioner tightening offers the best compromisebetween quality, reliability, safety and accuracy on the onehand, and ease of use, cost and time-saving on the other.

Particularly with regard to costs, this method is not onlycompetitive from the tooling-investment viewpoint but italso provides further savings through optimised design ofassemblies and the low cost of site operations.

To summarise, hydraulic bolt-tensioning offers the followingadvantages:

- no “parasite” torsion stress in the bolts,

- high load precision simply by controlling the hydraulicload,

- fast, safe and easy operation,

- the integrity and the reliability of threads and bearingsurfaces are maintained,

- readily compatible with simultaneous tensioning of all thebolts in an assembly (or even several assemblies),

- tightening and measuring operations can be partially orfully automated,

compatible ith a ide ariet of bolt materials (stainless

- greater homogeneity due to exc(very important for simultaneous

- easy implementation of simple min particular the “smart washer”,the final tightening load.

Hydraulic bolt tensioners can generwhich were not initially designed tensioners, provided that it is possufficient protrusion on top of the strongly recommended to make abolt tensioners from the beginning

process: under such circumstancmethod can be achieved.

Bolted assemblies specifically dewith bolt tensioners are particularrequirements of applications requand safety, and they are optimiseweight.

The SKF HYDROCAM General C

- products and services,- applications,- tensioner selection criteria- general instructions and operati

A full list of these products and seon the following page.

SKF Equipements• sells the SKF Linear Motion product range: driving systems (ball & roller screws linear motors)



• sells the SKF Linear Motion product range: driving systems (ball & roller screws, linear motors),

actuation systems, positioning systems, precision bearings and spindles for machine tools.

• offers top-quality components and industrial equipment parts such as:

- modular solutions for automation- gas springs

- shock absorbers- bushings- precision chains

- seals- wheel and ball transfer units

• designs and manufactures the HYDROCAM range of bolt tensioners

• backs up the quality of its services and products in the form of ISO 9001 certification

Bolt-tightening products and servicesprovided by SKF Equipements

HYDROCAM bolt tensioners:- complete standard range including six different types covering a large range of bo

M8 to M160 and loads of 50 kN to 8500 kN

- standard tensioners adaptable according to application interface

- special tensioners designed for dedicated applications, extending the range of bo

Sensor washers for measuring tightening loads

Accessories:- manual pumps delivering various pressure ranges: 700, 1000 or 1500 bar

- air-driven hydraulic power units delivering various pressure ranges: 700, 1000, 15- electrically driven hydraulic power units delivering various pressure ranges: 700,

- high-pressure hoses of all lengths; distribution blocks

- pressure intensifiers.

Simultaneous tightening machines and systems(with optional automation)

Standard and customised automatic remote-control sy

Services:- assistance in design of bolted joints,

- assistance in selecting most appropriate tightening method,

- expertise and experimentation

Linear Motion & Precision Technologies



http://www.linearmotion.skf.com

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