FEM 1.001 Rules for the Design of Hoisting Appliances

File is licenced for Sarawak Energy Berhad - Order-no: 401995 - 1License(s)

FEDERATION EUROPEENNE DE LA MANUTENTION

F.E.M.

SECTION I

3rd EDITION REVISED

HEAVY LIFTING APPLIANCES

1.001

1998.10.01

RULES FOR THE DESIGN OF HOISTING APPLIANCES BOOKLET 1

OBJECT AND SCOPE

The total 3rd Edition revised comprises booklets 1 to 5 and 7 to 9 Copyright by FEM Section I Also available in French and German


Booklet 1

OBJECT AND SCOPE

1.1.

PREFACE.............................................................................................................................. 2

1.2.

INTRODUCTION ................................................................................................................... 3

1.3.

OBJECT OF THE RULES ..................................................................................................... 5

1.4

SCOPE .................................................................................................................................. 6

LIST OF SYMBOLS AND NOTATIONS .......................................................................................... 7


1.1. PREFACE The Rules for the Design of Hoisting Appliances set up by the Technical Committee of the Section I of the F.E.M., which have been published so far in two Editions, the first one in 1962 and the second in 1970, have been increasingly widely used in many countries all over the world. Taking accourt of this enlarged audience, Section I of the FEM decided to change the format of these Design Rules and to facilitate updating by abandoning the single volume form and dividing the work into a number of separate booklets as follows : Booklet 1 - Object and Scope Booklet 2 - Classification and loading on structures and mechanisms Booklet 3 - Calculating the stresses in the structure Booklet 4 - Checking for fatigue and choice of mechanism components Booklet 5 - Electrical equipment Booklet 6 - Stability and safety against movement by the wind Booklet 7 - Safety rules Booklet 8 - Test loads and tolerances Although not directly a part of these Design Rules, the opportunity is taken to draw attention to the new Terminology of Section I.

1-2


1.2. INTRODUCTION To facilitate the use of these Rules by the purchasers, manufacturers and safety organizations concerned, it is necessary to give some explanation in regard to the two following questions. 1. How should these Rules be applied in practice to the different types of appliance whose construction they cover ? 2. How should a purchaser use these Rules to define this requirements in relation to an appliance which he desires to order and what conditions should he specify in this enquiry to ensure that the manufacturers can submit a proposal in accordance with tris requirements ? 1. It is necessary first to recognize the great variety of appliances covered by the Design Rules. It is obvious that a crane having very high speeds and a rapid working cycle is not designed in the same manner as a small overhead crane for infrequent duty. For such a machine there can be no question of making all the verifications which would appear to be required, from reading through the Rules, because one would clearly finish with a volume of calculations which would be totally out of proportion to the objective in view. The manufacturer must therefore decide in each particular case which parts of the machine, which he is designing, should be analysed and those for which calculation is unnecessary, not because he must accept that the results for the latter would not be in accordance with the requirements of the Rules, but because on the contrary he is certain in advance that the calculations for the latter would only confirm a favourable outcome. This may be because a standard compornent is being used which has been verified once and for all or because it has been established that some of the verifications imposed by the Rules cannot in certain cases have an unfavourable result and therefore serve no purpose. If one takes, for exemple, the fatigue calculations, it is very easy to see that certain verifications are unnecessary for appliances of light or moderate duty because they always lead to the conclusions that the most unfavourable cases are those resulting from checking safety in relation to the elastic limit. These considerations show that calculations made in accordance with the Rules can take a very different form according to the type of appliance which is being considered, and may in the case of a simple machine or a machine embodying standard components be in the form of a brief summary without prejudicing the compliance of the machine with the principles set out by the Design Rules. 2. As far as the second question is concerned, some explanation is first desirable for the purchaser, who may be somewhat bewildered by the extent of the document and confused when faced with the variety of choice which it presents, a variety which is, however, necessary if one wishes to take account of the great diversity of problems to be resolved.

1-3


In fact, the only important matter for the purchaser is to define the duty which he expects from his appliance and if possible to give some indication of the duty of the various motions. As regards the service to be performed by the appliance, two factors must be specified, i.e. : - the class of utilization, as defined in 2.1.2.2 ; - the load spectrum, as defined in 2.1.2.3. In order to arrive at the number of hoisting cycles determining the class of utilization, the purchaser may, for instance, find the product of : - the number of hoisting cycles which the appliance will have to average each day on which it is used ; - the average number of days of use per year ; - the number of years after which the appliance may be considered as having to be replaced. Similarly, the load spectrum may be calculated by means of the simplified formula set out in the above mentioned paragraph. In neither case do the calculations call for a high degree of accuracy, being more in the nature of estimates than of precise calculations. Moreover, the numbers of hoisting cycles determining the classes of utilization do not constitute guaranteed values : they are merely guide values, serving as a basis for the fatigue calculations and corresponding to an average life which can be expected with a reasonable degree of safety, provided the appliance, designed in accordance with the present design rules, is used under the conditions specified by the customer in his call for tender and also that it is operated and maintained regularly in compliance with the manufacturer's instructions. If he is unable to determine the class of utilization and the load spectrum, the purchaser may confine himself to stating the group in which the appliance is to be classified. A guide as to the choice of group is provided by Table 2.1.2.5., which is not binding but gives simple exemples which, by way of comparison, may facilitate selection. In the case of mecanisms, the following should also be specified : - the class of utilization, as specified in 2.1.3.2. : - the load spectrum, as defined in 2.1.3.3. : the same observation apply as were made concerning the appliance as a whole.

1-4


The tables in Appendix A.2.1.1. may be used to facilitate determination of the class of utilization. On the basis of the class of utilization of the appliance, they make it possible to determine a total number of working hours for the mechanism, according to the average duration of a working cycle and the ratio between the operating time of the mechanism and the duration of the complete cycle. Table T.2.1.3.5. may be used as a guide by a purchaser wishing simply to choose a group for each of the mechanisms with which the appliance he wants to order is to be fitted. As a general rule, the purchaser has no other information to supply in connection with the design of the appliance, except in certain cases : - the area of hoisted loads presented to the wind, if this area is larger than those defined in 2.2.4.1.2. ; - the value of the out-of-service wind, where local conditions are considered to necessitate design for an out-of-service wind greater than that defined in 2.2.4.1.2.

1.3. OBJECT OF THE RULES The purpose of these rules is to determine the loads and combinations of loads which must be taken into account when designing hoisting appliances, and also to establish the strength and stability conditions to be observed for the various load combinations.

1-5


1.4

SCOPE

The Rules apply to the design of lifting appliances or parts of lifting appliances which appear in the illustrated terminology for cranes and heavy lifting appliances of Section I of the FEM. Appliances not covered by Section I 1) Lifting appliances included in Section V, for exemple : - mobile jib cranes on pneumatic or solid rubber tyres, crawler tracks, lorries, trailers and brackets. 2) Lifting equipment which according to the internal regulations of FEM, are included in Section IX, that is to say : - various items of series lifting equipment, - electric hoists, - pneumatic hoists, - accessories for lifting, - hand operated chain blocks, - elevating platforms, work platforms, dock levellers, - winches, - jacks, tripods, combined apparatus for pulling and lifting, - stacker cranes. For series lifting equipment, those chapters of the Design Rules of Section I which have been accepted by Section IX should be used. These rules comprise eight booklets. In addition some booklets contain appendices which give further information on the method of application.

1-6


LIST OF SYMBOLS AND NOTATIONS Symbol

Unit

Désignation

Paragraphe

A

m2

Area exposed to wind

2.2.4.1.

A

-

Combined influence of residual tensile stresses with dead weight stresses

3.1

A1 à A8

-

Crane groups

2.1.2

Ae

m2

Enveloped area of lattice

2.2.4.1.4

a

mm

Wheelbase of crane : Dimension of lattice in wind load calculation : Length of strip of plate in buckling calculation : Size of fillet weld in notch case 2.33

2.2.2.3 2.2.4.1.4 A-3.4 T.A.3.6.-2.33

a

m/s2

Acceleration

5.8.3.1

B

-

Influence of thickness of structural member

3.1.1.2

B

mm

Width of lattice in wind load calculation

2.2.4.1.4

B0 à B10

-

Classes of utilization of structural members

2.1.4.2

b

mm

Breadth of section across wind front : Largest dimension of rectangular steel section : Length of plate in buckling calculation : Useful width of rail in wheel calculation

2.2.4.1.4 3.1.1.2 A-3.4 4.2.4.1

C

-

Influence of cold : Coefficient used to calculate the tightening torque of bolts : Selection coefficient for choice of running steel wire ropes

3.1.1.3 A-3.2.2.2.2.3 4.2.2.1.3.1

Cf

-

Shape coefficient in wind load calculation

2 2.4.1.4

c, c'

-

Factors characterising the slope of Wöhler curves

4.1.3.5

c1, c1max

-

Rotation speed coefficients for wheel calculation

4.2.4.1

c2, c2max

-

Group coefficient for wheel calculation

4.2.4.1

cos ϕ

-

Power factor

5.2.3.3.2

D

-

Symbol used in plate inspection for lamination defects

T.A.3.6

D

m

Section diameter in shape factor determination

2.2.4.1.4

D

mm

Rope winding diameter : Wheel diameter : Shaft diameter in fatigue verification of mechanism parts .

4.2.3.1 4.2.4.1 A-4.1.3

1-7


Dt

mm

Diameter of bolt holes

3.2.2.2.1

d

mm

d2

mm

Depth of section parallel to wind direction in wind load calculation : Nominal diameter of bolt : Nominal diameter of rope : Shaft diameter in fatigue verification of mechanism parts Bolt diameter at thread root

2.2.4.1.4 A-3.2 .2 .2 .3 4.2.2.1.3 A-4.1.3 3.2.2.2.1

dc

-

Number of completed starts per hour

5.8.1.4

di

-

Number of impulses or incomplete starts per hour

5.8.1.4

dmin

mm

Minimum rope diameter

A-4.2.2

dt

mm

Nominal bolt diameter

3.2.2.2.1

E

N/mm2 Elastic modulus of steel

A-3.4

E1 à E8

-

Groups of components

2.1.4.1

ED

%

Duty factor

5.8.1.4

e

mm

Thickness of strip of plate in buckling calculation : Thickness of plate in welded joints

A-3.4 T.A-3.6-2.31

e1, e2

mm

Plate thicknesses in welded joints

A-3.4

F

N

Wind force : Horizontal force during acceleration : Tensile load in bolts : Compressive force on member in crippling calculation

2.2.4.1.2 A-2.2.3 3.2.2.2.2 A-3.3

F0

N

Minimum breaking load of rope

4.2.2.1.2

F1

N

Permissible working load on bolts

3.2.2.2.1

Fc

N

Projection of rope load on the x axis during travelling

A-2.2.3

Fcm

N

Inertia force due to the load during travelling

A-2.2.3

Fcmax

N

Maximum value of Fc

A-2.2.3

f

-

Fill factor of rope

4.2.2.1.3

Number of electrical brakings

5.8.1.4

fcy g

m/s2

Acceleration due to gravity. according to ISO 9.80665 m/S2

A-2.2.3

H

-

Coefficient depending on group for choice of rope drums and pulleys

4.2.3.1.1

Moment of inertia of mass in slewing motion

A-2.2.3.-3

I

kgm2

1-8


I 1, I 2

mm4

Moment of inertia of stiffeners

A-3.4

ID

A

Starting current of motor

5.2.3.3.2

IN

A

Nominal current of motor

5.2.3.3.1

Itot

A

Sum of currents IA and IN

5.2.3.3.2

IZ

mm4

Moment of inertia of stiffeners

A-3.4

Ii

kgm2

Moment of inertia of mass of a part in rotation

A-2.2.3

Im

kgm2

Moment of inertia of mass of all parts in rotation

A-2.2.3.-2.1

JM

kgm2

Moment of inertia of mass of motor and brake

5.8.1.4

j

-

Group number in component groups E1 to E8

4.1.3.6

j0

m/s2

Acceleration in horizontal motions

A-2.2.3.-2.2

jm

m/s2

Average acceleration/deceleration in horizontal motions

A - 2.2.3

K’

-

Empirical coefficient for determining minimum breaking strength of rope

4.2.2.1.3

K0 à K4

-

Stress concentration classes for welded parts

A-3.6

K2

-

Coefficient for calculating force in the direction of the wind for lattice girders and towers

2.2.4.1.4.4

KL

N/mm2 Pressure of wheel on rail

4.2.4.2

Km

-

Mn med / M max

4.2.1.2

k

-

Spinning loss coefficient

4.2.2.1.3

kc

-

Corrosion coefficient in fatigue verification of mechanism parts

A-4.1.3

kd

-

Size coefficient in fatigue verification of mechanism parts

A-4.1.3

km

-

Spectrum coefficient for mechanisms

2.1.3.3

kp

-

Spectrum coefficient for cranes

2.1.2.3

ks

-

Shape coefficient in fatigue verification of mechanism parts

4.1.3.3

ksp

-

Spectrum coefficient for components

2.1.4.3

k’sp

-

Spectrum coefficient for mechanism parts

4.1.3.5

ku

-

Surface finish (machining) coefficient in fatigue verification of mechanism parts

4.1.3.3

1-9


Kσ Kτ

-

Buckling coefficients used in buckling calculations

A-3.4

L

N

Maximum permissible lifting force

5.8.2.1

L1 à L4

-

Spectrum classes for mechanisms

2.1.3.3

l

m

Length of suspension/length of load pendulum

A-2.2.3.-2

l

m

Equivalent length of fine

5.2.3.3.2

l

m

Length of members in wind force calculations : Overall width or rail head

2.2.4.1.4.1 4.2.4.1.2

lk

m

Length of parts tightened in bolted joints

3.2.2.2.1

M

N.m

External moment in bolted joints

3.2.2.2.2

M1 à M8

-

Mechanism groups ;

2.1.3.1

M1,M2, M3 -

Motor torques required during a cycle of operations

5.8.1.3.1

MF

N.m

Braking torque of motor

5.8.2.1

MNmax

N.m

Maximum running torque required to lift the load

5.8.2.1

Ma

N.m

Torque required to tighten bolts

A-3.2.2.2.2.3

MF

N.m

Bending moment in member in crippling calculation

A-3.3

Mmax

N.m

Maximum value of motor torque

5.8.2.1

Mmed

N.m

Mean value of torque M during motor running time fiT

5.8.2.1

Mmin

N.m

Minimum motor torque during starting

5.8.2.1

m

-

Number of friction surfaces in bolted joints

3.2.2.2.2

m

kg

Equivalent mass for calculating loads due to horizontal motions : Total mass of crane

A-2.2.3.-1 A-2.2.3.-2

m0

kg

Mass of crane without load

A-2.2 .3.-1

ml

kg

Mass of the load

A-2.2 .3.-1

mL

kg

Mass of the hook load

5.8 .3.1

me

kg

Equivalent mass in calculation of loads due to horizontal motion

A-2.2.3.-2.1

m

kg

Load

2.1.2.3

mlmax

kg

Safe working load

2.1.2.3

1 - 10


N

-

Number of hoisting cycles

A-2.1.1

N

N

Force perpendicular to joint plane in bolted joints

3.2.2.2.2

NG

-

Ordinary quality in welding

3-57 $$$$$

NM

N

Tensile force due to external moment in bolted joints

3.2.2.2.2

n

-

Number of hoisting cycles : Number of stress cycles

2.1.2.3 4.1.3.5

n

min-1

Nominal rotation speed of motors in rpm

5.8.1.4

nmax

-

Number of hoisting cycles determining the total duration of use

2.1.2.3

P

N

Load on wheel

4.2.4.2

P1 à P4

-

Spectrum classes for components

2.1.4.3

Symbols indicating welding tests

T.A-3.6

P10, P100 PL

N/mm2 Limiting pressure in wheel calculation

4.2.4.1

PN

W

Nominal power of motor

5.8.1.4

PNmax

W

Maximum power requirement of motor

5.8.2.1

Pmoy I, II

N

Mean load on wheel in loading cases I and II

4.2.4.1

Pmoy III

N

Mean load on wheel in loading case III

4.2.4.1

Pmin I, II, III

N

Minimum load on wheel in loading cases I, II and III

4.2.4.1

Pmax I, II,III

N

Maximum load on wheel in loading cases I, II and III

4.2.4.1

Pmed

kW

Equivalent mean power

5.8.1.3.2

p

mm

Span of crane

2.2.3.3

pa

mm

Pitch of thread

3.2.2.2.1

Q1 à Q4

-

Spectrum classes for cranes

2.1.2.3

q

-

Correction factor for shape coefficient ks

A-4.1.3

q

N/mm2 Dynamic pressure of the wind

2.2.4.1.1

R0

N/mm2 Minimum ultimate tensile strength of the wire of a rope

4.2.2.1.3

RE

N/mm2 Apparent elastic limit σE according to ISO 3800/1

3.2.2.2.1

r

-

2.1.3.3 3.5

Number of levers of loading : Ratio of stresses for large deformations

1 - 11


r

mm

Radius of cylindrical shells in buckling calculations : Radius of rope groove : Radius of rail head : Blending radius

A-3.4 4.2.3.2 4.2.4.1.2 A-4.1.3

r

Ω/km

Ohmic resistance per unit length

5.3.2

S

N

Stress : Maximum tensile force in rope

2.1.3.3 4.2.2.1.1.2

S

m2

Area of all members of lattice girders and towers

2.2.4.1.4.4

S

mm2

Cross sectional area of conductor

5.2.3.3.2

S1

mm

Bearing diameter under bolt head

3.2.2.2.1

SG

N

Load due to dead weight. constant load .

2.2.1 & 3.5

SH

N

Load due to horizontal motions

2.2.3

SL

N

Load due to working load

2.2.1

SM

N

Load due to torques

2.5

SMmoy

N

Mean type M load in bearing calculation

4.2.1.2

SMmin

N

Minimum type M load in bearing calculation

4.2.1.2

SMmax I

N

Maximum type M load in load case I

2.6.1.1

SMmax II

N

Maximum type M load in load case II

2.6.2.1

SMmax III

N

Maximum type M load in load case III

2.6.3.1

SMA

N

Load due to acceleration or braking

2.5.1

SMCmax

N

Load due to maximum motor torque

2.6.4.3

SMF

N

Load due to frictional forces

2.5.1

SMG

N

Load due to vertical displacement of moveable parts of a lifting appliance. excluding the working load

2.5.1 2.5.1

SML

N

Load due to vertical displacement of the working load

SMW

N

Load due to the effect of limiting wind for appliance in service

SMW 8

N

Load due to wind effect for q - 80 N/mm2

2.5.1 2.6.2.1

SMW 25

N

Load due to wind effect for q - 250 N/m2

2.6.2.1

SR

N

Load due to forces not reacted by torques

2.5

1 - 12


SRmax I

N

Maximum type R load in loading case I

2.6.1.1

SRmax II

N

Maximum type R load in loading case II

2.6.2.1

SRmax III

N

Maximum type R load in loading case III

2.6.3.1

SRmin

N

Minimum type R load in bearing calculation

4.2.1.3

SRmoy

N

Mean type R load in bearing calculation

4.2.1.3

SRA

N

Load due to accelerations/decelerations

2.5.2

SRG

N

Load due to self weight of crane parts

2.5.2

SRL

N

Load due to working load

2.5.2

SRW

N

Load due to wind

2.5.2

SRWmax

N

Load due to out of service wind

2.5.2

SRW25

N

Wind load for q - 250 N/m2

2.6.2.2

ST

N

Load due to buffer effect

2.3.3

SV

N

Variable load when calculating structural members subject to large deformations

3.5

SW

N

Load due to in service wind

2.3.2

SWmax

N

Load due to out of service wind

2.3.3

Sb

mm2

Root sectional area of bolt

3.2.2.2.1

Seq

mm2

Equivalent sectional area of tightened bolt

3.2.2.2.1

Sp

mm2

Area of members of lattice girders and towers

2.2.4.1.4.4

s

m

Span of lifting appliance : Rail centres of crab : Distance between travel rails of lifting appliance

8.2.2.1 8.2.2.4 8.2.3

T

h

Total duration of use of lifting appliance

2.1.3.3

T

J

Total kinetic energy in luffing motion

A-2.2.3.-4

T

°C

Ambient temperature at place of erection

3.1.1.3

T

N

Force parallel to joint plane in bolted joint

T

s

Duration of cycle

3.2.2.2.2 5.8.1.4

T0 à T9

-

Classes of utilization of mechanisms

2.1.3.2

1 - 13


T1

s

Period of oscillation

A-2.2.3.-2

Ta

N

Permissible load per bolt which can be transmitted by friction

Tc

°C

Test temperature for impact test

3.2.2.2.2 3.1.3

Ti

h

Total duration of use of mechanism

A-2.1.1

Tm

s

Mean duration of acceleration or deceleration

A-2.2.3-2

t

s

Time when calculating loads due to horizontal motion

t

mm

Thickness of structural member when choosing steel quality : Thickness of cylindrical shell wall in buckling analysis : Thickness of web of trolley rail girder

A-2.2.3.-2.1 3.1.1.2

t1, t2... ti, tr

s

Duration of different levers of loading

A-3.4 8.2.2.7 2.1.3.3

t1, t2, t3

s

Duration of action of couples M1, M2 and M3

5.8.1.3.1

t*

mm

Ideal section thickness when choosing steel quality

3.1.1.2

td

s

Duration of deceleration when calculating loads due to horizontal motion

2.2

tmc

s

Average duration of a hoisting cycle

A-2.1.1

U0 à U9

-

Classes of utilization of lifting appliances

2.1.2.2

∆u

V

Permissible voltage drop

5.3.2

VL

m/s

Hoisting speed :

Vs

m/s

Theoretical wind speed

2.2.2.1.1 5.8.2 .1 2.2.4.1.1

Vt

m/s

Nominal travel speed of appliance

2.2.3.4.1

v

m/s

Steady horizontal speed of point of suspension of load

A-2.2.3.-2

v

mm

Distance of extreme fibre from centre of gravity of section in crippling calculation

A-3.3

Travel speed

5.8.3.1

Notch cases of unwelded members

A-3.6

Angular velocity of a mechanism part about its centre of rotation when calculating loads due to horizontal motion

A-2.2.3

v

m/s

W0, W1, W2 Wi

s-1

1 - 14


x

Ω/km

Reactance per unit length

5.3.2

x

m

Coordinate of point of suspension of hoist rope along an axis parallel to the direction of travel

2.1

Coordinate of position of centre of gravity of suspended load along an axis having the same direction. sense and origin as the axis of x

2.1

x1

m

ZA

-

Assessing coefficient for influence A

3.1.1.1

ZB

-

Assessing coefficient for influence B

3.1.1.2

ZC

-

Assessing coefficient for influence C

3.1.1.3

Zp

-

Minimum practical factor of safety for choice of steel wire ropes

z

m

Coordinate expressing horizontal displacement of load relative to crane

4.2.2.1 A-2.2.3.-2.1

zd

m

Displacement of load during travel motion of crane

zm

m

Displacement of load during travel motion of crane

α

-

Ratio of sides of panel in buckling calculation

A-2.2.3.-2.2 T.A-3.4.1

αi

-

Ratio of duration of use of mechanism during a hoisting cycle to average duration of cycle

A-2.1.1

A-2.2.3.-2.2

αm

°

Angle of inclination of rope during acceleration of crane A-2.2.3.-2.1

β

-

Time coefficient relating to acceleration of crane

βcrit

-

Critical value of β

γc

-

Amplifying coefficient of loading depending on crane group

γm

-

Amplifying coefficient of loading depending on mechanism group

A-2.2.3 A-2.2.3.-2.2

2.3 2.6 ∆l1

mm

Shortening of joined elements under the tightening force in bolted joints

3.2.2.2.1

∆l2

mm

Extension of bolt under tightening force

3.2.2.2.1

∆s

mm

Divergence in span of crane : Divergence in crane rail centres

8.2.2.1 8.2.3

δb

-

Elastic coefficient of bolted joints

3.2.2.2.1

η

-

Shielding coefficient in calculation of wind force : Poisson's ratio : Overall efficiency of mechanism

2.2.4.1.4.2 A-3.4 5.8.3.1

1 - 15


θ

°

Angle of wind relative to longitudinal axis of member

2.2.4.1.4.4

κ, κ’, κ’’

-

Safety coefficients applying to bolted joints

3.2.2.2.1

κ

-

Ratio of the extreme stress values in fatigue calculation

κ

m/Ωmm2Electric conductivity

κx, κy, κxy

-

3.6 5.2.3.3.2

Ratio of extreme individual stresses σx, σy , τxy in fatigue calculation

A-3.6

λ

-

Coefficient applied to horizontal forces in travel motions : Slenderness of column in crippling calculation

2.2.3.3 A-3.3

µ

-

Mass constant in calculation of loads due to acceleration of horizontal motion : Coefficient of friction in threads : Coefficient of friction of contact surfaces in bolted joints

A-2.2.3.-2 3.2.2.2.1 3.2.2.2.2.-3

ν

-

Safety coefficient for critical stresses in structural members

3.Intro]

ν‘

-

Dead weight coefficient in calculation of structural members subjected to significant deformation

3.5

Safety coefficient for calculation of structural members depending on case of loading

3.2.1.1

Safety coefficient for calculation of mechanism parts depending on case of loading

4.1.1.1

= νE , safety coefficient for calculation of bolted joints depending on case of loading

3.2.2.2.2

νE νR νT

-

-

-

νV

-

Safety coefficient for buckling

3.4

νK

-

Safety coefficient for verification of fatigue strength of mechanism parts

4.1.3.7

Experimentally determined coefficient depending on crane type for calculating dynamic coefficient

2.2.2.1.1

Reducing coefficient applied to critical stresses in buckling calculation

A-3.4

ξ ρ

-

-

ρ1

-

Coefficient used to determine the dynamic test load

2.3.3

ρ2

-

Coefficient used to determine the static test load

2.3.3

σ

N/mm2 Calculated stress in structures in general

σ0

N/mm2 Tensile stress for κ =0 in calculation of fatigue strength

3.2.1.1

A-3.6

1 - 16


σ1

N/mm2 Working stress in the root section of bolts

3.2.2.1

σ‘1

N/mm2 Equivalent stresses permissible for bolts

3.2.2.1

σ+1

N/mm2 Permissible tensile stress for κ =+1 in fatigue calculation

σA

N/mm2 Amplitude of the permissible maximum stress in bolts for fatigue calculations

3.2.2.1

σE

N/mm2 Apparent elastic limit of steel

3.2.2.1

σG

N/mm2 Tensile stress due to permanent load : Stress due to dead weight

3.1.1.1 3.5

σR

N/mm2 Ultimate tensile strength

3.2.2.1

σER

N/mm2 The EULER Stress

A-3.4

σV

N/mm2 Stress due to variable loads

3.5

σa

N/mm2 Permissible tensile stress for structural members : Permissible stress for mechanism parts

3.1.1.1 4.1.1.1

σaf

N/mm2 Permissible normal stress for verification of fatigue strength of mechanism parts

4.1.3.7

σb

N/mm2 Initial stress in calculating bolted joints

3.2.2.2.1

σbw

N/mm2 Endurance limit of materials of mechanism parts under alternating bending

4.1.3.2

N/mm2 Permissible fatigue strength in compression for structural members : Calculated compressive stress for mechanism parts

A-3.6 4.1.1.3

σcg

N/mm2 Compression stress in wheel and rail

4.2.4.2

σcp

N/mm2 Equivalent stress used in calculating structural members

3.2.1.3

σcr

N/mm2 Critical stress used in calculating structural members subjected to large deformations

3.5

σvcr

N/mm2 Critical buckling stress

A-3.4

σvcr.c

N/mm2 Critical comparison stress used in buckling calculation

A-3.4

σd

N/mm2 Endurance limit of materials of mechanism parts

4.1.3.4

σf

N/mm2 Calculated bending stress in mechanism parts

4.1.1.3

σvi

N/mm2 Ideal buckling stress for thin walled circular cylinders

A-3.4

A-3.6

σc

1 - 17


σinf

N/mm2 Lower stress in determination of stress spectrum

2.1.4.3

σk

N/mm2 Fatigue strength of mechanism parts

4.1.3.6

σkx

N/mm2 Fatigue strength for normal stresses in the x direction

4.1.3.7

σky

N/mm2 Fatigue strength for normal stresses in the y direction

4.1.3.7

σm

N/mm2 Arithmetic mean of all upper and lower stresses during the total duration of use : Permissible stress in conformity tests to ISO 3600/1

2.1.4.3 3.2.2.2.1

σmax

N/mm2 Maximum stress in fatigue calculation for structural members

3.6

σmin

N/mm2 Minimum stress in fatigue calculation for structural members

3.6.4

σn

N/mm2 Bearing pressure in riveted joints

3.2.2.1

σp

N/mm2 Theoretical tensile stress in bolt due to tightening

3.2.2.2.1

σsup

N/mm2 Upper stress in determination of stress spectrum

2.1.4.3

σsup max

N/mm2 Maximum upper stress in determination fo stress spectrum

2.1.4.3

σsup min

N/mm2 Minimum upper stress in determination of stress spectrum

2.1.4.3

σt

N/mm2 Permissible tensile stress in fatigue verification of structural members : Calculated tensile stress in mechanism parts : Tensile stress in rope

A-3.6 4.1.1.3 A-4.2.2

σv

N/mm2 Reduced buckling stress of thin walled circular cylinders

A-3.4

σw

N/mm2 Permissible stress in alternating tension/compression in fatigue verification of mechanism parts

A-3.6

N/mm2 Permissible alternating stress in fatigue verification of mechanism parts

4.1.1.3

N/mm2 Normal stress in the x direction when calculating structural members

3.2.1.3

σxa

N/mm2 Permissible stress in fatigue verification of structural members

A-3.6

σx max

N/mm2 Maximum stress in fatigue verification of structural members

A-3.6

σx min

N/mm2 Minimum stress in fatigue verification of structural members

A-3.6

σy

N/mm2 Normal stress in the y direction when calculating structural members

3.2.1.3

N/mm2 Permissible stress in fatigue verification of structural members

A-3.6

σwk σx

σya

1 - 18


σy max

N/mm2 Maximum stress in fatigue verification of structural members

σy min

N/mm2 Minimum stress in fatigue verification of structural members

A-3.6 A-3.6

τ

N/mm2 Shear stress in general : Calculated shear stress for mechanism parts

3.2.1.3 4.1.1.3

τa

N/mm2 Permissible shear stress when calculating structural members

3.2.1.2

τaf

N/mm2 Permissible shear stress in fatigue verification of mechanism parts

4.1.3.7

τb

N/mm2 Torsional stress in bolts due to tightening

3.2.2.2.1

τvcr

N/mm2 Critical buckling shear stress

A-3.4

τd

N/mm2 Endurance limit of materials of mechanism parts

4.1.3.4

τk

N/mm2 Fatigue strength of mechanism parts

4.1.3.6

τmax

N/mm2 Maximum shear stress in fatigue verification of mechanism parts

3.6.4

τmin

N/mm2 Minimum shear stress in fatigue verification of mechanism parts

3.6.4

τw

N/mm2 Endurance limit under alternating shear of materials of mechanism parts

4.1.3.2

N/mm2 Endurance limit under alternating shear in fatigue verification of mechanism parts

4.1.3.3

τxy

N/mm2 Shear stress when calculating structural members

3.2.1.3

τxya

N/mm2 Permissible shear stress in fatigue verification of structural members

A-3.6

τxy max

N/mm2 Maximum shear stress in fatigue verification of structural members

A-3.6

τxy min

N/mm2 Minimum shear stress in fatigue verification of structural members

A-3.6

ϕ, ϕ‘

-

Slope of Wöhler curve

4.1.3.5

ψ

-

Dynamic coefficient for hoist motion : Ratio of stresses at plate edges in buckling calculation

2.2 .2.1.1 3.4

ψh

-

Dynamic coefficient when calculating loads due to acceleration of horizontal motions

A-2.2.3.-2

τwk

Ω

-

Tolerance factor in bolted joints

3.2.2.2.1

ω

-

Crippling coefficient

3.3

ω

s-1

Angular velocity of shaft when calculating loads due to horizontal

1 - 19


motion

A-2.2.3.-3

ω1, ω2, ωr

s-1

Frequencies of oscillation during load swing

A-2.2.3.-2.2

ωm

s-1

Angular velocity of motor

A-2.2.3.-2.1

1 - 20


FEDERATION EUROPEENNE DE LA MANUTENTION SECTION I HEAVY LIFTING APPLIANCES

F.E.M. 1.001

3rd

EDITION REVISED

1998.10.01


CLASSIFICATION AND LOADING ON STRUCTURES AND MECHANISMS



Booklet 2

CLASSIFICATION AND LOADING ON STRUCTURES AND MECHANISMS

2.1 GROUP CLASSIFICATION OF HOISTING APPLIANCES AND THEIR COMPONENT PARTS .......................................................................................................................................................................4 2.1.1. GENERAL PLAN OF CLASSIFICATION .................................................................................4 2.1.2. CLASSIFICATION OF HOISTING APPLIANCES AS A WHOLE..........................................4 2.1.2.1. 2.1.2.2. 2.1.2.3. 2.1.2.4. 2.1.2.5.

CLASSIFICATION SYSTEM ...................................................................................................... 4 CLASSES OF UTILIZATION...................................................................................................... 4 LOAD SPECTRUM .................................................................................................................... 5 GROUP CLASSIFICATION OF HOISTING APPLIANCES............................................................ 7 GUIDANCE ON GROUP CLASSIFICATION OF AN APPLIANCE................................................. 7

2.1.3. CLASSIFICATION OF INDIVIDUAL MECHANISMS AS A WHOLE .....................................7 2.1.3.1. 2.1.3.2. 2.1.3.3. 2.1.3.4. 2.1.3.5.

CLASSIFICATION SYSTEM ...................................................................................................... 7 CLASSES OF UTILIZATION...................................................................................................... 9 LOADING SPECTRUM............................................................................................................... 9 GROUP CLASSIFICATION OF INDIVIDUAL MECHANISMS AS A WHOLE............................... 10 GUIDANCE FOR GROUP CLASSIFICATION OF INDIVIDUAL MECHANISMS AS A WHOLE .... 10

2.1.4. CLASSIFICATION OF COMPONENTS .................................................................................12 2.1.4.1. 2.1.4.2. 2.1.4.3. 2.1.4.4.

CLASSIFICATION SYSTEM .................................................................................................... 12 CLASSES OF UTILIZATION.................................................................................................... 12 STRESS SPECTRUM............................................................................................................... 13 GROUP CLASSIFICATION OF COMPONENTS ........................................................................ 14


2.2. LOADS ENTERING INTO THE DESIGN OF STRUCTURES ..........................................................15 2.2.1. PRINCIPAL LOADS...................................................................................................................15 2.2.2. LOADS DUE TO VERTICAL MOTIONS..................................................................................16 2.2.2.1. LOADS DUE TO HOISTING OF THE WORKING LOAD............................................................. 16

2.2.2.1.1. VALUES OF THE DYNAMIC COEFFICIENT ψ ................................................................... 16

2.2.2.2. LOADS DUE TO ACCELERATION (OR DECELERATION) OF THE HOISTING MOTION AND TO VERTICAL SHOCK LOADINGS WHEN TRAVELLING ALONG RAIL TRACKS ...................................... 17 2.2.2.3. SPECIAL CASE........................................................................................................................ 17

2.2.3. LOADS DUE TO HORIZONTAL MOTIONS SH......................................................................19 2.2.3.1.

HORIZONTAL EFFECTS DUE TO ACCELERATION (OR DECELERATION)........................ 19

2.2.3.1.1. TRAVERSE AND TRAVEL MOTIONS................................................................................ 19 2.2.3.1.2. SLEWING AND LUFFING (DERRICKING) MOTIONS ......................................................... 20

2.2.3.2. EFFECTS OF CENTRIFUGAL FORCE....................................................................................... 20 2.2.3.3. TRANSVERSE REACTIONS DUE TO ROLLING ACTION.......................................................... 21 2.2.3.4. BUFFER EFFECTS ST............................................................................................................... 21 2.2.3.4.1. BUFFER EFFECTS ON THE STRUCTURE ........................................................................ 21 2.2.3.4.2. BUFFER EFFECTS ON THE SUSPENDED LOAD ............................................................. 22

2.2.4. LOADS DUE TO CLIMATIC EFFECTS...................................................................................22 2.2.4.1 WIND ACTION........................................................................................................................... 22 2.2.4.1.1. WIND PRESSURE......................................................................................................... 22 2.2.4.1.2. DESIGN WIND CONDITIONS ............................................................................................ 23 2.2.4.1.2.1. In-service wind ................................................................................................................ 23 2.2.4.1.2.2. Wind out of service .......................................................................................................... 24

2.2.4.1.3. WIND LOAD CALCULATIONS........................................................................................... 25 2.2.4.1.4. SHAPE COEFFICIENTS.................................................................................................... 25 2.2.4.1.4.1. 2.2.4.1.4.2. 2.2.4.1.4.3. 2.2.4.1.4.4.

Individual members, frames, etc. ........................................................................................ 25 Multiple frames of members : shielding factors...................................................................... 28 Lattice towers.................................................................................................................. 29 Parts inclined in relation to the wind direction ....................................................................... 29

2.2.4.2. SNOW LOAD........................................................................................................................... 30 2.2.4.3. TEMPERATURE VARIATIONS.................................................................................................. 30

2.2.5 MISCELLANEOUS LOADS.......................................................................................................30 2.2.5.1. LOADS CARRIED BY PLATFORMS......................................................................................... 30

2.3. CASES OF LOADING ........................................................................................................................31 2.3.1. CASE I : APPLIANCE WORKING WITHOUT WIND.............................................................31 2.3.2. CASE II : APPLIANCE WORKING WITH WIND ....................................................................31 2.3.3. CASE III : APPLIANCE SUBJECTED TO EXCEPTIONAL LOADINGS .............................32 2.3.4. CHOOSING THE AMPLIFYING COEFFICIENT γC ...............................................................33 2.4. SEISMIC EFFECTS ............................................................................................................................33 2.5. LOADS ENTERING INTO THE DESIGN OF MECHANISMS..........................................................34 2.5.1. TYPE SM LOADS........................................................................................................................34 2.5.2. TYPE SR LOADS........................................................................................................................34

2-2


2.6. CASES OF LOADING ........................................................................................................................35 2.6.1. CASE I - NORMAL SERVICE WITHOUT WIND....................................................................35 2.6.1.1. TYPE SM LOADS..................................................................................................................... 35 2.6.1.2. TYPE SR LOADS ..................................................................................................................... 35

2.6.2. CASE II - NORMAL SERVICE WITH WIND ...........................................................................36 2.6.2.1. TYPE SM LOADS..................................................................................................................... 36 2.6.2.2. TYPE SR LOADS ..................................................................................................................... 36

2.6.3. CASE III - EXCEPTIONAL LOADS ..........................................................................................37 2.6.3.1. TYPE SM LOADS..................................................................................................................... 37 2.6.3.2. TYPE SR LOADS ..................................................................................................................... 37

2.6.4. APPLICATION OF THE ABOVE CONSIDERATIONS FOR CALCULATING SM ..............37 2.6.4.1. HOISTING MOTIONS................................................................................................................ 38 2.6.4.2. HORIZONTAL MOTIONS.......................................................................................................... 38 2.6.4.3. COMBINED MOTIONS .............................................................................................................. 39

.....................................................................................................................................................................39 APPENDIX..................................................................................................................................................40 A.2.1.1. - HARMONISATION OF THE CLASSES OF UTILIZATION OF APPLIANCES AND MECHANISMS.......................................................................................................................................40 A.2.2.3. - CALCULATION OF LOADS DUE TO ACCELERATIONS OF HORIZONTAL MOTIONS .................................................................................................................................................................45

2-3


2.1 GROUP CLASSIFICATION OF HOISTING APPLIANCES AND THEIR COMPONENT PARTS 2.1.1. GENERAL PLAN OF CLASSIFICATION In the design of a hoisting appliance and its component parts, account must be taken of the duty which they will be required to perform during their duration of use ; for this purpose group classification is employed of : - the appliance as a whole ; - the individual mechanisms as a whole ; - the structural and mechanical components. This classification is based on two criteria, namely : - the total duration of use of the item considered ; - the hook load, loading or stress spectra to which the item is subjected.

2.1.2. CLASSIFICATION OF HOISTING APPLIANCES AS A WHOLE 2.1.2.1. CLASSIFICATION SYSTEM Appliances as a whole are classified in eight groups, designated by the symbol A1, A2, ..., A8 respectively (see section 2.1.2.4.), on the basis of ten classes of utilization and four load spectra. 2.1.2.2. CLASSES OF UTILIZATION By duration of use of a hoisting appliance is meant the number of hoisting cycles which the appliance performs. A hoisting cycle is the entire sequence of operations commencing when a load is hoisted and ending at the moment when the appliance is ready to hoist the next load. The total duration of use is a computed duration of use, considered as a guide value, commencing when the appliance is put into service and ending when it is finally taken out of service. On the basis of the total duration of use, we have ten classes of utilization, designated by the symbol U0, U1, ..., U9. They are defined in table T.2.1.2.2.

2-4


Table T.2.1.2.2. - Classes of utilization

Symbol U0 U1 U2 U3 U4 U5 U6 U7 U8 U9

Total duration of use (number nmax of hoisting cycles) ≤ 16 000 nmax nmax ≤ 16 000 32 000 < nmax ≤ 32 000 63 000 < nmax ≤ 63 000 125 000 < nmax ≤ 125 000 250 000 < nmax ≤ 250 000 500 000 < nmax ≤ 500 000 1 000 000 < nmax ≤ 1 000 000 2 000 000 < nmax ≤ 2 000 000 4 000 000 < nmax 4 000 000 <

2.1.2.3. LOAD SPECTRUM The load spectrum characterises the total number of loads hoisted during the total duration of use (see 2.1.2.2.) of an appliance. It is a distribution function (summed) y = f(x), expressing the fraction x (O ≤ x ≤ 1) of the total duration of use, during which the ratio of the hoisted load to the safe working load attains at least a given value y (O ≤ y ≤ 1). Examples of a load spectrum are given in figs. 2.1.2.3.1. - a and b.

Figure 2.1.2.3.1. - a

Figure 2.1.2.3.1. - b

ml = loads ; ml max = safe working load ; n = number of hoisting cycles in respect of which the hoisted load is greater than or equal to ml; nmax = number of hoisting cycles determining the total duration of use. Each spectrum is assigned a spectrum factor kp, defined by : 1

KP = ∫ yd dx 0

2-5


For the purposes of group classification the exponent d is taken by convention as equal to 3. In many applications the function f(x) may be approximated by a function consisting of a certain number r of steps (see fig. 2.1.2.3.2.), comprising respectively n1, n2, ..., nr hoisting cycles, the th load may be considered as practically constant and equal to mli during the ni cycles of the i step. If nmax represents the total duration of use and ml max the greatest among the mli loads, there exists a relation : r

n1 + n2 + ..... + nr =

∑

ni = nmax

i =1

or in approximated form : kp = ( ml1/mlmax)3.(n1/nmax) + ( ml2/mlmax)3.(n2/nmax) + .....+ ( mlr/mlmax)3.(nr/nmax) r

kp =

∑

[ ( mli/mlmax)3.(ni/nmax) ]

i =1

Figure 2.1.2.3.2. According to its load spectrum, a hoisting appliance is placed in one of the four spectrum classes Q1, Q2, Q3, Q4 defined in table T.2.1.2.3. Table T.2.1.2.3. - Spectrum classes Symbol Q1 Q2 Q3 Q4

0,125 0,250 0,500

Spectrum factor kp ≤ kp ≤ < kp ≤ < kp ≤ < kp

2-6

0,125 0,250 0,500 1,000


2.1.2.4. GROUP CLASSIFICATION OF HOISTING APPLIANCES Group classification of hoisting appliances as a whole is determined from table T.2.1.2.4. Table T.2.1.2.4. - Appliance groups Load spectrum class Q1 Q2 Q3 Q4

Class of utilization U0 A1 A1 A1 A2

U1 A1 A1 A2 A3

U2 A1 A2 A3 A4

U3 A2 A3 A4 A5

U4 A3 A4 A5 A6

U5 A4 A5 A6 A7

U6 A5 A6 A7 A8

U7 A6 A7 A8 A8

U8 A7 A8 A8 A8

U9 A8 A8 A8 A8

2.1.2.5. GUIDANCE ON GROUP CLASSIFICATION OF AN APPLIANCE Directions concerning the classification of hoisting appliances are given in table T.2.1.2.5. Since appliances of the same type may be used in a wide variety of ways, the grouping shown in this table can only be taken as a model. In particular, where several groups are shown as appropriate to an appliance of a given type, it is necessary to ascertain, on the basis of the appliance's computed total duration of use and load spectrum, in which classes of utilization and load spectrum it has to be placed, and consequently in which group.

2.1.3. CLASSIFICATION OF INDIVIDUAL MECHANISMS AS A WHOLE 2.1.3.1. CLASSIFICATION SYSTEM Individual mechanisms as a whole are classified in eight groups, designated respectively by the symbol M1, M2, ..., M8 (see 2.1.3.4.), on the basis of ten classes of utilization and four classes of loading spectrum.

2-7


Table T.2.1.2.5. - Guidance for group classification of appliances Type of appliance Reference 1 2 3 4 5 6 7 8 9 10 11 12.a 12.b 13 14 15 16 17

18 19 20 21 22

Particulars concerning nature of use (1)

Designation Hand-operated appliances Erection cranes Erection and dismantling cranes for power stations, machine shops, etc. Stocking and reclaiming transporters Hook duty Stocking and reclaiming transporters Grab or magnet Workshop cranes Overhead travelling cranes, pigbreaking cranes, scrapyard cranes Ladle cranes Soaking-pit cranes Stripper cranes, open-hearth furnacecharging cranes Forge cranes Bridge cranes for unloading, bridge cranes for containers Other bridge cranes (with crab and/or slewing jib crane) Bridge cranes for unloading, bridge cranes (with crab and/or slewing jib crane) Drydock cranes, shipyard jib cranes, jib cranes for dismantling Dockside cranes (slewing, on gantry), floating cranes and pontoon derricks Dockside cranes (slewing, on gantry), floating cranes and pontoon derricks Floating cranes and pontoon derricks for very heavy loads (usually greater than 100 t) Deck cranes Deck cranes

Appliance group (see 2.1.2.4.) A1 - A2 A1 - A2 A2 - A4 A5 A6 - A8 A3 - A5

Grab or magnet A6 - A8 A6 - A8 A8 A8 A6 - A8 Hook or spreader duty Hook duty Grab or magnet

A5 - A6 A4 A6 - A8

Hook duty

A3 - A5

Hook duty Grab or magnet

A5 - A6 A6 - A8


Tower cranes for building Derricks Railway cranes allowed to run in train

A2 - A3 A3 - A4 A4 - A5 A3 - A4 A2 - A3 A4

(1) Only a few typical cases of use are shown, by way of guidance, in this column.

2-8


2.1.3.2. CLASSES OF UTILIZATION By duration of use of a mechanism is meant the time during which the mechanism is actually in motion. The total duration of use is a calculated duration of use, considered as a guide value, applying up to the time of replacement of the mechanism. It is expressed in terms of hours. On the basis of this total duration of use, we have ten classes of utilization, TO, T1, T2, ..., T9. They are defined in table T.2.1.3.2. Table T.2.1.3.2. - Classes of utilization Symbol T0 T1 T2 T3 T4 T5 T6 T7 T8 T9

Total duration of use T T ≤ 200 < T ≤ 400 < T ≤ 800 < T ≤ 1 600 < T ≤ 3 200 < T ≤ 6 300 < T ≤ 12 500 < T ≤ 25 000 < T ≤ 50 000 < T

(h) 200 400 800 1600 3200 6300 12 500 25 000 50 000

2.1.3.3. LOADING SPECTRUM The loading spectrum characterizes the magnitude of the loads acting on a mechanism during its total duration of use. It is a distribution function (summed) y = f(x), expressing the fraction x (0 ≤ x ≤ 1) of the total duration of use, during which the mechanism is subjected to a loading attaining at least a fraction y (0 ≤ y ≤ 1) of the maximum loading (see figure 2.1.2.3.1.). Each spectrum is assigned a spectrum factor km defined by : 1

km = ∫ yd dx 0

For the purposes of group classification, d is taken by convention as equal to 3. In many applications the function f(x) may be approximated by a function consisting of a certain number r of steps (see fig. 2.1.2.3.2.), of respective durations tl, t2, ..., tr, the loadings S may be considered as practically constant and equal to Si during the duration ti. If T represents the total duration of use and Smax the greatest of the loadings S1, S2, ..., Sr, there exists a relation : t1 + t2 + ... + tr =

r

∑ i =1

2-9

ti = T


and in approximated form : km = (S1/Smax)3 ( t1/T) + (S2/Smax)3 ( t2/T) + ..... +(Sr/Smax)3 ( tr/T) =

r

∑

[ (Si/Smax)3 ( ti/T) ]

i =1

Depending on its loading spectrum, a mechanism is placed in one of the four spectrum classes L1, L2, L3, L4, defined in table T.2.1.3.3. Table T.2.1.3.3. - Spectrum classes Symbol L1 L2 L3 L4

0,125 0,250 0,500

Spectrum factor km km ≤ < km ≤ < km ≤ < km ≤

0,125 0,250 0,500 1,000

2.1.3.4. GROUP CLASSIFICATION OF INDIVIDUAL MECHANISMS AS A WHOLE On the basis of their class of utilization and their spectrum class, individual mechanisms as a whole are classified in one of the eight groups M1, M2, ..., M8, defined in table T.2.1.3.4. Table T.2.1.3.4. - Mechanism groups Class of load spectrum L1 L2 L3 L4

Class of utilization T0 M1 M1 M1 M2

T1 M1 M1 M2 M3

T2 M1 M2 M3 M4

T3 M2 M3 M4 M5

T4 M3 M4 M5 M6

T5 M4 M5 M6 M7

T6 M5 M6 M7 M8

T7 M6 M7 M8 M8

T8 M7 M8 M8 M8

T9 M8 M8 M8 M8

2.1.3.5. GUIDANCE FOR GROUP CLASSIFICATION OF INDIVIDUAL MECHANISMS AS A WHOLE Guidance for group classification of an individual mechanism as a whole is given in table T.2.1.3.5. Since appliances of the same type may be used in a wide variety of ways, the grouping directions in this table can only be taken as a model. In particular, where several groups are shown as appropriate to a mechanism of a given type, it is necessary to ascertain, on the basis of the mechanism's calculated total duration of use and loading spectrum, in which class of utilization (see 2.1.3.2.) and spectrum (see 2.1.3.3.) it has to be placed, and consequently in which group of mechanisms (see 2.1.3.4.).

2 - 10


Table T.2.1.3.5. - Guidance for group classification of a mechanism Type of appliance Reference 1 2 3 4

Designation Hand-operated appliances Erection cranes Erection and dismantling cranes for power stations, machine shops, etc. Stocking and reclaiming transporters

5

Stocking and reclaiming transporters

6 7

Workshop cranes Overhead travelling cranes, pigbreaking cranes, scrapyard cranes Ladle cranes Soaking-pit cranes Stripper cranes, open-hearth furnace-charging cranes Forge cranes Bridge cranes for unloading, bridge cranes for containers Other bridge cranes (with crab and/or slewing jib crane) Bridge cranes for unloading, bridge cranes (with crab and/or slewing jib crane) Drydock cranes, shipyard jib cranes, jib cranes for dismantling Dockside cranes (slewing, on gantry, etc.), floating cranes and pontoon derricks Dockside cranes (slewing, on gantry, etc.), floating cranes and pontoon derricks Floating cranes and pontoon derricks for very heavy loads (usually greater than 100 t) Deck cranes Deck cranes

8 9 10 11 12-a 12-b 13 14 15 16 17 18 19 20 21 22

Particulars concerning nature of use (1)

Hoisting M1 M2-M3

Type of mechanism SleLuffing Traver- Travel wing se M1 M1 M2-M3 M1-M2 M1-M3 M2-M3

M2

-

-

M2

M2

M5-M6

M4

-

M4-M5 M5-M6

M7-M8 M6

M6 M4

-

M8

M6

-

M6-M7 M7-M8 M4 M5 M7-M8 M6-M7

M7-M8 M8

M6

-

M8 M8

M6 --


Grab or magnet

a - Hook or spreader duty b - Hook duty Grab or magnet

M4-M5 M6-M7 M7 M8 M7 M5

-

M8 M6

M6-M7

M5-M6 M3-M4

M6-M7 M4-M5

M4-M5

M4-M5

M4-M5 M4-M5

-

M8

M5-M6 M3-M4

M7-M8 M4-M5

M5-M6

M4-M5 M4-M5

M4-M5 M5-M6

M6-M7

M5-M6 M5-M6

-

M3-M4

M7-M8

M6-M7 M6-M7

-

M4-M5

M3-M4

M3-M4 M3-M4

-

-

M4

M3-M4 M3-M4

M2

M3

M5-M6 M4 M2-M3

M3-M4 M3-M4 M5 M4 M1-M2 M1-M2

M3-M4

M2-M3 M2-M3

Hook duty Hook duty Grab or magnet


Tower cranes for building Derricks Railway cranes allowed to run in train

M4-M5 M3-M4 M3 M3 -

(1) Only a few typical cases of use are shown, by way of guidance, in this column.

2 - 11

-

-


2.1.4. CLASSIFICATION OF COMPONENTS 2.1.4.1. CLASSIFICATION SYSTEM Components, both structural and mechanical, are classified in eight groups, designed respectively by the symbol E1, E2, ..., E8, on the basis of eleven classes of utilization and four classes of stress spectrum. 2.1.4.2. CLASSES OF UTILIZATION By duration of use of a component is meant the number of stress cycles to which the component is subjected. A stress cycle is a complete set of successive stresses, commencing at the moment when the stress under consideration exceeds the stress σm defined in fig. 2.1.4.3. and ending at the moment when this stress is, for the first time, about to exceed again σm in the same direction. Fig. 2.1.4.3. therefore represents the trend of the stress σ over a duration of use equal to five stress cycles. The total duration of use is a computed duration of use, considered as a guide value, applying up to the time of replacement of the component. In the case of structural components the number of stress cycles is in a constant ratio with the number of hoisting cycles of the appliance. Certain components may be subjected to several stress cycles during a hoisting cycle depending on their position in the structure. Hence the ratio in question may differ from one component to another. Once this ratio is known, the total duration of use of the component is derived from the total duration of use which determined the class of utilization of the appliance. As regards mechanical components, the total duration of use is derived from the total duration of use of the mechanism to which the component under consideration belongs, account being taken of its speed of rotation and/or other circumstances affecting its operation. On the basis of the total duration of use, we have eleven classes of utilization, designated respectively by the symbol BO, B1, ..., B10. They are defined in table T.2.1.4.2. Table T.2.1.4.2. - Classes of utilization Symbol B0 B1 B2 B3 B4 B5 B6 B7 B8 B9 B10

Total duration of use (number n of stress cycles) n ≤ 16 000 16 000 < n ≤ 32 000 32 000 < n ≤ 63 000 63 000 < n ≤ 125 000 125 000 < n ≤ 250 000 250 000 < n ≤ 500 000 500 000 < n ≤ 1 000 000 1 000 000 < n ≤ 2 000 000 2 000 000 < n ≤ 4 000 000 4 000 000 < n ≤ 8 000 000 8 000 000 < n

2 - 12


2.1.4.3. STRESS SPECTRUM The stress spectrum characterizes the magnitude of the load acting on the component during its total duration of use. It is a distribution function (summed) y = f(x), expressing the fraction x (O ≤ x ≤ 1) of the total duration of use (see 2.1.4.2.), during which the component is subjected to a stress attaining at least a fraction y (O ≤ x ≤ 1) of the maximum stress. Each stress spectrum is assigned a spectrum factor ksp, defined by ksp =

∫

1

0

yc dx

Where c is an exponent depending on the properties of the material concerned, the shape and size of the component in question, its surface roughness and its degree of corrosion (see booklet 4). In many applications the function f(x) may be approximated by a function consisting of a certain number r of steps, comprising respectively n1, n2, ..., nr stress cycles ; the stress σ may be considered as practically constant and equal to σi during ni cycles. If n represents the total duration of use and σmax the greatest of the stresses σ1, σ2, ..., σr there exists a relation : r

n1 + n2+ ..... + nr =

∑

ni = n

i =1

and in approximated form : ksp = (σ1/ σmax)c (n1 / n) + (σ2/ σmax)c (n2 / n) + ..... +(σr/ σmax)c (nr / n) =

r

∑

[ (σi/ σmax)c (ni / n) ]

i =1

Depending on its stress spectrum, a component is placed in one of the spectrum classes P1, 1 P2, P3, P4, defined in table T.2.1.4.3. Table T.2.1.4.3. - Spectrum classes Symbol P1 P2 P3 P4

Spectrum factor ksp ≤ 0,125 ksp ≤ 0,250 0,125 < ksp ≤ 0,500 0,250 < ksp ≤ 1,000 0,500 < ksp

1

There are components, both structural and mechanical, such as spring loaded components, which are subjected to loading that is quite or almost independent of the working load. Special care shall be taken in classifying such components. In most cases ksp = 1 and they belong to class P4. 2 - 13


For structural components, the stresses to be taken into consideration for determination of the spectrum factor are the differences σsup - σm between the upper stresses σsup and the average stress σm, these concepts being defined by fig. 2.1.4.3. representing the variation of the stress over time during five stress cycles.

Fig. 2.1.4.3. - Variation of stress as a function of time during a five stress cycles σsup = upper stress

σsup max = maximum upper stress

σsup min = minimum upper stress

σinf = longer stress

σm = arithmetic mean of all upper and lower stresses during the total duration of use In the case of mechanical components, we can put σm = 0 the stresses to be introduced into the calculation of the spectrum factor then being the total stresses occurring in the relevant section of the component. 2.1.4.4. GROUP CLASSIFICATION OF COMPONENTS On the basis of their class of utilization and their stress spectrum class, components are classified in one of the eight groups E1, E2, ..., E8, defined in table T.2.1.4.4. Table T.2.1.4.4. - Component groups Stress Spectrum class

Class of utilization B0

B1

B2

B3

B4

B5

B6

B7

B8

B9

B10

P1 P2 P3 P4

E1 E1 E1 E1

E1 E1 E1 E2

E1 E1 E2 E2

E1 E2 E3 E4

E2 E3 E4 E5

E3 E4 E5 E6

E4 E5 E6 E7

E5 E6 E7 E8

E6 E7 E8 E8

E7 E8 E8 E8

E8 E8 E8 E8

2 - 14


2.2. LOADS ENTERING INTO THE DESIGN OF STRUCTURES The structural calculations shall be conducted by determining the stresses developed in an appliance during its operation. These stresses shall be calculated on the basis of the loads defined below : a) The principal loads exerted on the structure of the appliance, assumed to be stationary, in the most unfavourable state of loading ; b) Loads due to vertical motions ; c) Loads due to horizontal motions ; d) Loads due to climatic effects. The various loads, the factors to be applied, and the practical method of conducting the calculations are examined below. In what follows, the definitions given below are used : Working load : Weight of the useful load lifted, plus the weight of the accessories (sheaves blocks, hooks, lifting beams, grab, etc.). Dead load : Dead weight of components acting on a given member, excluding the working load.

2.2.1. PRINCIPAL LOADS The principal loads include : - the loads due to the dead weight of the components : SG - the loads due to the working load : SL all movable parts being assumed to be in their most unfavourable position. Each structural member shall be designed for the position of the appliance and magnitude of the working load (between zero and the safe working load) which gives rise to the maximum 2 stresses in the member in question.

2

(1) In certain cases, the maximum stress may be obtained with no working load.

2 - 15


2.2.2. LOADS DUE TO VERTICAL MOTIONS These loads stem from picking up the working load more or less suddenly from accelerations (or decelerations) of the hoisting motion, and from vertical shock loadings due to travelling along rails tracks. 2.2.2.1. LOADS DUE TO HOISTING OF THE WORKING LOAD Account shall be taken of the oscillations caused when lifting the load by multiplying the loads due to the working load by a factor called the "dynamic coefficient" ψ. 2.2.2.1.1. VALUES OF THE DYNAMIC COEFFICIENT ψ

The value of the dynamic coefficient ψto be applied to the load arising from the working load is given by the expression : ψ = 1 + ξ VL Where : VL is the hoisting speed in m/s, and ξ an experimentally determined coefficient 3 The following values shall be adopted : ξ = 0,6 for overhead travelling cranes and bridge cranes

ξ = 0,3 for jib cranes.

The maximum figure to be taken for the hoisting speed when applying this formula is 1 m/s. For higher speeds, the dynamic coefficient ψ is not further increased. The value to be applied for the coefficient ψ in the calculations shall in no case be less than 1,15. The values of ψ are given in the curves of figure 2.2.2.1.1. in terms of hoisting speeds VL.

A : Overhead travelling cranes, Bridge cranes B : Jib cranes Figure 2.2.2.1.1. - Values of dynamic coefficient ψ 3

In certain cases, the maximum stress may be obtained with no working load.

2 - 16


Note - The above mentioned coefficient ξ is not the same for "overhead travelling cranes and bridge cranes" and for "jib cranes". The difference arises from the fact that the dynamic coefficient ψ is, other things being equal, smaller when the hoisting load is carried by a member having some flexibility, as in jib cranes where the jib is never rigid. In a similar way, use of the coefficient ψ as indicated for jib cranes may be extended to certain other appliances such as, for example, transporters for the design case corresponding to load on the cantilever boom ; the value of ψ indicated for overhead travelling cranes should, of course, be used for the design cases where the load is applied between the legs of the machine as the rigidity of the structure at this point is comparable with that one of an overhead travelling crane girder. 2.2.2.2. LOADS DUE TO ACCELERATION (OR DECELERATION) OF THE HOISTING MOTION AND TO VERTICAL SHOCK LOADINGS WHEN TRAVELLING ALONG RAIL TRACKS Since the coefficient ψ takes account of the degree of snatch on the working load which is the largest shock loading, loads due to acceleration (or deceleration) of the hoisting motion and the 4 vertical reactions due to travelling along tracks, assumed to be properly laid, shall be neglected 2.2.2.3. SPECIAL CASE In the case of certain appliances, the loads due to the dead loads are of opposite sign to those due to the working load, in which case a comparison must be made between the loading figure obtained in the "appliance under load" condition, with the dynamic coefficient ψ applied to the working load, and the loading figure obtained in the "no-load" condition, taking into account the oscillations resulting from setting clown the load, as follows : Let : S G be the algebraic value of the loads due to the dead load S L be the algebraic value of the loads due to the working load. S G - S L (ψ-1)/2 Which is compared with the load for the "appliance under load" condition determined by the expression : 4

This assumes that the rail joints are in good condition. The detrimental effect on hoisting appliances of rail tracks in poor condition is so great, both for the structure and the machinery, that it is necessary to stipulate that the rail joints must be maintained in good condition : no shock loading coefficient can allow for the damage caused by faulty joints. In so far as high speed appliances are concerned, the best solution is to butt-weld the rails, in order to eliminate entirely the shock loadings which occur when an appliance runs over joints.

The amplified total load, when setting clown the load is obtained by the expression : 2 - 17


SG + ψ SL

2 - 18


the component being finally designed on the basis of the more unfavourable of these two values. Note - This formula is based on the fact that the dynamic coefficient determines the maximum amplitude of the oscillations set up in the structure when the load is picked up. The amplitude of the oscillation is given by : S L (ψ-1) It is assumed that the amplitude of the oscillation set up in the structure when the load is set clown is half that of the oscillation caused when hoisting takes place. The ultimate state of loading is therefore : S G - S L (ψ-1)/2 which must be compared with the state of loading given by : SG + ψ SL

Hoisting and lowering curve when SL and SG are of opposite sign

2 - 19


2.2.3. LOADS DUE TO HORIZONTAL MOTIONS SH The loads due to horizontal motions are as follows : 1) The inertia effects due to acceleration (or deceleration) of the traverse, travel, slewing or luffing motions. These effects can be calculated in terms of the value of the acceleration (or deceleration). 2) The effects of centrifugal force. 3) Transverse horizontal reactions resulting from rolling action. 4) Buffer effects. 2.2.3.1.

HORIZONTAL EFFECTS DUE TO ACCELERATION (OR DECELERATION)

The loads due to the accelerations (or decelerations) imparted to the movable elements when starting or braking are calculated for the various structural members. 2.2.3.1.1. TRAVERSE AND TRAVEL MOTIONS

For these motions the calculation is made by considering a horizontal force applied at the tread of the driven wheels parallel to the rail. The loads shall be calculated in terms of the acceleration (or deceleration) time assumed according to the working conditions and the speeds to be attained. From it is deduced the value (in m/s2) of the acceleration to be used for calculating the horizontal force according to the masses to be set in motion. Note - If the speed and acceleration values are not specified by the user, acceleration times corresponding to the speeds to be reached may be chosen according to the three following working conditions : a) Appliances of low and moderate speed with a great length of travel ; b) Appliances of moderate and high speed for normal applications ; c) High speed appliances with high acceleration. In the latter case, it is almost always necessary to drive all the rail wheels. Table T.2.2.3.1.1. gives the values of acceleration times and accelerations for the three conditions.

2 - 20


Table T.2.2.3.1.1. - Acceleration time and acceleration value (a) low and moderate speed Speed with long travel to be reached Acceleration Acceleration time m/s s m/s2 4,00 3,15 2,5 0,22 9,1 2 0,19 8,3 1,60 0,15 6,6 1,00 0,12 5,2 0,63 0,098 4,1 0,40 0,078 3,2 0,25 0,064 2,5 0,16

(b) moderate and high speed (normal applications) Acceleration Acceleration time s m/s2 0,50 8,0 0,44 7,1 0,39 6,3 0,35 5,6 0,32 5,0 0,25 4,0 0,19 3,2 0,16 2,5

(c) high speed with high accelerations Acceleration Acceleration time s m/s2 6,0 0,67 5,4 0,58 4,8 0,52 4,2 0,47 3,7 0,43 3,0 0,33

The horizontal force to be taken into account shall be not less than 1/30th nor more than 1/4 of the load on the driven or braked wheels. 2.2.3.1.2. SLEWING AND LUFFING (DERRICKING) MOTIONS

For slewing and luffing motions the calculations shall be based on the accelerating (or decelerating) torque applied to the motor shaft of the mechanisms. The rates of acceleration will depend upon the appliance ; for a normal crane a value between 2 2 0.1 m/s and 0.6 m/s , according to the speed and radius, may be chosen for the acceleration at the jib head so that an acceleration time of from 5 to 10 s is achieved. Note - A method for calculating the effects of acceleration of horizontal motions is given in appendix A.2.2.3. 2.2.3.2. EFFECTS OF CENTRIFUGAL FORCE In the case of jib cranes, account shall be taken of the centrifugal force due to slewing. In practice, it is sufficient to determine the horizontal force exerted at the jib head as a result of the inclination of the rope carrying the load and in general to neglect the effects of centrifugal force on the other elements of the crane.

2 - 21


2.2.3.3. TRANSVERSE REACTIONS DUE TO ROLLING ACTION When two wheels (or two bogies) roll along a rail, the couple formed by the horizontal forces normal to the rail shall be taken into consideration. The components of this couple are obtained by multiplying the vertical load exerted on the wheels (or bogies) by a coefficient λ which depends upon the ratio of the span p to the wheel base a 5. As shown in the graph, this coefficient lies between 0.05 and 0.2 for ratios p/a between 2 and 8.

2.2.3.4. BUFFER EFFECTS ST The case must be considered when the impact due to collision with buffers is applied to the structure, and the case when it is applied to the suspended load. 2.2.3.4.1. BUFFER EFFECTS ON THE STRUCTURE

A distinction must be drawn between : 1) the case in which the suspended load can swing. 2) that in which rigid guides prevent swing. In the first case the following rules shall be applied : For horizontal speeds below 0.7 m/sec, no account shall be taken of buffer effects. For speeds in excess of 0.7 m/sec, account must be taken of the reactions set up in the structure by collisions with buffers. It shall be assumed that a buffer is capable of absorbing the kinetic energy of the appliance (without the working load) at a fraction of the rated speed Vt fixed at 0.7 Vt.

5

By "wheelbase" is understood the centre distance between the outermost pairs of wheels, or, in the case of bogies, the centre distance between the fulcrum pins on the crane structure of the two bogies or bogie systems. Where horizontal guilding wheels are provided, the wheelbase shall be the distance between the rail contact points of two horizontal wheels.

2 - 22


The resulting loads set up in the structure shall be calculated on the basis of the retardation imparted to the appliance by the buffer in use. However, for higher speeds (greater than 1 m/sec), the use of decelerating devices which act upon approach to the ends of the track is permitted provided the action of these devices is automatic and they produce an effective deceleration of the appliance which always reduces the speed to the predetermined lower value before the buffers are reached. In this case the reduced speed obtained after slowing down is used for the value of Vt when 6 calculating the buffer effect . In the second case where the load cannot swing the buffer effect is calculated in the same manner but taking account of the value of the working load. (1) It must be emphasised that a sure and effective device must be fitted. A mere end-of-travel limit switch cutting off the power supply to the motor is not sufficient reason to assume reduced speed for the buffer effect. 2.2.3.4.2. BUFFER EFFECTS ON THE SUSPENDED LOAD

Impacts due to collision between the load and fixed obstructions are taken into account only for appliances where the load is rigidly guided. In that case, the loads generated by such a collision are to be taken into consideration. The loads can be computed by considering that horizontal force applied at the lever of the load which is capable of causing two of the crab wheels to lift.

2.2.4. LOADS DUE TO CLIMATIC EFFECTS The loads due to climatic effects are those resulting from the action of the wind, from snow loads and from temperature variations. 2.2.4.1 WIND ACTION INTRODUCTION This clause relates to wind loads on crane structure. It gives a simplified method of calculation and assumes that the wind can blow horizontally from any direction, that the wind blows at a constant velocity and that there is a static reaction to the loadings it applies to the crane structure. 2.2.4.1.1.

WIND PRESSURE

The dynamic wind pressure is given by :

q = 0.613 Vs2

Where q is the dynamic pressure N/m2., and Vs is the design wind speed in m/s.

6

Where a wind speed measuring device is to be attached to an appliance it shall normally be placed at the maximum height of the appliance. In cases where the wind speed at a different lever is more significant to the safety of the appliance, the manufacturer shall state the height at which the device shall be placed.

2 - 23


2.2.4.1.2. DESIGN WIND CONDITIONS

Two design wind conditions are taken into account in calculating wind loads on cranes. 2.2.4.1.2.1. In-service wind

This is the maximum wind in which the crane is designed to operate. The wind loads are assumed to be applied in the least favourable direction in combination with the appropriate service loads. In-service design wind pressures and corresponding speeds are given in table 7 T.2.2.4.1.2.1. They are assumed to be constant over the height of the appliance . It is assumed that the operating speeds and nominal accelerations are not necessarily reached under extreme wind conditions. Table T.2.2.4.1.2.1. - In-service design wind pressure

Type of appliance Lifting appliance easily protected against wind action or designed for use exclusively in light wind. Erection operations. All normal types of crane installed in the open * Appliances which must continue to work in high winds

Wind pressure in service N/m2

Wind speed in service m/s

125

14

250

20

500

28

* For example appliances of type 12a in table T.2.1.2.5. Action of wind on the load The action of the wind on the hook load for a crane which handles miscellaneous loads shall be determined from the relationship : F = 2,5 . A . q where : F is the force exerted by the wind on the hook load in N,

7

(1) Where, exceptionally, a crane is required to handle loads of large surface area, it is admissible for the manufacturer to determine a wind speed less than that specified in table T.2.2.4.1.2.1. above which such loads shall not be handled.

2 - 24


q is the in-service wind pressure from table 2.2.4.1.2.1. in N/m2, A is the maximum area of the solid parts of the hook load in m2 8. Where this area is not 2 known, a minimum value of 0.5 m per tonne of safe working load shall be used. Where a crane is designed to handle loads of a specific size and shape only, the wind loading shall be calculated for the appropriate dimensions and configurations. 2.2.4.1.2.2. Wind out of service

This is a maximum (storm) wind for which the lifting machine is designed to remain stable in out of service conditions, as indicated, by the manufacturer. The speed varies with the height of the apparatus above the surrounding ground lever, the geographical location and the degree of exposure to the prevailing winds. For lifting appliances used in the open air, the normal theoretical wind pressure and the corresponding speed, for "out of service" conditions are indicated in the table T.2.2.4.1.2.2.

Table T.2.2.4.1.2.2. - Out of service wind Height above ground lever m 0 to 20 20 to 100 More than 100

Out of service design wind pressure 2 N/m 800 1 100 1 300

Approximate equivalent out of service design wind speed m/s 36 42 46

When calculating wind loads for out of service conditions the wind pressure may be taken as constant over the vertical height intervals in table T.2.2.4.1.2.2. Alternatively, the design wind pressure at the top of the crane may be assumed constant over its entire height. Where cranes are to be permanently installed or used for extended periods in areas where wind conditions are exceptionally severe, the above figures may be modified by agreement between the manufacturer and purchaser in the light of local meteorological data. For certain types of appliance of which the jib can be quickly lowered, (such as a tower crane which can be easily lowered by a built-in mechanism) the out of service wind need not be taken into consideration provided the machine is intended for lowering after each working day.

8

Where, execptionnally, a crane is required to handle loads of large surface area, it is admissible for the manufacturer to détermine a wind speed less than that spécifie in table T.2.2.4.1.2.1. above wich such loads shall not be handled. 2 - 25


2.2.4.1.3. WIND LOAD CALCULATIONS

For most complete and part structures, and individual members used in crane structures the wind load is calculated from : F = A . q . Cf Where : F is the wind load in N, 2 A is the effective frontal area of the part under consideration in m ,

q is the wind pressure corresponding to the appropriate design condition in N/m2, Cf is the shape coefficient in the direction of the wind for the part under consideration. The total wind load on the structure is taken as the sum of the loads on its component parts. In determining strength and stability requirements of the appliance the total wind load shall be considered. The magnitude of the wind load to be allowed for in the design of a mechanism, in determining the motor and brake requirements for the mechanism and to provide for the safety of the appliance in the wind, are given in the chapter dealing with the design of mechanisms. 2.2.4.1.4. SHAPE COEFFICIENTS 2.2.4.1.4.1. Individual members, frames, etc.

Shape coefficients for individual members, single lattice frames and machinery houses are given in table T.2.2.4.1.4.1. The values for individual members vary according to the aerodynamic slenderness and, in the case of large box sections, with the section ratio. Aerodynamic slenderness and section ratio are defined in figure 2.2.4.1.4.1. The wind load on single lattice frames may be calculated on the basis of the coefficients for the individual members given in the top part of table T.2.2.4.1.4.1. In this case the aerodynamic slenderness of each member shall be taken into account. Alternatively the overall coefficients for lattice frames constructed of flat sided and circular sections given in the middle part of the table may be used. Where a lattice frame is made up of flat-sided and circular sections, or of circular sections in 2 2 both flow regimes (D.VS < 6 m /s and D.VS ≥ 6 m /s) the appropriate shape coefficients are applied to the corresponding frontal areas. Where gusset plates of normal size are used in welded lattice construction no allowance for the additional area presented by the plates is necessary, provided the lengths of individual members are taken between the centres of node points.

2 - 26


Shape coefficients obtained from wind-tunnel or full-scale tests may also be used. Table T.2.2.4.1.4.1. - Force coefficients

Type

Description Rolled sections [ ]

≤5 1,15

Aerodynamic Slenderness l/b or l/D (1) 10 20 30 40 50 1,5 1,45 1,4 1,3 1,15

> 50 1,6

Rectangular hollow sections up to 356 mm square

1,4

1,45

1,5

1,55

1,55

1,55

1,6

1,05

1,05

1,2

1,3

1,4

1,5

1,6

1,30

1,35

1,60

1,65

1,70

1,80

1,80

0,60 0,60

0,70 0,65

0,80 0,70

0,85 0,70

0,90 0,75

0,90 0,80

0,90 0,80

1,55 1,40 1,0 0,80

1,75 1,55 1,20 0,90

1,95 1,75 1,30 0,90

2,10 1,85 1,35 1,0

2,20 1,90 1,40 1,0

and 254 x 457 mm rectangular Other sections Individual Circular sections where : members D.Vs < 6 m2/s D.Vs ≥ 6 m2/s b/d Rectangular 2 hollow sections 1 over 356 mm 0,5 square and 0,25 254 x 457 mm rectangular Wind →

b

d Single lattice frames

Flat-sided sections Circular sections where :

1,70

D.Vs < 6 m2/s D.Vs ≥ 6 m2/s Machinery Rectangular clad houses structures on ground etc. or solid base

1,10 0.80 1,10

(1) See figure 2.2.4.1.4.1.

2 - 27


(I) Aerodynamic slenderness : ( length of member ) / ( breadth of section across wind front ) = l/b * or l/D *

* In lattice construction the lengths of individual members are taken between the centres of adjacent node points. See diagram below (II) Solidity ratio : (area of solid parts) / (enclosed area) = A /Ae =

∑

n

1

[(li . bi)/(L . B)]

(III) Spacing ratio : (distance between facing sides) / (breadth of members across wind front ) = a/b or a/B

for "a" take the smallest possible value in the geometry of the exposed face.

(IV) Section ratio : (breadth of section across wind front) / (depth of section parallel to wind flow)= b/d Figure 2.2.4.1.4.1. - Definitions : Aerodynamic Slenderness, Solidity Ratio, Spacing Ratio, and Section Ratio

2 - 28


2.2.4.1.4.2. Multiple frames of members : shielding factors

Where parallel frames or members are positioned so that shielding takes place, the wind loads on the windward frame or member and on the unsheltered parts of those behind it are calculated using the appropriate shape coefficients. The wind load on the sheltered parts is multiplied by a shielding factor η given in table T.2.2.4.1.4.2. Values of η vary with the solidity and spacing ratios as defined in figure 2.2.4.1.4.1. Table T.2.2.4.1.4.2. - Shielding coefficients Spacing ratio a/b 0,5 1,0 2,0 4,0 5,0 6,0

0,1 0,75 0,92 0,95 1,0 1,0 1,0

0,2 0,40 0,75 0,80 0,88 0,95 1,0

Solidity ratio A/Ae 0,3 0,4 0,32 0,21 0,59 0,43 0,63 0,50 0,76 0,66 0,88 0,81 1,0 1,0

0,5 0,15 0,25 0,33 0,55 0,75 1,0

≥0,6 0,10 0,10 0,20 0,45 0,68 1,0

Where a number of identical frames or members are spaced equidistantly behind each other in such a way that each frame shields those behind it, the shielding effect is assumed to increase up to the ninth frame and to remain constant thereafter. The wind loads are calculated as follows : On the 1st. frame

F1 = A.q.Cf

in N

On the 2nd. frame

F2 = η.A.q.Cf

in N

On the n.th frame (where n is from 3 to 8)

Fn = η (n-1).A.q.Cf

On the 9th and subsequent frames

F9 = η8.A.q.Cf

in N

Ftotal = [1 + η + η2 + η3 + .... + η(n-1)].A.q.Cf = [(1 - ηn) / (1 - η)].A.q.Cf

in N

Ftotal = [1 + η + η2 + η3 + .... + n8 + (η - 9)8].A.q.Cf = [(1 - η9) / (1 - η) + (n - 9) η8].A.q.Cf

in N

in N

The total wind load is thus : Where there are up to 9 frames

Where there are more than 9 frames

Note - The term ηx used in the above formula is assumed to have a longer limit of 0.10. It is taken x as 0.10 whenever η < 0.10. 2 - 29


2.2.4.1.4.3. Lattice towers

In calculating the "face-on" wind load on square towers, in the absence of a detailed calculation, the solid area of the windward face is multiplied by the following overall force coefficient : For towers composed of flat sided sections 1,7 (1 + η) For towers composed of circular sections where D.Vs < 6 m2/s 1,1 (1 + η) 2 where D.Vs ≥ 6 m /s 1,4 The value of η is taken from table 2.2.4.1.4.2. for a/b = 1 according to the solidity ratio of the windward face. The maximum wind load on a square tower occurs when the wind blows on to a corner. In the absence of a detailed calculation, this load can be considered as 1.2 times that developed with "face-on" wind on one side. 2.2.4.1.4.4. Parts inclined in relation to the wind direction

Individual members. frames, etc. Where the wind blows at an angle to the longitudinal axis of a member or to the surface of a frame, the wind load in the direction of the wind is obtained from : F = A.q.Cf sin2 θ in N where F, A, q and Cf are as defined in 2.2.4.1.3. and θ is the angle of the wind (θ < 90°) to the longitudinal axis or face. Lattice trusses and towers Where the wind blows at an angle to the longitudinal axis of a lattice truss or tower, the wind load in the direction of the wind is obtained from : in N F = A.q.Cf.K2 where : F, A, q and Cf are as defined in 2.2.4.1.3. and K2 = θ / [50 (1,7 - Sp/S)] which cannot be less than 0,35 or greater than 1. θ is the angle of the wind in degrees (θ < 90°) to the longitudinal axis of the truss or tower. 2 Sp is the area in m of the bracing members of the truss or tower projected on to its windward plane.

S is the area in m2 of all (bracing and main) members of the truss or tower projected on to its windward plane. The value of K2 is assumed to have lower and upper limits of 0.35 and 1.0 respectively. It is taken as 0.35 whenever the calculated value < 0.35 and as 1.0 whenever the calculated value > 1.0. 2 - 30


2.2.4.2. SNOW LOAD Snow loads shall be neglected in the design calculations for overhead travelling cranes, bridge cranes and jib cranes. 2.2.4.3. TEMPERATURE VARIATIONS Stresses due to temperature variations shall be considered only in special cases such as when members are not free to expand. In such cases, the maximum temperature fluctuation shall be taken to be : - 20° C to + 45° C.

2.2.5 MISCELLANEOUS LOADS 2.2.5.1. LOADS CARRIED BY PLATFORMS Access gangways, driver 's cabine and platforms shall be designed to carry the following concentrated loads : 3000 N for maintenance gangways and platforms where materials may be placed, 1500 N for gangways and platforms intended only for access of personnel, 300 N as the horizontal force which may be exerted on handrails and toe-guards. These loads are not to be used in the calculations for girders.

2 - 31


2.3. CASES OF LOADING Three different cases of loading are to be considered for the purpose of the calculations : - the working case without wind, - the working case with limiting working wind, - the case of exceptional loadings. Having determined the various loads in accordance with section 2.2, account is taken of a certain probability of exceeding the calculated stress, which results from imperfect methods of calculation and unforeseen contingencies, by applying an amplifying coefficient γC, which varies according to the group classification of the appliance. The values of this coefficient γC are indicated in clause 2.3.4.

2.3.1. CASE I : APPLIANCE WORKING WITHOUT WIND The following shall be taken into consideration : the static loads due to the dead weight SG, the loads due to the working load SL multiplied by the dynamic coefficient ψ, and the two most unfavourable horizontal effects SH among those defined in clause 2.2.3., excluding buffer forces. All these loads must then be multiplied by the amplifying coefficient γC specified in clause 2.3.4., viz : γC (SG + ψ SL + SH) In cases where travel motion takes place only for positioning the appliance and is not normally used for moving loads the effect of this motion shall not be combined with another horizontal motion. This is the case for example with a dockside crane which, once it has been positioned, handles a series of loads at a fixed point.

2.3.2. CASE II : APPLIANCE WORKING WITH WIND The loads of case I are taken to which are added the effects of the limiting working wind SW defined under 2.2.4.1.2.1. (table T.2.2.4.1.2.1.) and, where, applicable the load due to temperature variation, viz : γC (SG + ψ SL + SH) + SW Note - The dynamic effects of acceleration and retardation do not have the same values in case II as in case I, for when a wind is blowing the accelerating or braking times are not the same as when still conditions prevail.

2 - 32


2.3.3. CASE III : APPLIANCE SUBJECTED TO EXCEPTIONAL LOADINGS Exceptional loadings occur in the following cases : - appliance out of service with maximum wind - appliance working and subjected to a buffer effect - appliance undergoing the tests indicated in booklet 8. The highest of the following combinations shall be considered : a) The loads SG due to the dead weight, plus the load Sw max due to the maximum wind as mentioned under clause 2.2.4.1.2.2. (including the reactions of the anchorages) b) the loads SG due to the dead weight and SL due to the working load plus the greatest buffer effect ST as envisaged in clause 2.2.3.4. c) the loads SG due to the dead weight plus the highest of the two loads ψρ1 SL and ρ2 SL ; ρ1 and ρ2 being the coefficients by which the safe working load is multiplied for the dynamic test (ρ1) and for the static test (ρ2) as in clauses 8.1.1. and 8.1.2. These three cases are expressed by the formulae : a) SG + Sw max 9 b) SG + SL + ST c) SG + ψρ1 SL or SG + ρ2 SL Note 1 - It should be noted that the checks under (c) are only to be made in cases where the working load, when assumed to act alone, produces stresses opposed in direction to those caused by the dead weight up to the point at which the static test load does not exceed 1,5 times the safe working load. Note 2 - When using decelerating devices in advance of buffer impact under the conditions mentioned in clause 2.2.3.4.1. ST will be taken to be the highest load resulting either from the retardation previously caused by the decelerating device or from that finally caused by the buffer.

9

Loadings resulting from the working load are taken into account but the effects of load swing resulting from the stock are neglected because this swing only loads the structure when the other effects have been practically absorbed. This comment does not apply to rigidly guide loads which cannot swing. 2 - 33


2.3.4. CHOOSING THE AMPLIFYING COEFFICIENT γ C The value of the amplifying coefficient γc depends upon the group classification of the appliance. Table T.2.3.4. - Values of amplifying coefficient γC Appliance group γc

A1 1,00

A2 1,02

A3 1,05

A4 1,08

A5 1,11

A6 1,14

A7 1,17

A8 1,20

2.4. SEISMIC EFFECTS In general the structures of lifting appliances do not have to be checked for European seismic effects. However, if official regulations or particular specifications so prescribe, special rules or recommendations can be applied in areas subject to earthquakes. This requirement shall be advised to the supplier by the user of the installation who shall also provide the corresponding seismic spectra.

2 - 34


2.5. LOADS ENTERING INTO THE DESIGN OF MECHANISMS Mechanisms are subjected to two kinds of loading : a) The loads, represented by the symbol SM, which are directly dependent upon the torques exerted on the mechanisms by the motors or the brakes. b) The loads, represented by the symbol SR, which are independent of motor or brake action but which are determined by the reactions which act upon the mechanical parts and which 10 are not balanced by a torque acting on the drive shafts .

2.5.1. TYPE SM LOADS The loads of this type to be considered are : a) SMG loads, corresponding to a vertical displacement of the centre of gravity of moving parts of the appliance other than the working load. b) SML loads, corresponding to a vertical displacement of the working load as defined in clause 2.2. for structures. c) SMF loads, corresponding to frictional forces which have not been allowed for in calculating the efficiency of the mechanism (see clause 4.2.6.1.1., booklet 4). d) SMA loads, associated with acceleration (or braking) of the motion. e) SMW loads, corresponding to the effect of the working wind assumed for the appliance.

2.5.2. TYPE SR LOADS The loads of this type to be considered are : a) SRG loads due to the weights of components which act on the part under consideration ; b) SRL loads due to the working load as defined in clause 2.2., for structures. c) SRA loads due to the accelerations or decelerations of the various motions of the appliance or its parts, as calculated according to clause 2. 2.3.1. for structures, insofar as the order of magnitude of these loads is not negligible compared to the SRG and SRL loads. d) SRW loads due to the limiting working wind SW or to the maximum wind SW max (see clause 2.2.4.1.), insofar as the order of magnitude of these loads is not negligible.

10

In a travel motion, for instance, the loads due to the vertical reaction on the rail wheels and the transverse loads that stress the wheel axle but are not transmitted to the components of the driving mechanism. 2 - 35


2.6. CASES OF LOADING Three cases of loading are to be considered in the calculations : Case I : Normal service without wind Case II : Normal service with wind Case III : Exceptional loadings. A maximum load must be determined for each case of loading which serves as the basis for the calculations. Note - Clearly, case I and II are one and the same in the case of appliances which are not exposed to wind. The various loadings being determined as indicated in paragraph 2.5., account is taken of a certain probability of exceeding the calculated stress, which results from imperfect methods of calculation and unforeseen contingencies, by applying an amplifying coefficient γm depending on the group in which the mechanism is classified. The values of this coefficient γm are indicated in table T.2.6. Table T.2.6. - Values of amplifying coefficient γm Mechanism group γm

M1 1,00

M2 1,04

M3 1,08

M4 1,12

M5 1,16

M6 1,20

M7 1,25

M8 1,30

2.6.1. CASE I - NORMAL SERVICE WITHOUT WIND 2.6.1.1. TYPE SM LOADS The maximum load SM max I of the SM type (see clause 2.5.) is determined by combining the loads SMG, SML, SMF, and SMA defined in clause 2.5.1. which can be expressed by the relation : SM max I = ( S MG + S ML + S MF + S MA ) γm Note - It must be pointed out that it is not the combination of the maximum values of each of the terms in this relation that must be considered, but the value resulting from the most unfavourable combination that could actually occur in practice. 2.6.1.2. TYPE SR LOADS The maximum load SR max I of the SR type (see clause 2.5.) is determined by combining the loads SRG, SRL, SRA, defined in clause 2.5.2. which can be expressed by the relation : SR max I = ( S RG + S RL + S RA ) γm The note in clause 2.6.1.1. above applies here also. 2 - 36


2.6.2. CASE II - NORMAL SERVICE WITH WIND 2.6.2.1. TYPE SM LOADS The maximum load SM max II of the SM type (see clause 2.5.) is determined by combining the loads SMG, SML and SMF defined in clause 2.5.1. with one of the following two combinations : a) the load SMA and the load SMW 8 corresponding to a 80 N/m2 wind. b) the load SMW 25 corresponding to a 250 N/m2 wind. The higher of the two values expressed by the relations set out below is taken : SM max II = ( S MG + S ML + S MF + S MA + S MW 8 ) γm or SM max II = ( S MG + S ML + S MF + S MW 25 ) γm The note in clause 2.6.1.1. applies here also. 2.6.2.2. TYPE SR LOADS The maximum load SR max II of the SR type (see clause 2.5.) is determined by combining the loads 2 SRG, SRL and SRA defined in clause 2.5.2. with SRW 25 which corresponds to a 250 N/m wind, as expressed by the relation : SR max II = ( S RG + S RL + S RA + S RW 25 ) γm The note in clause 2.6.1.1. applies here also.

2 - 37


2.6.3. CASE III - EXCEPTIONAL LOADS 2.6.3.1. TYPE SM LOADS The maximum load SM max III of the SM type defined under clause 2.5. is determined by considering the maximum load that the motor can actually transmit to the mechanism, allowing for limitations due to practical operating conditions. The values of SM max III are specified in clause 2.6.4. 2.6.3.2. TYPE SR LOADS Since the consequences of an overload due to collision with a buffer or fouling are far less serious for a mechanism than for the structure, the exceptional loading to be taken is that given under paragraph a) of clause 2.3.3. in the structures chapter. This gives : SR max III = ( S RG + S RW max ) In cases where additional mooring or guying means are used to ensure immobility or stability under maximum wind, the effect of these devices on the mechanism must be taken into account where applicable.

2.6.4. APPLICATION OF THE ABOVE CONSIDERATIONS FOR CALCULATING SM The mechanisms of hoisting appliances perform one of the following functions : - Purely vertical displacements of the centre of gravity of moving masses (e.g. hoisting motions). - Purely horizontal displacements in which the centre of gravity of the moving masses as a whole shifts horizontally (e.g. traverse, travel, slewing or counterbalanced luffing motions). - Movements combining an elevation of the centre of gravity of the moving masses with a horizontal displacement (e.g. non-counterbalanced luffing).

2 - 38


2.6.4.1. HOISTING MOTIONS For type SM loads, the formula reduces to the following : Case I and II :

SM max I = ( S ML + S MF ) γm

In this case the load due to the hoisting acceleration is neglected because it is small compared to SML. Case III :

SM max III = 1,6 ( S ML + S MF )

Bearing in mind the general rules of clause 2.6.3.1., it is assumed that the maximum loads that can be transmitted to hoisting mechanisms are limited in practice to 1,6 times the SM max I load 11.

2.6.4.2. HORIZONTAL MOTIONS Case I - The formula reduces to : SM max I = ( S MF + S MA ) γm Case II - The higher of the following two values is taken : SM max II = ( S MF + S MA + S MW 8 ) γm or SM max II = ( S MF + S MW 25 ) γm Case III - For SM max III the load corresponding to the maximum torque of the motor (or the brake) is taken unless operating conditions limit the torque actually transmitted, through wheel slip on the rails, or through the use of suitable limiting means (e.g. hydraulic coupling, torque limiter, 12 etc.). In this case the value actually transmitted must be taken .

11

In a hoisting motion it is impossible under normal working conditions to transmit to the mechanism loads greater than those due to the hoisting of the working load, as the effects of acceleration are negligible. A greater load could result only from mishandling (poor judgement of the load, etc.).

On the basis of experience gained over many years of practice with widely differing hoisting appliances it is now accepted that a coefficient of 1,6 gives adequate safety. It must be stressed that the use of excessively powerful motors should be avoided. 12

Whereas in the case of hoisting motions the loads normally transmitted to the mechanism are limited by the load lifted, in horizontal motions the maximum torque of the motor can always be transmitted to the mechanism if no mechanical limitation exists. This is why a different way of evaluating SM max III has been specified according to whether a hoist motion or other motion is being considered.

2 - 39


2.6.4.3. COMBINED MOTIONS Case I and II : For cases I and II, the load SM max II clauses 2.6.1.1. and 2.6.2.1.

13

is determined by applying the general formula defined in

Case III : The load caused by applying the maximum motor torque SMC max III can be taken for the maximum value SM max III This often unduly high value is always acceptable since it enhances safety. It must be used when the power involved for raising the centres of gravity of the moving masses is negligible compared to the power needed to overcome accelerations or wind effects. Conversely, when the effect of the accelerations or the wind is negligible in comparison with the effect of displacing the centres of gravity of the moving masses vertically, this value is too high and SM max III can be calculated from the formula : SM max III = 1,6 SM max II Between these two limiting values, each individual case should be examined according to the motor chosen, the method of starting and the relative magnitudes of the loads due to inertia and wind effects on the one hand and those due to raising of the centres of gravity on the other. Without exception, when operating conditions limit the torque actually transmitted to the mechanism (see clause 2.6.4.2.), this limiting torque will be taken as the value of SMC max if it is less than the values defined above.

.

13

or SM max I in the case of appliances not subjected to wind.

2 - 40


APPENDIX A.2.1.1. - HARMONISATION OF THE CLASSES OF UTILIZATION OF APPLIANCES AND MECHANISMS The present appendix sets out to demonstrate a method by which it is possible in many cases to derive the class of utilization of mechanisms from that of appliances as a whole and from certain parameters characterising the duty to be performed. The starting point is the average duration tmc (in seconds) of a hoisting cycle as defined in clause 2.1.2.2. This is therefore the time necessary to perform all the operations in such a cycle. The total duration of use T of the appliance, expressed in hours, is then given by the relation : T = N.tmc / 3600 Where N represents the number of hoisting cycles determining the class of utilization of the appliance. Table T.A.2.1.1.1. gives the values of T for cycle durations of 30 - 480 s in accordance with the class of utilization of the appliance. The number of hoisting cycles is the maximum number for this class of utilization ; these values are, however, adjusted to 15 625, 31 250 and 62 500 respectively for class U0, U1 and U2, in order to reduce the number of different values for T. The next step is to determine for each mechanism the ratio αi between the duration of use of the mechanism during a hoisting cycle and the average duration tmc of the cycle. Table T.A.2.1.1.2. gives the total durations of use Ti of the mechanism depending on the total duration of use of the appliance, and for various conventional values of the ratio αi . This table also shows the class of utilization of the mechanism. The various classes are represented by the stepped areas. It is thus sufficient to determine the class of utilization of the appliance by reference to table T.2.1.2.2., the average duration of the hoisting cycle and the values of αi in order to obtain the classes of utilization of the mechanisms. From the curves of the nomogram T.A.2.1.1.3. the classes of utilization for the mechanisms in terms of these three parameters can be found directly.

2 - 41


Table T.A.2.1.1.1. - Total duration of use (T) of lifting appliances in hours Average duration of a hoisting cycle tmc (s)

Class of utilization of appliances U0

U1

U2

U3

U4

U5

U6

U7

U8

U9

30

130

260

520

1 040

2 085

4 165

8 335

16 665

33 335

> 33 335

45

195

390

780

1 565

3 125

6 250

12 500

25 000

50 000

> 50 000

60

260

520

1 040

2 085

4 165

8 335

16 665

33 335

66 665

> 66 665

75

325

650

1 300

2 605

5 210

10 415

20 835

41 665

83 335

> 83 335

90

390

780

1 565

3 125

6 250

12 500

25 000

50 000

100 000 > 100 000

120

520 1 040

2 085

4 165

8 335

16 665

33 335

66 665

133 335 > 133 335

150

650 1 300

2 605

5 210 10 415

20 835

41 665

83 335

180

780 1 565

3 125

6 250 12 500

25 000

50 000

240

1 040 2 085

4 165

8 335 16 665

33 335

66 665

166 665 > 166 665 ________ 100 000 200 000 > 200 000 ________ 133 335 > 200 000

300

1 300 2 605

5 210

10 415 20 835

41 665

83 335

166 665 > 200 000

360

1 565 3 125

6 250

12 500 25 000

50 000

100 000

420

1 825 3 645

7 290

14 585 29 165

58 335

200 000 > 200 000 ________ 116 665 > 200 000

480

2 085 4 165

8 335

16 665 33 335

66 665

133 335 > 200 000

2 - 42


Table T.A.2.1.1.2. - Total duration of use Ti (in hours) of mechanisms in terms of T and αi Values of αi

T (h) 130 195 260 325 390 520 650 780 1 040 1 300 1 565 1 825 2 085 2 605 3 125 3 645 4 165 5 210 6 250 7 290 8 335 10 415 12 500 14 585 16 665 20 835 25 000 29 165 33 335 41 665 50 000 58 335 66 665 83 335 100 000 116 665 133 335 166 665 200 000 > 200 000

1,00 130 195 260 325 390 520 650 780 1 040 1 300 1 565 1 825 2 085 2 605 3 125 3 645 4 165 5 210 6 250 7 290 8 335 10 415 12 500 14 585 16 665 20 835 25 000 29 165 33 335 41 665 50 000 58 335 66 665 83 335 100 000 116 665 133 335 166 665 200 000 > 200 000

0,63

0,40

0,25

0,16

82 123 164 205 246 328 410 491 655 819 986 1 150 1 314 1 641 1 969 2 296 2 624 3 282 3 938 4 593 5 251 6 561 7 875 9 189 10 499 13 126 15 750 18 374 21 001 26 249 31 500 36 751 41 999 52 501 63 000 73 499 84 001 104 999 126 000 > 126 000

52 78 104 130 156 208 260 312 416 520 626 730 834 1 042 1 250 1 458 1 666 2 084 2 500 2 916 3 334 4 166 5 000 5 834 6 666 8 334 10 000 11 666 13 334 16 666 20 000 23 334 26 666 33 334 40 000 46 666 53 334 66 666 80 000 > 80 000

33 49 65 81 98 130 163 195 260 325 391 456 521 651 781 911 1 041 1 303 1 563 1 823 2 084 2 604 3 125 3 646 4 166 5 209 6 250 7 291 8 334 10 416 12 500 14 584 16 666 20 834 25 000 29 166 33 334 41 666 50 000 > 50 000

21 31 42 52 62 83 104 125 166 208 250 292 334 417 500 583 666 834 1 000 1 166 1 334 1 666 2 000 2 334 2 666 3 334 4 000 4 666 5 334 6 666 8 000 9 334 10 666 13 334 16 000 18 666 21 334 26 666 32 000 > 32 000

0,10 13 20 26 33 39 52 65 78 104 130 157 183 209 261 313 365 417 521 625 729 834 1 042 1 250 1 459 1 667 2 084 2 500 2 917 3 334 4 167 5 000 5 834 6 667 8 334 10 000 11 667 13 334 16 667 20 000 > 20 000

Class of utilization for mechanism

T0

T1

T2

T3

T4

T5

T6

T7

T8 T9

2 - 43


Table T.A.2.1.1.3. - Classes of utilization for appliances and mechanisms U - Class of utilization for appliances

T - Class of utilization for mechanisms

2 - 44


EXAMPLE OF APPLICATION Dockside cargo crane. The class of utilization for the appliance will be U5. A hoisting cycle comprises the following operations : - hoisting of load ; - travelling ; - slewing ; - lowering ; - unhooking of load ; - hoisting empty ; - slewing ; - travelling ; - lowering empty ; - hooking on of new load. The average time for completion of the cycle will be estimated at 150 s. The ratios αi will be estimated as follows : - hoisting (hoisting and lowering) : - slewing (2 directions) : - travelling (do.) :

αi = 0.63 αi = 0.25 αi = 0.10

Table T.A.2.1.1.1. gives us for class U5 and tmc = 150 s : T = 20 835 h For the various mechanisms, table T.A.2.1.1.2. gives us, for T = 20 835 h, the following total durations Ti and classes of utilization : - hoisting (αi = 0.63) : - slewing (αi = 0.25) : - travelling (αi = 0.10) :

Ti = 13 126 h Ti = 5 209 h Ti = 2 084 h

T7 T5 T4

From the curves in table T.A.2.1.1.3. the same conclusions are drawn on the basis of the ordinate tmc = 150 s (broken line).

2 - 45


A.2.2.3. - CALCULATION OF LOADS DUE TO ACCELERATIONS OF HORIZONTAL MOTIONS PART 1 - METHOD 1. - BASIC DATA Let v be the steady horizontal velocity of the point of suspension of the load, either at the end of the acceleration period, or at the beginning of the braking period, according to whether an acceleration or a braking process is being considered, and F an imaginary horizontal force in the same direction as v, applied at the point of suspension of the load and producing the same effect on the motion under consideration as the accelerating or decelerating torque applied by the motor or the brake. 2. - PROCEDURE The different quantities set out below must be calculated in succession. Equivalent mass (m) The inertia of all moving parts other than the load, in the motion under consideration, is replaced by a single equivalent mass m assumed to be concentrated at the point of suspension of the load and given by the relation : m = m0 + ∑ i [ ( Ii . wi2) / v2 ] Where : m0

is the total mass of all elements, other than the load, undergoing the same pure linear motion as the point of suspension of the load ;

Ii

the moment of inertia of a part undergoing a rotation during the motion under consideration, this moment of inertia being considered about the axis of rotation, and

wi

the angular velocity of the part referred to, about its axis of rotation, corresponding to the linear velocity v of the point of suspension of the load.

The sum Σ covers all parts in rotation (structure, mechanisms, motor) during the motion considered. However, in the case of mechanisms, the inertia of components other than those directly coupled to the motor shaft can be ignored. Mean acceleration or deceleration ( Jm ) : Jm = F / (m + m1 ) where m1 is the mass of the load.

2 - 46


Mean duration of acceleration or deceleration ( Tm ) : Tm = v / Jm Mean inertia forces : The acceleration corresponding to the acceleration Jm at the point of suspension of the load is calculated for each component part in motion. Multiplying this acceleration by the mass of the component considered gives the mean inertia force it sustains. In the particular case of the load itself, this force of inertia Fcm will be given by : Fcm = m1 . Jm Period of oscillation : Tl

T1 = 2 . π . ( l / g )0,5

l=

the length of suspension of the load when it is in its uppermost position (values of l below 2,00 m need not be taken into consideration) and, g = the acceleration due to gravity. Value of µ :

µ = m1 / m

When the system driving the motion controls the acceleration and the deceleration and maintains it at a constant value, µ is taken equal to 0 irrespective of the masses m and m1. Value of β :

β = Tm / T1

Value of ψh : With the values obtained for µ and β, the graph in figure A.2.2.1. is used to find the corresponding value ψh. Inertia forces to be considered in the design of the structure : The forces of inertia which take account of dynamic effects and which must therefore be considered in the structural calculations are obtained as follows : - Inertia force due to the load :

ψh . Fcm

- Inertia force on moving parts other than the load : twice the mean inertia forces. 3. - JUSTIFICATION A justification of the method given above follows in part 2 of this appendix.

2 - 47


PART 2 - EXPLANATION OF THE METHOD 1. - STATEMENT OF THE PROBLEM A hoisting appliance is a physical system consisting essentially of : - concentrated masses (hook load, counterweights, ...) and distributed masses (girders, ropes, ...), - elastic connections between these masses (girders, ropes, ...). If such a system, originally in a state of equilibrium, is subjected to a varying load, it does not tend progressively towards a new state of equilibrium even if the new load applied is itself constant. On the contrary, it is set in a more or less complex oscillating motion about this new state of equilibrium. During this motion, the various internal loads and stresses of the system can exceed sometimes to a marked extent - the values they would have assumed had the system been in static equilibrium under the influence of the new load. Such a situation arises during acceleration or deceleration (braking) of a horizontal motion of a hoisting appliance. Thus if, starting from a position of rest, an appliance or part of an appliance begins a motion of translation or rotation, the component parts of the system undergo accelerations and are therefore subjected to inertia forces. Once a steady speed is attained, the acceleration ceases, the inertia forces disappear and the external load undergoes a new variation. The angle through which a rotating system turns (e.g. the rotating part of a crane) during the time for which inertia forces are applied is generally relatively small. This being so, no appreciable error will be involved if one assumes that each point in the system follows a straight path during this time. Since, moreover, there is no difference of principle between the treatment used for linear motions and motions of rotation, in what follows the linear motion will be considered in greater detail (chapter 2), whereas only a short note (chapter 3) will cover rotation.

2 - 48


2. - CALCULATING THE LOADS IN THE CASE OF A LINEAR MOTION 2.1. - GENERAL DATA It is now proposed to examine the particular case of braking of the travel motion of a complete overhead travelling crane when it is carrying a load suspended from its hoisting rope. Other cases encountered in practice can be dealt with in similar fashion. Considering figure A.2.1. let : m 1 be the mass of the suspended load, m the total mass of the overhead travelling crane including the crab (see note below concerning the inertia of the motor and of the machinery driving the motion), x a coordinate defining the position of the crane along its track (more precisely, x represents the coordinate of the point of suspension of the hoisting rope along an axis parallel to the direction of travel), a coordinate defining the position of the centre of gravity of the suspended load along an x1 axis of the same direction, sense and origin as the axis of x , z = x1 - x a coordinate expressing the horizontal displacement of the load relative to the crane. Let us assume that at the instant t = 0 the overhead travelling crane is moving in the positive sense of the x axis at a velocity v, and that the load is at rest relative to the crane. ( z = z' = 0, with : z' = dz / dt ) If the brake is applied to the travel mechanism at the instant t = 0, it will give rise from that instant to a horizontal braking force parallel to, but of opposite sense to, the x axis at each point where a driving wheel is in contact with its rail. To simplify masters, let us assume that the crab is located at mid-span of the main girders of the overhead travelling crane. It follows by symmetry that the total force at each rail is the same. Let us designate its projection on the x axis by F/2 (with F > 0), so that the total braking force acting on the system in motion 2 (crane plus load) is equal to F in absolute value. If the system were composed of rigidly interconnected masses, this would result in a (2.1.1) deceleration of absolute value Jm given by the relation : Jm = F / ( m +m1 )

2 - 49


Figure A.2.1. It must not be forgotten however that F originates in the braking torque applied to the travel mechanism which must not only brake the travel inertia of the crane and the load but also the rotational inertia of the driving motor and the intervening machinery. Generally speaking, one can neglect the rotating inertia of all components other than those integral with the motor shaft. In many cases, however, the inertia of the latter must be taken into account and the relation (2.1.1.) holds good only provided that m incorporates an equivalent mass me given by the relation : me . v2 = Im . ωm2 (2.1.2.) where : Im

is the moment of inertia of all the components integral with the motor shaft (including the motor itself, of course) and

ωm

the angular velocity of the motor corresponding to the travelling speed v of the crane.

Under the effect of the deceleration Jm, the suspension rope cannot retain its vertical position. Its new position of equilibrium is inclined to the vertical at an angle αm given by the relation : (2.1.3.) αm = arctg( Jm / g ) where g is the acceleration due to gravity. In this case the rope exerts a horizontal force on the crane whose projection Fcm on the x axis is given by : Fcm = m1 . Jm 2 - 50

(2.1.4.)


In point of fact, the system is not rigid, the deceleration is not constant and is not therefore given by (2.1.1.), the load and its suspension rope adopt an oscillating motion, and the horizontal force developed by the rope on the crane can assume values differing greatly from (2.1.4.). By a similar reasoning, one may conclude that the deceleration of the system gives rise to inertia forces which act on each component part of the crane and the crab, but that because of the elasticity of the girders the system will undergo an oscillating motion in the course of which the stresses will be subject to fluctuations which must be estimated. The next two paragraphs deal in succession with the effect of the inertia forces on the load and on the girders.

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2.2. - EFFECT OF INERTIA FORCES ON THE LOAD In determining the motion which the load executes after the brake is applied, one can neglect the movement of the point of suspension due to girder flexibility in a horizontal plane. The amplitude of this movement is, in fact, very small compared with the amplitude of swinging of the load. Calculations can therefore be carried out with the crane considered as a system which is not subject to deformation. The projection FC on the x axis of the force exerted by the rope on the crane is given by the relation : (2.2.1) FC= m1 . g [( x1 - x ) / l] = m1 . g . z / l where l is the suspension length of the load. It will be noted that FC is proportional to the displacement z of the load with respect to its position of initial equilibrium, just as if it were an elastic restoring force. The equations of motion can be written : ml . x’’1 = m1 . g [( x1 - x ) / l]

(2.2.2.)

m . x’’ = m1 . g [( x1 - x ) / l] - F

(2.2.3.)

while, assuming x = 0, for t = 0, the initial conditions are as follows : for t = 0,

x1 = x = 0

(2.2.4.)

x'1 = x' = v

(2.2.5.)

z = x1 - x = 0

(2.2.6.)

z’ = x’1 - x’ = 0

(2.2.7.)

g / l = ω12

(2.2.8.)

(m1 / m) . (g / l) = ω22

(2.2.9.)

Let :

ω12 + ω22 = ωr2

(2.2.10.)

F / m = J0

(2.2.11.)

Equations (2.2.2.) and (2.2.3.) then become : x" + z" + ω12 . z = 0

(2.2.12.)

x" - ω22 . z = - J0

(2.2.13.)

whence 2 - 52


z’’ + ωr2 . z = J0

(2.2.14.)

With the initial conditions of (2.2.4.) to (2.2.7.), the solution to these equations is given by : 2 z = ( J0 / ωr ) . [1 - cos( ωr . t)]

x' = v -[ ( ω12./ ωr2 ) . J0 . t ] - [ ( ω22./ ωr2 ) . ( J0 / ωr ) ] . sin(ωr . t) The complete expression for x is of no direct interest to us. Let :

J0 / ωr2 = zm

(2.2.15.) (2.2.16.)

(2.2.17.)

it can then be seen without difficulty that zm is the position of equilibrium that can be assumed by the load during a constant deceleration of the crane equal to the value Jm defined by (2.1.1.), i.e. during the deceleration that would be obtained by applying the braking force F to the total mass (crane plus load) in motion, this mass being assumed to constitute a rigid system. The value z = zm defining the load displacement corresponds to the horizontal force Fcm, defined by (2.1.4.) exerted by the rope on the crane. Comparison between (2.2.1.), (2.2.15.) and (2.2.17.) then shows that : (2.2.18.) Fc = Fcm . [ 1 - cos(ωr . t) ] If the deceleration period of the crane lasts for a time td such that : ωr . td ≥ π

(2.2.19.)

it will be seen that Fc momentarily becomes twice Fcm, or in other words, that its maximum value Fc max is given by the relation : (2.2.20.) Fc max = 2 . Fcm If the condition (2.2.19.) is not satisfied, this means that the crane has stopped before the load has reached its maximum displacement z = 2 zm. However, after the crane stops, the load will usually continue to oscillate, so the rope will continue to exert a varying horizontal force on the crane, and the maximum value which this can attain must be sought. It is easy to verify that after the crane has stopped, the motion of the load is defined by the expression : (2.2.21.) z = zd . cos[ ω1 . (t - td) ] + (z’d / ω1) . sin[ ω1 . (t - td) ] with zd = zm . [ 1 - cos(ωr . td) ]

(2.2.22.)

z'd = ωr . zm . sin(ωr . td)

(2.2.23.)

where td is the smallest positive value of t that makes the expression (2.2.16.) for x' equal to zero. The maximum value Fc max assumed by Fc is then given by the relation :

2 - 53


Fc max = Fcm . { [ 1 - cos( ωr . td ) ]2 + ( ωr2 / ω12 ) . sin2( ωr . td ) }0,5

2 - 54

(2.2.24.)


Generally speaking, we may take : Fc max / Fcm = ψh

(2.2.25.)

The determination of ψh is simplified by introducing the following quantifies : Tm = v / Jm T1 = 2 π / ω1

the time for which the slowing-down phase of the crane would last if the deceleration were constant and the system in motion not subject to deformation. the period of oscillation of the pendulum system formed by the suspended load (crane stopped).

T1 = 2 π . ( l / g )0,5 It can be verified without difficulty that ψh depends only on two non-dimensional parameters µ and β defined by the ratios : µ = m1 / m

(2.2.26.)

β = Tm / T1

(2.2.27.)

which can be obtained very easily. It will be noted that (2.2.16.) can be written : 0,5

x’ = v . { 1 - [ ωr . t + µ . sin( ωr . t ) ] / [ 2 π . β ( 1 + µ )

]}

(2.2.28.)

and therefore : [ ωr . td + µ . sin( ωr . td ) ] / [ 2 π . β ( 1 + µ )0,5 ] = 1

(2.2.29.)

this equation makes it possible to determine the value of ωr . td to be introduced into (2.2.24.). The graph in figure (2.2.1.) plots the values of ψh against β for various values of µ. (The curve µ = 0 will be explained later in Chapter 5). If µ < 1 (which is generally the case with overhead travelling crane travel motions, such as that in the example dealt with), an analysis of the problem shows that ψh can in no case exceed 2. This value is reached during the crane deceleration phase if the condition (2.2.19.) is satisfied or, which is the same thing, if β reaches or exceeds a certain critical value, βcrit dependent upon µ. Above this critical value, ψh therefore remains constant and equal to 2, whatever the value of β. If µ > 1 (which could be the case with traverse motions, in which m essentially represents only the mass of the crab, or with slewing motions), the same analysis shows that, again provided β reaches or exceeds a certain critical value βcrit dependent upon µ, ψh can exceed 2 and reach a maximum given by : ψh = [ 2 + µ + ( 1 / µ ) ]0,5 (2.2.30.) 2 - 55


2 - 56


This maximum can actually be reached only during the swinging motion of the load subsequent to the bringing to rest of its point of suspension. The critical value βcrit is such that the crane is halted before the condition (2.2.19.) is satisfied, or before Fc reaches 2 Fcm. However, any value of β greater than βcrit leads to the condition (2.2.19.) being satisfied and Fc necessarily passes the value 2 Fcm , whence ψh > 2. It will also be noted that if β > βcrit has been calculated taking v equal to the maximum steady speed of the motion, braking applied starting from the initial speed v . βcrit / β will necessarily lead to the maximum value of ψh given by (2.2.30.). This is the reason why, in the graph of figure A.2.2.1., the values of ψh have been maintained for all values of β greater than βcrit.

Figure A.2.2.1. As regards the choice of T1, it should be noted that the danger of reaching high values of ψh is all the greater as the suspension length l of the load becomes shorter, because β then attains its critical value more rapidly. The calculations must therefore be made assuming that the load is near its uppermost position. In practice l will generally lie between 2 and 8 m. The table below gives the values of T1 for a few values of l. l

(m) 2 3 4 5 6 7 8

T1 (s) 2,84 3.47 4.01 4.49 4.91 5.31 5.67

2 - 57


It remains for us to examine the effect of the horizontal force Fc max on the loading conditions sustained by the structure. This force actually exist, so that any components such as the crab which transmit it directly must be designed to withstand it. The configuration of the load acting on the girder as a whole therefore deserves some attention. Let us first consider the case where Fc max occurs before the crane has come to halt. It would be incorrect to consider the latter as a beam supported at both ends and subjected at its centre to the force Fc max. One must not lose sight of the fact that each of the two supporting points can transmit only a reaction F/2. The successive diagrams in figure A.2.2.2. illustrate how the problem must be considered. Diagram "a" represents the ideal state of equilibrium, in which the system as a whole is subjected to a deceleration Jm (or an acceleration x" = - Jm) and in which the rope develops a force Fcm. Each material element dm of the system therefore sustains an inertia force Jm dm. Diagram "a" is a superimposition of diagram "b" and diagram "c". Diagram "b" relates to the load due to the inertia forces on the crane proper (this is dealt with in paragraph 2.3.), while diagram "c" shows the effect of the load due to the rope. In point of fact, the actual force developed by the rope is not the force Fcm represented in diagram "c" but the force : Fc max = ψh . Fcm (2.2.31.) Since the supporting points (braked wheels) are not capable of increasing their reaction, the excess force (ψh - 1)Fcm can only result in a supplementary acceleration x" expressed by : x" = ( ψh - 1) . Fcm / m (2.2.32.) which is translated into a distributed load - x" dm on all material elements of the crane. Diagram "d" consequently represents the loading configuration to be taken into account when designing the girders. Let us consider the case in which Fc max arises after the crane has halted. This time, the braked wheels no longer have to devote part of the reaction of which they are capable to taking up the inertia forces on the crane, and in general, must be regarded as being fixed. This being so, the girder must be designed as if it were supported at each end and subjected to the force Fc max at its centre. This latter case is in point of fact the only one which needs to be considered, because when Fc reaches its maximum value 2 Fcm before the crane comes to a standstill this force can still arise in the course of the pendulum motion which follows after it has stopped. All the preceding considerations still hold good if, instead of considering a braking phase, one is dealing with an acceleration phase of the crane, in the course of which it is speeded up, by a constant driving torque, from rest to a given steady speed.

2 - 58


2.3. - EFFECT OF INERTIA FORCES ON THE STRUCTURAL STEELWORK In the previous chapter, the structure was assumed to be perfectly rigid. In fact, however, it possesses a degree of elasticity and consequently also assumes an oscillating motion during the braking period and after coming to rest. Because the structure is composed essentially of distributed rather than simple concentrated masses, it is usually very difficult to determine the motion theoretically, and such calculations would be justified only in the case of very large appliances in which the inertia forces play an appreciable part. LOADING

ACCELERATION

Figure A.2.2.2.

2 - 59


In almost all cases, it will suffice to represent a structure as being a simple oscillating system having restoring forces proportional to the extension and subjected to the overall acceleration of the reference system to which it is referred. In view of the note following expression (2.2.1.) the considerations developed in paragraph 2.2. can be applied here also. However, the natural period of the oscillations (comparable to the period T1 of paragraph 2.2.) is always appreciably shorter than that of a suspended load, not exceeding a few tenths of a second in most cases. The result of this is that the parameter corresponding to β always exceeds the critical value βcrit, so that ψh must always be taken equal to 2, this being the coefficient applicable to inertia loads calculated with the mean deceleration Jm. The only exception that could be made to this rule would be for extremely brief retardation phases, such as those resulting from braking a lowspeed travel motion, with the wheels slipping on the rails. Because the oscillating motions of the structure have a high frequency, the maximum resulting loadings are superimposed momentarily upon those due to the load.

2 - 60


3. - CALCULATING THE LOADS IN THE CASE OF A SLEWING MOTION In the case of a slewing motion, considerations similar to those of chapter 2 can be developed. To calculate the effects of the inertia forces on the load, it is only necessary to determine m from the relation : m v2 = I ω2 (3.1.) where v

is the horizontal linear velocity of the suspension point of the load ;

I

is the moment of inertia of all parts in motion (structure, machinery, motor) referred to a particular shaft,

ω

is the angular velocity of that shaft corresponding to the velocity v above.

4. - CALCULATING THE LOADS IN THE CASE OF A LUFFING MOTION In the case of a luffing motion, considerations similar to those of chapter 2 can be developed. It will suffice to determine m from the relation : 2 (4.1.) mv =2T where v is the horizontal linear velocity of the suspension point of the load. T

is the total kinetic energy of the masses in motion when the horizontal linear velocity of the suspension point of the load is equal to v.

5. - SYSTEMS WITH REGULATED ACCELERATION In some control systems, such as certain Ward-Leonard or hydraulically actuated systems, the magnitudes of the accelerations and decelerations are dictated by the characteristics of the system and are maintained constant regardless of external conditions. For this reason, load swing does not disturb the acceleration or deceleration conditions of the appliance or part of the appliance in motion. In the exempla dealt with in paragraph 2.2., this would be tantamount to assuming that x" is a given constant. Using equation (2.2.12.) and its derivations, it is easy to show that in this case for β ≤ 0,5 (5.1.) ψh = 2 sinβ π for β > 0,5 (5.1.) ψh = 2 Such a situation would also obtain if one assumed the mass m1 to be infinitely small compared with m and therefore unable to disturb its motion. This being so, (5.1.) is the limiting curve obtained by making µ tend to zero, and is represented on diagram 2.2.1. by the curve µ = 0. The considerations in paragraph 2.3. are in no way modified.

2 - 61


6. - GENERAL CONCLUSIONS Knowing the torque or the braking or accelerating force, the first step is to calculate the mean deceleration or acceleration Jm, obtained on the assumption that the various parts of the structure are perfectly rigid and the load is concentrated at its point of suspension. Using this acceleration, one calculates the inertia forces acting on both the load and the various elements of the structure. These forces are then-multiplied by a certain coefficient ψh in order to take account of the elasticity of the various connections. For the inertia forces acting on the structure ψh is always taken equal to 2, except possibly in the special case mentioned in the penultimate paragraph of 2.3., provided justification can be provided for the reduction. In the case of the inertia forces acting on the load, the mass m is calculated (incorporating, where necessary, the mass equivalent to the inertia of the motor and the mechanism) and the mean deceleration or acceleration time Tm is then determined on the basis of the maximum steady speed of the motion. The value of T1 depends on the suspension length of the load in its uppermost position, and is therefore known. Hence one can determine the parameters µ and β (µ = 0 in the case of a regulated-acceleration system), and figure A.2.2.1. furnishes the corresponding value ψh. In almost all cases, the maximum force appears or can appear after completion of the braking or starting phase under consideration. Its effects on the structure can be ascertained by applying the ordinary laws of statics. It will be noted that the calculations developed in Chapter 2 assume the load to be relatively at rest (z = z' = 0) at the initial time t = 0. If this is not so, the motion of the system is affected and ψh can reach values considerably higher than those we have fixed. Such a situation could arise for instance when a motion is braked by repeated intermittent applications of the brake, or when successive motions take place at fairly short intervals. The method of calculation indicated above is therefore not excessive in any way, and special cases exist in which it would be well to exercise some caution in applying it.

2 - 62



F.E.M.

SECTION I

3rd EDITION REVISED


1.001

1998.10.01


CALCULATING THE STRESSES IN STRUCTURES



Booklet 3

CALCULATING THE STRESSES IN STRUCTURES

INTRODUCTION ......................................................................................................................... 3 3.1. THE CHOICE OF STEEL QUALITIES ................................................................................. 4 3.1.1. ASSESSMENT OF THE FACTORS WHICH INFLUENCE BRITTLE FRACTURE ........ 4 3.1.1.1. INFLUENCE A : COMBINED EFFECT OF LONGITUDINAL RESIDUAL TENSILE STRESSES WITH STRESSES FROM DEAD WEIGHT .....................................................................4 3.1.1.2. INFLUENCE B : THICKNESS OF MEMBER t........................................................................6 3.1.1.3. INFLUENCE C : INFLUENCE OF COLD ...............................................................................7

3.1.2. DETERMINATION OF THE REQUIRED STEEL QUALITY GROUP............................. 7 3.1.3. QUALITY OF STEELS ................................................................................................... 8 3.1.4. SPECIAL RULES ......................................................................................................... 10 3.2. CHECKING WITH RESPECT TO THE ELASTIC LIMIT.................................................... 11 3.2.1. STRUCTURAL MEMBERS OTHER THAN JOINTS .................................................... 11 3.2.1.1. MEMBERS SUBJECTED TO SIMPLE TENSION OR COMPRESSION ...............................11 3.2.1.2. MEMBERS SUBJECTED TO SHEAR .................................................................................12 3.2.1.3. MEMBERS SUBJECTED TO COMBINED LOADS - EQUIVALENT STRESS ......................12

3.2.2. CASE OF JOINTS........................................................................................................ 13 3.2.2.1 RIVETED JOINTS................................................................................................................13 1 - Rivets in shear.................................................................................................................................................13 2 - Rivets in tension ..............................................................................................................................................13 3 - Rivets loaded in tension and shear ................................................................................................................14 4 - Limit of bearing pressure ................................................................................................................................14 5 - Notes concerning riveted joints ......................................................................................................................14

3.2.2.2. BOLTED JOINTS................................................................................................................14 3.2.2.2.0. GENERAL ............................................................................................................................................14 3.2.2.2.1. JOINTS MADE WITH TENSION BOLTS WITH CONTROLLED TIGHTENING...............................14 3.2.2.2.2. BOLTED JOINTS SUBJECTED TO FORCES ACTING PARALLEL TO THE JOINT PLANE ........18

3.2.2.3. WELDED JOINTS ...............................................................................................................21

3.3. CHECKING MEMBERS SUBJECT TO CRIPPLING ......................................................... 23 3.4. CHECKING MEMBERS SUBJECT TO BUCKLING.......................................................... 24 3.5. CASE OF STRUCTURES SUBJECTED TO SIGNIFICANT DEFORMATION .................. 25


3.6. CHECKING MEMBERS SUBJECTED TO FATIGUE ........................................................ 26 3.6.1. CONVENTIONAL NUMBER OF CYCLES AND STRESS SPECTRUM ...................... 26 3.6.2. MATERIAL USED AND NOTCH EFFECT ................................................................... 26 3.6.3. DETERMINATION OF THE MAXIMUM STRESS σmax ................................................ 26 3.6.4. THE RATIO κ BETWEEN THE EXTREME STRESSES.............................................. 27 3.6.5. CHECKING MEMBERS SUBJECTED TO FATIGUE .................................................. 27 .APPENDIX ............................................................................................................................... 28 A - 3.2.2.2.2.3. DESIGN OF JOINTS USING HIGH STRENGTH BOLTS WITH CONTROLLED TIGHTENING ............................................................................................... 28 A- 3.2.2.3. - STRESSES IN WELDED JOINTS...................................................................... 32 A - 3.3. - CHECKING STRUCTURAL MEMBERS SUBJECT TO CRIPPLING...................... 34 A - 3.6. - CHECKING STRUCTURAL MEMBERS SUBJECT TO FATIGUE.......................... 46 1 - VERIFICATION OF STRUCTURAL MEMBERS..........................................................................47 2 - VERIFICATION OF THE JOINING MEANS (welds, bolts, rivets).................................................54

EXAMPLES OF CALCULATING CHECKS .............................................................................. 67 EXAMPLES OF FATIGUE CHECKS FOR A WELDED WEB TO FLANGE JOINT STEEL St 37....................................................................................................................................... 67 CHECKING FOR FATIGUE AND ELASTIC LIMIT................................................................. 68 FIRST EXAMPLE - COMPONENT IN GROUP E4 WITH FILLET WELD (O Q )................................68 SECOND EXAMPLE - COMPONENT IN GROUP E6 - K WELD (S.Q.) ............................................71


INTRODUCTION The stresses set up in the various structural members are determined for the three cases of loading defined under section 2.3., and a check is made to ensure that there is a sufficient safety coefficient ν in respect of the critical stresses, considering the following three possible causes of failure : - exceeding the elastic limit ; - exceeding the critical crippling or buckling load ; - exceeding the limit of endurance to fatigue. The quality of the steels used must be stated and the physical properties, chemical composition and welding qualities must be guaranteed by the manufacturer of the material. The permissible stresses for the materials used are determined as stipulated in clauses 3.2., 3.3., 3.4. and 3.6. thereunder, with reference to the critical stresses for the material. These critical stresses are those which correspond either to the elastic limit (which in practice, involves establishing the stress corresponding to a critical limit for elongation), or to the critical stress for crippling or buckling, or, in the case of fatigue, to the stress for which the probability of survival, under tests, is 90 %. The stresses in the structural members shall be calculated on the basis of the different cases of loading envisaged under section 2.3. by applying conventional strength of materials calculation procedures. The sections of metal to be considered shall be the gross sections (i.e. without deducting the areas of holes) for all parts which are subjected to compression loads 1, and the net sections (i.e. with the areas of holes deducted) for all parts subjected to tensile loads. In the case of a member subjected to bending, a half-net section should be assumed, taking the net section in parts under tension and the gross section in parts under compression. To simplify the calculations, however, one may use either the section modulus of the net section or the section modulus computed for the half-net section, using as centre of gravity of the section that of the gross section.

1

The area of the holes shall be included in the cross-sectional area only when they are filled by a rivet or a bolt.


3.1. THE CHOICE OF STEEL QUALITIES The verifications required in the design rules for the safety of the structure against yielding, instability and fatigue failure do not guarantee safety against brittle fracture. In order to obtain sufficient safety against brittle fracture, a steel quality has to be chosen depending on the conditions influencing brittle fracture. The most important influences on the sensitivity to brittle fracture in steel structures are : A. Combined effect of longitudinal residual tensile stresses with tensile stresses from dead load. B. Thickness of the member. C. Influence of cold. Influences A, B, and C are valued with points. The required steel quality depends on the sum of these points.

3.1.1. ASSESSMENT OF THE FACTORS WHICH INFLUENCE BRITTLE FRACTURE In the following, influences A, B, and C in paragraph 3.1. are described and quantified. 3.1.1.1. INFLUENCE A : COMBINED EFFECT OF LONGITUDINAL RESIDUAL TENSILE STRESSES WITH STRESSES FROM DEAD WEIGHT Equations for lines I, II and III in figure 3.1.1.1. σa = permissible tensile stress with respect to the elastic limit, loading case I.

Line I : no welds, or only transverse welds ZA = σG / ( 0,5 . σa ) - 1 valid only for σG ≥ 0,5 . σa

σG = tensile stress from permanent load, e.g. from dead weight

Line II : longitudinal welds

ZA = assessing coefficient for influence A

Line III : accumulation of welds

ZA = σG / ( 0,5 . σa )

ZA = σG / ( 0,5 . σa ) + 1


The danger of brittle fracture is increased by high stress concentrations, in particular by 3-axial tensile stresses, as is the case with an accumulation of welds. If members with low-stresses are stress relieved after welding (approx. 600-650° C) line I can be used for all types of welds. Assessing coefficient

I : No welds or only tranverse welds II : Longitudinal welds III : Weld accumulation Figure 3.1.1.1. - ZA in terms of stresses and welds


3.1.1.2. INFLUENCE B : THICKNESS OF MEMBER t

t = Thickness of member in mm ZB = Assessing coefficient for influence B from t = 5 to t = 20 mm ZB = 9 . t2 / 2500. from t = 20 to t = 100 mm ZB = 0,65 . ( t - 14,81 )0,5 -0,05 Thickness of member Figure 3.1.1.2. - Assessing coefficient ZB = f (t)

t mm 5 6 7 8 9 10 12 15

ZB 0,10 0,15 0,20 0,25 0,30 0,40 0,50 0,80

t mm 16 20 25 30 35 40 45 50 55

ZB 0,9 1,45 2,0 2,5 2,9 3,2 3,5 3,8 4,0

t mm 60 65 70 75 80 85 90 95 100

ZB 4,3 4,55 4,8 5,0 5,2 5,4 5,6 5,8 6,0

For rolled sections an idealised thickness t* is to be used. This is : for round sections :

t* = d / 1,8

for square sections :

t* = t / 1,8

for rectangular sections :

t* = b / 1,8

where b represents the larger side of the rectangle and the ratio of the sides b / t ≤ 1,8 For b / t > 1,8 then t* = t.


3.1.1.3. INFLUENCE C : INFLUENCE OF COLD The lowest temperature at the place of erection of the crane determines the classification. This temperature is generally lower than the working temperature.

T = Temperature at the place of erection in °C ZC = Assessing coefft. for influence C from T = 0°C to T = - 30°C take ZC = 6 . T2 / 1600 from T = - 30°C to T = - 55°C take ZC = [ ( - 2,25 . T ) -33,75 ] / 10 T en ° C 0 -5 - 10 - 15 - 20 - 25

ZC 0,0 0,1 0,4 0,8 1,5 2,3

T en ° C - 30 - 35 - 40 - 45 - 50 - 55

ZC 3,4 4,5 5,6 6,7 7,9 9,0

Temperature T in °C Figure 3.1.1.3. - Assessing coefft. ZC = f (T)

3.1.2. DETERMINATION OF THE REQUIRED STEEL QUALITY GROUP It is the sum of assessing coefficients from paragraph 3.1.1. which determines the minimum required quality for the steel structure. Table T. 3.1.2. shows the classification of the quality group in relation to the sum of the assessing coefficients. If the sum of the assessing coefficients is higher than 16 or if the required steel quality cannot be obtained, special measures will be necessary to obtain the steel quality necessary for safety against brittle fracture which will have to be determined with material experts.


Table T.3.1.2. Classification of quality groups in relation to the sum of the assessing coefficients Sum of the assessing coefficients from paragraph 3.1.1. ΣZ = ZA + ZB + ZC

Quality group corresponding in table T.3.1.3

≤ 2

1

≤ 4

2

≤ 8

3

≤ 16

4

3.1.3. QUALITY OF STEELS The quality of steels in these design rules is the property of steel to exhibit a ductile behaviour at determined temperatures. The steels are divided into four quality groups. The group in which the steel is classified, is obtained from its notch ductility in a given test and temperature. Table T.3.1.3. comprises the notch ductility values and test temperatures for the four quality groups. The indicated notch ductilities are minimum values, being the mean values from three tests, where no value must be below 20 Nm/cm2. The notch ductility is to be determined in accordance with V-notch impact tests to ISO R 148 and Euronorm 45-63. Steels of different quality groups can be welded together. TC is the test temperature for the V-notch impact test. T is the temperature at the place of erection of the crane. TC and T are not directly comparable as the V-notch impact test imposes a more unfavourable condition than the loading on the crane in or out of service.


Table T.3.1.3. - Quality groups

Quality group

1

2

3

4

Notch ductility measured in ISO sharp notch test ISO R 148 in Nm/cm2

-

35

35

35

Test temperature

Steels, corresponding to the quality group

Standard

Tc °C

-

+ 20

0

-20

Designation of steels Fe 360 - A Fe 430 - A St 37 - 2 St 44 - 2 E 24 - 1 43 A 50 B * Fe 360 - B Fe 430 - B Fe 510 - B R St 37-2 St 44-2 E 24 (A37) - 2 E 26 (A42) - 2 E 36 (A52) - 2 40 B 43 B * Fe 360 - C Fe 430 - C Fe 510 - C St 37 - 3U St 44 - 3U St 52 - 3U E 24 (A37) - 3 E 26 (A42) - 3 E 36 (A52) - 3 40 C 43 C * 50 C 55 C Fe 360 - D Fe 410 - D Fe 510 - D St 37 - 3N St 44 - 3N St 52 - 3N E 24 (A37) - 4 E 26 (A42) - 4 E 36 (A52) - 4 40 D 43 D * 50 D 55 E

Euronorm 25 DIN 17100 NF A 35-501 BS 4360 1972 Euronorm 25 DIN 17100

NF A 35-501 BS 4360 1972 Euronorm 25

DIN 17100

NF A 35-501 BS 4360 1972

Euronorm 25

DIN 17100

NF A 35-501 BS 4360 1972

* The test requirements of steels to BS.4360 do not in all cases agree with the Euronorm and other national standards, and the guaranteed impact test properties for steels to BS.4360 may be different to other steels in the same quality group. Impact test properties are stated in BS. 4360 and where the requirements are different from those guaranteed in BS. 4360, agreement must be obtained from the steel suppliers.


3.1.4. SPECIAL RULES In addition to the above provisions for the choice of the steel quality, the following rules are to be observed : 1 - Non killed steels of group 1 shall be used for load carrying structures only in case of rolled sections and tubes not exceeding 6 mm thickness. 2 - Members of more than 50 mm thickness, shall not be used for welded load carrying structures unless the manufacturer has a comprehensive experience in the welding of thick plates. The steel quality and its testing has in this case to be determined by specialists. 3 - If parts are cold bent with a radius/plate thickness ratio < 10 the steel quality has to be suitable for folding or cold flanging.


3.2. CHECKING WITH RESPECT TO THE ELASTIC LIMIT For this check, a distinction is made between the actual members of the structure and the riveted, bolted or welded joints.

3.2.1. STRUCTURAL MEMBERS OTHER THAN JOINTS 3.2.1.1. MEMBERS SUBJECTED TO SIMPLE TENSION OR COMPRESSION 1) Case of steels for which the ratio between the elastic limit σE and the ultimate tensile strength σR is < 0,7. The computed stress σ must not exceed the maximum permissible stress σa obtained by dividing the elastic limit stress σE by a coefficient νE which depends upon the case of loading as defined under section 2.3. The values of νE and the permissible stresses are : Values of νE

Case I 1,5

Case II 1,33

Case III 1,1

Permissible stresses σa

σE / 1,5

σE / 1,33

σE / 1,1

For carbon steels of current manufacture A.37 - A.42 - A.52 (also called E.24 - E.26 - E.36 or Fe 360 - Fe 510) the critical stress σE is conventionally taken as that which corresponds to an elongation of 0,2 %. Table T.3.2.1.1. - Values of σE and σa for steels A.37 - A.42 - A.52

STEELS

Elastic limit σE N/mm2

Maximum permissible stresses : σa Case I Case II Case III 2 2 N/mm N/mm N/mm2

E.24 (A.37, Fe 360)

240

160

180

215

E 26 (A.42)

260

175

195

240

E.36 (A.52, Fe 510)

360

240

270

325


2) Case of steels with high elastic limit ( σE / σR > 0,7) For steels with high elastic limit where the ratio σE / σR is greater than 0.7, the use of the νE coefficients does not ensure a sufficient margin of safety. In this case a check can be made that the permissible stress σa given by the formula below is not exceeded : σa = [ (σE + σR ) / (σE.52 + σR.52 ) ] . σa 52 where : σE and σR are the elastic limit and the ultimate tensile strength of the steel considered σE.52 and σR.52 these same stresses for steel A.52, i.e. 360 N/mm2 and 510 N/mm2 σa 52 the permissible stress for steel A.52 in the case of loading considered. 3.2.1.2. MEMBERS SUBJECTED TO SHEAR The permissible stress in shear τa has the following value : τa = σa / 3 0,5 σa being the permissible tensile stress. 3.2.1.3. MEMBERS SUBJECTED TO COMBINED LOADS - EQUIVALENT STRESS σx , σy and τxy being respectively the two normal stresses and the shear stress at a given point, a check shall be made : 1 - that each of the two stresses σx and σy is less than σa and that τxy is less than τa 2 - that the equivalent stress σcp is less than σa, i.e. : σcp = ( σx2 + σy2 - σx . σy + 3 . τxy2 ) 0,5 ≤ σa When using this formula, a simple method is to take the maximum values σx , σy and τxy. But, in fact, such a calculation leads to too great an equivalent stress if it is impossible for the maximum values of each of the three stresses to occur simultaneously. Nevertheless, the simple calculation method, being conservative, is always acceptable. If it is desired to calculate more precisely, it is necessary to determine the most unfavourable practical combination that may occur. Three checks must then be made by calculating successively the equivalent stress resulting from the three following combinations : σx max and the corresponding stresses σy and τxy σy max and the corresponding stresses σx and τxy τxy maw and the corresponding stresses σx and σy


Note : It should be noted that when two out of the three stresses are approximately of the same value, and greater than half the permissible stress, the most unfavourable combination of the three values may occur in different loading cases from those corresponding to the maximum of each of the three stresses. Special case : - Tension (or compression) combined with shear The following formula should be checked : ( σ2 + 3 . τ2 )0,5 ≤ σa

3.2.2. CASE OF JOINTS 3.2.2.1 RIVETED JOINTS 1 - Rivets in shear

Taking the effect of the clamping force into account, the calculated shearing stress τ must not exceed : τ = 0,6 . σa in the case of single shear and τ = 0,8 . σa in the case of double or multiple shear where σa is the permissible tensile stress of the metal used for the rivet.

single shear

double or multiple shear

2 - Rivets in tension

The calculated tensile stress σ must not exceed the value : σ = 0,2 . σa


3 - Rivets loaded in tension and shear

The following conditions must be checked : σ ≤ 0,2 . σa and τ ≤ 0,6 . σa for single shear for double shear or τ ≤ 0,8 . σa 4 - Limit of bearing pressure

The bearing pressure in the walls of holes σn must not exceed : σn ≤ 1,5 . σa for single shear σn ≤ 2 . σa for double shear 5 - Notes concerning riveted joints

a) Rivets subjected to tension should be avoided, particularly for the main members ; b) all joints must have at least two rivets aligned in the direction of the force. 3.2.2.2. BOLTED JOINTS 3.2.2.2.0. GENERAL

Bolted joints may be subjected to stresses due to forces acting perpendicular to the joint (joints by tension bolts), due to forces acting parallel to the joint surfaces, and due to forces acting simultaneously perpendicular and parallel to the joint surface. 3.2.2.2.1. JOINTS MADE WITH TENSION BOLTS WITH CONTROLLED TIGHTENING 1 - General

A joint by tension bolts with controlled tightening is a joint in which the main tension is in the direction of the axis of the bolt, screw or threaded rod and which has been subject to a tightening effect, applied in the absence of any external load, which is recommended for all joints subjected to fatigue. Care must be taken to ensure that the bolt is not subjected to shear loading. These bolts do not come into the category of H.S. bolts but may be used if they fulfil the conditions of 3.2.2.2.2.3. Care should be taken to ensure that the bolts are correctly tightened and that the tightening is permanent (tolerance +/- 10 %). Factor Ω = 1,1 is introduced to take account of tolerances. During the application of the initial tightening on the bolt, under the combined effect of tension and torsional loading the stress should not exceed 80 % of the elastic limit, taking account of the scatter in applying the initial tightening.


2 - Calculation of the permissible load on joints A - Calculation of the initial tightening force to be used a) Tightening with twist

σb = (σp 2 + 3 . τb2 )0,5 ≤ 0,8 σE τb = [ ( 2 . d2 . σp ) / dt ] . [ pa / ( π . d2 ) + 1,155 . µ]

where : σp = τb = d2 = dt = = pa µ = σE =

theoretical tensile stress under the tightening effect torsional stress under the tightening effect diameter of the root of the thread nominal diameter of the bolt thread pitch friction coefficient in the threads elastic limit of the bolt metal σb ≤ 0,8 σE

b) Tightening without twist B - Permissible load F1 on the joint Two checks are to be made :

a) Under the maximum load, taking into account the safety coefficient κ and κ‘, the elastic limit of the bolt must not be exceeded. σ‘1 = ( σb2 - 3 . τb2 )0,5

determine : check that :

F1 / Sb ≤ ( σ‘1 - σp ) / ( κ . κ‘ . δb ) where :

Sb = section of the root < section of the shank. δb = ∆l1 / (∆l1 + ∆l2 ) ∆l1 = shortening of the elements to be tightened under the action of the tightening force ∆l2 = lengthening of bolt under the action of the tightening force. For assembled steel parts, the section to be considered for ∆l1 : Seq = 0,25 . π . [ ( S1 + 0,1 . lk )2 - Dt2 ] where : S1 = bearing diameter under head lk = length of tightened parts Dt = diameter of bolt holes


For bolts whose shank diameter differs considerably from the root diameter of the thread and where there is an appreciable threaded length remaining in the part submitted to stress, a complete calculation of ∆l2 should be made. b) Under the maximum load with application of coefficients Ω, κ' and κ " separation of the parts should not occur. σ1 = F1 / Sb ≤ σp / [ κ' . κ'‘ . ( 1 - δb ) . Ω ] Safety coefficients κ , κ' and κ'‘ κ depends on the surface state of the parts to be tightened (machined surface κ = 1 ) κ' corresponds to safety in relation to the elastic limit in accordance with table T.3.2.2.2. κ'' corresponds to safety against separation of the parts. Table T.3.2.2.2.

Case I

Case II

Case III

κ‘

1,50

1,33

1,1

κ‘’

1,3

1,0

1,0

Note : The coefficients κ‘ and κ‘’ should be applied to the most unfavourable condition arising from the scatter in applying the initial tightening effort. C - Checking for fatigue Checking bolts for fatigue is carried out solely for case I loads. Under the effect of the service load F1, the true tensile stress varies between the values : σp

and

σp + ( F1 . δb ) / Sb

The following equation must be verified : σ1 = F1 / Sb ≤ 2 . σA / δb σA is the amplitude of the maximum permissible stress for fatigue given in the following graph. For any other type of bolt or design method the σA value should ensure at least an equivalent lever of safety against fatigue. Any conformity tests should be carried out according to ISO specification 3800/1 with σm = 0,8 . RE . ( R E = σE )


Amplitude of maximum permissible fatigue stress Graph for ISO bolts - standard thread - classes 8.8, 10.9, 12.9 - cold rolled thread with heat treatment after rolling


3.2.2.2.2. BOLTED JOINTS SUBJECTED TO FORCES ACTING PARALLEL TO THE JOINT PLANE

1 - Bolts subjected to shear (fitted bolts) Preferably for non-fluctuating stresses with and without preload. The following checks presuppose that the bolting has been effected under proper conditions, i.e. using fitted bolts (turned or cold finished) with ISO tolerances and the shanks of which bear against the full length of holes drilled in the parts being assembled. Holes must be drilled and reamed with the ISO tolerances. Black bolts are permitted only for secondary joints which do not transmit heavy loads. They are prohibited for joints subject to fatigue. The calculated stress τ on the shank shall not exceed the values given for rivets in clause 3.2.2.1.1-. The bearing pressure shall not exceed the values indicated in clause 3.2.2.1.4. 2 - Bolts subjected to combined tension and shear A check shall be made that : σ ≤ 0,65 . σa and

τ ≤ 0,6 . σa

for single shear

or

τ ≤ 0,8 . σa

for double shear

and that

( σ2 + 3 . τ2 )0,5 ≤ σa

The permissible stress in a bolt is limited to : σa = 0,7 . σE(0,2) for normal execution σa = 0,8 . σE(0,2) for a construction which prevents stripping the thread. where σE(0,2) is the 0,2 % proof stress of the metal constituting the bolt.


3 - Joints using high strength bolts with controlled tightening (H.S.) This type of joint is recommended for assemblies subjected to fatigue and whose main loads are parallel to joint faces. Members joined by H.S. bolts are subjected to the following types of loads : A - Loads acting in the plane of the joint (symbol T) In this case, the loads tend to make the parts in contact slip and the force is transmitted by friction. To determine the permissible load per bolt Ta which can be transmitted by friction, the tensile force F which exists in the bolt after tightening must be considered. This is multiplied by the coefficient of friction µ of the contact surfaces, and the safety coefficients νT which are the same as those in clause (3.2.1.1.) are applied to this limiting force, i.e. νT = 1,5

for case I loading

= 1,33

for case II loading

= 1,1

for case III loading

This may be expressed : Ta = m . ( µ . F )/ νT m being the number of friction surfaces. The tension, F, in a bolt depends upon the tightening torque ; the value of µ depends upon the metal constituting the members, the state of the surfaces in contact, and the method of preparation. Appendix A 3.2.2.2.2.3. gives information on this subject. B - Forces perpendicular to the plane of the joint (symbol N) The checking by calculation of the forces perpendicular to the assembly surface shall be carried out in accordance with clause 3.2.2.2.1. If the bolted joint is subjected to an external couple M, the tensile loading has to be determined at the bolt which is subjected to the maximum loading and, where applicable, added to the existing tensile load N.


C - Combined loads of the T, N and M types Two checks must be made : a) That, for the most highly stressed bolt, the sum of the tensile forces due to N and M loadings remains less than the permissible tensile force as defined in 3.2.2.2.2. b) That the mean load which is transmitted by friction is less than the following value : T = µ . ( F - N ) . m / νT D - Determination of the stresses in the members joined For members subject to compression, the stress is calculated on the gross section (cross sectional area of the holes not deducted). For members subjected to tension there are two cases : 1st case : Bolts set in a single row, perpendicular to the direction of the load ; the following conditions must be checked : a) the total load on the gross section b) 60 % of the total load on the net section (cross-sectional area of holes deducted) 2nd case : Several rows of bolts perpendicular to the direction of the load.

The most heavily loaded section (corresponding to row 1 for the member A - see figure) must be analysed and the following two conditions checked : a) the total load on the gross section, and b) on the net section, the total load from rows 2 and 3 (i.e. in the case of the figure, 2/3 of the total load of the joint) to which 60 % of the load taken by row 1 is added. This assumes that the load is equally divided amongst all the bolts and that the number of rows of bolts is small because if there are too many, the last bolts carry little load. It is therefore recommended that not more than two rows of bolts should be used or exceptionally three.


E - Execution of joints with high-strength bolts It must be emphasised that the above calculations to check the adequacy of joints with high strength bolts are valid only for joints made in accordance with accepted practice which requires controlled tightening of the bolts and preparation of the contact surfaces to obtain suitable coefficients of friction. See appendix A - 3.2.2.2.2.3. for further guidance. 3.2.2.3. WELDED JOINTS In welded joints, it is assumed that the deposited metal has at least as good characteristics as the parent metal. It must be verified that the stresses developed, in the cases of longitudinal tension and compression, do not exceed the permissible stresses σa given in clause 3.2.1.1. For shear in the welds, the permissible stress τa is given by : τa = σa / 20,5 However, for certain types of loading, particularly transverse stresses in the welds, the maximum permissible equivalent stress is reduced. Table T. 3.2.2.3. summarises the values not to be exceeded, for certain steels, according to the type of loading.


Table T. 3.2.2.3. - Maximum permissible equivalent stresses in welds (N/mm2) : steels A.37 (Fe 360) - A.42 - A.52 (Fe 510) Types of loading Case I

A.37 Case II Case III

Case I


Case I


Longitudinal equivalent stresses for all types

160

180

215

175

195

240

240

270

325

160

180

215

175

195

240

240

270

325

140

158

185

153

170

210

210

236

285

113

127

152

124

138

170

170

191

230

160

180

215

175

195

240

240

270

325

130

146

175

142

158

195

195

220

265

113

127

152,

124

138

170

170

191

230

of welds

Transverse tensile stresses 1) Butt-welds and special quality K welds 2) Ordinary quality Kwelds 3) Fillet welds

Transverse compressive stresses 1) Butt-welds and K welds 2) Fillet welds

Shear All tapes of welds

Appendix A-3.2.2.3. gives some additional information on welded joints.


3.3. CHECKING MEMBERS SUBJECT TO CRIPPLING The guiding principle shall be that parts subject to crippling must be designed with the same safety margin as that adopted in respect of the elastic limit ; in other words, having determined the practical crippling stress, the maximum permissible stress shall be the crippling stress divided by the appropriate coefficient 1,5 or 1,33 or 1,1 specified in 3.2.1.1. The choice of a practical method of calculation is left to the manufacturer who must state the origin of the method chosen. If the method chosen involves multiplying the computed stress by a crippling coefficient ω depending upon the slenderness ratio of the member and then checking that this amplified stress remains less than a certain allowable stress, the value to be chosen for this allowable stress shall be as specified in 3.2.1.1. Note : Appendix A-3.3. shows how to apply various classical methods of calculation in accordance with the above requirements.


3.4. CHECKING MEMBERS SUBJECT TO BUCKLING In determining the new buckling safety coefficients, stated below, it was considered that flat plates under compressive stresses equally distributed over the plate width, are exposed to a greater danger of buckling than plates under stresses changing from compression to tension over the plate width. In consequence, safety against buckling was made dependent on the ratio ψ of stresses at the plate edges (appendix A-3.4.) In addition it was found necessary to determine the critical buckling stress for circular cylinders and the spacing and moment of inertia of the transverse stiffeners in order to avoid too great divergences in the effective safety due to the use of highly divergent data in technical literature. It shall be verified that the calculated stress is not higher than the critical buckling stress divided by the following coefficients νV :

Buckling of plane members

Buckling of curved members ; Circular cylinders (e.g. tubes)

Case I II III I II III

Buckling safety νV 1,70 + 0,175 ( ψ - 1) 1,50 + 0,125 (ψ - 1) 1,35 + 0,075 (ψ - 1) 1,70 1,50 1,35

The edge-stresses ratio ψ varies between + 1 and - 1. Appendix A.3.4. gives the procedure for determining the critical buckling stress.


3.5. CASE OF STRUCTURES SUBJECTED TO SIGNIFICANT DEFORMATION In this case the stresses in the members may not be proportional to the forces which cause them due to the deformation of the structure as a result of the application of these forces. This is the case, for exempla, with the stresses produced in the column of a crane (illustrated diagrammatically) where it is clear that the moment in the column is not proportional to the forces applied because of deformations which increase their moment arm. In this case the calculation is made as follows : 1 - First make the checks required by clauses 3.2. - 3.3. - 3.4. calculating the stresses resulting from the various cases of loading and checking that there is a sufficient safety margin in relation to the critical stresses (elastic limit, crippling, buckling). In the calculation of the stresses account is taken of the deformation due to the loads on the structure. 2 - A further check is also carried out by calculating the stresses resulting from the application of the loads multiplied by the coefficient ν of the case of loading considered and taking into account the deformations resulting from the application of these increased loads and checking that the stresses thus calculated remain less than the critical stresses for the elastic limit, for crippling and for buckling. However, to take account of the fact that the variable loads Sv (loads due to the hoisted load multiplied by ψ, to the wind and to horizontal movements) are more dangerous than the constant load due to the dead weight SG, a check can be made in practice by considering two cases as follows : 1 - When the effects of the dead weight SG and of the variable load Sv lead to deformation in opposite directions : Determine the stress σG resulting from the application of the dead weight SG (without amplification) and σv resulting from the variable loads SV, multiplied by the coefficient ν corresponding to the case considered (clause 3.2. elastic limit, 3.3. crippling, 3.4. buckling) and check that this stress is less than the critical value i.e. : σ resulting from ( SG + ν . Sv ) ≤ σcr 2 - when the dead weight and the variable load lead to deformations in the same direction : determine the stress resulting from the application of the variable load multiplied by the coefficient ν and of the dead weight multiplied by the coefficient : ν‘ = 1 + ( ν - 1 ) . r where r = σG / (σG + σV ) calculated in the initial stage of the deformations. We then have : σ resulting from (ν' . SG + ν . Sv ) ≤ σcr


3.6. CHECKING MEMBERS SUBJECTED TO FATIGUE Danger of fatigue occurs when a member is subjected to varying and repeated loads. Fatigue strength is calculated by considering the following parameters : 1 - the conventional number of cycles and the stress spectrum to which the member is subjected ; 2 - the material used and the notch effect at the point being considered ; 3 - the extreme maximum stress σmax which can occur in the member ; 4 - the ratio κ between the values of the extreme stresses.

3.6.1. CONVENTIONAL NUMBER OF CYCLES AND STRESS SPECTRUM The number of cycles of variations of loading and the spectrum of stresses to be taken into consideration are discussed in clause 2.1.2.2. and in clause 2.1.2.3. These two parameters are taken into account when considering solely the group in which the member is classified in accordance with clause 2.1.4.

3.6.2. MATERIAL USED AND NOTCH EFFECT The fatigue strength of a member depends upon the quality of the material used and upon the shape and the method of making the joints. The shapes of the parts joined and the means of doing it have the effect of producing stress concentrations (or notch effects) which considerably reduce the fatigue strength of the member. Appendix A-3.6. gives a classification of various joints according to their degree of stress concentration (or notch effect).

3.6.3. DETERMINATION OF THE MAXIMUM STRESS σmax The maximum stress, σmax is the highest stress in absolute value (i.e. it may be tension or compression) which occurs in the member in loading case I referred to in clause 2.3.1. without the application of the amplifying coefficient γC When checking members in compression for fatigue the crippling coefficient, ω given in clause 3.3. should not be applied.


3.6.4. THE RATIO κ BETWEEN THE EXTREME STRESSES This ratio is determined by calculating the extreme values of the stresses to which the component is subjected under case I loadings. The ratio may vary depending upon the operating cycles but it errs on the safe side to determine this ratio κ by taking the two extreme values which can occur during possible operations under case I loadings. If cmax and cmin are the algebraic values of these extreme stresses, cmax being the extreme stress having the higher absolute value, the ratio κ may be written : κ = σmin / σmax or τmin / τmax

in the case of shear.

This ratio, which varies from + 1 to - 1, is positive if the extreme stresses are both of the same sense (fluctuating stresses) and negative when the extreme stresses are of opposite sense (alternating stresses).

3.6.5. CHECKING MEMBERS SUBJECTED TO FATIGUE Using the parameters defined in clauses 3.6.1. to 3.6.4. the adequacy of the structural members and of the joints subjected to fatigue is ensured by checking that the stress σmax , as defined in clause 3.6.3. is not greater than the permissible stress for fatigue of the members under consideration. This permissible stress for fatigue is derived from the critical stress, defined as being the stress which, on the basis of tests made with test pieces, corresponds to a 90 % probability of survival to which a coefficient of safety of 4/3 is applied thus : σa for fatigue = 0,75 . c at 90 % survival. The determination of these permissible stresses having regard to all these considerations is a complex problem and it is generally advisable to refer to specialised books on the subject. Appendix A-3.6. gives practical indications, based on the results of research in this field, on the determination of permissible stresses for A.37 - A.42 and A.52 steels, according to the various groups in which the components are classified, and the notch effects of the main types of joints used in the manufacture of hoisting appliances.


.APPENDIX A - 3.2.2.2.2.3. DESIGN OF JOINTS USING HIGH STRENGTH BOLTS WITH CONTROLLED TIGHTENING Clause 3.2.2.2.2.3. determines the general requirements to be observed for the execution of joints with high strength bolts. This appendix gives some directions on the preparation of the surfaces to be joined, the friction coefficients obtained and the tightening methods. Coefficient of friction µ The coefficient of friction used for the calculation of the force transmitted by friction depends upon the joined material and upon the preparation of the surfaces. A minimum preparation before jointing will consist in removing every trace of dust, rust, oil and paint by energetic brushing with a clean metallic brush. Oil stains must be removed by flame cleaning or by the application of suitable chemical products (carbon tetrachloride, for instance). A more careful preparation will increase the coefficient of friction. This could be sandblasting, shotblasting or oxy-acetylene flame cleaning clone not more than five hours before tightening ; brushing must be alone just prior to jointing. The coefficients of friction are given in the following table. Table T.A.3.2.2.2.2.3.1. - Values of µ Normally prepared surfaces (degreasing and brushing)

Specially prepared surfaces (flame-cleaned, shot or sand-blasted)

E-24 (A.37) Fe 360

0,30

0,50

E-26 (A.42)

0,30

0,50

E-36 (A.52) Fe 510

0,30

0,55

Joined material

It is necessary to insert two washers, one under the boit head, the other under the nut. These washers must have a 45° bevel, at least on the internal rim, and turned towards the boit head or the nut. They must be heat-treated in order that their hardness shall be at least equal to that of the metal constituting the boit


Bolt tighteninq The value of the tension induced in the bolt must reach the value determined by calculation. This tension, resulting from tightening, can be measured by calculation of the torque to be applied to the boit and given by the formula : Ma = 1,10 . C . d .F where : Ma is the torque to be applied in Nm d

is the nominal diameter of the bolt in mm

F

is the nominal tension to be induced in the bolt (kN)

C

is a coefficient depending on the thread form,the friction coefficient or the threads and between the nut and the washer.

With metric-threaded bolts and washers as delivered (slightly oiled, without rust or dust) : C = 0,18 The tensile stress in the boit must not exceed that defined under clause 3.2.2.2.2. Value of the tensile stress area of the bolts When determining the stress in the bolt, the tensile stress area shall be calculated by taking the arithmetic mean of the core (minor) diameter and the effective thread diameter. These values are given in the following table : Nominal diameter ( mm ) Tensile stress area ( mm2 )

8

10

12

14

16

18

20

22

24

27

30

36,6

58

84,3

115

157

192

245

303

353

459

561


Qualitv of the bolts Bolts used for this type of joint have a high elastic limit. The ultimate tensile strength σR must be greater than the values given hereunder : σE 0,2 N/mm2 < 700 700 to 850 > 850

σR N/mm2 > 1,15 . σE > 1,12 . σE > 1,10 . σE

The diameter of holes shall not exceed by more than 2 mm the diameter of the bolt. The following table gives per bolt and per friction surface, the values of the transmissible forces in the plane parallel to that of the joint for bolts of 1000 - 1200 N/mm2 with an elastic limit of σE = 90O N/mm2 for various friction coefficients for the steels A. 37, A.42 and A. 52. To apply these figures, the number of effective friction surfaces as indicated in the drawing below must be determined.

1 friction surface m = 1

2 friction surfaces m = 2

3 friction surfaces m = 3

Effective friction surfaces


Table T.A.3.2.2.2.2.3.2. Transmissible forces in the plane of the joint per bolt and per friction surface Bolts of 1000/1200 N/mm2 : σE = 90O N/mm2 with means of preventing stripping of the threads : σa = 0,8 . σE Boltdiameter

Tensile stress area

mm

mm

10

Normally prepared surfaces Steels A-37, A-42, A-52 µ = 0,30 Case Case Case III II I kN kN kN

Specially prepared surfaces Steels A-37, A-42 µ = 0,50 Case Case Case III II I kN kN kN

Steel A-52 µ = 0,55 Case Case III II kN kN

Clamping force

Applied torque

kN

Nm

58

41,7

82,7

8,3

9,4

11,4

13,9

15,7

18,9

15,2

17,2

20,8

12

84,3

60,6

144,0

12,1

13,6

16,5

20,2

22,8

27,5

22,2

25,0

30,3

14

115

82,7

229,0

16,5

18,6

22,5

27,5

31,0

37,6

30,2

34,2

41,4

16

157

113,0

358,0

22,6

25,5

30,8

37,7

42,5

51,4

41,5

46,8

56,5

18

192

138,0

492,0

27,6

31,0

37,6

46,0

51,8

62,7

50,6

57,0

69,0

20

245

176,0

697,0

35,2

39,7

48,0

58,5

66,1

80,0

64,5

72,7

88,0

22

303

218,0

950,0

43,6

49,3

59,7

72,5

82,0

99,0

80,0

90,2

109,0

24

353

254,0

1200,0

50,8

57,1

69,4

84,5

95,5

115,5

93,1

105,0

127,0

27

459

330,0

1760,0

66,0

74,2

90,0

110,0

124,0

150,0

121,0

136,0

165,0

2

Case I kN

For a bolt with an elastic limit of σE’ the values of the forces and of the torques indicated in this table are to be multiplied by the ratio σE /9oo. Where no special measures are taken to avoid stripping of the threads ( σa = 0,7 . σE ) these values are to be divided by 1,14.


A- 3.2.2.3. - STRESSES IN WELDED JOINTS Determining the stresses in welds is a highly complex problem primarily because of the great number of possible configurations welded joints can assume. For this reason it is not possible, as the master stands at present, to lay clown precise directives in these Rules for the Design of Hoisting Appliances. Indeed, both the volume and the subject master of rules relating to welding would be difficult to fit into the general context of the present design rules. It was consequently decided to include only the following general indications : 1 - All methods of calculation assume of necessity a properly executed joint, i. e. a weld with correct penetration and a good shape, so that the joint between the components to be assembled and the weld seam is free from discontinuity or sudden change of section as well as from craters or notches due to undercutting. The design of the weld must be adapted to the forces to be transmitted, and specialised literature on the subject should be consulted. It should be noted that the strength of a welded joint is significantly improved if the surface of the weld is finished by careful grinding. 2 - There is no need to take into consideration stress concentrations due to the design of the joint or residual stresses. 3 - The permissible stresses in welds are those determined under clause 3.2.2.3. and the equivalent stress σcp in the case of combined stresses (tensile or compressive) σ and shear stress τ is given by the formula : σcp = ( σ2 + 2 . τ2 )0,5 In cases involving dual stresses σx and σy and the shearing stress τxy the following formula is applied : σcp = ( σx2 + σy2 - σx . σy + 2 . τxy2 )0,5 4 - In a fillet weld, the width of the section considered is the depth of the weld to the bottom of the throat and its length is the effective length of the weld less the end craters. The length need not be reduced if the joint closes on to itself or if special precautions are taken to limit the effect of the craters. Attention is drawn to the fact that it seems to be reliably established that fatigue failures in welded joints seldom occur in the weld seam itself but usually beside it in the parent metal.


The stresses σmin and σmax for the fatigue strength calculations for the parent metal beside the weld seam, can therefore in general be computed using the classical methods for calculating the strength of meterials. In order to verify the fatigue strength of the weld itself, it is generally held that it suffices to confirm that it is capable of transmitting the same loads as the adjacent parent metal. This rule is not obligatory however when the parts jointed are generously dimensioned in relation to the forces actually transmitted. When this is the case it suffices to dimension the weld seam in accordance with those forces, with the proviso that a fatigue check should then be performed in accordance with appendix A-3.6. Whatever the case it is emphasised that the size of a weld should invariably be in proportion to the thickness of the assembled parts. Special cases In certain cases of assembly by welding, particularly when there is a transverse load (i.e. perpendicular to the weld seam), the permissible stresses must be reduced (see clause 3.2.2.3.).


A - 3.3. AND A - 3.4. - CHECKING STRUCTURAL MEMBERS SUBJECT TO CRIPPLING AND BUCKLING The aim of these two appendices is not to adopt any specific stand on the problem but merely to give some general indications and enable reference to be made to existing works. A number of different methods are at present in use, among which the following are cited : 1 - in Germany, DIN 4114 2 - in Belgium, regulation NBN 1 3 - in France, the CM 1966 Rules 4 - in the United Kingdom, BS 2573.

A - 3.3. - CHECKING STRUCTURAL MEMBERS SUBJECT TO CRIPPLING While not wishing to adopt any particular standpoint on this problem, the FEM recommends the use of a practical method in the simpler cases, consisting in amplifying the calculated stress in the various loading cases defined in clauses 2.3.1., 2.3.2., and 2.3.3., by a crippling coefficient ω dependent upon the slenderness ratio of the member, and checking that, in each of these cases, the stress thus augmented remains less than the stresses given in table T.3.2.1.1. The values of ω are given in the tables below for the following cases, as a function of the slenderness ratio λ : Table T.A.3.3.1. : rolled sections in St 37 steel (Fe 360) Table T.A.3.3.2. : rolled sections in St 52 steel (Fe 510) Table T.A.3.3.3. : tubes in St 37 steel (Fe 360) Table T.A.3.3.4. : tubes in St 52 steel (Fe 510) Determination of effective lengths for calculating the slenderness ratio λ 1 - In the ordinary case of bars hinged at both ends and loaded axially, the effective length is taken as the length between points of articulation. 2 - For an axially loaded bar encastered at one end and free at the other the effective length is taken as twice the length of the bar.


3 - Because of the uncertainty which exist at present about the effect of fixity on bars in compression between two connections, the effects of fixity are not taken into consideration and the bar is designed as if it were hinged at both ends, the effective length therefore being taken as the length between points of intersection of axes. The case of bars subjected to compression and bending : In the case of bars loaded eccentrically or loaded axially with a moment causing bending in the bar : - either check the following two formulae : F / S + ( M f . v ) / I ≤ σa and ω . F / S + 0,9 . Mf . v / I ≤ σa where : F is the compressive load applied to the bar, S is the section area of the bar, Mf is the bending moment at the section considered, v is the distance of the extreme fibre from the neutral axis, I

is the moment of inertia ;

- or perform the precise calculation in terms of the deformations sustained by the bar under the combined effect of bending and compression, the necessary calculation being effected either by integration or by successive approximations.


Table T.A.3.3.1. - Value of the coefficient ω in terms of the slenderness ratio for rolled sections in St 37 steel (Fe 360) λ

0

1

2

3

4

5

6

7

8

9

20

1,04

1,04

1,04

1,05

1,05

1,06

1,06

1,07

1,07

1,08

30

1,08

1,09

1,09

1,10

1,10

1,11

1,11

1,12

1,13

1,13

40

1,14

1,14

1,15

1,16

1,16

1,17

1,18

1.19

1,19

1,20

50

1,21

1,22

1,23

1,23

1,24

1,25

1,26

1,27

1,28

1,29

60

1,30

1,31

1,32

1,33

1,34

1,35

1,36

1,37

1,39

1,40

70

1,41

1,42

1,44

1,45

1,46

1,48

1,49

1,50

1,52

1,53

80

1,55

1,56

1,58

1,59

1,61

1,62

1,64

1,66

1,68

1,69

90

1,71

1,73

1,74

1,76

1,78

1,80

1,82

1,84

1,86

1,88

100

1,90

1,92

1,94

1,96

1,98

2,00

2,02

2,05

2,07

2,09

110

2,11

2,14

2,16

2,18

2,21

2,23

2,27

2,31

2,35

2,39

120

2,43

2,47

2,51

2,55

2,60

2,64

2,68

2,72

2,77

2,81

130

2,85

2,90

2,94

2,99

3,03

3,08

3,12

3,17

3,22

3,26

140

3,31

3,36

3,41

3,45

3,50

3,55

3,60

3,65

3,70

3,75

150

3,80

3,85

3,90

3.95

4,00

4,06

4,11

4,16

4,22

4,27

160

4,32

4,38

4,43

4,49

4,54

4,60

4,65

4,71

4,77

4,82

170

4,88

4,94

5,00

5,05

5,11

5,17

5,23

5,29

5,35

5,41

180

5,47

5,53

5,59

5,66

5,72

5,78

5,84

5,91

5,97

6,03

190

6,10

6,16

6,23

6,29

6,36

6,42

6,49

6,55

6,62

6,69

200

6,75

6,82

6,89

6,96

7,03

7,10

7,17

7,24

7,31

7,38

210

7,45

7,52

7,59

7,66

7,73

7,81

7,88

7,95

8,03

8,10

220

8,17

8,25

8,32

8,40

8,47

8,55

8,63

8,70

8,78

8,86

230

8,93

9,01

9,09

9,17

9,25

9,33

9,41

9,49

9,57

9,65

240

9,73

9,81

9,89

9,97

10,05

10,14

10.22

10,30

10,39

10,47

250

10,55


Table T.A.3.3.2. - Value of the coefficient ω in terms of the slenderness ratio λ for rolled sections in St 52 steel (Fe 510) λ

0

1

2

3

4

5

6

7

8

9

20

1,06

1,06`

1,07

1,07

1,08

1,08

1,09

1,09

1,10

1,11

30

1,11

1,12

1,12

1,13

1,14

1,15

1,15

1,16

1,17

1,18

40

1,19

1,19

1,20

1,21

1,22

1.23

1,24

1,25

1,26

1,27

50

1,28

1,30

1,31

1,32

1,33

1,35

1,36

1,37

1,39

1,40

60

1,41

1,43

1,44

1,46

1,48

1,49

1,51

1,53

1,54

1,56

70

1,58

1,60

1,62

1,64

1,66

1,68

1,70

1,72

1.74

1,77

80

1.79

1,81

1,83

1,86

1,88

1,91

1.93

1,95

1,98

2,01

90

2,05

2,10

2,14

2.19

2,24

2,29

2,33

2,38

2,43

2,48

100

2,53

2,58

2,64

2,69

2,74

2,79

2,85

2,90

2,95

3,01

110

3,06

3,12

3,18

3,23

3,29

3,35

3,41

3,47

3.53

3.59

120

3,65

3,71

3,77

3,83

3,89

3,96

4,02

4,09

4,15

4,22

130

4,28

4,35

4,41

4,48

4,55

4,62

4,69

4,75

4,82

4,89

140

4,96

5,04

5,11

5,18

5,25

5,33

5,40

5,47

5.55

5,62

150

5,70

5,78

5,85

5,93

6,01

6,09

6,16

6,24

6,32

6,40

160

6,48

6,57

6,65

6,73

6,81

6,90

6,98

7,06

7,15

7,21

170

7,32

7,41

7,49

7,58

7,67

7,76

7,85

7,94

8,03

8,12

180

8,21

8,30

8,39

8,48

8,58

8,67

8,76

8,86

8,95

9,05

190 200

9,14 10,13

9,24 10,23

9,34 10,34

9,44 10,44

9,53 10,54

9,63 10,65

9,73 10,75

9,83 10,85

9.93 10,96

10,03 11,06

210

11,17

11,28

11,38

11,49

11,60

11,71

11,82

11,93

12,04

12,15

220

12,26

12,37

12,48

12,60

12,71

12,82

12,94

13,05

13,17

13,28

230

13,40

13,52

13,63

13,75

13,87

13,99

14,11

14,23

14,35

14,47

240

14,59

14,71

14,83

14,96

15,08

15,20

15,33

15,45

15,58

15,71

250

15,83


Table T.A.3.3.3. - Value of the coefficient ω in terms of the slenderness ratio λ for tubes in St 37 steel (Fe 360) λ

0

1

2

3

4

5

6

7

8

9

20

1,00

1,00

1,00

1,00

1,01

1,01

1,01

1,02

1,02

1,02

30

1,03

1,03

1,04

1,04

1,04

1,05

1,05

1,05

1,06

1,06

40

1,07

1,07

1,08

1,08

1,09

1,09

1,10

1,10

1,11

1,11

50

1,12

1,13

1,11

1,14

1,15

1,15

1,16

1,17

1,17

1,18

60

1,19

1,20

1,20

1,21

1,22

1,23

1,24

1,25

1,26

1,27

70

1,28

1,29

1,30

1,31

1,32

1,33

1,34

1,35

1,36

1,37

80

1,39

1,40

1,41

1,42

1,44

1,46

1,47

1,48

1,50

1,51

90

1,53

1,54

1,56

1,58

1,59

1,61

1,63

1,64

1,66

1,68

100

1,70

1,73

1,76

1,79

1,83

1,87

1,90

1,94

1,97

2,01

110

2,05

2,08

2,12

2,16

2,20

2,23

For λ > 115 take the value of ω in T.A.3.3.1. Table T.A.3.3.4. - Value of the coefficient ω in terms of the slenderness λ ratio for tubes in St 52 steel (Fe 510) λ

0

1

2

3

4

5

6

7

8

9

20

1,02

1,02

1,02

1,03

1,03

1,03

1,04

1,04

1,05

1,05

30

1,05

1,06

1,06

1,07

1,07

1,08

1,08

1,09

1,10

1,10

40

1,11

1,11

1,12

1,13

1,13

1,14

1,15

1,16

1,16

1,17

50

1,18

1,19

1,20

1,21

1,22

1,23

1,24

1,25

1,26

1,27

60

1,28

1,30

1,31

1,32

1,33

1,35

1,36

1,38

1,39

1,41

70

1,42

1,44

1,46

1,47

1,49

1,51

1,53

1,55

1,57

1,59

80

1,62

1,66

1,71

1,75

1,79

1,83

1,88

1,92

1,97

2,01

90

2,05

For λ > 90 take the value of ω in T.A.3.3.2. Note : The values of ω in table T.A.3.3.3. and T.A.3.3.4. are valid for calculating the case of an axially loaded bar consisting of a single tube whose diameter is equal to at least six times its thickness.


A - 3.4. - CHECKING STRUCTURAL MEMBERS SUBJECT TO BUCKLING From the theoretical standpoint, the critical buckling stress σvcr is regarded as a multiple of the EULER Stress given by the formula : σER = π2 . E . ( e / b )2 / [ 12 . ( 1 - η2 ) ] representing the critical buckling stress for a strip of thickness e, having a width equal to b, this being the plate dimension measured in the direction perpendicular to the compression forces (see sketch below).

In this formula, E is the modulus of elasticity and η Poisson's Ratio. For normal steels in which E = 210 000N/mm2 and η = 0,3, the EULER Stress becomes : σER = 189 800 . ( e / b )2 The critical buckling stress σvcr must be a multiple of this value, whence : σvcr = κσ . σER in the case of compression. In the case of shear the critical stress is : τvcr = κτ . σER The coefficients κσ and κτ, known as the buckling coefficients, depend on : - the ratio α = a / b of the two sides of the plate - the manner in which the plate is supported along the edges - the type of loading sustained by the plate in its own plane - any reinforcement of the plate by stiffeners. Value of coefficients κσ and κτ Without wishing to enter into the details of this problem, which is the subject of specialised works and of particular standards, we give hereafter values of κσ and κτ for a few simple cases (see table T.A.3.4.1.). For more complex cases, reference should be made to specialised literature.


Combined compression and shear Taking σ and τ to be calculated stresses in compression and in shear the critical comparison stress σvcr.c is determined from the expression : σvcr.c = (σ2 + 3 τ2 )0,5 /

{ { [ ( 1 + ψ) / 4 ] . ( σ / σ

v

cr

) + { [ 0,25 . ( 3 - ψ ) . σ / σvcr ]2 . [ τ / τvcr ]2 }0,5

}

ψ being defined in the table T.A.3.4.1. Important note : It is essential to note that the formulae above giving the critical stresses σvcr and σvcr.c apply only when the values determined thus are below the limit of proportionality (i.e. 190 N/mm2 for A.37 steel, 290 N/mm2 for A.52 steel). Similarly, the formula giving τvcr applies only when the value 30,5 . τvcr is below the limit of proportionality. Whenever the formulae give values above these limits, it is necessary to adopt a limiting critical value, obtained by multiplying the calculated critical value by the coefficient ρ given in the table T.A.3.4.2., which also indicates the reduced values corresponding to various calculated values of σvcr and τvcr .


Table T.A.3.4.1. Value of the buckling coefficients κσ and κτ for plates supported at their four edges No. 1

2

3

α=a/b α≥1

CASE Simple uniform compression

Non-uniform compression

Pure bending ψ = - 1 or bending with tension preponderant

4

Bending with compression preponderant - 1 < ψ< 0

5

Pure shear

κσ = 4

κσ or κτ

α≤1

κσ = (α + 1 / α)2

α≥1

κσ = 8,4 / ( ψ + 1,1 )

α≤1

κσ = 2,1 . (α + 1 / α)2 / ( ψ + 1,1 )

α≥2/3

κσ = 23,9

α≤2/3

κσ = 15,87 + 1,87 / α2 + 8,6 α2

κσ = ( 1 + ψ ) . κ‘ - ψ . κ‘’ + 10 . ψ . ( 1 + ψ ) where : κ‘ = value of κσ for ψ = 0 in case n° 2 κ‘’ = value of κσ for pure bending (case n°. 3) α≥1

κτ = 5,34 + 4 / α2

α≤1

κτ = 4 + 5,34 / α2

Table T.A.3.4.2. - Values of ρ and the reduced critical stresses σvcr , σvcr.c and τvcr (N/mm2) σvcr σvcr σvcr σvcr v v v or or or or τ cr τ cr τ cr τvcr v v v v σ cr.c σ cr.c σ cr.c σ cr.c ρ ρ calculated

190 200 210 220 230 240 250 260 280 300 340

calculated

reduced

Steel St 37 (Fe 360) 110 1,00 190 116 0,97 194 121 0,94 197 127 0,91 200 133 0,88 202 139 0,85 204 145 0,82 206 150 0,80 208 162 0,76 212 173 0,72 215 197 0,65 221

reduced

110 113 114 116 117 118 119 120 122 124 128

calculated

290 300 310 320 330 340 350 360 380 400 440

calculated

reduced

Steel St 52 (Fe 510) 168 1,00 290 173 0,98 294 179 0,96 297 185 0,94 300 191 0,92 303 196 0,90 306 202 0,88 308 208 0,86 309 220 0,82 312 231 0,79 316 254 0,73 322

reduced

168 169 172 174 175 176 177 178 180 182 185


Determination of permissible buckling stresses After the critical buckling stresses have been determined as indicated above, the permissible stress is obtained by dividing the critical stress by the coefficient νV determined in clause 3.4. The calculations are then performed as follows : The stresses are determined for each case of loading, in accordance with clause 3.4., after which a check is made to ensure that these calculated stresses do not exceed the permissible stresses determined as indicated above. Note : In the case of combined compression and shear, the critical comparison stress σvcr.c must be compared with the comparison stress calculated from the formula in clause 3.2.1.3. : σcp = ( σ2 + 3 . τ2 )0,5 Example of checking for buckling Take the case of a plate girder in St 37 steel, having a span of 10 m, a depth of 1,50 m, a web thickness of 0,010 m, a uniformly distributed load of 162 kN/m and stiffeners 1,25 m apart.

Reactions on supports : A = B = 810 kN Moment of inertia of the beam = 1 419 000 cm4 Checking at section MN, located 0,625 m from A Bending moment at MN : Mf = 810 x 0 625 - ( 162 x 0,6252 ) / 2 = 474,7 kNm Upper stress (compression) : σ1 = - ( 474,7 x 106 x 0,84 x 103 ) / ( 1 419 000 . 104 ) = - 28 N/mm2 Lower stress (tension) : σ2 = ( 474,7 x 106 x 0,66 x 103 ) / ( 1 419 000 . 104 ) = 22 N/mm2 These stresses are calculated at the upper and longer edges of the web.


shear stress : ( 810 . 103 - 162 . 0,625 . 103 ) / ( 10 . 1500 ) = 47 N/mm2 Bending (case 4 - compression preponderant) : ψ = 0,22 / -0,28 = -0,79

α = 1,25 / 1,50 = 0,83 (< 1 )

giving

κσ = ( 1 + ψ ) . κ‘ - ψ . κ‘’ + 10 . ψ . ( 1 + ψ )

in which

κ‘ = ( α+ 1 / α )2 . 2,1 / ( 0 + 1,1 ) = ( 0,83+ 1 / 0,83 )2 . 2,1 / 1,1 = 7,90

and

κ‘’ = 23,9

whence

κσ = ( 1- 0,79 ) . 7,90 + 0.79 . 23,9 - 10 . 0,79 . ( 1 - 0,79 ) = 18,88

The Euler Stress : σE R = 189 800 . ( e / b )2 = 189 800 ( 10 / 1500 )2 = 8,4 N/mm2 giving a critical buckling stress : σvcr = κσ . σER = 18,88 . 8,4 = 158,6 N/mm2 Shear :

κτ = 4 + 5,34 / σ2 = 4 + 5,34 / 0,832 = 11,75

and

τvcr = κτ . σE R = 11,75 . 8,4 = 99 N/mm2

The critical comparison stress then becomes : σvcr.c = ( 282 + 3 . 472 )0,5 ________________________________________________________________________________________________________________ 2 2 0,5

{ [ ( 1 - 0,79 ) / 4 ] . ( 28 / 158,5 ) + { [ 0,25 . ( 3 + 0,79 ) . 28 / 158,5 ] . [ 47 / 99 ] } σvcr.c = 168 N/mm2


Conclusion : The comparison stress in the case of tension (or compression) combined with shear is given in clause 3.2.1.3. ( σ2 + 3 τ2 )0,5 = 86 N/mm2. This value is smaller than the critical buckling stress given in 3.4. (with νV = 1,4) 168 / 1,4 = 120 N/mm2 for loading case I. The permissible buckling stress is therefore not exceeded in loading case I. Naturally, a check must also be made to ensure that the permissible buckling stresses are not exceeded in loading cases II and III. Checking of buckling for circular cylinders : Thin wall circular cylinders such as, for exempla, large tubes, which are subject to central or eccentric axial compression have to be checked for local buckling if : t / r ≤ 25 . σE / E where : t = thickness of the wall ; r = radius from the middle of the wall thickness ; σE = elastic limit of the steel type, as in table T. 3.2.1.1. E = modulus of elasticity, see A-3.4. The ideal buckling stress σvi can be determined from : σvi = 0,2 E . t / r In all cases where σvi is situated above the limit of proportionality of the structural steel, the ideal buckling stress σvi has to be reduced to σv by means of the factor ρ. At a maximum spacing of 10 . r, transverse stiffeners have to be provided whose moment of inertia has to be at least : I = 0,5 . r . t3 / ( r / t )0,5 The moment of inertia is calculated from the following formulae : 1 - Central disposition of the stiffener F (centre of gravity of the stiffener section in the median plane of the wall thickness).


2 - Eccentric disposition of the stiffener F (centre of gravity of the stiffener section F2 outside the median plane of the wall 1). .

I = I1 + I2 + F1 . e12 + F2 . e22 It is accepted that this calculation of σvi and σv respectively takes account of geometrical divergences between the real and the ideal cylinder surfaces due to local construction defects up to a dimension of t / 2.


A - 3.6. - CHECKING STRUCTURAL MEMBERS SUBJECT TO FATIGUE It must be remembered that fatigue is one of the causes of failure envisaged in clause 3.6. and therefore checking for fatigue is additional to checking in relation to the elastic limit of permissible crippling or buckling. If the permissible stresses for fatigue, as determined thereunder, are higher than those allowed for other conditions then this merely indicates that the dimensions of the components are not determined by considerations of fatigue. Clause 3.6. enumerates the parameters which must be considered when checking structural components for fatigue. The purpose of this appendix is firstly to classify the various joints according to their notch effect, as defined in clause 3.6.2. and, then, to determine for these various notch effects and for each classification group of the component as defined in clause 2.1.4. the permissible stresses for fatigue as a function of the coefficient κ defined in clause 3.6.4. These permissible fatigue stresses were determined as a result of tests carried out by the F.E.M. on test pieces having different notch effects and submitted to various loading spectre. They were determined on the basis of the stress values which, in the tests, assured 90 % survival including a factor of safety of 4/3. In practice, a structure consists of members which are welded, riveted or bolted together and experience shows that the behaviour of a member differs greatly from one point to another ; the immediate proximity of a joint invariably constitutes a weakness that will be vulnerable to a varying extent according to the method of assembly used. An examination is therefore made in the first sections, of the effect of fatigue on structural members both away from any joint and in immediate proximity to the usual types of joint. The second section examines the resistance to fatigue of the means of assembly themselves, i.e. weld seams, rivets and bolts.


1 - VERIFICATION OF STRUCTURAL MEMBERS The starting point is the fatigue strength of the continuous metal away from any joint and, in general, away from any point at which a stress concentration, and hence a lessening of the fatigue strength, may occur. In order to make allowance for the reduction in strength near joints, as a result of the presence of holes or welds producing changes of section, the notch effects in the vicinity of these joints, which characterize the effects of the stress concentrations caused by the presence of discontinuities in the metal, are examined. These notch effects bring about a reduction of the permissible stresses, the extent of which depends upon the type of discontinuity encountered, i.e. upon the method of assembly used. In order to classify the importance of these notch effects, the various forms of joint construction are divided into categories as follows : Unwelded parts These members present three cases of construction. Case WO concerns the material itself without notch effect. Cases W1 and W2 concern perforated members (see table T.A.3.6.(1)) Welded parts These joints are arranged in order of the severity of the notch effect increasing from K0 to K4, corresponding to structural parts located close to the weld fillets. The table T.A.3.6. (1) gives some indications as to the quality of the welding and a classification of the welding and of the various joints that are most often used in the construction of lifting appliances. Determination of the permissible stresses for fatigue Tensile and compressive loads The basis values which have been used to determine the permissible stresses in tension and compression are those resulting from the application of a constant alternating stress ± σW (κ = - 1 ) giving a survival rate of 90 % in the tests, to which a factor of safety of 4/3 has been applied. To take account of the number of cycles and of the stress spectrum, the σW values have been set for each classification group of the member the latter taking account of these two parameters. For unwelded parts, the values σW are identical for steel St 37, and St 44. They are higher for St 52. For welded parts, the σW values are identical for the three types of steel.


Table T.A.3.6.1. Values of σ W depending on the component group and construction case (N/mm2)

Component group

E1

Welded components Unwelded components Construction cases Construction cases (Steels St 37 to St 52, Fe 360 to Fe 510) W0 W1 W2 Fe 360 Fe 360 Fe 360 K1 K2 K3 K4 St 37 St 52 St 37 St 52 St 37 St 52 K0 St 44 Fe 510 St 44 Fe 510 St 44 Fe 510 249,1 298,0 211,7 253,3 174,4 208,6 (361,9) (323,1) (271,4) 193,9 116,3

E2

224,4

261,7

190,7

222,4

157,1

183,2

(293,8)

262,3

220,3

157,4

94,4

E3

202,2

229,8

171,8

195,3

141,5

160,8

238,4

212,9

178,8

127,7

76,6

E4

182,1

201,8

154,8

171,5

127,5

141,2

193,5

172,8

145,1

103,7

62,2

E5

164,1

177,2

139,5

150,6

114,9

124,0

157,1

140,3

117,8

84,2

50,5

E6

147,8

155,6

125,7

132,3

103,5

108,9

127,5

113,8

95,6

68,3

41,0

E7

133,2

136,6

113,2

116,2

93,2

95,7

103,5

92,4

77,6

55,4

33,3

E8

120,0

120,0

102,0

102,0

84,0

84,0

84,0

75,0

63,0

45,0

27,0

The values in brackets are greater than 0,75 times the breaking stress and are only theoretical values (see note 2 at the end of this clause). The following formulae give for all values of κ the permissible stresses for fatigue a) κ ≤ 0 - for tension : - for compression :

σt = 5 . σw / ( 3 - 2 . κ )

(1)

σc = 2 . σw / ( 1 - κ )

(2)

σw is given in table above. b)

κ>o - for tension - for compression

σt = σO / [ 1 - κ . (1 - σO / σ+1 ) ]

(3)

σc = 1,2 . σt

(4)


where σO = tensile stress for κ = 0 is given by the formula (1) that is : σO = 1,66 . σw σ+1 = tensile stress for κ = + 1 that is the ultimate strength σR divided by the coefficient of safety 4/3 : σ+1 = 0,75 . σR σt is limited in every case to 0,75 . σR . By way of illustration, fig. A.3.6.1. shows curves giving the permissible stress as a function of the ratio κ for the following cases : - steel A.52 ; - predominant tensile stress ; - group E6 ; - construction cases W0, W1, W2 for unwelded components and cases of construction for joints K0 to K4. The permissible stresses have been limited to 240 N/mm2, i.e. to the permissible stress adopted for checking for ultimate strength. [permissible stress)

Ratio between the extreme stresses Figure A.3.6.1. - (A 52; tension; group E6)


Shear stresses in the material of structural parts For each of the group from E1 to E8 the permissible fatigue stress in tension of the case wO divided by 30,5 is taken : τa = σt of case wO / 30,5 Combined loads in tension (or compression) and shear In this case the permissible stresses for fatigue for each normal load in tension (or compression) σxa and σya and shear τxya are determined by assuming that each acts separately taking respectively the following values of κ in accordance with clause 3.6.4. : κx = σx min / σx max

κy = σy min / σy max

κxy = τxy min / τxy max

Then the following three conditions are checked : σx max < σxa

σy max < σya

τxy max < τxya

None of the calculated sheares should exceed the permissible value of σa in case I loading (see table T.3.2.1.1.). a) If any one stress is markedly greater than the other two in any given case of loading, it will suffice to check the member for fatigue under the corresponding load, neglecting the effect of the other two. b) In the other cases, in addition to checking for each loading assumed to act alone, it is recommended that the following relationship be checked : 2 (σx max / σxa )2 + (σy max / σya )2 - σx max . σy max / ( | σxa | . | σya | ) + ( τxy max / τxya )2 ≤ 1 (5) where the stress values σxa, σya and τxya are those resulting from the application of formulae (1), (2), (3) and (4) limited to 0,75 . σR.

2

As this inequality constitutes a severe requirement, values slightly higher than 1 are acceptable, but in this case it is necessary to check the relation :

[ (σx max / σxa )2 + (σy max / σya )2 - σx max . σy max / ( | σxa | . | σya | ) + ( τxy max / τxya )2 ]0,5 ≤ 1,05

It should also be noted that the values | σxa | and | σya | in the denominator for the third term should be taken as absolute values, σx max and σy max being assigned their algebraic values.


In applying this formula, reference should be made to the directions given in clause 3.2.1.3. In other words : - either perform the check by combining the maximum values σx max , σy max and τxy max and comparing with the permissible stresses σxa, σya and τxya computed on the basis of the most unfavourable values of κ. - or seek the most unfavourable combination actually possible by making the check with the following values : a) σx max and κx min with the corresponding values of σy, τxy, κy and κxy b) σy max and κy min with the corresponding values of σx, τxy, κx and κxy c) τxy max and κxy min with the corresponding values of σx, σy, κx and κy In this connection, see note in clause 3.2.1.3. In order to facilitate the calculations, table T.A.3.6.2. gives the permissible values of : τxy max / τxya as a function of σx max / σxa and of σy max / σya In this table, the values of σx max / σxa are given in the left hand column with the following convention : the ratio is considered to be positive if σx max and σy max have the same sign, and negative otherwise.


Table T.A.3.6.2. - Values of τxy max / τxya in terms of σx max / σxa and σy max / σya σx max ________

σy max / σya

σxa

+ 1,0

0

0,9 0,300

+ 0,9

0,300

0,436

0,520

0,575

0,608

0,625

0,625

0,608

0,575

0,520

0,436

+ 0,8

0,400

0,520

0,600

0,656

0,693

0,714

0,721

0,714

0,693

0,656

0,600

+ 0,7

0,458

0,575

0,656

0,714

0,755

0,781

0,794

0,781

0,781

0,755

0,714

+ 0,6

0,490

0,608

0,693

0,755

0,800

0,831

0,849

0,854

0,849

0,831

0,800

+ O,5

0,500

0,625

0,714

0,781

0,831

0,866

0,889

0,900

0,900

0,889

0,866

+ 0,4

0,490

0,625

0,721

0,794

0,849

0,889

0,917

0,933

0,938

0,933

0,917

+ 0,3

0,458

C,608

0,714

0,794

0,854

0,900

0,933

0,954

0,964

0,964

0,954

+ 0,2

0,400

0,575

0,693

0,781

0,849

0,900

0,938

0,964

0,980

0,985

0,980

+ 0,1

0,300

0,520

0,656

0,755

0,831

0,889

0,933

0,964

0,985

0,995

0,995

0

0,436

0,600

0,714

0,800

0,866

0,916

0,9S4

0,980

0,99S

1,000

0,300

0,520

0,656

0,755

0,831

0,889

0,933

0,964

0,98S

0,995

- 0,2

0,400

0,575

0,693

0,781

0,849

0,900

0,938

0,964

0,980

- 0,3-

0,173

0,458

0,608

0,714

0,794

0,854

0,900

0,933

0,954

0,265

0,490

0,625

0,721

0,781

0,849

0,889

0,917

0,300

0,500

0,625

0,714

0,781

0,831

0,866

0,300

0,490

0,608

0,693

0,755

0,800

0,265

0,458

0,575

0,656

0,714

0,173

0,400

0,520

0,600

0,300

0,436

0 - 0,1

- 0,4 - 0,5 - 0,6 - 0,7 - 0,8 - 0,9

1,0

0,8 0,400

0,7 0,458

0,6 0,490

0,5 0,500

0,4 0,490

0,3 0,458

0,2 0,400

0,1 0,300

0

- 1,0

If σx max and σy max are of opposite sign (tension or compression) read the values of τxy max / τxya starting from the negative values of σx max / σxa

0

0


General notes Note 1 - In applying the above considerations, it is essential to take into account the secondary bending effects which a particular method of assembly may cause in the members of the structure. Note 2 - If reference is made to the table of values of σw it can be seen that in group E1 and E2 much higher stresses than those usually permitted in structures are quoted. These values are in fact only theoretical values obtained by extrapolation of the test results on higher group (E3 to E8) with medium or severe notch cases (K2, K3 and K4). Therefore there is no need to attach any material significance to these values in brackets, consideration of which could in some cases lead to the conclusion that an assembly of type K0 or K1 could resist fatigue better than the unwelded metal (case WO). This apparent anomaly illustrates the well known fact that it is not always necessary to carry out fatigue checks for the longer group with slight or moderate notch cases. With respect to the calculations it must be remembered that these theoretical σw values are used only to determine the permissible fatigue stresses σxa , σya and τxya for use in formula (5) which covers the case of combined loads. Examples of calculations are given at the end of the Appendix.


2 - VERIFICATION OF THE JOINING MEANS (welds, bolts, rivets) Welds a) Tensile and compressive loads in the welds : Welds subjected to fatigue under tensile and compressive loads are checked using the same permissible stresses as those of the metal joined. Note - The limit indicated under 3.2.2.3. for certain particular cases of transverse tension and compression in weld seams must be observed. Appendix A-3.2.2.3. gives, in addition, some indications for the determination of the stresses in the weld seams. b) Shear loads in the welds : The permissible shear fatigue stresses in the welds are determined by dividing the permissible stresses in tension for case K0 by 20,5 c) Combined loads : The method set out above for structural members is used when considering the effect of fatigue in weld seams subjected to variable combined loads. Bolts and rivets a) Tensile loads : Fatigue due to variable tensile loads in bolts and rivets need not be considered. In this connection, it should be noted that bolts and, even more important, rivets working in tension should be avoided as far as possible. b) Shear loads and bearing pressure : Single and multiple shear loads as defined under 3.2.2.1.1. must be distinguished. The permissible shear stresses for fatigue for bolts and rivets are fixed by multiplying the permissible stresses in tension for case W2 by : 0,6 for single shear

0,8 for multiple shear

The permissible bearing pressure values are obtained by multiplying the permissible shear values in the bolts and rivets by 2,5.


Table T.A.3.6.(1) - Classification of cases of construction for joints Joints may be riveted, bolted or welded. The types of weld most commonly used for hoisting appliances are butt welds, double bevel butt welds (K welds) and fillet welds, of ordinary quality (O.Q.) or special quality (S.Q.) as specified below. Weld testing is also stipulated for certain types of joint.


A - Weld qualities

Type of weld

Weld quality

Special quality (S.Q.) Full depth butt weld

3

Execution of weld Root of weld scraped (or trimmed) before making sealing run. No end craters. Weld ground flush with plate parallel to direction of forces

Ordinary Root of weld scraped quality (or trimmed) before (O.Q.) making sealing run. No end craters

K-weld in angle formed by two parts with bevel on one of the parts to be joined at location of seam

Fillet welds in the angle formed by two parts

Root of weld scraped (or trimmed) before making weld on other Special side. Weld edges quality without undercutting (S.Q.) and ground if necessary. Full penetration welds Width clear of weld penetration between Ordinary the two welds < 3 mm quality (O.Q.)

Special quality (S.Q.)

Welded edges without undercutting and ground if necessary

Symbol

Weld testing Check (e.g. with Xrays) over 100 % of seam length

If the calculated stress > 80 % times the permissible stress Otherwise random check over at least 10 % of seam length

Check that for tensile loads the plate perpendicular to the direction of the forces is free from lamination.

Check that for tensile loads the plate perpendicular to the direction of the forces is free from lamination

Symbol

P 100

P 100

P 10

D

D

Ordinary quality (O.Q.) Table T.A.3.6.(1) (continued) B - Cases of construction for joints

In the tables below the various cases of means of assembly are classified in terms of the magnitude of the notch effect they produce.

3

It is forecasted that the symbol shall be adapted to the ISO standard 2553 at the next edition of the Design Rules, when the addition of this standard will be definitively adopted.


It should be noted that, with a given weld, the notch effect differs according to the type of loading to which the joint is-subjected. For exempla, a fillet welded joint is classified under case K0 for longitudinal tension or compression loads (0,31) or longitudinal shear (0,51), and under cases K3 or K4 for transverse tension or compression loads (3,2 or 4,4). 1 - Non welded parts

Reference WO

Case WO Description Parent metal, homogeneous surface. Part without joints or breaks in continuity (solid bars) and without notch effects unless the latter can be calculated.

Figure

Symbol

Figure

Symbol

Case W1 Reference

W1

Reference W2.1

W2.2

W2.3

Description Parts drilled. Parts drilled for riveting or bolting with rivets and bolts loaded up to 20 % of permissible values. Parts drilled for joints using high strength bolts (C1 3.2.2.2.2.3.) loaded up to 100 % of permissible values(C1 3.2.2.2.2.2.)

Case W2 Description Parts drilled for riveting or bolting in which the rivets or bolts are loaded in multiple shear Parts drilled for riveting or bolting, in which the rivets or bolts are loaded in single shear (allowing for eccentric loads), the parts being unsupported Parts drilled for assembling by means of rivets or bolts loaded in single shear, the parts being supported or guided

Figure

Symbol


2 - welded parts Case K0 - Slight stress concentration Reference 0,1

0,11

0,12

0,13

0,3

0,31

0,32

0,33

Description Parts butt-welded (S.Q.) at right angles to direction of forces Parts of different thickness buttwelded (S.Q.) at right angles to direction of forces. Asymmetrical slope : 1/4 to 1/5; Symmetrical slope : 1/3 Butt weld (S.Q.) in transverse joint of web plate

Gusset secured by butt-welding (S.Q.) at right angles to the direction of the forces Parts joined by butt-welding (O.Q.) parallel to the direction of the forces

Figure

Symbol

P 100

P 100

P 100

P 100

P 100 or P10

Parts joined by fillet welds (O.Q.) parallel to the direction of the forces (longitudinal to the joined parts) Butt weld (O.Q.) between section forming flange and web of a beam

K- or fillet weld (O.Q.) between flange and web of a beam calculated for the equivalent stress for combined forces (C1 3.2.1.3.)

P 100 or P 10


Case K0 - Slight stress concentration (continued) Reference

0,5

0,51

Description

Figure

Butt weld (O.Q.) in the case of longitudinal shear

Symbol P 100 or P10

K-weld (O.Q.) or fillet weld (O.Q.) in the case of longitudinal shear

Case K1 - Moderate stress concentration Reference

Description

Figure

Symbol

1,1

Parts joined by butt welding (O.Q.) at right angles to the direction of the forces

P 100 or P10

1,11

Parts of different thickness butt welded (O.Q.) at right angles to the direction of the forces. Asymmetrical slope : 1 in 4 to 1 in 5 (or symmetrical slopes : 1 in 3)

P 100 or P10

1,12

1,13

1,2

Butt weld (O.Q.) executed for transverse joint of web plate

Gusset joined by butt welding (O.Q.) at right angles to the direction of the forces Continuous main member to which are joined by continuous K-welds (S.Q.) parts at right angles to the direction of forces

P 100 or P10

P 100 or P10


Case K1 - Moderate stress concentration (continued) Reference

1,21

1,3

1,31

Description Web plate to which stiffeners are joined at right angles to the direction of the forces by means of fillet welds (S.Q.) which extend round the corners of the web stiffeners Parts joined by butt welding parallel to the direction of the forces (without checking the welding)

Figure

Symbol

K-weld (S.Q.) between curved flange and web

Case K2 - Medium stress concentration Reference

2,1

Description Parts of different thickness butt welded (O.Q.) at right angles to the direction of the forces. Asymmetrical slope : 1 in 3 (or symmetrical slopes : 1 in 2)

Figure

Symbol

2,11

Sections joined by butt welds (S.Q.) at right angles to the direction of the forces

2,12

Section joined to a gusset by a butt weld (S.Q.) at right angles to the direction of the forces

P100

2,13

Butt weld (S.Q.) at right angles to the direction of the forces, made at intersection of flats, with welded auxiliary gussets. The ends of the welds are ground, avoiding notches

P100

P 100 or P10


Case K2 - Medium stress concentration (continued) Reference

2,2

2,21

2,22

2,3

2,31

2,33

2,34

Description Continuous main member to which transverse diaphragm, web stiffeners, rings or hubs are fillet welded (S.Q.) at right angles to the direction of the forces Web in which fillet welds (S.Q.) are used to secure transverse web stiffeners with cut corners, the welds not extending round the corners Transverse diaphragm secured by fillet welds (S.Q.) with cut corners, in which the welds do not extend round the corners Continuous main member to the edges of which are butt welded (S.Q.) parts parallel to the direction of the forces. These parts terminal in bevels or radii. The ends of the welds are ground avoiding notches Continuous main member to which are welded parts parallel to the direction of the forces. These parts terminal in bevels or radii. Valid where the ends of the welds are K-welds (S.Q.) over a length equal to ten times the thickness provided that the ends of the welds are ground avoiding notches Continuous member to which a flat (1 in 3 bevel) is joined by a fillet weld (S.Q.), the fillet weld being executed in the X area, with a = 0,5 e

K-weld (O.Q.) made between curved flange and web

Figure

Symbol

P100


Case K2 - Medium stress concentration (continued) Reference

2,4

2,41

2,5

Description

Figure

Cruciform joint made with K-welds (S.Q.) perpendicular to the direction of the forces

Symbol D

K-weld (S.Q.) between flange and web in the case of load concentrated in the plane of the web at right angles to the weld K-weld (S.Q.) joining parts stressed in bending or shear

Case K3 - Severe stress concentration Reference

3,1

3,11

3,12

3,13

Description Parts of different thickness connected by butt welds (O.Q.) at right angles to the direction of the forces. 1 in 2 asymmetrical slope, or symmetrical position without blend slope Butt weld with backing strip and no backing run. Backing strip secured by intermittent tack welds Tubes joined by butt welds whose root is supported by a backing piece and not covered by a backing run Butt weld (O.Q.) at right angles to the direction of the forces at the intersection of flats with welded auxiliary gussets. The ends of the welds are ground, avoiding notches

Figure

Symbol P 100 or P10


Case K3 - Severe stress concentration (continued) Reference

3,2

3,21

3,3

3,31

3,32

3,33

Description Continuous main member to which parts are fillet welded (O.Q.) at right angles to the direction of the forces. These parts take only a small portion of the loads transmitted by the main member Web and stiffener or transverse diaphragm secured by uninterrupted fillet weld (O.Q.) Continuous member to the edges of which are butt welded (O.Q.) parts parallel to the direction of the forces. These parts terminal in bevels and ends of the welds are ground avoiding notches Continuous member to which are welded parts parallel to the direction of the forces. These parts terminal in bevels or radii. Valid where the ends of the welds are fillet welds (S.Q.) over a length equal to 10 times the thickness, provided that the ends of the welds are ground, avoiding notches Continuous member through which extends a plate, terminating in bevels or radii parallel to the direction of the forces, secured by K-weld (O.Q.) over a length equal to 10 times the thickness Continuous member to which is welded a flat parallel to the direction of the forces, by means of fillet weld (S.Q.) in the indicated area when e1 < 1,5 . e2

Figure

Symbol


Case K3 - Severe stress concentration (continued) Reference

3,34

3,35

3,36

3,4

3,41

Description Members at the extremity of which connecting gussets are secured by a fillet weld (S.Q.) when e1 ≤ e2 In case of unilateral gusset allow for eccentric load Continuous member to which stiffeners parallel to the direction of the forces are welded. The ends of the welds are fillet welds (S.Q.) over a length equal to ten times the thickness and are ground avoiding notches Continuous member to which stiffeners parallel to the direction of the forces are secured by fillet welds (O.Q.) which are intermittent or made between indentations Cruciform joint made with K-weld (O.Q.) at right angles to the direction of the forces

Figure

Symbol

D

K-weld (O.Q.) between flange and web in case of concentrated load in the plane of the web at right angles to the weld D

3,5

K-weld (O.Q.) joining parts stressed in bending and shear

3,7

Continuous member to which sections or tubes are fillet welded (S.Q.)


Case K4 - Very severe stress concentration Reference 4,1

4,11

4,12

4,3

4,31

4,32

4,33

Description Parts of different thickness butt welded (O.Q.) at right angles to the direction of the forces. Asymmetrical position without blend slope Butt welds (O.Q.) at right angles to the direction of the forces, at the intersection of flats (no auxiliary gussets) Single bevel weld at right angles to the direction of the forces, between intersecting parts (cruciform joint) Continuous member to the sides of which are welded parts ending at right angles, parallel to the direction of the forces Continuous member to which parts, ending at right angles, parallel to the direction of the forces, and receiving a large proportion of the loads transmitted by the main member, are secured by fillet weld (O.Q.) Continuous member through which extends a plate ending at right angles and secured by fillet welding (O.Q.) Continuous member on which a flat is secured by means of a fillet weld (O.Q.) parallel to the direction of the forces

Figure

Symbol

D


Case K4 - Very severe stress concentration (continued) Reference 4,34

Description Joint plate secured by (O.Q.) fillet welds (e1 = e2). In case of unilateral joint plate allow for eccentric loads

4,35

Parts welded one on the other secured by fillet welds (O.Q.) in a slot or in holes

4,36

Continuous members between which connecting gussets are secured by fillet welds (O.Q.) or butt welds (O.Q.)

Figure

Symbol

D 4,4

4,41

4,5

4,7

Cruciform joint made with fillet weld (O.Q.) at right angles to the direction of the forces Fillet weld (O.Q.) between flange and web in the case of concentrated load in the plane of the web at right angles to the weld Fillet welds (O.Q.) joining parts stressed in bending and shear Continuous member to which sections or tubes are connected by fillet welds (O.Q.)

D


EXAMPLES OF CALCULATING CHECKS EXAMPLES OF FATIGUE CHECKS FOR A WELDED WEB TO FLANGE JOINT STEEL St 37 TOP FLANGE OF GIRDER OF AN OVERHEAD TRAVELLING CRANE ON WHICH A CRAB RUNS (Combined check for fatigue and elastic limit) The results of stress calculations in the top fringe of the girder are as follows :

Longitudinal compression : σx max = - 140 N/mm2 σx min = - 28 N/mm2 from which κ = 0,2 Lateral compression when the crab wheel passes : σy max = - 100 N/mm2 σy min = 0 from which κ = 0 Shear : changing sign when passing from one side to the other of the section: τxy max = ± 40 N/mm2 from which κ = -1 Equivalent stress : [ ( -140 )2 + ( -100 )2 - 140 . 100 + 3 . 402 ]0,5 = 144 < 160 N/mm2 ( σa ) acceptable (See clause 3.2.1.3.).


CHECKING FOR FATIGUE AND ELASTIC LIMIT FIRST EXAMPLE - COMPONENT IN GROUP E4 WITH FILLET WELD (O Q ) 1 - CHECKING MATERIAL ADJACENT TO THE WELDING

a) Longitudinal compression : case K0 (reference 0,31) Checking for elastic limit : σa σx max

= 160 = - 140

N/mm2 N/mm2

(table T.3.2.1.1.)

σw

= 193,5

N/mm2

(table T.A.3.6.1.)

σa = 5 / 3 . σ w

= 322,5

N/mm2

σ+1 = 0,75 . σR

= 270

N/mm2

σt is limited to

270

N/mm2

σc = - 1,2 . σt

= - 324

N/mm2

σxa

= - 324

N/mm2

from which | σx max l < σa Checking for fatigue :

l σx max l < l σxa l b) Lateral compression : case K4 (reference 4,41) Checking for elastic limit : σa

= 160

N/mm2

σy max

= - 100

N/mm2

| σy max | < σa

(table T. 3.2.1.1.)


Checking for fatigue : σw

=

62,2

N/mm2

σa = 5 / 3 . σ w

= 103,7

N/mm2

σt = σ0

= 107,7

N/mm2

σc = 1,2 . σt

= 124,4

N/mm2

σya

= - 124,4

N/mm2

l σy max l

<

l σya |

τxya = 160 / 30,5

=

92,4

τxy max

= + 40

τxy max

< τa

(table T.A.3.6.1.)

(formula (3))

c) Shear in the material Checking for elastic limit : N/mm2

(table T. 3.2.1.1.)

N/mm2

(formula (1))

(table T.A.3.6.1.)

Checking for fatigue : τw = 182,1 / 3 0,5

= 105,1

N/mm2

τa = τw

= 105,1

N/mm2

τxya

= 105,1

N/mm2

τXY max

=

40

N/mm2

| τXY max l

<

τa

d) Checking for combined loads : Use formula (5) : Condition to be checked : ( -140 / -324 )2 + ( -100 / -124,4 )2 - ( -140 ) . ( -100 ) / ( 324 . 124,4 ) + ( 40 / 92,4 )2 = 0,672 < 1 therefore satisfied.


2 - CHECKING IN THE WELD

If the thickness of the two welds is equal to the thickness of the web, the stresses σx max, σy max and τxy max have the same values as in 1 - above. The permissible tensile and compressive stresses are the same as for 1 above (in the material), with respect to both checking for elastic limit and checking for fatigue. It follows that we can dispense with a check for the cases corresponding to a) and b) above. The permissible shear stresses, as regards checking for elastic limit, are obtained by dividing the permissible tensile stress by 20,5 , instead of 30,5 in the case of the material itself. They are therefore more favourable than those used in cases c) and d) above. To sum up, we may confine ourselves to checking for fatigue the cases corresponding to c) and d) above. c) Shear in the weld : τxya = 193,5 / 20,5 τxy max

= 136,8 = 40

N/mm2 N/mm2

(table T.A.3.6.1.)

from which | τxy max l < | τxya l d) Checking for combined loads : Using formula (5) Condition to be checked : (-140 / -324 )2 + ( -100 / -124,4 )2 - ( -140 ) . (-100 ) / ( 324 . 124,4 ) + ( 40 / 136,8 )2 = 0,571 < 1 therefore satisfied. Note : If the component had been classified in group E6, the stress σy max = - 100 N/mm2 would be too high, since the permissible fatigue stress for case K4 and κ = 0 is only : σya = 1,2 . 5 / 3 . 41 = 82 N/mm2


SECOND EXAMPLE - COMPONENT IN GROUP E6 - K WELD (S.Q.) The loads - and therefore the stresses - will be assumed to be the same as in the first example. As the permissible stresses for the elastic limit checks are not affected by the change of group, nor by the type of weld, the calculations in the first example may, in this respect, be reproduced as they stand. We shall therefore confine ourselves to checking for fatigue. 1 - CHECKING MATERIAL ADJACENT OT THE WELD

a) Longitudinal compression, case K0 (reference 0,33) σw

= 127,5

N/mm2

σO = 5 / 3 . σw

= 212,5

N/mm2

σ+1 = 0,75 . σR

= 270

N/mm2

σt = 212,5 / [ 1 - ( 1 - 212,5 / 270 ) . 0,2] = 222,0

N/mm2

σc = - 1,2 . σt

=

266

N/mm2

σxa

= - 266

N/mm2

σx max

= - 140

N/mm2

(table T.A.3.6.1.)

(formula (3)) (formula (4))

l σx max l < l σxa l

from wihch

b) Lateral compression : case K2 (reference 2,41)

from wihch

σw

= 95,6

N/mm2

σa = 5 / 3 . σ w

= 159,3

N/mm2

σt = σO

= 159,3

N/mm2

σc = - 1,2 . σt

= 191,2

N/mm2

σya

= - 191,2

N/mm2

σy max

= - 100

N/mm2

l σy max l < l σya l

(table T.A.3.6.1.)

(formula (3))


c) Shear in the material : τxya = 147,8 / 30,5

=

τxy max

= ± 40

85,3

N/mm2

(table T.A.3.6.1.)

N/mm2

| τxy max l < | τxya l

from which d) Checking for combined loads : Use formula (5) Condition to be checked :

(-140 / -266 )2 + ( -100 / -191,2 )2 - ( -140 ) . (-100 ) / ( 266 . 191,2 ) + ( 40 / 85,3 )2 = 0,495 < 1 therefore satisfied. 2 - CHECKING IN THE WELD

Same reasoning as for first example. Leaving cases c) and d) to be checked for fatigue. c) Shear in the weld : τxya = 127,5 / 20,5

=

τxy max

= ± 40

from which

90,2

N/mm2

(table T.A.3.6.1.)

N/mm2

| τxy max l < | τxya l

d) Checking for combined loads : Use formula (5) Condition to be checked : (-140 / -266 )2 + ( -100 / -191,2 )2 - ( -140 ) . (-100 ) / ( 266 . 191,2 ) + ( 40 / 90,2 )2 = 0,472 < 1 therefore satisfied.



F.E.M. 1.001

3rd

EDITION REVISED

1998.10.01


CHECKING FOR FATIGUE AND CHOICE OF MECHANISM COMPONENTS



Booklet 4

CHECKING FOR FATIGUE AND CHOICE OF MECHANISM COMPONENTS

4.1. CALCULATION PROCEDURE.....................................................................................................3 4.1.1. CHECKING FOR ULTIMATE STRENGTH .............................................................................3 4.1.1.1. VALUE OF THE PERMISSIBLE STRESS..................................................................................... 3 4.1.1.2. VALUES OF THE COEFFICIENT νR ............................................................................................ 4 4.1.1.3. RELATIONS BETWEEN THE CALCULATED STRESSES AND THE PERMISSIBLE STRESSES.... 4

4.1.2. CHECKING FOR CRIPPLING...................................................................................................5 4.1.3. CHECKING FOR FATIGUE........................................................................................................5 4.1.3.1. GENERAL METHOD................................................................................................................... 5 4.1.3.2. ENDURANCE LIMIT UNDER ALTERNATING LOADING OF A POLISHED SPECIMEN................... 6 4.1.3.3. INFLUENCE OF THE SHAPE, SIZE, SURFACE CONDITION AND CORROSION........................... 7 4.1.3.5. WÖHLER CURVE....................................................................................................................... 9 4.1.3.6. FATIGUE STRENGTH OF A MECHANICAL COMPONENT ........................................................ 11 4.1.3.7. PERMISSIBLE STRESSES AND CALCULATIONS .................................................................... 11

4.1.4. CHECKING FOR WEAR...........................................................................................................13 4.2. DESIGN CALCULATIONS FOR PARTICULAR COMPONENTS ..................................................14 4.2.1. CHOICE OF ANTI-FRICTION BEARINGS.............................................................................14 4.2.1.1. THEORETICAL LIFE ................................................................................................................. 14 4.2.1.2. MEAN LOADING OF BEARINGS SUBJECTED TO TYPE SM LOADS........................................ 14 4.2.1.2.1. Determination of the mean load SM mean on anti-friction bearings for combined motions ....... 14

4.2.1.3. MEAN LOADING OF BEARINGS SUBJECTED TO TYPE SR LOADS ........................................ 15 4.2.1.4. MEAN LOADING OF BEARINGS SIMULTANEOUSLY SUBJECTED TO TYPE SM AND TYPE SR LOADS............................................................................................................................................ 15

4.2.2. CHOICE OF ROPES.................................................................................................................15 4.2.2.1. CHOICE OF ROPE DIAMETER .................................................................................................. 16 4.2.2.1.1. Common bases for the two methods ................................................................................. 16 4.2.2.1.2. Method using the minimum practical factor of safety Zp ...................................................... 17 4.2.2.1.3. C-factor method ................................................................................................................ 18


4.2.3. CHOICE OF PULLEYS, DRUMS AND ROPE ATTACHMENT MEANS.............................19 4.2.3.1. MINIMUM WINDING DIAMETER................................................................................................. 19 4.2.3.1.1. Values of H ....................................................................................................................... 20 4.2.3.1.2. Note.................................................................................................................................. 20

4.2.3.2. RADIUS OF THE BOTTOM OF THE GROOVE.......................................................................... 20 4.2.3.3. ROPE ATTACHMENT MEANS.................................................................................................. 21

4.2.4. CHOICE OF RAIL WHEELS ....................................................................................................21 4.2.4.1. RAIL WHEEL SIZE ................................................................................................................... 21 4.2.4.1.1. 4.2.4.1.2. 4.2.4.1.3. 4.2.4.1.4. 4.2.4.1.5.

Determining the mean load ............................................................................................... 22 Determining the useful rail width b ..................................................................................... 22 Determining the limiting pressure PL................................................................................... 23 Determining the coefficient c1 ............................................................................................ 24 Determining the coefficient c2 ............................................................................................ 25

4.2.4.2. NOTES..................................................................................................................................... 25

4.2.5. DESIGN OF GEARS..................................................................................................................26 APPENDIX..................................................................................................................................................27 A- 4.1.3. - DETERMINATION OF PERMISSIBLE STRESSES IN MECHANISM COMPONENTS SUBJECTED TO FATIGUE....................................................................................27 EXAMPLE OF APPLICATION................................................................................................................. 30 LIST OF SOME WORKS DEALING WITH FATIGUE PROBLEMS............................................................ 32

A- 4.2.2. - COMMENTS ON THE CHOICE OF ROPES AND ON THE PROBLEM OF THE FACTOR OF SAFETY...........................................................................................................................33 A- 4.2.3. - CONSIDERATIONS ON THE DETERMINATION OF MINIMUM WINDING DIAMETERS FOR ROPES..................................................................................................................37

4-2


4.1. CALCULATION PROCEDURE Mechanism components are designed by checking that they offer adequate safety against failure due to fracture, crippling, fatigue or excessive wear. Other factors must also be taken into consideration and it is particularly important to avoid overheating or deflection which could interfere with correct functioning of the mechanism.

4.1.1. CHECKING FOR ULTIMATE STRENGTH 1 Mechanism components are checked for ultimate strength by verifying that the calculated stress does not exceed a permissible stress dependent on the breaking strength of the material used. 4.1.1.1. VALUE OF THE PERMISSIBLE STRESS The value of the permissible stress σa is given by the following formula : σa = σR / νR where : σR is the ultimate stress for the material νR is a safety coefficient corresponding to each case of loading (clause 2.3.)

1

It might seem logical to check against the elastic limit, in line with the structures booklet, as this figure is, in principle, the limit not to be exceeded in the use of a material. The steels normally used for structures have a wide gap between the yield strength and the ultimate strength and this gap affords protection against sudden failure even when the yield strength is considerably exceeded. On the other hand, the use in mechanisms of certain steels with a very high, elastic limit as compared to the ultimate strength would result in fragile parts being produced if the permissible limit stress were to be based on the elastic limit, and any accidental overstepping of this limit would lead to immediate failure. This explains why the ultimate strength is chosen as the criterion for verification.

4-3


4.1.1.2. VALUES OF THE COEFFICIENT νR The values to be adopted for νR are given in table T.4.1.1.2. Table T.4.1.1.2. Values of νR Cases of loading Value of νR

I and II 2,2

III 1,8

In the case of grey cast iron, the values of νR are to be amplified by 25 %. 4.1.1.3. RELATIONS BETWEEN THE CALCULATED STRESSES AND THE PERMISSIBLE STRESSES According to the type of loading considered, the following relations must be verified in which : σt is the calculated tensile stress σc is the calculated compressive stress σf is the calculated bending stress τ is the calculated shear stress. 1) Pure tension :

1,25 . σt ≤ σa

2) Pure compression :

σc ≤ σa

3) Pure bending :

σf ≤ σa

4) Combined bending and tension :

1,25 . σt + σf ≤ σa

5) Combined bending and compression :

σc + σf ≤ σa

6) Pure shear :

30,5 τ ≤ σa

7) Combined tension, bending and shear :

[ ( 1,25 . σt + σf )2 + 3 τ2 ]0,5 ≤ σa

8) Combined compression, bending and shear :

[ ( σc + σf )2 + 3 τ2 ]0,5 ≤ σa

4-4


4.1.2. CHECKING FOR CRIPPLING Parts subject to crippling are designed in compliance with clause 3.3. for booklet 3, checking that the calculated stress does not exceed a limit stress determined as a function of the critical stress above which there is a risk of crippling occurring. For this check, the coefficient γm is taken into account, its value depending on the group in which the mechanism is classified (see table T.2.6.) Some general considerations relating to the checking of parts for crippling are given in appendix A-3.3.

4.1.3. CHECKING FOR FATIGUE 4.1.3.1. GENERAL METHOD The fatigue strength of a given component is mainly determined by : - the material from which the component is constructed ; - the shape, surface condition, state of corrosion, size (scale effect) and other factors producing stress concentration ; - the ratio κ between the minimum and maximum stresses which occur during the various stress cycles ; - the stress spectrum ; - the number of stress cycles. The fatigue strength of a mechanical component is known only in exceptional cases. Generally speaking, it is to be derived from the characteristics of the material and of the component and from accepted laws concerning their behaviour. The starting point is provided by the endurance limit under alternating tensile fatigue loading ( κ = - 1) of a polished specimen, made from the material under consideration. The diminution of this fatigue strength as a result of the geometric shape of the piece, its surface condition, its state of corrosion and its size is allowed for by introducing appropriate factors. From the endurance limit under alternating loading the corresponding limit with respect to other ratios κ between extreme stresses is obtained with the aid of a SMITH diagram, in which certain hypotheses are made as to the shape of the strength curve. The endurance limit thus determined for the actual component, and with respect to a given ratio κ between extreme stresses, is taken as the basis for the plotting of the WÖHLER curve, concerning which certain hypotheses are also made. From this WÖHLER curve (fatigue strength 4-5


under the effect solely of stress cycles, all having the same ratio κ between extreme stresses), the PALMGREN-MINER hypothesis on fatigue damage accumulation can be used to determine the fatigue strength of a component according to a group in which the component is classified. The method described in 4.1.3. for determining the fatigue strength is applicable only to components in which the structure of the material is homogenous over the entire section being considered. It cannot, therefore, be used in the case of components which have undergone a surface treatment (e.g. hardening, nitriding, casehardening). In such cases the fatigue strength can be derived from the WÖHLER curve only if the latter has itself been determined with respect to components which have been made from the same material, have a comparable shape and size and have undergone exactly the same surface treatment. Checking for fatigue strength only needs to be performed for case of loading I. Where the number of stress cycles is less than 8 000, such checking is not necessary. 4.1.3.2. ENDURANCE LIMIT UNDER ALTERNATING LOADING OF A POLISHED SPECIMEN The specialised works on the subject (see also appendix A-4.1.3.) provide the endurance limit value σbw under alternating rotational bending of a polished specimen in the case of materials used regularly in construction of mechanisms. By approximation, the same values of σbw may be accepted for the endurance limit under alternating rotary bending. To obtain the endurance limit under alternating axial tension and compression, the values of σbw have to be decreased by 20 % 2.

2

An element of material, when subjected to the same stress as an adjacent element, supports the latter less effectively than if it were subjected to a lower stress, as is the case with bending. A stress gradient, i.e. : ( difference in stress between two adjacent elementary parts ) ( distance between these two elementary parts )

which is higher, produces a strengthening effect.

4-6


The endurance strength τw under alternating shear (pure shear or torsion) is derived from σbw by the relation : τw = σbw / 30,5 The values given for σbw are generally those corresponding statistically to a 90 % survival probability. In the case of carbon steels in common use in mechanisms, it is permissible to adopt : σbw = 0,5 . σR σR being the minimum ultimate strength. 4.1.3.3. INFLUENCE OF THE SHAPE, SIZE, SURFACE CONDITION AND CORROSION The shape, size, surface condition (machining) and state of corrosion of the component under consideration entail a decrease in the endurance limit under alternating loading in relation to the ideal case of a polished specimen. This influence is allowed for by introducing factors ks, kd, ku and kc respectively, concerning the determination of which, directions will be found in appendix A-4.1.3. The endurance limit under alternating loading σwk or τwk of the component under consideration is given for tension, compression, bending and torsional shear by the relation : σwk = σbw / ( ks . kd . ku . kc ) or τwk = τw / ( ks . kd . ku . kc ) In the case of pure shear we take : τwk = τw4.1.3.4. ENDURANCE LIMIT AS A FUNCTION OF κ, σR AND σwk (or τwk) Fig. 4.1.3.4. expresses, in the form of a SMITH diagram, the hypotheses made concerning the relations between the endurance limit σd (or τd), the ratio κ between extreme stresses, the tensile strength σR and the endurance limit under alternating loading σwk (or τwk) This gives the following relations : Normal stresses : -1 ≤ κ < 0

0

≤ κ < 1

Alternating stresses σd = 5 . σwk / ( 3 - 2 . κ ) Pulsating stresses σd = [ 5 . σwk / 3 ] / { 1 - [(1 - 5 . σwk / ( 3 . σR )) . κ ] }

4-7


Shear stresses : -1 ≤ κ < 0

0

≤ κ < 1

Alternating stresses τd = 5 . τwk / ( 3 - 2 . κ ) Pulsating stresses τd = [ 5 . τwk / 3 ] / { 1 - [(1 - 5 . 30,5 . τwk / ( 3 . σR )) . κ ]

Tension, Compression

or Shear or

Average

Figure 4.1.3.4.

4-8


4.1.3.5. WÖHLER CURVE "WÖHLER curve" in this context means an endurance curve representing the number n of stress cycles which can be withstood before fatigue failure, as a function of the maximum stress σ (T), when all stress cycles present the same amplitude and the same ratio κ between extreme values. With regard to this WÖHLER curve, the following hypotheses are made respectively : - for n = 8 * 103 : σ = σR or τ = σR / 30,5 - for 8 * 103 ≤ n ≤ 2 * 106 , the area of limited endurance, the function is represented by a straight fine TD in a reference system comprising two logarithmic scale axes (figure 4.1.3.5.) The slope of the WÖHLER curve, in the interval considered, is characterised by the factor : c = tan( ϕ ) = [ log(2 * 106 ) - log(8 * 103 ) ] / ( log σR - log σd ) or c = tan( ϕ ) = [ log(2 * 106 ) - log(8 * 103 ) ] / [ ( log ( σR / 30,5 ) - log τd ) ] 6 - for n = 2 * 10 :

σ = σd or τ = τd - for n > 2 * 106 , the so-called region of endurance limit, the function is represented, in the same reference system as above, by the straight fine DN, bisector of the angle formed by the extension of TD and a fine parallel to the axis of the n values, passing through D. 6 The slope of the WÖHLER curve for n > 2 * 10 is characterised by the factor :

c’ = tan( ϕ‘ ) = c + ( c2 + 1 )0,5

4-9


Figure 4.1.3.5. The spectrum factor ksp of the component is determined by means of the above mentioned value of c. In the case of certain components in group E8 (see 4.1.3.6.), the calculation must also be performed in exactly the same way, but after replacing c by c'. To distinguish between the two spectrum factors thus found, the second will be designated k’sp . A c value below 2,5 is an indication of faulty design of the component concerned. Such a component must not be put into service.

4 - 10


4.1.3.6. FATIGUE STRENGTH OF A MECHANICAL COMPONENT The fatigue strength σk or τk of a given mechanical component is determined by the following expressions respectively : σk = 2[ ( 8 - j ) / c ] . σd or τk = 2[ ( 8 - j ) / c ] . τd where j is the component's group number. In the case of group E8 components, of which the total duration of use n and the spectrum factor k'sp (see 4.1.3.5.) satisfy the inequality : n . k'sp > 2 * 106 σk or τk must, however, be determined by the expression : σk = [ (2 * 106 / n ) . ( 1 / k'sp ) ]1/c’ . σd or τk = [ (2 * 106 / n ) . ( 1 / k'sp ) ]1/c’ . τd The group classification of components, on the basis of their total duration of use n and their spectrum factor ksp, as well as the critical fatigue stresses associated with each group, are illustrated in figure 4.1.3.6. where σjk represents the stress applying to group Ej. For the critical shear stresses, the letter σ must be replaced by τ. 4.1.3.7. PERMISSIBLE STRESSES AND CALCULATIONS The permissible stresses σaf and τaf are obtained by dividing the stresses σk and τk, defined in 4.1.3.6., respectively by a safety factor νk. One takes : νk = 3,21/c or νk = 3,21/c’

σaf and τaf will therefore be obtained by the relations : σaf = σk / νk τaf = τk / νk 4 - 11

for group E8 components satisfying the inequality in the penultimate paragraph of 4.1.3.6.


and one verifies that : σ ≤ σaf τ ≤ τaf with : σ maximum calculated normal stress ; τ maximum calculated shear stress.

Figure 4.1.3.6. In the case of components acted upon simultaneously by normal stresses and shear stresses with different ratios κ between extreme stresses, the following condition must be satisfied : ( σx / σkx )2 + ( σy / σky )2 - [ σx . σy / ( σkx . σky ) ] + (τ / τk )2 ≤ 1,1 / νk2 in which : σx, σy = maximum normal stresses in the directions x and y respectively ; τ = maximum shear stress ; σkx , σky = fatigue strength for normal stresses, in the directions x and y respectively ; 4 - 12


τk = shear fatigue strength. If it is not possible to determine the most unfavourable case of the foregoing relation from the corresponding stresses σx, σy and τ, calculations must be performed separately for the loads σx max, σy max and τmax and the most unfavourable corresponding stresses. It should be noted that the checks described above do not guarantee safety against brittle fracture. Such safety can be ensured only by a suitable choice of material quality.

4.1.4. CHECKING FOR WEAR In the case of parts subjected to wear, the specific physical quantifies which affect this, such as the surface pressure or the circumferential velocity must be determined. The figures must be such that, on the basis of present experience, they will not lead to excessive wear.

4 - 13


4.2. DESIGN CALCULATIONS FOR PARTICULAR COMPONENTS 4.2.1. CHOICE OF ANTI-FRICTION BEARINGS To select anti-friction bearings, it is first necessary to check that the bearing is capable of withstanding : - the static load to which it can be subjected under whichever of loading cases I, II or III is the most unfavourable, and - the maximum dynamic load in the more unfavourable of loading cases I or II. 4.2.1.1. THEORETICAL LIFE In addition, anti-friction bearings must be selected to give an acceptable theoretical life in hours (see table T.2.1.3.2.) as a function of the class of operation of the mechanism under a constant mean load as defined in clauses 4.2.1.2. and 4.2.1.3. below. 4.2.1.2. MEAN LOADING OF BEARINGS SUBJECTED TO TYPE SM LOADS In order to allow for variations in the loads of type SM during the cycles of operation, an equivalent mean loading SM mean is determined which is supposed to be applied constantly during the theoretical life determined by clause 4.2.1.1. SM mean is obtained by multiplying SM max II 3 , defined by clause 2.6.4.1. and 2.6.4.2., by the cube root of the spectrum factor km defined in 2.1.3.3. SM mean = km . SM max II (1)

4.2.1.2.1. Determination of the mean load SM mean on anti-friction bearings for combined motions

In the case of motions which combine an elevation of the centre of gravity of the moving masses with a horizontal displacement (e.g. unbalanced luffing), the mean load SM mean is determined by combining : - the mean load due to the accelerations and the effect of the wind, as determined by applying clause 4.2.1.2. with,

3

or SM max I for components not subjected to wind.

4 - 14


- the mean load due to the vertical displacement of the centre of gravity of the moving masses, as determined from the expression : SM mean = ( 2 . SM max + SM min ) / 3 where SM max and SM min are the maximum and minimum values of the corresponding loads. 4.2.1.3. MEAN LOADING OF BEARINGS SUBJECTED TO TYPE SR LOADS The extreme loads SR max and SR min developed in loading case I for appliances not subjected to wind or loading case II for appliances subjected to wind (see clause 2.6.) are considered and the bearing is designed for a constant mean load given by the expression : SR mean = ( 2 . SR max + SR min ) / 3 and applied for the theoretical life in accordance with clause 4.2.1.1. 4.2.1.4. MEAN LOADING OF BEARINGS SIMULTANEOUSLY SUBJECTED TO TYPE SM AND TYPE SR LOADS On the basis indicated above the equivalent mean loads are determined for each of the type SM and SR loads, assumed to be acting alone and the bearing is selected for an equivalent mean load resulting from combination of the two mean loads SM and SR.

4.2.2. CHOICE OF ROPES The following rules aim at defining reasonable minimum requirements for the choice of ropes used on hoisting appliances covered by these Design Rules. They do not purport to resolve all the problems nor to serve as a substitute for the dialogue which is essential between the rope manufacturer and the manufacturer of hoisting appliances. They apply to preferred ropes conforming to ISO Recommendation 2408 : "Steel ropes for general use - Characteristics". They do not exclude, however, ropes which are not specified in ISO Recommendation 2408. For such ropes, it is incumbent upon the rope manufacturer to validate for the user the minimum values of parameters detailed in the ISO Recommendation. The terminology of the rope parameters complies with that used in ISO Recommendation 2408.

4 - 15


The methods stated hereafter assume that the ropes are greased correctly, that the winding diameters of the pulleys and the drums are suitably selected in compliance with 4.2.3. and that, when in service, the ropes are properly maintained, inspected and periodically replaced in conformity with ISO Recommendation 4309 "Rope inspection" The selection of rope diameter (and winding diameters in 4.2.3.) is based on the group of the hoisting mechanism. However, for appliances which require frequent dismantling (such as builders tower cranes), in which ropes have to be changed more frequently, it is permissible to select a hoist rope from the group immediately below that of the hoisting mechanism but not inferior to group M3. Whenever hoisting appliances are used for dangerous handling operations (e.g. molten metal, highly radioactive or corrosive products, etc.) the choice of the ropes and pulleys must take account of the mechanism group next above that resulting from the normal classification of the hoisting appliances. Group M5 is the minimum group to be used for the handling of dangerous loads for the choice of rope and pulley diameters. 4.2.2.1. CHOICE OF ROPE DIAMETER Two methods can be used at the choice of the manufacturer : - the method using the minimum practical factor of safety Zp (see 4.2.2.1.2.) which is valid for running ropes and static ropes (guy ropes, stays, etc.). - the C factor method (4.2.2.1.3.) applicable to running ropes only. 4.2.2.1.1. Common bases for the two methods 4.2.2.1.1.1. Definition of the maximum tensile force S in the hoist rope (grab ropes excepted)

This is obtained by taking account of the following factors : - maximum safe working load of the appliance, - weight of the pulley block and the hoist accessories, the dead weights of which are added to the load effect so as to increase the rope tension, - mechanical demultiplication due to the rope reeving, - efficiency of the rope reeving, - loads due to accelerations if they exceed 10 % of the vertical loads, - rope inclination at the upper extreme position if the angle of the rope with the hoist axis exceeds 22,5°. 4 - 16


4.2.2.1.1.2. Definition of the maximum tensile force S in ropes other than hoist ropes

The determination of the maximum tensile force S in the various ropes which are not exclusively used for the vertical hoisting of loads is based on the loads determined in load cases I or II taking account of the most unfavourable case which can occur repeatedly in normal use. For ropes which produce horizontal movement of loads, account must be taken of the loading resulting from rolling motion and friction, together with the maximum inclination that the support, on which the load is moved, can assume locally under the influence of the normal loading. 4.2.2.1.1.3. Determination of the maximum tensile force S in the ropes of multirope grabs (holding and closing)

In the case of appliances with grabs, where the weight of the load is not always equally distributed between the closing ropes and the holding ropes during the whole of a cycle, the value of S to be applied should be determined as follows : 1) If the system used automatically ensures an equal distribution of the hoisted load between the closing and holding ropes, or any difference between the loads carried by the ropes is limited to a short period at the end of closing or at the beginning of opening, S should be determined as follows : a) closing ropes : S = 66 % of the weight of the loaded grab, divided by the number of closing ropes. b) holding ropes : same pourcentage. 2) If the system used does not automatically secure an equal distribution of the load between the closing ropes and the holding ropes during the hoisting motion and, in practice, almost all the load is applied to the closing ropes, S should be determined as follows : a) closing ropes : S = total weight of the loaded grab divided by the number of closing ropes. b) holding ropes : S = 66 % of the total weight of the loaded grab divided by the number of holding ropes. 4.2.2.1.2. Method using the minimum practical factor of safety Zp

Definitions The minimum practical factor of safety Zp is the ratio between : - the minimum breaking load F0 of the rope (minimum load which must be attained when carrying out the rope breaking test),

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- and the maximum tensile force S in the rope Zp = F0 / S 4.2.2.1.2.1. Rope selection

The chosen rope must have a minimum practical factor of safety at least equal to the minimum value Zp for the mechanism group to which the rope in question belongs (see table T.4.2.2.1.2.). Table T.4.2.2.1.2. - Factor of safety Zp Group of mechanism M1 M2 M3 M4 M5 M6 M7 M8

Minimum value Zp Running ropes Static ropes 3,15 2,5 3,35 2,5 3,55 3 4 3,5 4,5 4 5,6 4,5 7,1 5 9 5

4.2.2.1.3. C-factor method

Definitions : C = rope selection factor, S = maximum tensile force exerted on the rope when in use, d = nominal diameter of the rope (dimension by which the rope is designated), f = fill factor of the rope, k = spinning loss factor due to the rope construction, RO = minimum ultimate tensile stress of the wire composing the rope, K' = empirical factor for the minimum breaking load for a given rope construction such that k’ = π / 4 .f . k

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4.2.2.1.3.1. Rope selection

For a rope of a given construction, having a given minimum steel strength, and for a given mechanism group there is a factor C which is expressed by the formula : 0,5

C = [ Zp / (π . k . f . RO / 4 ) ]

= [ Zp / (k’ . RO ) ]0,5

where Zp is the minimum value for running ropes in table T.4.2.2.1.2., corresponding to the mechanism group chosen for the rope. d ≥ C . S0,5

The nominal diameter must be such that : 4.2.2.1.3.2. Calculation of the factor C - guarantees

The values of C are calculated taking account of : - the factor Zp corresponding to the mechanism group, - the breaking strength under tension of the steel of the rope wires, - the factor k' (or factors k and f) which can either : • be taken from ISO Recommendation 2048 for normal ropes covered therein (see Appendix), • or be guaranteed by the rope manufacturer if the rope is of a special construction. In this case, the certificate supplied by the rope manufacturer must clearly state the guaranteed values of k'.

4.2.3. CHOICE OF PULLEYS, DRUMS AND ROPE ATTACHMENT MEANS 4.2.3.1. MINIMUM WINDING DIAMETER The minimum winding diameter for the rope is determined by checking the relationship : D ≥ H.d where : D is the winding diameter on pulleys, drums or compensating pulleys measured to the axis of the rope. H is a coefficient depending upon the mechanism group. d is the nominal diameter of the rope. Note - Refer to 4.2.2. for the mechanism group in which the rope should be classified.

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4.2.3.1.1. Values of H

The minimum values of the coefficient H, depending upon the group in which the mechanism is classified, are given in table T.4.2.3.1.1. for drums, pulleys and compensating pulleys. They correspond to ropes currently used and known and are based on experience concerning their working conditions. They do not however serve as a substitute for the dialogue which is indispensable between the rope manufacturer and the manufacturer of hoisting appliances, especially when the use of new ropes with varying flexibility characteristics is being considered. Table T.4.2.3.1.1. Values of H Mechanism group M1 M2 M3 M4 M5 M6 M7 M8

Drums 11,2 12,5 14 16 18 20 22,4 25

Pulleys 12,5 14 16 18 20 22,4 25 28

Compensating pulleys 11,2 12,5 12,5 14 14 16 16 18

4.2.3.1.2. Note

When the formula given in clause 4.2.2.1. has been used to determine a minimum rope diameter from which in turn the minimum diameters for drums and pulleys have been determined, a rope of diameter greater than the minimum calculated diameter can be used with these latter diameters, provided that the diameter of the rope used does not exceed the minimum diameter by more than 25 % and that the pull in the rope does not exceed the value S used for calculating this minimum diameter. 4.2.3.2. RADIUS OF THE BOTTOM OF THE GROOVE The useful life of the rope depends not only on the diameter of the pulleys and drums, but also on the pressure exerted between the rope and the groove supporting the rope. The winding ratios above are given on the assumption of a radius of supporting groove r where : r = 0,53 . d d being the nominal diameter of the rope.

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4.2.3.3. ROPE ATTACHMENT MEANS Rope attachments must be so designed as to withstand a tensile force 2,5 times the maximum tensile force S without showing permanent deformation. The means attaching the rope to the drum must be of such a design that, taking account of the friction of the turns which remain around the drum, the sum of the frictional and fixing forces withstands a tensile force 2,5 times the maximum tensile force S. The coefficient of friction between the rope and the drum used in the calculations shall be : µ = 0,1 when the rope is completely unwound from the drum, at least two complete turns of rope must remain on the drum before the rope end attachment.

4.2.4. CHOICE OF RAIL WHEELS In order to choose a rail wheel, its diameter is determined by considering : - the load on the wheel, - the quality of the metal from which it is made, - the type of rail on which it rues, - the speed of rotation of the wheel, - the group classification of the mechanism. 4.2.4.1. RAIL WHEEL SIZE To determine the size of a rail wheel, the following checks must be made : - that it is capable of withstanding the maximum load to which it will be subjected, and - that it will allow the appliance to perform its normal duty without abnormal wear. The two requirements are checked by means of the following two formulae : Pmean III ≤ PL . C1max . C2max < 1,38 . PL ≈ 1,4 PL taking

C1max = 1,2 and C2max = 1,15

and

Pmean I, II / ( b . D ) ≤ PL . C1 . C2

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where : D is the wheel diameter in mm b the useful width of the rail in mm PL a limiting pressure dependent upon the metal used for the wheel, in N/mm2 C1 a coefficient depending on the speed of rotation of the wheel C2 is a coefficient depending on the group of the mechanism Pmean III is the mean load to be withstood by the wheel, in loading case III, in N, calculated according to the formulae in clause 4.2.4.1.1. Pmean I,II is the mean load in case I or II. 4.2.4.1.1. Determining the mean load

In order to determine the mean loads, the procedure is to consider the maximum and minimum loads withstood by the wheel in the loading cases considered, i.e. with the appliance in normal duty but omitting the dynamic coefficient ψ when determining Pmean I,II and with the appliance not in use for Pmean III .The values of Pmean are determined by the formula below in the three cases of loading I, II and III : Pmean I,II,III = ( Pmin I,II,III + 2 . Pmax I,II,III ) / 3 4.2.4.1.2. Determining the useful rail width b

For rails having a flat bearing surface and a total width l with rounded corners of radius r at each side, we have: b=l-2.r for rails with a convex bearing surface, we have: b=l-4.r/3

4

4

(1) For the same width of rail head, these formulae give a greater useful bearing width for convex rails than for flat rails. This allows for the superior adaptation of a slightly convex rail to the rolling motion of the wheel. 4 - 22


4.2.4.1.3. Determining the limiting pressure PL

The value of PL is given in table T.4.2.4.1.3. as a function of the ultimate strength of the metal of which the rail wheel is made; Table T.4.2.4.1.3. - Values of PL Ultimate strength for metal used for rail wheel σR > 500 N/mm2 σR > 600 N/mm2 σR > 700 N/mm2 σR > 800 N/mm2

PL in N/mm2 5,0 5,6 6,5 7,2

The qualifies of metal refer to cast, forged or rolled steels, and spheroidal graphite cast iron. In the case of rail wheels with tyres, consideration must obviously be given to the quality of the tyre, which should be sufficiently thick not to roll itself out. In the case of wheels made of high tensile steel and treated to ensure a very high surface hardness, the value of PL is limited to that for the quality of the steel composing the wheel prior to surface treatment, according to table T.4.2.4.1.3., since a higher value would risk causing premature wear of the rail. For a given load, however, wheels of this type have a much longer useful life than wheels of lesser surface hardness, which makes their use worthwhile in the case of appliances performing intensive service. Alternatively, it is possible to use wheels of ordinary cast iron, especially chilled cast iron, which has good surface hardness. It must be remembered that such wheels are brittle and that their use should be avoided for high speed motions or when shock loadings are to be feared. 2 When these are used their diameter is determined by taking PL equal to 5 N/mm .

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4.2.4.1.4. Determining the coefficient c1

The values of c1 depend on the speed of rotation of the wheel and are given in table T.4.2.4.1.4.a. These same values are also given in table T.4.2.4.1.4.b. as a function of the wheel diameter and the speed in m/min. Table T.4.2.4.1.4.a. - Values of c1 Wheel rotation speed in R.P.M. 200 160 125 112 100 90 80 71 63 56

c1 0,66 0,72 0,77 0,79 0,82 0,84 0,87 0,89 0,91 0,92

Wheel rotation speed in R.P.M. 50 45 40 35,5 31,5 28 25 22,4 20 18

c1 0,94 0,96 0,97 0,99 1,00 1,02 1,03 1,04 1,06 1,07

Wheel rotation speed in R.P.M. 16 14 12,5 11,2 10 8 6,3 5,6 5

c1 1,09 1,10 1,11 1,12 1,13 1,14 1,15 1,16 1,17

Table T.4.2.4.1.4.b Values of c1 as a function of the wheel diameter and the speed of travel wheel diameter in mm 200 250 315 400 500 630 710 800 900 1 000 1 120 1 250

Values of c1 for travel speeds in m/min 10 1,09 1,11 1,13 1,14 1,15 1,17 -

12,5 1,06 1,09 1,11 1,13 1,14 1,15 1,16 1,17 -

16 1,03 1,06 1,09 1,11 1,13 1,14 1,14 1,15 1,16 1,17 -

20 1 1,03 1,06 1,09 1,11 1,13 1,13 1,14 1,14 1,15 1,16 1,17

25 0,97 1 1,03 1,06 1,09 1,11 1,12 1,13 1,13 1,14 1,14 1,15

31,5 0,94 0,97 1 1,03 1,06 1,09 1,1 1,11 1,12 1,13 1,13 1,14

40 0,91 0,94 0,97 1 1,03 1,06 1,07 1,09 1,1 1,11 1,12 1,13

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50 0,87 0,91 0,94 0,97 1 1,03 1,04 1,06 1,07 1,09 1,1 1,11

63 0,82 0,87 0,91 0,94 0,97 1 1,02 1,03 1,04 1,06 1,07 1,09

80 0,77 0,82 0,87 0,91 0,94 0,97 0,99 1 1,02 1,03 1,04 1,06

100 0,72 0,77 0,82 0,87 0,91 0,94 0,96 0,97 0,99 1 1,02 1,03

125 0,66 0,72 0,77 0,82 0,87 0,91 0,92 0,94 0,96 0,97 0,99 1

160 0,66 0,72 0,77 0,82 0,87 0,89 0,91 0,92 0,94 0,96 0,97

200 0,66 0,72 0,77 0,82 0,84 0,87 0,89 0,91 0,92 0,94

250 0,66 0,72 0,77 0,79 0,82 0,84 0,87 0,89 0,91


4.2.4.1.5. Determining the coefficient c2

The coefficient T.4.2.4.1.5.

c2

depends on the group classification of the mechanism and is given in table

Table T.4.2.4.1.5. - Values of c2 Group classification of mechanism M 1 to M 4 M5 M6 M7-M8

c2 1,12 1,00 0,90 0,80

4.2.4.2. NOTES Note 1 The formulae apply only to wheels whose diameters do not exceed 1,25 m. For larger diameters experience shows that the permissible pressures between the rail and the wheel must be lowered. The use of wheels of greater diameter is not recommended. Note 2 It should be noted that the limiting pressure PL is a notional pressure determined by supposing that contact between the wheel and the rail takes place over a surface whose width is the useful width defined earlier (clause 4.2.4.1.2.) and whose length is the diameter of the wheel. The method of calculating set out above is derived from application of the HERTZ formula, which may be written : 2 σcg / ( 0,35 . E ) = P / ( b . D ) where : σcg is the compressive stress in the wheel and the rail N/mm2 E the modulus of elasticity of the metal in N/mm2 P the wheel load in N b and D in mm, being as defined above (clause 4.2.4.1.). Taking KL to represent the value σcg2 / ( 0,35 . E ) which has the dimension of a pressure in 2 N/mm , the relation may be written :

KL = P / ( b . D )

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and characterizes the wheel pressure on the rail. The formula of clause 4.2.4.1. is obtained by putting : KL = PL . c1 . c2

4.2.5. DESIGN OF GEARS The choice of the method of making design calculations for gears is left to the manufacturer, who must indicate the origin of the method adopted, the loads to be taken into account being determined in accordance with the directions given in 2.6. In the case of a calculation which takes account of the operating time the conventional hours determined in 2.1.3.2. should be used.

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APPENDIX A- 4.1.3. - DETERMINATION OF PERMISSIBLE STRESSES IN MECHANISM COMPONENTS SUBJECTED TO FATIGUE The endurance limit for a polished specimen is a laboratory value, which is practically never attained in parts actually used. Numerous factors - shape, size, surface condition (machining quality) and possible corrosion - induce discontinuities resulting in "notch effects", which decrease the permissible stresses in the part, when these stresses are calculated by conventional elementary methods for the strength of materials. These factors are taken into account by coefficients, called ks, kd, ku, kc, respectively all greater than or equal to unity, by the product of which the endurance limit for a polished specimen is devided. Guidelines concerning the determination of these coefficients are set out below : a. Determination of ks This coefficient specifies the stress concentrations caused by changes of section with radii, annular grooves, transverse holes and the method of securing hubs. Figures A.4.1.3.1. a. and b. give the values of the shape coefficient ks as a function of the ultimate strength of the metal, valid for diameter D of 10 mm. The curves a. give the coefficient ks for changes of section of ratio D/d = 2 with a correction table T.A.4.1.3.1. for other values of D/d. The b curves give, for guidance some values of ks for holes, annular grooves and keyways. Diameters in excess of 10 mm are taken care of by introducing the size coefficient kd.

Figure A.4.1.3.1.a. Shape coefficient ks (Diameter D = 10 mm) Change of section D/d = 2 4 - 27


For other values of D/d read ks from the curve (r/d) + q with the following values for q : Table T.A.4.1.3.1. - Correction factors q for D/d ≤ 2 D/d q

1,05 0,13

Curve

1,1 0,1

1,2 0,07

1,3 0,052

1,4 0,04

1,6 0,022

2 0

I Transverse hole d1 = 0,175 d II Annular groove : depth 1 mm III Keyed hub IV Press-fitted hub Figure A.4.1.3.1.b. Shape coefficient ks (Diameter D =10 mm) Hole, annular groove, keyway

b. Determination of size coefficient kd For diameters greater than 10 mm the stress concentration effect increases and this increase is allowed for by introducing the size coefficient kd. The values of the coefficient kd are given in table T.A.4.1.3.2. for values of d from 10 mm to 400 mm. Table T.A.4.1.3.2. - Values of kd d mm kd

10 1

20 1,1

30 1,25

50 1,45

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100 1,65

200 1,75

400 1,8


c. Determination of surface condition (method of machining) coefficient ku Experience shows that parts produced with a rough finish have a longer endurance limit than carefully polished parts. This is allowed for by applying a machining coefficient ku given in figure A.4.1.3.2. for the case of a surface which is ground or finely polished with emery and for the case of a surface which is rough machined. d. Determination of corrosion coefficient kc Corrosion can have a very appreciable effect on the endurance limit of steels ; this is allowed for by applying a coefficient kc. Figure A.4.1.3.2. gives the values of this coefficient kc for the cases of corrosion due to fresh water and due to sea water.

Figure A.4.1.3.2. - Values of the machining coefficient ku , corrosion coefficient kc Values of ku Curve

I II

Surface ground or finely polished Surface rough machined

III IV

Surface corroded by fresh water Surface corroded by sea water

Values of kc

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EXAMPLE OF APPLICATION

Shaft in A-550 steel with change of section. Diameters D = 70 mm and d = 5D mm with transition radius r = 5 mm. Turned on lathe, with keyed wheel.

The component will be deemed to be classified in group E4. We shall assume alternating loading (κ = - 1) and the shaft to be of A 550 steel 2 (minimum σR = 550 N/mm ). We can therefore adopt : σbw = 0,5 . 550 = 275 N/mm2 Section A-B D/d = 70 / 50 = 1,4 r/d = 5 / 50 = 0,1 Determination of ks (shape) For D/d = 1.4 we have : q = 0,04

(Table T.A.4.1.3.1.)

From the curve (r/d) + q = 0,1 + 0,14 we find by interpolation : ks = 1,4

(Figure A.4.1.3.1.a.)

Determination of kd (size) For d = 50 we have : kd = 1,45

(Table T.A.4.1.3.2.)

Determination of ku (machining) For a part turned on a lathe we have : ku = 1,15

(Figure A.4.1.3.2., curve II)

From the foregoing values we derive : σwk = 275 / ( 1,4 . 1,45 . 1,15 ) = 117,8 N/mm2

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For κ = - 1 we have : σd = σwk = 117,8 N/mm2 c = log(2 000 000/8 000) / log(550 / 117,8) = 3,58 For group E4 the critical stress is therefore : σk = σd . 2(8-4)/c = 117,8 . 2(4 / 3,58) = 255,4 N/mm2 The safety coefficient νk is given by : νk = 3,21/c = 3,2(1 / 3.58) = 1,38 The permissible stress σaf is therefore : σaf = 255,4 / 1,38 = 184,6 N/mm2 Section C-D We have : ks = 2,2

(Figure A.4.1.3.1.b.)

kd = 1,45

(same value as above)

ku = 1,15

(same value as above)

Hence : σwk = 275 / ( 2,2 . 1,45 . 1,15 ) = 75,0 N/mm2 σd = σwk

= 75,0 N/mm2

c =log(2 000 000 / 8 000) / log(550 / 75) = 2,77 σk =

75 . 2(4/2,77) = 204 N/mm2

νk =

3,21/2.77 = 1,52

σaf =

204 / 1,52 = 134 N/mm2

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LIST OF SOME WORKS DEALING WITH FATIGUE PROBLEMS

(1) Niemann, G. Maschinenelemente Band 1 Springer Verlag Berlin/Gottingen/Heidelberg 1975 (2) Niemann, G. Maschinenelemente Band 2 Springer Verlag Berlin/Gottingen/Heidelberg 1983 (3) Decker, K.-H. Maschinenelemente Carl Hanser Verlag, Munchen 1982 (4) "Metal Fatigue" by J.A. POPE - Ph D, D.Sc - Wh.Sch. I. Mech. E. Chapman and Hall Ltd., 37, Essex Street, London, W.C.2. (5) "La Fatigue des Metaux" by R. CAZAUD - Ingenieur CNAM - Doctor of the University of Paris Lecturer at the Higher Institute for Mechanical Engineering Materials, Consulting Engineer Dunod 92, rue Bonaparte - Paris (6) "Fatigue of Metals and Structures" by H.J. GROVER, S.A. GORDON, R.L. JACKSON Thames and Hudson London

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A- 4.2.2. - COMMENTS ON THE CHOICE OF ROPES AND ON THE PROBLEM OF THE FACTOR OF SAFETY The useful life of a hoisting rope depends on factors inherent both in the construction of the rope (internal factors) and in the characteristics of the hoisting appliance and the way in which the rope is reeved (external factors). The main external factors are the tensile load, the pulley diameters and the type and number of working cycles. These factors will in principle determine the rope winding diameters. Figure F.A.4.2.2. shows the relationships between the tensile stress, the pulley diameters and the useful life (number of bending reversals causing failure) for a given rope (of diameter 16 mm). According to this diagram, the larger the pulley diameter and the smaller the tensile stress, the longer the useful life of the rope. Winding diameters must be determined so as to ensure a reasonable useful life of the rope before its replacement. The old method of calculation defined fixed minimum factors of safety and minimum winding diameters (as a function of the rope diameter) for certain uses, such as for example in hoisting appliances. This method, which is still used and sometimes even stipulated by law in many countries, does not enable the conditions for a reasonable useful life to be met. The loads and the type of utilisation of hoisting appliance mechanisms differ considerably from one appliance to another with the result that such minimum values may be too high in certain cases (e.g. for a power station overhead travelling crane) or too low in other cases (e.g. for a grabbing crane on arduous duty). Even if the safety factors were chosen according to the group of the mechanisms, the calculation of winding diameters on the basis of these safety factors could not result in a good design for the following reasons : The manufacturer wishes to use ropes of the smallest possible diameter in order to obtain the smallest possible diameters for the pulleys and drums. For a given safety factor, he achieves this by using wire of the greatest possible strength and using a rope with the greatest possible fill factor.

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Yet the useful life of a rope whose diameter has been calculated in this way is not always optimum. A rope of identical diameter, of which the wire is of longer strength (e.g. 1600 N/mm2 instead of 2 2200 N/mm ) and the fill factor is lower (e.g. eight strands instead of six) may have a much longer useful life in spite of the lower factor of safety. Another difficulty stems from the fact that the safety factor is related to breaking strengths whose definitions differ from one country to another, and safety factors do not have the same meaning if they relate to a breaking strength defined in different ways. Four definitions exist for the breaking strength of ropes : - Calculated breaking strength : this is the cross-section of the rope multiplied by the strength of the wire of which it is composed ; - Theoretical breaking strength : given by the sum of the breaking strengths of the wires used in the rope ; - Actual breaking strength : this is the load obtained by an ultimate tensile strength test on the rope ; - Nominal breaking strength : this is the minimum breaking strength guaranteed by the rope manufacturer. When a rope is determined by using a safety factor which is related either to the actual breaking strength or to the nominal breaking strength, the manufacturer tends to adopt ropes in which the spinning loss (difference between the theoretical breaking strength and the actual breaking strength) is as small as possible in order to obtain a smaller rope diameter. The spinning loss, however, is not related to the resistance of a rope to repeated bending. A satisfactory useful life for ropes cannot therefore be obtained with this method of calculation. This shows that the safety factor is not an adequate basis for determining the winding diameters required to ensure satisfactory rope life under bending reversals ; indeed this method will often prevent the optimum solution from being obtained. Because it is difficult for the manufacturer to allow for the influence of these different factors, it is preferable to determine the rope diameter dmin simply as a function of the tensile load S, from the formula : 0,5 dmin = C . S where C is a coefficient depending solely on the group of the mechanism.

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In cases where non rotating ropes are used (e.g. tower cranes in which the load is suspended from a single part) and for dangerous handling operations (e.g. molten materials), the values of C are increased above the normal values in order to compensate for the more unfavourable rope construction or the greater risk. The value of C, the safety factor Zp referred to the theoretical breaking strength and the rope fill factor f (ratio of the metal cross section of the rope to the area of the circle circumscribing the rope) are linked by the following relation : 0,5

C = [ Zp / (π . k . f .RO / 4 ) ]

where RO is the ultimate strength (in N/mm2) of the wire used in the rope. The values of C apply to ropes made with wire having a strength of 1600, 1800, 2000 or 2200 N/mm2 . where, exceptionally, use is made of a rope composed of wires with a strength of 1400 N/mm , the rope diameter must be increased accordingly. The rope manufacturer or crane maker must choose the composition and cross section of the rope, for the calculated minimum diameter d, to suit the reeving conditions of the particular rope and in the light of the latest technical progress.

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Number of bending reversals required to cause failure (useful life) Number of bending reversals

Tensil stress

Langs lay rope, diameter 16 mm, six strands of nineteen 1 mm dia. wires, Ro = 1400 N/mm2 Cast-iron pulleys with machined groove of radius r = 8,5 mm. Figure A.4.2.2. Influence of pulley diameter D and tensile stress σt on the useful life of a rope

4 - 36


A- 4.2.3. - CONSIDERATIONS ON THE DETERMINATION OF MINIMUM WINDING DIAMETERS FOR ROPES There are no absolute minimum pulley and drum diameters below which a rope could no longer operate. Nor is there a required absolute minimum diameter for different types of rope. The useful life of a rope is progressively reduced with decreasing pulley and drum diameters, if other conditions remain unchanged. Figure A.4.2.3. shows the pattern of behaviour of a particular rope. In order to ensure an adequate useful rope life, the minimum winding diameters D must be determined, as a function of the group classification of the mechanism involved, from the formula : D/d ≥ H where d is the nominal rope diameter and H is a coefficient which is chosen according to the group in which the mechanism is classified and which becomes higher when the duty is more arduous. The coefficient H is higher for pulleys than for drums because, in the course of a cycle of operations, the rope is subjected to twice as many bending reversals over a pulley (rope straight, rope bent, rope straight) as on a drum (rope straight, rope bent). The coefficient H is lower for equalising pulleys because the rope undergoes fewer bending reversals and the movement is normally very limited. Such pulleys must nonetheless be dimensioned with reference to the number of bending reversals. Unfavourable winding conditions, e.g. reeving around several pulleys, reverse bends, or the use of non-rotating ropes whose construction is less favourable for bending reversals, must be compensated for by a suitable increase to ensure a useful rope life commensurate with the group classification of the mechanism.

4 - 37


Number of bending reversals required to cause failure (useful life) Number of bending reversals

Pulley diameter

Ratio

Langs lay rope, diameter 16 mm, six strands of nineteen 1 mm dia. wires, σR = 1400 N/mm2 Cast-iron pulleys with machined groove radius r = 8,5 mm. Figure A.4.2.3. Influence of pulley diameter D and tensile stress σt on the useful life of a rope

4 - 38



F.E.M.

SECTION I

3rd EDITION REVISED


1.001

1998.10.01


ELECTRICAL EQUIPMENT Revised in 2003



FEDERATION EUROPEENNE DE LA MANUTENTION Section I HEAVY LIFTING APPLIANCES

F.E.M. 1.001 rd

3 EDITION REVISED 1998.10.01

RULES FOR THE DESIGN OF HOISTING APPLIANCES BOOKLET 5 ELECTRICAL EQUIPMENT Revised in 2003

rd

The total 3 revised comprises booklets 1 to 5 and 7 to 9 Copyright by FEM Section 1 Also available in French and German


The third edition of the "Rules for the design of hoisting appliances" dated 1987.10.01') included 8 booklets. An addition to this edition was compiled in 1998. This addition is incorporated in booklet 9, which also replaces booklet 6. This booklet forms part of the "Rules for the design of hoisting appliances" 3rd edition revised, consisting of 8 booklets: Booklet 1 - Object and scope Booklet 2 - Classification and loading on structures and mechanisms Booklet 3 - Calculating the stresses in structures Booklet 4 - Checking for fatigue and choice of mechanism components Booklet 5 - Electrical equipment revised in 2003 Booklet 6 – Stability and safety against movement by the wind Booklet 7 - Safety rules Booklet 8 - Test loads and tolerances Booklet 9 - Supplements and comments to booklets 1 to 8 NOTE: Booklet 9 must not therefore be used separately

Booklet 5 has been revised in 2003 to take into account various European standards. The main modifications concern the following points: 5.3 Installation of cables and conductors This clause has been re-arranged; the new version includes some parts of old 5.2 and references to EN 60204-32 EN 60364-5-52 5.5 Limiting and indicating devices (old title "End limit switches") This clause has been replaced by references to requirements coming from European standards on motor over speed protection

5.6 Type of control The revision of this clause makes references to requirements coming from European standards on braking resistors of inverters

Booklet 5 - 1/ 35


5.7 Environment The existing text has been deleted and replaced by the reference to EN 60204-32. Clauses on EMC and on potentially explosive atmospheres have been introduced

5.8 Selection of motors The revision of this clause introduced minor changes. A new clause 5.8.4. on inverter drives has been added

5.9 to 5.11 The revision of this clause introduced minor changes. The reference to IEC 60034-1 has been added.

Booklet 5 - 2/ 35


BOOKLET 5 - ELECTRICAL EQUIPMENT 5.1 FOREWORD..................................................................................................................................................................................... 5 5.2 POWER SUPPLY .......................................................................................................................................................................... 6 5.2.1.

G E N E R A L.........................................................................................................................................................................................................................................................6

5.2.2.

CUT-OUT AND S AFETY DEVICES..............................................................................................................................................................................................6

5.2.3.

CONDUCTOR BAR S, CABLE REELS AND FLEXIBLE CABLES ..........................................................................................................................6

5.3 INSTALLATION OF CABLES AND CONDUCTORS ................................................................................................................. 7 5.3.1.

CALCULATION O F C R O S S - S E C T I O N O F C O N D U C T O R S .....................................................................................................................................7 5.3.1.1.

CALCULATION OF THE CROSS-SECTION IN RELATION TO THE ADMISSIBLE VOLTAGE DROP.........................7

5.3.1.2. CALCULATION OF THE MINIMUM CROSS-SECTION IN RELATION TO THE THERMAL CAPACITY OF THE CONDUCTORS ...............................................................................................................................................................8

5.3.2.

INSTALLATION C O N D I T I O N S.........................................................................................................................................................................................................9

5.4 ELECTRICAL PROTE CTIVE AND SAFETY EQUIPMENT..................................................................................................... 10 5.4.1.

SAFEGUARDING MOTORS AGAINST OVERHEATING ..........................................................................................................................................10

5.4.2.

SAFEGUARDING WIRING...............................................................................................................................................................................................................10

5.4.3.

SAFEGUARDING AGAINST ABSENCE OR INVERSION OF PHASES ........................................................................................................10

5.4.4.

ACTION OF SAF ETY DEVICES...................................................................................................................................................................................................10

5.4.5.

PROTECTION AG AINST THE EFFECTS OF LIGHTNING ......................................................................................................................................11

5.5 LIMITING AND INDICATING DEVICES..................................................................................................................................... 12 5.5.1.

GE N E R A L R E Q U I R E M E N T S.........................................................................................................................................................................................................12

5.5.2.

M O T O R O V E R S P E ED PROTECTION....................................................................................................................................................................................12

5.6 CONTROLS .................................................................................................................................................................................... 13 5.6.1.

C O M P O N E N T S.........................................................................................................................................................................................................................................13 5.6.1.1.

RELAYS AND CONTACTORS.......................................................................................................................... 13

5.6.1.2.

RESISTOR UNITS.......................................................................................................................................... 13

5.6.2.

CONTROLGEAR, LOCATION, MOUNTING AND ENCLOSURES.....................................................................................................................14

5.6.3.

T Y P E O F C O N T R OL.............................................................................................................................................................................................................................14

5.6.4.

5.6.3.1.

ENERGISATION ............................................................................................................................................. 14

5.6.3.2.

CAB CONTROL .............................................................................................................................................. 14

5.6.3.3.

FLOOR CONTROL.......................................................................................................................................... 14

5.6.3.4.

CABLE-LESS CONTROLS............................................................................................................................... 14

C O N T R O L O F M E CHANICAL BRAKES.................................................................................................................................................................................15 5.6.4.1.

SECOND BRAKE............................................................................................................................................ 15

5.7 ENVIRONMENT............................................................................................................................................................................. 16 5.7.1.

POTENTIALLY E XPLOSIVE ENVIRONMENTS...............................................................................................................................................................16

5.7.2.

E L E C T R O M A G N E TIC COMPATIBILITY................................................................................................................................................................................16

5.8 SELECTION OF MOTORS ........................................................................................................................................................... 17 5.8.1.

CRITERIA FOR MOTOR SELECTION (IEC 60034-1) .................................................................................................................................................17 5.8.1.1.

REMARKS ON THE SELECTION OF MOTORS................................................................................................. 17

5.8.1.2.

DEGREE OF PROTECTION (IEC 60034-5)........................................................................................................ 18

5.8.1.3.

THERMAL CALCULATION OF THE MOTOR ..................................................................................................... 18

5.8.1.4.

SQUIRREL CAGE MOTORS WITH DIRECT STARTING..................................................................................... 20

Booklet 5 - 3/ 35


5.8.2.

5.8.3.

5.8.4.

5.8.1.5.

POWER CORRECTION IN FUNCTION OF AMBIENT TEMPERATURE AND ALTITUDE........................................ 21

5.8.1.6.

CYCLIC DURATION FACTOR AND NUMBER OF WORKING CYCLES PER HOUR.............................................. 22

MOTORS FOR VE R T I C A L M O T I O N S.....................................................................................................................................................................................23 5.8.2.1.

DETERMINATION OF REQUIRED TORQUE..................................................................................................... 23

5.8.2.2.

CYCLIC DURATION FACTOR AND NUMBER OF CYCLES PER HOUR............................................................... 24

M O T O R S F O R H O R I Z O N T A L M O T I O N S.............................................................................................................................................................................25 5.8.3.1.

DETERMINING THE NECESSARY TORQUE .................................................................................................... 25

5.8.3.2.

CYCLIC DURATION FACTOR AND NUMBER OF CYCLES PER HOUR............................................................... 27

5.8.3.3.

ROTATION..................................................................................................................................................... 27

5.8.3.4.

SPAN VARIATION .......................................................................................................................................... 27

DEVICE SELECT IONS FOR INVERTER DUTY...............................................................................................................................................................28 GENERAL..................................................................................................................................................................... 28 5.8.4.2.

THERMAL DIMENSIONING............................................................................................................................. 29

5.8.4.3.

SELECTION CRITERIA FOR VERTICAL MOTIONS ........................................................................................... 29

5.8.4.4.

SELECTION CRITERIA FOR HORIZONTAL MOTIONS ...................................................................................... 30

5.9 LOAD LIFTING MEANS................................................................................................................................................................ 31 5.9.1.

CURRENT SUPPL Y...............................................................................................................................................................................................................................31

5.9.2.

LOAD HOLDING DEVI C E S..............................................................................................................................................................................................................31 5.9.2.1.

5.10

LIFTING MAGNETS ........................................................................................................................................ 31

5.9.3.

G R A B S.............................................................................................................................................................................................................................................................31

5.9.4.

LOAD TURNING EQUIPMENT......................................................................................................................................................................................................32

MAINTENANCE AND CHECKS........................................................................................................................................... 33 5.10.1.

M A I N T E N A N C E........................................................................................................................................................................................................................................33

5.10.2.

C H E C K S.........................................................................................................................................................................................................................................................33 5.10.2.1. REGULAR CHECKS........................................................................................................................................ 33 5.10.2.2. CHECKS BEFORE COMMISSIONING.............................................................................................................. 34

5.11

AUXILIARY ELECTRICAL EQUIPMENT............................................................................................................................ 34 5.11.1.

LIGHTING......................................................................................................................................................................................................................................................34 5.11.1.1. CABIN ........................................................................................................................................................... 34 5.11.1.2. WORKING AREA LIGHTING............................................................................................................................ 34 5.11.1.3. ACCESS AND MACHINERY CABINET LIGHTING.............................................................................................. 34 5.11.1.4. EMERGENCY LIGHTING................................................................................................................................. 34

5.11.2.

HEATING AN D AIR-CONDITIONING .................................................................................................................................................................................35 5.11.2.1. MACHINERY CABINETS............................................................................................................................... 35 5.11.2.2. CABIN ........................................................................................................................................................... 35

5.11.3.

AUXILIARY CI RCUIT............................................................................................................................................................................................................................35

LIST OF SYMBOLS AND NOTATIONS

See booklet

Booklet 5 - 4/ 35

1


5.1

FOREWORD

The electrical equipment for hoisting appliances should conform to EN 60204-32 and other applicable EN-standards published by CEN and CENELEC. In case the relevant ENstandards do not specify particular requirements, the recommendations given in this document should be followed. Many details specified in the previous versions of this Booklet 5 are now covered by the EN 60204-32.

Booklet 5 - 5/ 35


5.2

POWER SUPPLY

5.2.1.

GENERAL

This document deals with a.c. low-voltage systems up to 1000 V. The power supply system should comply with EN 60204-32 clause 4.3. Voltage variation at the point of supply (4.3 of EN 60204-32) should not exceed –5% ... +5%. of the rated voltage under normal operating conditions. In applications with very long cabling distances it may be necessary to limit the voltage variation further. The allowed per cent voltage drop in the different parts of the power supply need to considered case by case. It is recommended that the details of the power supply system are agreed between the supplier and the purchaser e.g. by using Annex A of EN 60204-32. The type of the power supply grounding (see IEC 60364-1) may have significant effects on the requirements of the crane electrification. The type should always be agreed between the purchaser and supplier. In cases where the power supply type is TN with neutral directly earthed, equipment will usually perform correctly. If the power supply type is TN with other than neutral earthing, TT or IT some restrictions may need to be considered, e.g. devices designed for neutral earthed TN supply may incorporate components for EMC filtering, which might − not withstand the phase to earth voltage of corner earthed systems or IT systems during ground fault − introduce unacceptably high leakage currents to earth. 5.2.2.

CUT-OUT AND SAFETY DEVICES

The requirements for crane-supply-switch and crane disconnector are given in clause 5 of EN60204-32. Additional requirements may apply, if these devices are used in emergency operations (see 9.2.5.4 of EN 60204-32). 5.2.3.

CONDUCTOR BARS, CABLE REELS AND FLEXIBLE CABLES

The following clauses in EN 60204-32 set the basic requirements: − 13.7. Flexible cables − 13.8. Collector wires, collector bars and slip-ring assemblies − 14.4.3. Connections to cranes and between moving parts of the crane According to 13.8.2. of EN 60204-32 the continuity of any protective bonding connection with sliding contacts shall be ensured e.g. by duplication of the current collector. Additionally, when sliding contacts are used to supply power to electronic drives, double collectors should be used also in phase collectors to avoid noise and hardware failures which may occur if a contact cuts off momentarily.

Booklet 5 - 6/ 35


5.3

INSTALLATION OF CABLES AND CONDUCTORS

General requirements for cables and wiring practices are given in clauses 13 and 14 of EN 60204-32.

5.3.1.

CALCULATION OF CROSS-SECTION OF CONDUCTORS

These requirements apply both to the power supply to the crane and also to wiring within the crane. The voltage drop must be considered, paying attention to the fluctuation and the voltage drop within the power supply. For very long supply lines, not only the resistive but also the inductive part of the supply impedance need to be taken into account. The cross-section of the conductors should be determined by taking into account the mechanical strength required and the electrical load to be carried. 5.3.1.1.

CALCULATION OF THE CROSS-SECTION IN RELATION TO THE ADMISSIBLE VOLTAGE DROP

When calculating the voltage drop, the most unfavourable position of the hoisting appliance in relation to the supply point must be considered. When calculating the admissible voltage drop on a supply line used by several hoisting appliances, the start-up (ID) and rated (IN) currents of the motors operating simultaneously must be taken into account. Notes for the calculation: − In this clause, the rated current (IN) should be considered not necessarily to mean the nameplate current of the motor but the current drawn by the motor at full rated load. − For squirrel-cage rotor motors ID (start-up current), refer to the manufacturer's catalogue. In case the motor is controlled by an electronic drive (soft-starter, frequency converter etc), the maximum current during any phase of operation should be considered as start-up current, although the highest current does not necessarily occur when starting the motion. With direct starting the ID is typically 5 to 10 times IN. With electronic drives the start-up current depends on the converter type and on its’ adjustments; with frequency converters the ID is typically below 2 times IN. − For slip-ring rotor motors, consider ID to be approx. 2 * IN. − For drive with n motors in parallel, apply n * ID or n * IN. − In case two or more hoisting appliances are working together, they should be considered as one appliance by using the sum current ( ID or IN ) of each joint motion. − In case the power supply also feeds other (continuos) loads such as lighting, hydraulic pumps, lifting magnets or other cranes, the current drawn by these devices need to be taken into account.

Booklet 5 - 7/ 35


For a three phase power supply, the required minimum cross-section (S) of copper conductors can be calculated with the formula: S

= √3 * l * Itot * cos ϕ / ( ∆u * κ ) l

= Effective length of the line [m]

Itot

= Sum of the above calculated (ID and IN) currents [A]

∆u

= Admissible voltage drop [V]

κ

= Electric conductivity [Ω * mm2 * m-1]-1

cos ϕ

5.3.1.2.

[mm2], where

= Power factor.

CALCULATION OF THE MINIMUM CROSS-SECTION IN RELATION TO THE THERMAL CAPACITY OF THE CONDUCTORS

When calculating the cross-section for the conductor bar, which supplies several hoisting appliances, the actual simultaneous operation of the drive motors must be taken into account. Notes for the calculation: − In this clause, the rated current (IN) should be considered not necessarily to mean the nameplate current of the motor but the current drawn by the motor at full rated load. − In case n>1 motors are driven in parallel, consider : IN = n * IN’ (IN’ = nominal current for one motor). − In case two or more hoisting appliances are working together, they should be handled as one by using the sum current of each joint motion. − In case the power supply also feeds other (continous) loads such as lighting, hydraulic pumps, lifting magnets or other cranes, the current drawn by these devices need to be taken into account. The maximum allowed conductor temperature shall not be exceeded during normal operation. Conductor cross-section should be selected according to manufacturers’ specifications or according to IEC60364-5-52. The tables of IEC 60364-5-52 contain a number of parameters, the most important of which are: - conductor type - installation method - correction factors (ambient temperatures, bunching of cables,...); for cable drums see also EN 60204-32 clause 13.7.3, The current limits given in IEC 60364-5-52 are for continuous current. If the conductor manufacturer does not provide more detailed guidance for intermittent use consisting of −

“active periods (Ta)” and

− “idle periods (Ti)”, the information provided in 5.3.1.2.1 or in 5.3.1.2.2 can be used.

Booklet 5 - 8/ 35


5.3.1.2.1.

Overloading capacity of conductors with 10 minute cycle time

Table 5. 3.1.2-A

The allowed increase of the current depends on the duration of the load cycle compared to the thermal time constant of the conductor. Table 5.3.1.2-a defines the overloading capacity ( fED ) of conductors in intermittent duty when duration of the operating cycle is 10 minutes. These fED values have been calculated using the formulas in 5.3.1.2.2. 5.3.1.2.2.

Overloading capacity of conductors with any cycle time

If the duration of the operating cycle is very short compared with the time constant of the cable (T) it may be possible to apply higher factors than shown in 5.3.1.2.1. The allowed overloading factor fED for the particular duty cycle can be expressed as a function of Ta, Ti, and T as shown below.

fED for a 10 minute cycle

cross section

Ta / (Ta + Ti)

mm2

0,6

0,4

0,25

0,15

1,5

1,044

1,120

1,265

1,505

2,5

1,058

1,150

1,315

1,580

4

1,075

1,183

1,369

1,660

6

1,092

1,215

1,421

1,737

10

1,116

1,260

1,493

1,842

16

1,139

1,303

1,561

1,942

25

1,161

1,344

1,626

2,037

35

1,177

1,373

1,673

2,105

50

1,193

1,403

1,719

2,173

70

1,207

1,429

1,760

2,231

95

1,219

1,450

1,793

2,280

120

1,227

1,464

1,816

2,314

150

1,234

1,477

1,836

2,343

185

1,240

1,488

1,854

2,369

240

1,247

1,501

1,874

2,397

300

1,252

1,510

1,888

2,419

The thermal time constant T to be used in the f ED calculation is given in Table 5.3.1.2-b.

f ED =

1-e

1- e

5.3.2.

Table 5. 3.1.2-B

T a+ T i  −    T 

2

1,5

2,5

4

6

10

16

25

35

T/min

2,7

3,1

3,6

4,2

5,2

6,4

7,9

9,4

mm 2

50

70

95

120

150

185

240

300

mm

 Ta  −   T 

T/min 11,3 13,6 16,1 18,4 21,0 23,7 27,7 31,8

INSTALLATION CONDITIONS

Type of protection for connection and distribution equipment must be suitable for surrounding conditions following the guidelines given in EN 60204-32 clause 12.3. The connections and linking terminals should be placed in cabinets or boxes. Plug-in arrangements whose accidental connection could be dangerous should be clearly separated unless their design precludes this risk. In order to ensure continual mechanical protection, the protective covering of the cables and conductors should enter housings through packing glands or such similar devices. The wires or conductors belonging to electrical circuits with different rated voltages may be arranged within a single enclosure or may form part of the same cable provided that these wires or conductors are insulated against the highest rated voltage. Conductors having single insulation can only be installed in conduits or trunking whose ends are fitted with adequate protection. Non-sheathed conductors and cables, which are fixed to parts of the framework should be protected, if necessary, against any mechanical wear and tear. Booklet 5 - 9/ 35


5.4 5.4.1.

ELECTRICAL PROTECTIVE AND SAFETY EQUIPMENT SAFEGUARDING MOTORS AGAINST OVERHEATING

Motor overload protection shall fullfill the requirements of 7.3 of EN 60204-32. Normal thermal relays or protection methods based on calculating methods may not perform properly in short time duty (S2 of IEC 60034-1) and in intermittent duty, (S3...S8 of IEC 60034-1) because e.g. − their time constant may not be comparable to that of the load cycle − initial temperature during power up is not known − when self-ventilated motors are operated at low speeds the cooling is not efficient Therefore protection methods based on actual temperature measurement of the motor are preferred. The overload protection can also be implemented by electronic devices, which may be either separate devices or integrated in the control or drive unit.

5.4.2.

SAFEGUARDING WIRING

The cross-section of a conductor should be determined according to the current intensity to which it is subject during both normal running of the motor and starting-up or electrical braking, see clause 5.3.2. Whether the loads (motors) are overloadprotected or not, all wires should be safeguarded according to 7.2 of EN 60204-32 against any overcurrent, which could result from a shortcircuit or faulty insulation. The protective device shall be rated for the anticipated short-circuit currents.

5.4.3.

SAFEGUARDING AGAINST ABSENCE OR INVERSION OF PHASES

Where an incorrect phase sequence of the supply voltage can cause a hazardous condition or damage to the hoisting machine, protection should be provided according to 7.8 of EN 60204-32. If the absence of phases may occasion a danger, the appropriate safety measures must be taken.

5.4.4.

ACTION OF SAFETY DEVICES

When several motors drive the same motion, the action of a safety device should stop all of the motors for this movement. After a safety device has been activated, it should be possible for the equipment to be started up again only manually.

Booklet 5 - 10/ 35


5.4.5.

PROTECTION AGAINST THE EFFECTS OF LIGHTNING

For very tall pieces of hoisting equipment which are erected in particularly exposed locations, the effects of lightning must be considered − on pieces of vulnerable structure (for example : jib support cable) − on anti-friction bearings or runners which form a link between large parts of the frame (for example : slewing ring, travel runner). When necessary, safeguarding against the effects of lightning should be carried out e.g. by following IEC 61024-1. For the safety of personnel, it is recommended that the runner rails for the lifting equipment are earthed.

Booklet 5 - 11/ 35


5.5

LIMITING AND INDICATING DEVICES

In addition to the requirements given later in this clause, the limiting and indicating devices should comply − with 7.7 of FEM 1.001 Booklet 7, and − with EN 12077-2, and − with the EN-standards for the particular crane type, and − in case there is no published EN-standards for the the particular crane type, the requirements given in ISO 10245 for the particular crane type should be followed.

5.5.1.

GENERAL REQUIREMENTS

A limit switch should bring about the arrest of motion by opening the electric circuit and keeping it open as long as safety conditions are not restored. If it is unavoidable to by-pass a safety device, this operation should only be able to be effected with the aid of a device which, when no longer actuated, automatically re-inserts the safety device. The provisions of 9.2.4 of EN 60204-32 should be followed, and if necessary a permanent warning signal should be given. In general, the safety functions should comply with safety category 1 if hardwired or safety category 2 of EN 954-1 if not hardwired. 5.5.2.

MOTOR OVERSPEED PROTECTION

All electronically controlled hoisting motions shall be equipped with overspeed protection. The overspeed protection shall prevent − uncontrolled and unintentional motions and − all parts of the mechanism from reaching its mechanical limit speed. The trigger limit of the overspeed protection shall be set so that the mechanical brake is capable to stop the motion safely in all conditions. In general, the trigger limit should not exceed 1,2 times the specified speed at nominal load. When utilising field weakening with a reduced load, the trigger limit can be adjusted to a higher value, in general not exceeding 1,2 times the specified speed for that load. The emergency stopping with any possible speed and load combination shall not cause any hazard including the cases where the drive is intended to operate above the nominal speed. NOTE : Topics to be considered include allowed speed/load combinations, time delays in the system (particularly brake delays), mechanical speed limit of the machinery, reliability of load measurement/estimation.

Booklet 5 - 12/ 35


5.6

CONTROLS

5.6.1.

COMPONENTS

5.6.1.1.

RELAYS AND CONTACTORS

Relays and contactors must comply with the requirements of EN6024-32 clause 4.2.2. In case the crane will be used at an altitude in excess of 1000 m, this shall be considered by selecting the contactors and relays. Reversing contactors should be of the electrically or mechanically interlocking type. 5.6.1.2. 5.6.1.2.1.

RESISTOR UNITS General

Resistor units should be accommodated in suitable enclosures according to EN 60204-32 cl. 12.3. The temperature limit of the resistance material is defined in 7.2.2.8 of IEC 60947-1. Temperature rise of the touchable surfaces of the resistor enclosures shall comply with 7.2.2.2 of IEC 60947-1. When designing the resistor units, the equivalent torque, cyclic duration factor and switching rate have to be considered. Cooling fans for resistors should be used only if proper monitoring is arranged (air flow detection and/or temperature measurement). In dirty environments fans should not be used to ensure reliability. 5.6.1.2.2.

Braking resistors for inverter drives

The braking resistor of a hoisting drive shall be capable of absorbing the generative energy when lowering the maximum load at maximum speed. For horizontal motions, the braking resistor shall be capable of absorbing the generative energy during deceleration of the motion also taking into account the possibility of a swinging load and the wind push (the load and load attachment included). The braking resistor shall be thermally capable to absorb the generative energy during successive drive cycles of the application. The failure of the braking resistor shall neither lead to the loss of stopping capability nor to any uncontrolled acceleration of the motion. The braking resistors and their enclosure may heat up to hazardous temperatures. Protection or warning of the hazard shall be provided depending on the application. The requirements of clause 5.1.4.1 of prEN 13135-2 shall be fulfilled. NOTE : The braking resistor may also be common for several hoist and travel drives. In such cases the dimensioning of the braking resistor shall be made accordingly. NOTE : The braking resistor is usually connected to the DC-bus of the inverter and due to this it becomes live always when electric power is supplied to the inverter – also during not running the motor.

Booklet 5 - 13/ 35


5.6.2.

CONTROLGEAR, LOCATION, MOUNTING AND ENCLOSURES

Switching devices, controlgear and panels housing electrical equipment may be enclosed: − in cabinets or housings, − in special enclosed spaces, − in the supporting structure (principally the crane girder) of the hoisting appliance. These enclosures and the equipment installed should comply with cl. 12 of EN 60204-32.

5.6.3.

TYPE OF CONTROL

5.6.3.1.

ENERGISATION

The lifting appliance can only be energised when all the control devices are in the off position. This off position can be determined either by a checking circuit or by using hold-torun controls. As required in EN 60204-32 clause 10.7.1, a device for emergency stop or for emergency switching off shall be located at each operator control station.

5.6.3.2.

CAB CONTROL

The controls should be so arranged that the operator has an adequate view of the crane’s working area. The control for hoisting appliances should preferably be arranged on the right-hand side of the operator’s seat.

5.6.3.3.

FLOOR CONTROL

Push-buttons or other switching devices, which automatically return to their “off” position as soon as they are released, should be provided for the control of all motions by pendant control units. Housings of pendant control units should preferably be of fully insulating material or of material with protective insulaion. Metal parts accessible from the outside, which pass through the insulation, should be separately earthed. The surface of the housing must be a vivid colour. For indoor operation, the degree of protection should be at least IP43, and for outdoor operation at least IP55 Pendant control units should be suspended with a strain relief arrangement.

5.6.3.4.

CABLE-LESS CONTROLS

The requirements of 9.2.7 of EN 60204-32 and Annex C of prEN 13557 apply to cable-less controls incl. radio and infrared controls. The transmitter must have a minimum protection class IP 43 for indoor use and IP 55 for outdoor use. Booklet 5 - 14/ 35


5.6.4.

CONTROL OF MECHANICAL BRAKES

The requirements of EN 60204-32 clauses 9.3.4, 15.6, and 15.7 should be fulfilled. Measures should be taken to ensure that no unintentional movements occur when starting.

5.6.4.1.

SECOND BRAKE

Cranes, which require particular safety, e.g. in steel works or with dangerous or melted loads, should be provided with a second brake. The operation of the second brake shall be arranged according to the design of the drive. It is recommended that under normal operating conditions, the second brake is always be applied on stopping, after the motion has been brought to a halt by the main brake. In some applications – for example if waiting for the releasing of the second brake would cause unacceptable time delay at each starting - it may be necessary to apply the second brake only when the crane switch (main contactor) is de-energized. In the event of an emergency stop, the second brake should be applied immediately.

Booklet 5 - 15/ 35


5.7

ENVIRONMENT

The electrical equipment shall be suitable for use in the physical environment and operating conditions specified in 4.4.2 to 4.4.8 of EN 60204-32. When the physical environment or the operating conditions are outside those specified, an agreement may be needed between the supplier and the user (see annex A of EN 60204-32).

5.7.1.

POTENTIALLY EXPLOSIVE ENVIRONMENTS

In potentially explosive atmospheres, the electrical equipment including motors shall comply with relevant additional requirements (e.g. EN 50014...50020).

5.7.2.

ELECTROMAGNETIC COMPATIBILITY

The requirements of 4.4.2 of EN 60204-32 apply. The manufacturer should perform an EMC analysis to define what essential safety and/or protection requirements apply to his apparatus and how to conform to them. The manufacturer should apply − harmonised specific product family standards and − harmonised generic standards for the residential, commercial, and light industrial environment, and industrial environment to bring the hoisting machine into compliance with the EMC Directive. When using only CE marked apparatus and components (complying with the EMC Directive) and following strictly the instructions and limitations of use of the manufacturer of these products, the finished hoisting machine could be considered to comply with the EMC Directive and no further verification is then needed. The manufacturer shall provide clear instructions for assembly/installation/operation/maintenance in the instructions for use to ensure that the compliance with the Directive can be reached in all expected electromagnetic environments. In cases where the manufacturer does not restrict himself to only using CE marked apparatus, a thorough EMC analysis and verification of compliance is needed.

Booklet 5 - 16/ 35


5.8

SELECTION OF MOTORS

5.8.1.

CRITERIA FOR MOTOR SELECTION (IEC 60034-1)

When selecting motors for an application, at least the following details need to be considered: - required powers-the thermal power is also included in these required powers, - maximum rated torque and maximum acceleration torque, - cyclic duration factor, - number of cycles/hour, - type of control (type of braking), - speed regulation, - type of power feed, - degree of protection, (environment conditions), - ambient temperature, - altitude. The motor has to comply with the following two dimensioning requirements: - the thermal calculation according to clause 5.8.1.3. - the required maximum torque: • for hoisting mechanisms according to clause 5.8.2.1. • for horizontal motions according to clause 5.8.3.1. NOTE: Additional or different criteria may be needed depending on the driving system. NOTE: Selection of motors for inverter drives is defined in clause 5.8.4, which also covers the dimensioning of inverter drives. If the required torque diagrams, in order to define the mean equivalent torque (cl. 5.8.1.3.1.) are not available, these can be assessed respectively with the help of tables T 5.8.2.2a. and T 5.8.3.2a 5.8.1.1.

REMARKS ON THE SELECTION OF MOTORS

The selection of the motor should be agreed with the manufacturer in taking into account the torque and powers calculated in the following clauses and the real operating conditions of the motor. In the event of electronic power control, the definition of the motors has to be made in cooperation with the manufacturer, taking into account the cooling system and the speed range. In cases where two or more mechanisms drive the same motion, the following shall be considered: − both static and dynamic synchronisation of the motions according to the needs of the application − necessary interlocks between the mechanisms to ensure safe operation − both static and transitional asymmetrical loading of the mechanisms and consequently needed adequate dimensioning of motors and other drive components Booklet 5 - 17/ 35


5.8.1.2.

DEGREE OF PROTECTION (IEC 60034-5)

The degree of protection for all motors shall be according to EN 60204-32 clause 15.2. In case of water condensation risk, care should be taken that the water condensation drain holes remain open. 5.8.1.3.

THERMAL CALCULATION OF THE MOTOR

5.8.1.3.1.

Mean equivalent torque

In order to carry out the thermal calculation, the mean equivalent torque must be determined as a function of the required torque during the working cycles, by the formula:

Mmed =

M12 * t1 + M22 * t 2 + ... + Mn 2 * tn t 1 + t 2 + ... + t n

Where : t1, t2, ...,tn

are the durations of the time periods during which the different torque values are produced; periods of rest are not taken into account.

M1, M2 , .... Mn are the calculated torque values taking into account all the inertia forces including the one of the rotor mass of the motor. In case of variable loads at least 10 successive working cycles must be taken into account (see definition 2.1.2.2.). Diagram 5.8.1.3.1. shows an example of the torque for 2 different operating cycles. 5.8.1.3.2.

Mean equivalent power

Starting from the mean equivalent torque, the mean equivalent power P med [kW] is defined by the formula: Pmed = (M med * nm / 9 550 ) where: Mmed

= mean equivalent torque [Nm]

nm

= speed of motor [1/min]

If the motor is rated for S3-duty and the rating corresponds to the actual use in the particular application, then the motor can be selected according to the calculated mean equivalent power. For S1-rated squirrel cage motors, the thermal dimensioning shall be carried out according to the method described in clause 5.8.1.4. (NOTE: applies only for direct starting motors). For the motor selection, the mean equivalent power P med should be corrected as a function of altitude if it exceeds 1000 m and the ambient temperature if it deviates from 40 °C (See 5.8.1.5.).

Booklet 5 - 18/ 35


Torque raising with load

lowering with load

M1

raising without load M6

lowering without load

M7

M12

M2 M5

hoisting motion

M8 t1

t2

t3

t4

t5

t6

t8

tr1 M3

tr2

t9

t10

t7

tr3

time

M11 t11

t12

tr4

M4 M9

M10

M1

M2 M6

with load and resisting wind horizontal motion

t1

t2

t3

t4

tr1

t5

without load with driving wind

t6

tr2

tr3

M5

CYCLE 1

M3

raising with partial load

lowering with partial load

M4

raising without load

lowering without load

M6

M1

M12

M2 M5

hoisting motion

M8 t1

t2

t3

t4

t5

t6

t8

tr1 M3

tr2

t9

t7

t10 tr3

M11 t11

time t12

tr4

M4 M9

M1

M10

M2 M6

with partial load and resisting wind horizontal motion

t1

t2

t3

t4

tr1

t5

tr2

without load with driving wind

t6 tr3

M5

CYCLE 2

M3

M4

Diagram 5.8.1.3.1.

Typical torques for 2 different operating cycles: Hoisting motion tr: rest time M1,M4,M7,M10, M2,M8, M3,M6,M9,M12 M5,M11,

starting torque hoisting torque raising braking torque hoisting torque lowering

Horizontal motion tr: rest time M1,M4 starting torque M2 working torque with wind M3, M6 braking torque M5 torque without load with wind

Booklet 5 - 19/ 35


5.8.1.4.

SQUIRREL CAGE MOTORS WITH DIRECT STARTING

The following inequality has to be fulfilled for the thermal dimensioning of squirrel cage motors: Ck ( 1 - η N ) * P N * T > ( 1 - ηmcy ) * P mcy * tN + ( P N * t E * ID / IN ) - ( J * nmcy2 * 10-3 / 182 ) NOTE: Subscript “cy” refers to cycle. loss energy of the motor working at its rated power (S1) during a time T loss energy of the motor during the time t N ( 1 - ηmcy ) * P mcy * t N (constant speed) in a cycle 2 -3 ( P N * tE * ID / IN ) - ( J * n mcy * 10 / loss energy of the motor during the starting 182 ) and braking phases Ck correction factor linked to the type of motor nominal power [kW] of the motor in PN continuous (S1) duty ηN efficiency of the motor at P N Pmcy. Mmcy * nmcy / 9 550 [kW] nmcy speed of motor [1/min] for power P mcy mean resisting torque [Nm] calculated in the same manner as M med (see clause 5.8.1.3.1), Mmcy but not including the starting and braking phases. ηmcy efficiency of the motor at power P mcy total time of cycle [s], T = t N + t E + tr cyclic duration factor (see clause 5.8.1.6.) ED = 100 * ( tN + t E ) / T operating time [s] at constant speed during tN one cycle. equivalent time [s] of starting and braking during one cycle, tE = ( π / 30 ) * nmcy * J / Macc / ( dccy + 0,5 * d icy + 3 * fcy ) tr total rest (idle) time [s] during one cycle. total inertia of masses in motion referred to J the motor shaft [kgm2]. the number of complete starts during one dccy cycle dicy the number of impulses during one cycle the number of electrical brakings during one fcy cycle the mean accelerating torque [Nm], Macc = M Dmcy - M mcy MDmcy mean starting torque of motor [Nm] ( 1 - ηN ) * P N * T

Booklet 5 - 20/ 35


The following data has to be indicated by the motor manufacturer: nominal power [kW] of motor in continuous (S1) duty efficiency for 1/4 P N ... 5/4 P N powers moment of inertia of motor [kgm2] speed of motor at 1/4 P N ... 5/4 P N [1/min] mean starting torque of motor in [Nm] ratio between the starting current and the current at P N correction factor linked to the type of motor

PN η1/4 ... 5/4 JM n1/4 ... 5/4 MDmcy ID/IN Ck

In case the C k factor is not mentioned in the manufacturer's catalogue, C k shall be taken equal to 1 for motors of polarity equal or above 4.

5.8.1.5.

POWER CORRECTION IN FUNCTION OF AMBIENT TEMPERATURE AND ALTITUDE

These corrections are depending from the type of motor, the cooling method and the insulation class. The precise calculation can only be made by the motor manufacturer in supplying them with the following indications: - P med without correction - value of ambient temperature - altitude The thermal dimensioning can be based on the formulas below and on the values of k indicated in Diagram 5.8.1.5: P’med = P med / k or P'mcy = P mcy / k

(for squirrel cage motors)

P'mcy or P'med = required nominal power of motor as function of altitude and ambient temperature.

Booklet 5 - 21/ 35


ALTITUDE

Ambient Temperature

Diagram 5.8.1.5 = Correction factor k as function of ambient temperature and altitude Note 1: The k > 1 coefficient values are to be applied only by agreement between the motor manufacturer and the hoisting appliances manufacturer. Note 2: The ambient temperature shall be indicated above an altitude of 1000 m.

5.8.1.6.

CYCLIC DURATION FACTOR AND NUMBER OF WORKING CYCLES PER HOUR

The cyclic duration factor is given by the following formula: ED = Operating time / (Operating time + idle time ) * 100

(%)

The operating time and the number of operations per hour of the motors as well as the number of working cycles of the crane, are an important base for the thermal definition of the motors and which should be agreed between the user and the manufacturer of the crane. In case it is not possible to give these indications in a precise manner, it should be referred to tables T 5.8.2.2 a and T 5.8.3.2 a.

Booklet 5 - 22/ 35


5.8.2.

MOTORS FOR VERTICAL MOTIONS

5.8.2.1.

DETERMINATION OF REQUIRED TORQUE

For a hoisting motor, the required power to raise the maximum nominal load (P Nmax) is defined in kW in taking account of the configuration of the transmission and of the reeving according to the following formula: PNmax = L * V L * 10-3 * η Where: L = maximum nominal permissible lifting force [N] VL = lifting speed [m/s] η = efficiency of machinery The required torque to raise the maximum nominal load is: MNmax = P Nmax * 9 550 / n, where nm = rotating speed of the motor [1/min]. In order to be able to develop the necessary torque for acceleration, for lifting the test load or for compensating for variations in the mains voltage and frequency, the torque developed by the motor must satisfy the following minimum condition: -

For squirrel cage motors with direct starting: Mmin / MNmax ≥ 1,6, where Mmin is the minimum torque of the motor during starting.

-

For slip ring motors: Mmax / M Nmax ≥ 1,9 with M max being the maximum torque of the motor.

-

For all types of motors which are fed by voltages and /or variable frequencies: Mmax / M Nmax ≥ 1,4

The mechanical braking torque at the motor shaft (MF) should at least be equal to: MF ≥ 2 * M Nmax * η2

-

Static

-

Dynamic MF ≥ 1,6 * M Nmax * η2

Definition of the braking torque: -

Static: is the required minimum torque to prevent the SWL (safe working load) rotating the machinery.

-

Dynamic: is the braking torque produced by the brake during the whole duration of a braking cycle.

In case electrical braking is applied, it shall be capable to slow down the load in complete safety.

Booklet 5 - 23/ 35


5.8.2.2.

CYCLIC DURATION FACTOR AND NUMBER OF CYCLES PER HOUR

In the case where no precise indications are given, the values mentioned in Table T 5.8.2.2.a can be chosen. Table T.5.8.2.2.a. Indications for the number of cycles per hour and the cyclic duration factor for the vertical motions Type of appliance

Reference 1 2 3

4

Hand-operated (=not motorised) appliances Erection cranes Erection and dismantling cranes for power stations, machine shops, etc. Stocking and reclaiming transporters Stocking and reclaiming transporters

6 7

Workshop cranes Overhead travelling cranes, pigbreaking cranes Scrapyard cranes Ladle cranes Soaking-pit cranes Stripper cranes, open-hearth furnace-charging cranes Forge cranes Bridge cranes for unloading, bridge cranes for containers Other bridge cranes (with crab and/or slewing jib crane)

11 12-a 12-b

13

14

Numbe r of concerning cycles nature of use per (1) hour

Designation

5

8 9 10

Particulars

Lifting

2-25 2-15

25-40 15-40

Hook duty

20-60

40

Grab or magnet

25-80

60-100

10-50 40-120

25-40 40-100

3-10 30-60 30 10 6 20-60

60 40-60 40-60 60 60 40 40-60

b – Hook duty

20-60

40-60

Grab or magnet

20-80

40-100

20-50

60 40

Grab or magnet

a – Hook or spreader duty

Bridge cranes for unloading, bridge cranes (with crab and/or slewing jib crane) Drydock cranes, shipyard jib cranes, jib cranes for dismantling

Type of mechanism ED%

Hook duty

Booklet 5 - 24/ 35

Derrickin Derricki g hinged ng boom boom

S2 (2) 15-30 min S2 (2) 15-30 min

S2 (2) 15-30 min S2 (2) 15-30 min S2 (2) 15-30 min 40


18 19

Dockside cranes (slewing, on gantry, etc.), floating cranes and pontoon derricks Dockside cranes (slewing, on gantry, etc.), floating cranes and pontoon derricks Floating cranes and pontoon derricks for very heavy loads (usually greater than 100 t) Deck cranes Deck cranes

20 21

Tower cranes Derricks

20 10

22

Railway cranes allowed to run in train

10

15

16

17

Hook duty

Grab or magnet


40

60

40-60

20 25-60

40 60-100

40-60

2-10

S1 (2) or S2 30 min 40 60

S2 (2) 15-30 min 40 60

40-60 S1 (2) or S2 30 min 40

25-40 S1 or S2 (2) 30 min

30-60 30-80

1) This column comprises only some indicatory typical cases of utilisation 2) it is recommended for S1 and S2 to refer to the definition IEC 60034-1

5.8.3.

MOTORS FOR HORIZONTAL MOTIONS

In order to select travel motors correctly, all the necessary torque (or power) values must be considered, taking into account the starting time, the number of starting cycles per hour and the cyclic duration factor. The maximum transmissible torque of the travel motors is limited by the adhesion of the driven travel wheels on their tracks. 5.8.3.1.

DETERMINING THE NECESSARY TORQUE

- Speed maintaining torque To determine the torque necessary for maintaining the speed, account has to be taken of the sum of forces (w) resisting to travel resulting from the dead-weight, the load and operating conditions such as: - deformation of the running surface, - friction of the wheels on straight sections and in curves, - wind force, - gradients in the track, - necessary traction of power supply cable. - Acceleration torque (running up to speed) The acceleration torque shall take into account the sum of the acceleration forces of the mass of lifted load and of the other masses put into motion. The recommended acceleration values are given in Table T 2.2.3.1.1 (booklet 2).

Booklet 5 - 25/ 35


The travel motors must deliver the necessary torque in the following operating conditions: -

Case I for cranes not exposed to wind

-

Case II for cranes exposed to wind The necessary torque can be calculated by the following formulae (see diagram 5.8.1.3.1) Case I M1, ... Mn = [ a * ( m + mL ) + w0 ]* v * 60 / ( 2 * π * nm * η ) Case II The largest of the values from the results of the following formula shall be taken into account: M1, ... Mn = [ a * ( m + mL ) + w8 ]* v * 60 / ( 2 * π * nm * η ) and M1, ... Mn = w25 * v * 60 / ( 2 * π * nm * η ) where: a m

mL m0

mrot

acceleration [m/s2] (at constant speed a = 0) = m0 + mrot * η , equivalent mass [kg] of all parts put into motion, excluding the load, which is supposed to be concentrated at the suspension point of the load. mass of lifted load [kg] mass [kg] of the whole of the elements, excluding the load, undergoing the same horizontal motion as the suspension point of the load. = Σ( J * nχ2 / v2 ) / 91,2 , equivalent mass [kg] of the inertia of rotating parts reduced to linear motion, where: nχ speed of rotating masses [1/min]

moment of inertia of all rotating masses [kgm2] w0, w8, w25 total travel resistance [N] (w can also become negative in some cases) w0 at zero wind w8 at a wind of 80 N/m2 w25 at a wind of 250 N/m2 v travel speed [m/s] nm rotation speed of motors [1/min] η overall efficiency of mechanism J

The motor shall be selected based on the highest of the calculated torque values (M1 , ... Mn ) in case I and II. For slip ring motors used for the horizontal motions, the starting resistances shall be so defined that the minimum torque delivered by the motor is never less than 1,2 times the torque required to maintain the travel speed.

Booklet 5 - 26/ 35


5.8.3.2.

CYCLIC DURATION FACTOR AND NUMBER OF CYCLES PER HOUR

In the case where no precise indications are given, the values mentioned in Table T 5.8.3.2.a. can be chosen. 5.8.3.3.

ROTATION

The calculation is carried out in an analogous fashion to clause 5.8.3.1, angular speeds being substituted for the linear speeds. 5.8.3.4.

SPAN VARIATION

If the span variation in the case of luffing jibs, leads to an elevation or to a lowering of the centre of gravity of the masses put into motion, the calculation can be carried out in an analogous fashion to clause 5.8.3 in inserting into the factor (w) the forces required to the vertical displacement of the centre of gravity. Table T. 5.8.3.2.a Indications for the number of cycles per hour and the cyclic duration factor for the horizontal motions Type of appliance Particulars Numbe Type of mechanism ED% r Refeconcerning of rence nature of use cycles Rotatio Designation Crab Travel (1) per n hour 1 Hand-operated appliances 2 Erection cranes 2-25 25 25-40 25-40 Erection and dismantling 3 2-15 25 25 cranes for power stations, machine shops, etc. 4 Stocking and reclaiming Hook duty 20-60 15-40 40-60 25-40 transporters Stocking and reclaiming Grab or 5 25-60 40 60 15-40 transporters magnet 6 Workshop cranes 10-50 25-40 25-40 7 Overhead travelling cranes, Grab or 40-120 40-60 60-100 pigbreaking cranes, scrapyard magnet cranes 8 Ladle cranes 3-10 40-60 40-60 9 Soaking-pit cranes 30-60 40 40-60 40-60 10 Stripper cranes, open-hearth 30 40 60 furnace-charging cranes 10 40 40 11 Forge cranes 6 100 25 25 12-a Bridge cranes for unloading, a – Hook or 20-60 15-40 40-60 15-40 bridge cranes for containers spreader duty Other bridge cranes (with crab 12-b b – Hook duty and/or slewing jib crane) 20-60 25-40 40-60 25-40

Booklet 5 - 27/ 35


Bridge cranes for unloading, bridge cranes (with crab and/or slewing jib crane) Drydock cranes, shipyard jib cranes, jib cranes for dismantling Dockside cranes (slewing, on gantry, etc.), floating cranes and pontoon derricks Dockside cranes (slewing, on gantry, etc.), floating cranes and pontoon derricks Floating cranes and pontoon derricks for very heavy loads (usually greater than 100 t) Deck cranes Deck cranes

13

14

15

16

17

18 19 20 21 22

Grab or magnet

20-80

40

40-100

15-60

Hook duty

20-50

25

40

25-40

Hook duty

40 25-40

40

15-25

Grab or magnet


Tower cranes Derricks Railway cranes allowed to run in train

20 25-60

40-60

2-10

15-40

30-60 30-80

40 60

20 10 10

40-60 25 25

25-40

25

15-40

1) This column comprises only some indicatory typical cases of utilisation 5.8.4.

DEVICE SELECTIONS FOR INVERTER DUTY

5.8.4.1.

GENERAL

When feeding asynchronous motor by inverters, the motor speed is always close to the no-load point (= between the maximum torque point and zero-torque point) on the torque curves shown in the figure 5.8.4.1. The speed adjustment as realized by changing the supply frequency to the motor. In the field-weakening region above the nominal motor speed nN, the motor’s maximum torque (“pull-out torque” Mpo) decreases inversely proportional to the square of the speed, which sets limitations to the use of field-weakening. The figure also shows the motor current at rated speed (thick line), which increases very rapidly if the motor is loaded excessively.

5

I IN 4

M MN

3

2

1

n nN 0 0.0

0.2

0.4

0.6

0.8

1.0

1.2

1.4

1.6

1.8

2.0

2.2

2.4

NOTE 1: If the motor is loaded up to the pullFigure 5.8.3.4 out torque or too close to it, there is a high risk of getting into an unstable situation. The required minimum torque margins are defined in 5.8.4.3 and 5.8.4.4. An even higher margin may be needed depending on

Booklet 5 - 28/ 35

2.6


• the characteristic of the inverter (e.g. vector control or U/f control, with or without speed feedback), and • the foreseen environment (e.g. variation of supply voltage, load swaying, transient phenomena). NOTE 2: The curves shown in the figure are indicative only, and actual values for a particular motor type need to be asked from the motor supplier. NOTE 3:

The advice given in 5.8.1.1 shall be followed throughout this clause.

5.8.4.1.1.

Use of field weakening

As with field weakening the motor's pull-out torque decreases, only a limited load can be handled. The allowed maximum percent load for different speeds can not be specified as such, but must be analyzed in more detail. The main points affecting are the load to be handled and the rotational inertia of the system (motors, brakes, couplings, gears) together with the desired acceleration rate. When applying field weakening is considered, all restrictions caused by the mechanical components (motor, brake, gear) have to be checked. Note also that increasing the speed increases the stopping distance in quadrature. 5.8.4.2.

THERMAL DIMENSIONING

The thermal calculation of the motor can be done as shown in 5.8.1.3. When using S1-rated motors, instead of the inequality of 5.8.1.4, the below formula shall be used. Ck (1 - ηN ) * P N * T > ( 1 - η med ) * P med * (T - tr ) The operational class of the crane as well as the load spectrum has only a negligible effect on the inverter selection. The following items may vary depending on application and should be checked: • dimensioning of the braking resistor • need of enclosure cooling. 5.8.4.3. 5.8.4.3.1.

SELECTION CRITERIA FOR VERTICAL MOTIONS Motor selection

The required torque (MNmax ) to raise the maximum nominal load is calculated as in 5.8.2.1. The torque M max developed by the motor shall comply with 5.8.2.1. In order to be able to develop the necessary torque for acceleration, for lifting the test load or for compensating for variations in the mains voltage, the largest of the torques M1 ... Mn ( =Mi,max ) during the load cycle (see 5.8.1.3.1) should comply with Mpo / Mi,max ≥ 1,3, unless it is ensured by the control system that exceeding the pull out torque Mpo can be avoided. See also NOTE 1 in 5.8.4.1. NOTE: In the field weakening region, a lower safety margin according to the inverter manufacturer’s instructions may be appropriate.

Booklet 5 - 29/ 35


5.8.4.3.2.

Inverter selection

The continuous current rating of the inverter shall be at least equal to the motor current at load torque M Nmax. The current required by the motor at any foreseen loading including dynamic situations shall not exceed the short time overload rating of the inverter. 5.8.4.4. 5.8.4.4.1.

SELECTION CRITERIA FOR HORIZONTAL MOTIONS Motor selection

The required torques (M1 ... Mn) are calculated as in 5.8.3.1. The largest of them (M i,max ) must satisfy the following minimum condition: Mpo / Mi,max ≥ 1,2, unless it is ensured by the control system that exceeding the pull out torque Mpo can be avoided. The possibility to apply field weakening should be checked case by case. Usually there are very limited possibilities to apply field weakening in horizontal motions due to the fact that very often a major portion of the torque requirement comes from the acceleration (and deceleration) dependent term. Typical applications for field weakening are cranes exposed to wind during low wind condition. 5.8.4.4.2.

Inverter selection

The continuous current rating of the inverter shall be at least equal to the motor current at the highest speed maintaining torque. The current required by the motor at any foreseen loading including dynamic situations shall not exceed the short time overload rating of the inverter.

Booklet 5 - 30/ 35


5.9

LOAD LIFTING MEANS

5.9.1.

CURRENT SUPPLY

In view of the arduous duty to which current supply systems are subjected, the electrical equipment must be selected and installed with special care. Supply cables should be able to be wound on cable winders and their mechanical strength, resistance to external influences and heat-resistance, must be suitable for the service conditions. Cable fixing means should be so selected that all strain on the connections or damage to the cables is avoided. Cables should be installed and guided in such a way as to exclude the possibility of damage in normal service. 5.9.2.

LOAD HOLDING DEVICES

The requirements given in this clause apply to all load holding devices such as lifting magnets and vacuum lifters. Load holding devices are normally designed for a cyclic duration factor of 50 %. Other cyclic duration factors should be agreed between the manufacturer and user. The tear-off force should be at least twice the lifting capacity. If there is a stand-by power supply from batteries, the holding time should be at least 20 minutes. In this case, an automatic charging unit and a charge level indicator should be provided. Use of the stand-by supply should be indicated visually and audibly for general warning. If the battery voltage level is not adequate, a device preventing the installation from being used should come into effect. 5.9.2.1.

LIFTING MAGNETS

The insulation class of the windings should be selected according to the power loss, the ambient temperature and, if necessary, the heating caused by the goods handled. The lifting capacity for a lifting magnet should be specified for a precise load at rated voltage and operating temperature of the magnet coil. 5.9.3.

GRABS

The drive motors (electro-hydraulic or electro-mechanical drive) should be designed for S3, S4 or S6 duty depending on type and application. In normal service, the motors and electrical equipment located on the grab must comply with IP 55 at least. For underwater operation the degree of protection must be IP 57 at least. Due to the special service conditions of this equipment, jolts and vibrations must be given particular attention.

Booklet 5 - 31/ 35


5.9.4.

LOAD TURNING EQUIPMENT

Load turning equipment should be so designed that loads can be accelerated and braked without the ropes twisting. The arrangement of the lifting ropes, the load, the lifting height, the centre of gravity and the moment of inertia of the load and loading beam if applicable should be taken into account in the design of the equipment. The installation of guides such as telescoping or articulated systems may be used in order to prevent the twisting of ropes. All electrical connections to turning parts should be designed in accordance with the turning range. If the turning motor is mounted on the supporting structure of the hoisting appliance, it must comply with the degree of protection of the other motors on the structure at least. If the turning motor is mounted on the load lifting means, it must comply with IP 44 at least for indoor operation and IP 55 for outdoor operation.

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5.10

MAINTENANCE AND CHECKS

5.10.1. MAINTENANCE The electrical equipment of a hoisting appliance should be maintained in good condition. Maintenance should be based on the duty class and load spectrum of the hoisting appliance and carried out in accordance with the instructions of the supplier or manufacturer. In addition to the checklists in 5.10.2, requirements for maintenance instructions and practices can be found e.g. in ISO 12480-1 and EN 12644-1. 5.10.2. CHECKS A distinction is made between regular checks and checks made before the appliance is commissioned. Regular checks are subdivided into simple checks and comprehensive checks. 5.10.2.1. 5.10.2.1.1.

REGULAR CHECKS Simple checks

The safety devices which can be checked from the control position are to be checked regularly, in principle before the start of each workday, for their proper electric functioning. In particular, the following, at least, must be checked : − emergency limit switches, − brake functions, − emergency stop. 5.10.2.1.2.

Comprehensive checks

At least once a year, the electrical equipment of a hoisting appliance should be given a comprehensive check. Besides the above simple checks, the following should be checked thoroughly : — the settings and conditions of the electrical safety devices, — integrity of protective earth systems, — integrity of equipotential circuits, — insulation of all the electrical equipment, — tightness of all connections, — predetermined resistance values, if any, — physical condition of cables and cable inlets, — physical condition of safety devices, — presence and condition of devices protecting against direct contact, — the technical performance of replaced parts is compatible with the proper functioning of the hoisting appliance.

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5.10.2.2.

CHECKS BEFORE COMMISSIONING

In addition to the comprehensive checks, the checks before commissioning include, at least : — checking that all the hoisting appliance’s electrical equipment is in conformity with regulations and standards, — checking that the electrical equipment agrees with the circuit diagrams, — checking the switching sequence of the safety and control circuits, — checking the proper functioning of the electrical components — checking that the control system does not permit any uncontrolled excess speeds in normal operation, — checking the correct settings for all the electrical equipment and its proper functioning.

5.11

AUXILIARY ELECTRICAL EQUIPMENT

In addition to the requirements of EN 60204-32 clause 16 the following clauses apply. 5.11.1. LIGHTING 5.11.1.1.

CABIN

A fixed non-dazzling service lighting should be provided, so arranged that only the necessary illumination for the lighting of the control equipment is provided. When the general area lighting equipment is not sufficient to permit access and exit out of the cabin in safety, supplementary portable lighting should be provided ; this equipment must be able to work, even if the principal electrical circuits of the crane are isolated. 5.11.1.2.

WORKING AREA LIGHTING

When the working area lighting is provided by the appliance, projectors should be suitably placed on the crane, so that a minimum illumination of 30 lux at ground level is guaranteed. This lighting circuit should be independent of the principal circuits of the hoisting appliance. Precautions must be taken to avoid voltage drops produced by starting the motors cutting out the gas discharge lamps. 5.11.1.3.

ACCESS AND MACHINERY CABINET LIGHTING

When the general area lighting does not permit sufficient illumination, supplementary lighting independent of the principal circuits of the hoisting appliance should be provided. The minimum illumination should be 30 lux. 5.11.1.4.

EMERGENCY LIGHTING

When the lighting of the area does not permit exit out of the appliance in safety, a portable lamp, equipped with batteries should be provided. A battery charger must be provided in the cabin.

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5.11.2. 5.11.2.1.

HEATING AND AIR-CONDITIONING MACHINERY CABINETS

Natural or forced ventilation should be provided to disperse thermal power generated by the machinery and its equipment. Where electronic equipment is used and working conditions do not guarantee an ambient temperature for proper functioning of the electronic equipment, an air conditioning unit should be provided. 5.11.2.2.

CABIN

If necessary heating appliances should be provided in the cabin. This apparatus of black heat/non-radiant type shall be securely fixed. It must be provided with a thermostat and must have such a power to assure a minimum temperature of 15° C, taking into account the environment in which the equipment is installed. This apparatus must be fed independently of the principal circuits of the hoisting appliance. If required by the environment an air conditioning unit should be installed in the cabin to maintain a maximum acceptable temperature. This apparatus must be fed by a circuit independent of the principal circuits of the hoisting appliance. 5.11.3. AUXILIARY CIRCUIT If there is no possibility of supply in the proximity, auxiliary circuits should be provided for maintenance purposes, as follows : − A circuit for portable lighting, if the ambient lighting is not sufficient to carry out maintenance. − A circuit for portable tools according to agreement between customer and supplier. − These circuits should be protected by a differential circuit breaker of high sensitivity and they should be independent of the principal circuits of the hoisting appliance.

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F.E.M. 1.001

3rd

EDITION REVISED

1998.10.01


SAFETY RULES



Booklet 7

SAFETY RULES 7.1. SCOPE..................................................................................................................................................4 7.2. BASIS OF CALCULATIONS...............................................................................................................4 7.3. MARKING AND PLATES.....................................................................................................................4 7.3.1. RATING PLATE............................................................................................................................4 7.3.2. MANUFACTURER'S PLATE ......................................................................................................5 7.3.3. WARNING NOTICES ..................................................................................................................5 7.4. CONSTRUCTION REQUIREMENTS..................................................................................................6 7.4.1. CLEARANCES.............................................................................................................................6 7.4.1.1. Safety distances....................................................................................................................... 6 7.4.1.2. Longer clearance...................................................................................................................... 6 7.4.1.3. Upper clearance ....................................................................................................................... 6

7.4.2. DRIVER'S CABS IN GENERAL.................................................................................................6 7.4.2.1. Driver’s visibility ........................................................................................................................ 6 7.4.2.2. Cab’s dispositions..................................................................................................................... 7 7.4.2.3. Heat-insulating material............................................................................................................. 7 7.4.2.4. Glazed windows and entrances.............................................................................................. 7 7.4.2.5. Anti-glare lighting and heating................................................................................................... 7 7.4.2.6. Heat-proof................................................................................................................................. 8 7.4.2.7. Clean air supply ........................................................................................................................ 8

7.4.3. ADDITIONAL REGULATIONS REGARDING HOIST-SUSPENDED DRIVER'S CABS....8 7.4.3.1. Plate indication .......................................................................................................................... 8 7.4.3.2. Inopportune motions.................................................................................................................. 8 7.4.3.3. Suspended driver's cabins....................................................................................................... 8 7.4.3.4. Lowering speed limit................................................................................................................. 8 7.4.3.5. Controls..................................................................................................................................... 9 7.4.3.6. Limit switches........................................................................................................................... 9 7.4.3.7. Buffers...................................................................................................................................... 9 7.4.3.8. Distress signal .......................................................................................................................... 9 7.4.3.9. Safety headroom ...................................................................................................................... 9 7.4.3.10. Remote control........................................................................................................................ 9

7.4.4. GANGWAYS AND PLATFORMS..............................................................................................10 7.4.4.1. Cab access............................................................................................................................. 10 7.4.4.2. Indirect access ....................................................................................................................... 10 7.4.4.3. Access to gangways, stairways and platforms.................................................................... 10 7.4.4.4. Maintenance areas ................................................................................................................. 10 7.4.4.5. Raised locations areas ........................................................................................................... 10 7.4.4.6. Erection, dismantling, testing, repairs and maintenance work’s areas................................... 11 7.4.4.7. Lowered work’s areas........................................................................................................... 11 7.4.4.8. Headroom................................................................................................................................ 11 7.4.4.9. Guard rails .............................................................................................................................. 11 7.4.4.10. Slip-proof surfaces............................................................................................................... 12 7.4.4.11. Power lines protections........................................................................................................ 12


7.4.5. STAIRWAYS AND LADDERS ..................................................................................................12 7.4.5.1. General ................................................................................................................................... 12 7.4.5.2. STAIRWAYS........................................................................................................................... 12 7.4.5.3. LADDERS................................................................................................................................ 13

7.5. MECHANICAL DEVICES...................................................................................................................14 7.5.1. ROPE AND CHAIN DRIVES ....................................................................................................14 7.5.1.1. Drums...................................................................................................................................... 14 7.5.1.2. Lowest hook position.............................................................................................................. 14 7.5.1.3. Ropes...................................................................................................................................... 14 7.5.1.4. Chains..................................................................................................................................... 14

7.5.2. HOOK BLOCKS, PULLEYS AND OTHER LOAD CARRYING DEVICES ........................15 7.5.2.1. Prevent to jumping................................................................................................................... 15 7.5.2.2. Hand guard ............................................................................................................................. 15 7.5.2.3. Maintenance............................................................................................................................ 15 7.5.2.4. Accidental un-hooking protection ........................................................................................... 15 7.5.2.5. Working load plate................................................................................................................... 15

7.5.3. BRAKES......................................................................................................................................15 7.5.3.1. General ................................................................................................................................... 15 7.5.3.2. Hoisting brake ......................................................................................................................... 16 7.5.3.3. Travel and crab brakes........................................................................................................... 16 7.5.3.4. Slewing brake......................................................................................................................... 17 7.5.3.5. Luffing brake........................................................................................................................... 17

7.6. HYDRAULIC EQUIPMENT ................................................................................................................18 7.6.1. Pipes ...........................................................................................................................................18 7.6.2. Cylinders.....................................................................................................................................18 7.6.3. Working pressure .....................................................................................................................18 7.6.4. Pollution......................................................................................................................................18 7.6.5. Pressure gauge ........................................................................................................................18 7.6.6. Breathers....................................................................................................................................18 7.6.7. Limit positions...........................................................................................................................19 7.6.8. Bursting pressure.....................................................................................................................19 7.6.9. Hydraulic fluid ............................................................................................................................19 7.6.10. Unintentional start-up ............................................................................................................19

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7.7. SAFETY DEVICES..............................................................................................................................20 7.7.1. DEVICES FOR LIMITING WORKING MOTIONS ..................................................................20 7.7.1.1. HOIST MECHANISMS............................................................................................................... 20 7.7.1.2. TRAVEL DRIVES..................................................................................................................... 20 7.7.1.3. LUFFING AND SLEWING MECHANISMS........................................................................... 21

7.7.2. SAFETY AGAINST OVERLOADING AND OVERTURNING ................................................21 7.7.2.1. Derailment safety devices ...................................................................................................... 21 7.7.2.2. Overload protection ................................................................................................................ 22 7.7.2.3. Load chart............................................................................................................................... 22

7.8. AGEING OF APPLIANCES................................................................................................................23

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7.1. SCOPE These rules are applicable to cranes and heavy lifting appliances covered by clause 1.4. of 1 booklet 1 - "Object and Scope" .

7.2. BASIS OF CALCULATIONS The calculation of crane structures and mechanisms shall be in accordance particularly with booklet 3-"Design stresses in the structure" and booklet 4 "Design and choice of mechanism components".

7.3. MARKING AND PLATES Lifting appliances shall bear the following markings or plates, in the language of the country in which the appliance will operate, or in a language accepted by the user.

7.3.1. RATING PLATE The lifting capacity (and radius where applicable) shall be permanently marked in a visible position and shall be easily legible from the ground. The lifting capacity shall be the heaviest mass which may be hoisted by the crane, or by any hoisting accessory, either permanent or incorporated under certain conditions ; in the case of grabbing cranes, the lifting capacity shall be the permissible total weight of the grab and contents. In the case of luffing cranes, the lifting capacity corresponding to each radius shall be indicated in durable form showing appropriate graduations, and shall be clearly legible from the ground. More detailed indications of permissible loads at different radii shall be obtainable from the manufacturer's operating manual. In the case of cranes with more than one hoist, the lifting capacity of each hoist shall be indicated on the relevant hook block. It should furthermore be indicated if all the hoists can be used at the same time.

1

For builders tower cranes, the safety measures in preparation by the E.E.C. are also accepted by the F.E.M.

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7.3.2. MANUFACTURER'S PLATE Each lifting appliance, independent crab or winch shall be fitted at a convenient point with the maker's plate, detailing the following : - name of manufacturer, - year of manufacture, - manufacturer's serial number, - lifting capacity in kgs and/or tonnes, - type.

7.3.3. WARNING NOTICES A notice reading : "Do not stand under the load" shall be suitably located so as to be clearly visible. Crane access points shall be marked with a notice reading : "No access for unauthorised personnel". Particularly dangerous areas shall be marked with a notice reading : "Danger - Crane" and, where necessary, by means of warning colour stripes.

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7.4. CONSTRUCTION REQUIREMENTS 7.4.1. CLEARANCES 7.4.1.1. Safety distances All moving parts of lifting appliances, with the exception of handling and grabbing devices, in their most unfavourable position and under the most unfavourable loading conditions shall be at least 0,05 m from any fixed part of a building, at least 0,1 m from any guard rail or handrail and a minimum of 0.5 m from access areas. Access areas are all access ways authorised to personnel. This does not apply to working platforms. For railway loading profiles, the appropriate loading gauge shall be used and there shall be a minimum clearance of 0,5 m in access areas. Under no circumstances shall fixed parts of any lifting appliance encroach upon the railway clearance gauge. 7.4.1.2. Longer clearance The minimum vertical distance between the longer clearance line of a lifting appliance and areas of general access below (from the floor as well as from fixed or movable equipment belonging to the building, with the exception of working or service platforms or similar shall be at least 1.8 m in areas of general working access. From parts of stationary or mobile installations with limited walk-on or step-on access (such as roofs, heaters, machinery parts and cranes travelling below etc.) as well as from guard rails, the minimum vertical distance shall be 0,5 m. 7.4.1.3. Upper clearance The minimum vertical distance between the upper clearance line of a lifting appliance and fixed or moving parts above (e.g. between crab structures or guard-rails on the one hand and roof joists, pipelines or lifting appliances travelling overhead on the other) shall be not less than 0,5 m in maintenance areas and in the vicinity of platforms. This distance may be reduced to 0,1 m in the case of individual structural members, provided no danger to personnel results or that adequate precautions are taken to eliminate the risks.

7.4.2. DRIVER'S CABS IN GENERAL 7.4.2.1. Driver’s visibility Driver's cabs shall be designed so that the driver has a clear view of all work areas or so that he may adequately follow all operations with the aid of suitable equipment.

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7.4.2.2. Cab’s dispositions The driver's cab shall have sufficient room for the driver to be able to reach or leave the controls without hindrance. Controls shall preferably be operated from a sitting position, but also from a standing position when necessary. Driver 's cabs shall have a minimum headroom of 1,9 m and shall be fitted with a guard rail of at least 1,0 m high. Outdoor cabs, or those operating in unheated bays, shall be of enclosed construction, except in warm climates. Driver's cabs in heated bays, or which are seldom used or of an auxiliary nature may be of open construction. A protective shield shall be installed above the driver 's cab when there is any danger of falling objects. The layout of the cab and controls shall be ergonomically designed. 7.4.2.3. Heat-insulating material The structural framework of the driver's cab shall be of non-combustible material, and the side panels and roof may optionally be of fire resistant materials. The floor of the driver's cab shall be covered with non-metallic, heat-insulating material. 7.4.2.4. Glazed windows and entrances In cabs with windows less than 1,0 m from the floor and glazed areas in the floor, the glazing shall be constructed or protected to a height of 1,0 m so that personnel cannot fall through them ; walkover windows shall be tread proof. Entrances shall be protected against accidental opening. Sliding doors and outward opening doors of driver's cabs must lead on to landings. It must be possible to clean the windows of the driver's cab without risk. Glazed openings let into the floor of the driver's cab and those which are exposed to an increased risk of breakage or subjected to heat radiation when the crane is in operation shall be made of suitable safety glass. 7.4.2.5. Anti-glare lighting and heating Driver's cabs shall be provided with adequate anti-glare lighting to allow handling of the controls and, if necessary, can be ventilated. Enclosed driver's cabs for outdoor operation and cabins located in unheated bays must be provided with heating.

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7.4.2.6. Heat-proof Driver's cabs which are exposed to radiant heat shall be protected against heat radiation and of a heat-proof design, and they shall be air-conditioned in order to ensure tolerable working conditions. 7.4.2.7. Clean air supply Driver's cabs exposed to conditions presenting a health hazard, such as dust, vapours or gases, shall be protected against their entry and provided with a clean air supply.

7.4.3. ADDITIONAL REGULATIONS REGARDING HOIST-SUSPENDED DRIVER'S CABS 7.4.3.1. Plate indication The permitted number of persons and maximum load of the driver's cab shall be permanently and clearly indicated. Additional "Operating and maintenance instructions for hoist-suspended driver's cabs" shall be posted in the cab. 7.4.3.2. Inopportune motions It must not be possible for the driver's cab to spin or swing dangerously. 7.4.3.3. Suspended driver's cabins Hoist suspended driver's cabins shall be provided with an anti-fall device. Alternatively, there may be two independent means of suspension, provided that the driver's cabin remains secure should one means of suspension break, or should the drive or service brake fail. Each individual means of suspension shall be designed with a safety factor of not less than five times the full working load. If there is an anti-fall device and only one means of suspension, a minimum safety factor of eight is necessary. Rope drives shall be designed as a minimum in accordance with mechanism group M8. The diameter of the rope shall be not less than 6 mm. Ropes for outdoor duty shall be made of galvanised steel wire. 7.4.3.4. Lowering speed limit On reaching a speed of 1.4 times the nominal lowering speed the driver's cab shall automatically be brought to a halt. The driver's cab must be able to move independently of the load.

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7.4.3.5. Controls All controls shall stop automatically as soon as the driver releases them. 7.4.3.6. Limit switches Normal and emergency limit switches shall be provided for the highest and lowest positions of the cab, with separate switching and operating systems. Emergency limit switches shall directly switch off the main power circuit and activate an audible warning signal. In the event of the driver 's cab striking an obstacle or a suspension means becoming slack, all crane motions shall automatically shut down. Devices for returning the crane to service shall not be of self re-setting type. 7.4.3.7. Buffers If the travel speed of the driver's cab is greater than 40 m/min, devices shall be provided to reduce the speed promptly so that the buffers cannot be struck at a speed greater than 40 m/min. If the impact velocity is greater than 20 m/min, energy absorption buffers shall be provided. 7.4.3.8. Distress signal The driver 's cab shall be provided with a distress signal system independent of the electrical supply of the crane. It shall also be provided with a mean of emergency descent, e.g. a rope ladder or escape apparatus, which is always in the cab. 7.4.3.9. Safety headroom The user shall ensure that with the maximum stacking height of goods there is a safety headroom of 0,5 m to the underside of the driver's cab in its working position. 7.4.3.10. Remote control It must only be possible to remotely operate the crane from the ground with the driver's cab in its highest working position (see also 7.7.).

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7.4.4. GANGWAYS AND PLATFORMS 7.4.4.1. Cab access Easy and safe access to the driver's cab must be possible with the lifting appliance in any position under normal working conditions. If the floor of the driver's cab is less than 5 m above ground level, access may be restricted to certain positions of the lifting appliance, provided the driver's cab is fitted with appropriate emergency exit means, e.g. a rope ladder. Entry to the driver's cab should be preferably from a platform at the same level as the floor of the cab and provided with guard rails. Entry through the floor, or through the ceiling of the driver's cab shall be permissible only when necessitated by virtue of space restrictions. Where entry is made directly via a staircase, a platform or a gangway, the horizontal gap to the driver's cab entrance shall not exceed 0,15 m and the difference in level between the platform and the driver's cab floor shall not exceed 0,25 m. 7.4.4.2. Indirect access When the driver's cab cannot be reached directly from ground level in any position of the crane, and where the driver's cab floor is higher than 5 m from ground level, the crane installation shall be provided with appropriate gangways. For certain appliances, such as overhead travelling cranes, access may be limited to certain positions of the crane, if appropriate devices are provided which enable the driver to leave the cab. 7.4.4.3. Access to gangways, stairways and platforms Gangways, stairways, and platforms must have safe access with the lifting appliance in any position. Stairways and ladders in frequent use shall lead on to platforms or gangways. For such access, stairways are preferable to ladders. 7.4.4.4. Maintenance areas All operating locations and all equipment of the crane which requires regular inspection or maintenance must be provided with safe access, or be reached by means of portable work platforms. 7.4.4.5. Raised locations areas The above-mentioned locations when more than 2 m above floor level, and also crane jibs must be accessible via stairways, platforms, walkways or ladders. Steps shall be fitted with guard rails on both sides (see also 7.4.5.2.).

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7.4.4.6. Erection, dismantling, testing, repairs and maintenance work’s areas When work is carried out during erection, dismantling, testing, repairs and maintenance at points situated more than 2 m above floor level, appropriate measures shall be taken on cranes and jibs to ensure the safety of personnel (such as handrails, handgrips, safety devices, etc.) and to permit access of personnel to these points. Pulleys and moving parts placed at the end of jibs shall be designed so that no lubrication is necessary between erecting and dismantling the crane. If this is not the case, the jib must be provided with access. 7.4.4.7. Lowered work’s areas The above-mentioned access to the jib may be omitted when the latter can be lowered for the purpose of a complete visual check or when other constructions permit a visual check. 7.4.4.8. Headroom Stairways, gangways and platforms shall have a headroom of not less than 1,8 m. A clear passageway not less than 0,5 m wide must be provided in the vicinity of driven parts which move relative to gangways and platforms ; this dimension may be reduced to 0,4 m up to a height of 0,6 m by providing a handrail. The clear width of passage way between stationary parts shall be not less than 0,4 m. The clear headroom of little used access ways inside crane structures may be reduced to a minimum of 1,3 m, whilst at the same time the width shall be increased to 0,7 m, varying linearly with the reduction in height. The headroom above platforms used only for maintenance purposes may be reduced to 1,3 m. 7.4.4.9. Guard rails Parts of crane installations with access shall be provided with continuous guard rails on those sides where there is a danger of falling from a height of over 1 m. Toe guards shall be not less than 0,1 m high. Openings in guard rails shall be permissible where adequate protective measures against falling are provided. Guard rails shall, as a rule, be not less than 1 m high and shall be provided with toe guards and intermediate rails. The height of the guard rails may be reduced to 0,8 m for passage ways where a clear height of 1,3 m is permissible. Along gangways there shall be at least one handrail. For gangways alongside building walls or a solid wall construction, handrails shall be permissible in lieu of guard rails. Interruptions in the length of these shall not exceed 1 m (e.g. for building columns, door openings).

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7.4.4.10. Slip-proof surfaces Platform surfaces shall be suitably slip-proof. Holes, gaps and openings in the flooring shall be restricted in size so that a 0,02 m diameter ball cannot pass through. Gangways which are less than 0,5 m above exposed power lines must be provided with solid flooring in those areas. 7.4.4.11. Power lines protections When gangways are located adjacent to power lines, these lines must be protected against accidental contact.

7.4.5. STAIRWAYS AND LADDERS 7.4.5.1. General Stairways and ladders shall be provided wherever the difference in level exceeds 0,5 m. Footholds provided with hand grips may be installed on vertical surfaces where the height does not exceed 2 m (e.g. end carriages). Ladders shall be interrupted by intermediate landings if they exceed a height of 8 m. For great heights as, for example, tower cranes for building, additional intermediate platforms may be provided for which the vertical interval must be a maximum of 8 m. Where there are space problems, a single continuous ladder with platforms alongside may be installed. 7.4.5.2. STAIRWAYS The slope of stairways shall not exceed 65°, the height of individual steps shall not exceed 0,25 m (0,2 m for tower cranes) and their depth shall not be less than 0,15 m. If possible, the following ratios shall be observed : 2 . step height + 1 tread width = 0,63 m The height interval between steps shall be constant. In the case of main stairways the guard rail posts shall be spaced not less than 0,6 m apart, but in the case of other stairways 0,5 m shall suffice. Surface of treads shall be anti-slip. Stairways shall be provided with guard rails on each side ; where there is a wall on one side of the stairway a handrail shall be sufficient on the wall side.

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7.4.5.3. LADDERS The length of rungs between side frames shall be no less than 0,3 m ; their pitch shall be constant and not exceeding 0,3 m. The rungs shall be at least 0,15 m away from fixed structural members. A rung shall be able to withstand a force of 1200 N at the centre without suffering permanent deformation. Climb-through openings shall not be smaller than 0,63 m x 0,63 m or less than 0,8 m in diameter. Ladders over 5 m in height shall be provided with safety hoops starting at a height of 2,5 m. The distance between safety hoops shall be not greater than 0,9 m. They must be interconnected by at least three equally spaced longitudinal stringers. In all cases, one longitudinal stringer must be placed at a point which is exactly opposite the vertical centre line of the ladder. The strength of safety hoops, reinforced by the longitudinal stringers, must be adequate to withstand a force of 1000 N distributed over 0,1 m acting on any point of the hoop, without any visible deformation. The sides of ladders shall extend at least 1 m above the top rung, unless some other appropriate handhold is provided. Where space is limited, 0,8 m shall be acceptable. Safety hoops are not necessary on ladders placed on the inside of structures which can act as a safety guard and where there is a clearance of 0,7 m to 0,8 m between the ladder and the opposite side. Structural members can be considered equivalent to safety hoops provided they are arranged so that the perpendicular distance between bars in the danger zone is always less than 0,75 m and the inscribed circle between the ladder and the struts is less than 0,75 m. The ladders must be provided with rest platforms spaced so that the first stretch does not exceed 10 m and there are then rest platforms at every 8 m.

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7.5. MECHANICAL DEVICES 7.5.1. ROPE AND CHAIN DRIVES 7.5.1.1. Drums Rope drums shall be provided with rope grooves. The rope, as a rule, shall be wound in one layer. If winding is done in more than one layer, an appropriate spooling device or winding system shall be provided ; such a device is not necessary in the case of two layers where the rope is self-guided during winding. If there is a possibility of the rope becoming slack on the drum during operation or of not being wound on properly, means shall be provided to prevent this. Drums shall be provided with end flanges unless other measures are taken to prevent the ropes from overriding the ends or falling. The diameter of the flanges of the drum shall be such that, with the rope fully wound on, the flange shall project a distance of not less than one-and-a-half rope diameters above the top layer of the rope (for builders cranes, twice the rope dia). 7.5.1.2. Lowest hook position At the lowest permissible hook position, there shall still be at least two turns on the drum before the rope anchorage. If the rope end is fastened to the drum with bolted clamps, there shall be at least two separate clamps held by bolting fitted with positive locking devices. 7.5.1.3. Ropes Ropes shall be protected wherever possible against the influence of direct radiant heat and against being sprayed with molten and other dangerous substances. Special ropes shall be used when operating under conditions of excessive influence of heat, corrosive materials etc... 7.5.1.4. Chains Chain drives shall be provided with a device ensuring smooth running of the chain on the sprocket wheel and preventing it from jumping. An effective chain guard shall be provided.

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7.5.2. HOOK BLOCKS, PULLEYS AND OTHER LOAD CARRYING DEVICES . 7.5.2.1. Prevent to jumping Adequate means shall be provided to prevent the rope or chain from jumping off the pulleys. 7.5.2.2. Hand guard An adequate guard shall be provided where there is any danger of a hand being trapped between the rope and the pulley of the hook block. 7.5.2.3. Maintenance Rope pulleys shall be designed so as to be accessible for maintenance. 7.5.2.4. Accidental un-hooking protection Safety hooks or specially designed hooks shall be required where the method of operation induces increased danger of accidental un-hooking of the load or of the hook becoming snagged. 7.5.2.5. Working load plate Interchangeable load carrying devices, such as grabs, lifting magnets, buckets, tongs and beams, shall be permanently marked with their safe working load and dead weight and also, in the case of grabs and buckets for bulk materials, with their capacity and name of manufacturer.

7.5.3. BRAKES The provisions of this clause shall not apply to cylinder operated mechanisms e.g. hydraulic jack hoists. 7.5.3.1. General Drives shall be provided with mechanical brakes. If in exceptional cases the drive is through a self-locking gear, the brake may be omitted, provided it has been ensured that no excessive stresses or movements can occur. Brake mechanisms shall be easy to inspect. Brake springs shall be of the compression type. Brakes must be adjustable and brake linings must be replaceable.

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7.5.3.2. Hoisting brake Hoist units must be provided with brakes which are automatically applied and which can safely hold the test load in the event of switching off or failure of the hoist drive. Brake systems shall be designed for 1,6 times the hoist load and they shall be capable of braking the dynamic test load without a damaging snatch effect and without unacceptable overheating. Brakes of hoist units shall be arranged so that there is a positive mechanical link between the winch components which, on the one hand, generate the braking moment and, on the other hand, support the load. The electrical and mechanical gear shall make it possible to keep the lowering speed under load within the permissible limits. Hoist units for carrying molten materials shall be provided with two mechanical brakes which operate independently of each other, each of which shall meet the preceding requirements ; the second brake shall be applied with a time lag in relation to the first one. In special critical cases where failure of a driving unit must be catered for the second brake shall act on the rope drum ; this brake shall be so controlled that it is applied automatically, not later than the instant a speed of 1,5 times the nominal lowering speed has been reached. In such cases the control gear of the crane shall include an emergency stop which shall also activate the brake. 7.5.3.3. Travel and crab brakes Power driven travel drives of cranes and crabs shall be equipped with an automatic brake, or a brake which may be operated from the control position. Excluded from this category are cranes not subjected to wind, operating on a horizontal track at a speed not exceeding 40 m/min., or when on wheels with antifriction bearings, not exceeding 20 m/min. For cranes intended to carry molten materials, a brake is required independently of speed. The brakes must be so designed that the crane or crab can be brought to rest in a suitable time and held stationary in all operating conditions, under wind load when applicable and also in the case of power failure. Non-automatic travel brakes of cranes and crabs exposed to the wind shall be provided with a clamping device. Automatic travel brakes or anchoring devices shall be designed with a factor of safety not less than 1.1 against the maximum forces in out of service conditions.

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7.5.3.4. Slewing brake Power driven slewing drives shall be provided with brakes designed to bring to a halt, in a suitable time and to hold the slewing part stationary in all service conditions, under wind load when applicable and in the event of power failure. 7.5.3.5. Luffing brake Luffing systems shall be provided with brakes designed so that in the event of shut down or failure of the luffing gear drive they shall be applied automatically and safely hold the jib with the test load in the most unfavourable position. Brake mechanisms shall be designed for a minimum braking moment equivalent to 1,6 times the moment due to the hook load and the dead weight of the jib system plus 1,0 times the moment due to the wind load, in the most unfavourable operating configuration (i.e. maximum wind load in service). For the crane out of service this shall be at least 1,1 times the moment due to the dead weight of the jib system and due to the wind (max. out of service storm wind) in the most unfavourable jib position or in a specified out of service jib position.

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7.6. HYDRAULIC EQUIPMENT 7.6.1. Pipes Precision seamless steel pipe shall be used for pressure lines up to 3 cm outer diameter ; no welding shall be done on these except for welding on the flanges of bolted connections.

7.6.2. Cylinders When hoisting and luffing mechanisms are driven by hydraulic jacks, automatic devices (burst protection valves) shall be installed immediately adjacent to the connections to pressure lines to avoid any undesirable lowering of the load, particularly in the event of pipe failure. When there is a risk of dangerous lowering of the load due to oil leakage or leaking components, mechanical devices shall be provided to prevent this. With other hydraulic drives, the above motions must be stopped by means of automatic brakes, actuated by self-resetting controls, as specified under paragraph 7.5.3.

7.6.3. Working pressure Exceeding of the maximum specified working pressure shall be prevented by means of pressure relief valves. Appropriate provisions or constructional measures shall be taken to prevent the working pressure from being exceeded by more than 1.6 times, including the case of transient peak pressure.

7.6.4. Pollution Prior to start-up, the hydraulic system shall be free from foreign bodies such as turnings, sprinters or scale. The system shall be designed so that such foreign bodies can be readily removed when making repairs.

7.6.5. Pressure gauge Each hydraulic circuit shall have at least one connection outlet for a pressure gauge, enabling the measurement of pressure without any dismantling of pipework.

7.6.6. Breathers Hydraulic systems shall be fitted with breathers at suitable points.

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7.6.7. Limit positions Overrunning of limit positions shall be prevented by means of appropriate devices.

7.6.8. Bursting pressure Pipework and hoses must be designed with a factor of safety of four against bursting pressure ; this applies also to connections and flange joints. For stationary lifting appliances which are free from hydraulic shocks and vibration , a factor of safety of 2,5 shall suffice.

7.6.9. Hydraulic fluid Hydraulic fluid used in the hydraulic installations of cranes and winches shall comply with the requirements of service conditions and with the technological and safety requirements. The hydraulic fluid shall be specified to the user. It must be possible to check the maximum and minimum levels of the hydraulic tank.

7.6.10. Unintentional start-up Unintentional start-up of drives following the renewal of power supply after failure or the switching on of the isolator switch or of the main crane switch must be prevented ; e.g. by means of electrical interlocks or by means of an automatic mechanical return of the controller.

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7.7. SAFETY DEVICES 7.7.1. DEVICES FOR LIMITING WORKING MOTIONS 7.7.1.1. HOIST MECHANISMS The range of power driven hoist mechanisms shall be restricted at the highest and the lowest permissible positions of the load supporting means by automatic limit switches (emergency limit switches), having regard to the distance required to slow clown. The return from the limit positions must be possible by means of the controller. If a limit position is reached during normal operation, there shall be an additional and independent service limit switch. In this case, when the service limit switch has been tripped, it shall be possible to effect the return movement by use of the controller, but if the emergency limit switch has been tripped, this return movement shall not be possible. Hoists powered by an internal combustion engine and mechanically coupled with non intermediate electrical, hydraulic or pneumatic link may be provided with visual or acoustic warning devices instead of limit switches. 7.7.1.2. TRAVEL DRIVES Power driven cranes and crabs shall be provided with devices such as shoe brakes, rubber, spring or hydraulic buffers or other special devices, which are capable of absorbing one half of the energy of the moving masses at normal travelling speed and such that the maximum 2 deceleration in the driver 's cab does not exceed 5 m/sec . If the limit of travel is frequently reached during normal operation, the maximum deceleration in the driver 's cab shall not exceed 2,5 m/sec2. Cranes and crabs with radio control shall be provided with limit switches when the travel speeds are in excess of 40 m/min. Cranes and crab mounted driver's cabs, which are subject to wind, shall be provided with storm anchors for "Out of Service" conditions. When operating conditions require that certain wind conditions be taken into account in the operation of the crane, a wind indicator and alarm must be provided on the crane. Cranes shall be fitted with rail sweeps where material obstructions can come to rest in the track.

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When two or more cranes run on the same track special devices shall be provided to prevent a dangerous collision. Under no circumstances shall the deceleration in the driver's cab exceed 5 2 m/sec . In areas which are dangerous due to their being within the operational area of cranes or crabs, adequate measures shall be taken to protect personnel ; e.g. by the use of warning notices, flashing lights, acoustic warnings or, if necessary, automatic stopping devices. 7.7.1.3.

LUFFING AND SLEWING MECHANISMS

With power driven luffing mechanisms the movement of the jib at the limit of the travel shall be restricted by means of automatic limit switches (emergency limit switches) having regard to the distance required to slow clown. The return from the limit positions must be possible by means of the controller. Luffing mechanisms powered by an internal combustion engine and mechanically coupled with no intermediate electrical, hydraulic or pneumatic link, may be provided with visual or acoustic warning devices instead of limit switches. Similarly, power driven slewing mechanisms with-limited slewing range shall have the slewing movement limited by means of an automatic emergency limit switch. Furthermore, there shall be devices in accordance with the spirit of the provisions of paragraph 7.7.1.2. at the limits of travel of restricted slewing range or luffing range.

7.7.2. SAFETY AGAINST OVERLOADING AND OVERTURNING 7.7.2.1. Derailment safety devices Cranes and crabs shall be so designed, or provided with such additional safety devices that, even in the event of derailment or failure of a runner wheel or a wheel shaft or bearing, the maximum drop is limited to 3 cm and falling or overturning is prevented. In addition, exceptional forces such as those due to impact on buffers, collision and erection shall not cause the crane or the crab to overturn or fall.

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7.7.2.2. Overload protection Cranes and trolleys equipped with jibs and outriggers which can overturn due to an overload, and having a load lifting capacity independent of radius, shall be provided with an overload protection switch ; however, when the load lifting capacity varies with the outreach, the switch shall also act as a load moment switch. It should be possible to return into the permissible range of the load moment by reversing the movement or, where the overload has been caused by lifting the load, by setting it down using the controller. Cranes with hoists and/or luffing mechanism powered by an internal combustion engine and mechanically coupled with no intermediate electrical, hydraulic or pneumatic link may be provided with a visual or acoustic warning device instead of an overload switch. 7.7.2.3. Load chart Cranes and lifting appliances with the lifting capacity dependent on radius shall be provided with a permanent notice, clearly visible from the driving position and stating, in suitable graduations, the hook loads corresponding to the various radii.

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7.8. AGEING OF APPLIANCES Like other machines and lifting appliances, those pertaining to Section I of the FEM are also designed for a certain duration of life. They are, in addition, covered by Design Rules which have been developed from the scientific knowledge and experience of users and manufacturers for application to the various types of appliances. This notion of ageing applies mainly to the structure and the mechanisms, and not so much to the consumable components (such as : ropes, brake linings, brushes, heat engines, etc...). The principal factors contributing adversely to the ageing of appliances are : -

fatigue phenomena corrosion operational, assembly and dismantling accidents overloading inadequate maintenance.

The user must always bear in mina the importance of ageing.

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F.E.M. 1.001

3rd

EDITION REVISED

1998.10.01


TEST LOADS AND TOLERANCES



Booklet 8

TEST LOADS AND TOLERANCES

8.1. TESTS...................................................................................................................................................2 8.1.1. DYNAMIC TEST ...........................................................................................................................2 8.1.2. STATIC TEST ...............................................................................................................................2 8.1.3. NOTE 1..........................................................................................................................................2 8.1.4. NOTE 2..........................................................................................................................................2 8.2. TOLERANCES OF CRANES AND TRACKS .....................................................................................3 8.2.1. MEASURING PROCEDURE .....................................................................................................3 8.2.2. MANUFACTURING TOLERANCES FOR CRANES...............................................................3 8.2.2.1. Span.......................................................................................................................................... 3 8.2.2.2. Crane girders camber............................................................................................................... 4 8.2.2.3. Inclination of the wheel............................................................................................................. 4 8.2.2.4. Trolley rail center distance........................................................................................................ 4 8.2.2.5. Difference in height of two opposite points.............................................................................. 5 8.2.2.6. Bearing surface........................................................................................................................ 5 8.2.2.7. Vertical axis of the trolley......................................................................................................... 6 8.2.2.8. Trolley rails linearity .................................................................................................................. 6 8.2.2.9. Axes of the wheel bores.......................................................................................................... 7 8.2.2.10. Axle bores of opposite wheels .............................................................................................. 7 8.2.2.11. Bushed wheels....................................................................................................................... 8 8.2.2.12. Guide rollers............................................................................................................................ 8 8.2.2.13. Bushed wheels diameter........................................................................................................ 8

8.2.3. TOLERANCES FOR CRANE TRACKS....................................................................................9 APPENDIX..................................................................................................................................................11 A - 8.1.3. - TEST LOADS FOR CRANES IN SOME EUROPEAN COUNTRIES .........................11


.

8.1. TESTS Prior to being placed in service, appliances must be tested under overload conditions, as follows :

8.1.1. DYNAMIC TEST The dynamic test shall be carried out with an overload coefficient ρ1 = 1,2 i.e. with a load equal to 120 % of the safe working load. All motions shall be carefully operated in turn, without checking speeds of temperature rises in the motors (see clause 2.3.3.c).

8.1.2. STATIC TEST The static test shall be carried out with an overload coefficient ρ2 = 1,4 i.e. with a load equal to 140 % of the safe working load. This test must be carried out under still conditions and consists in hoisting the safe working load to a small distance above the ground and then adding the required surplus without shock (see clause 2.3.3.c.).

8.1.3. NOTE 1 The figures given for these test loads represent minimum requirements. Where national legislation or rules call for higher values, these must be complied with insofar as appliances destined for such countries are concerned. The test to be used in certain countries are given in appendix A-8.1.3. for information.

8.1.4. NOTE 2 When making these tests, it is customary to measure the deflection of the structure of an appliance. The present rules impose no obligation as to the allowable deflections. Should the user wish to impose a deflection limit, he must specify this in his call for tenders 1. 1

(1) The custom of regarding small deflection under load as a measure of the strength of an appliance should be discontinued. Although an unduly large deflection can adversely affect lattice girders because of the danger of movement at the joints, no untoward effects are to be feared in the case of solid-web or box girders. In practice, the magnitude of the deflection should be limited only from the standpoint of convenience of operation, since vertical oscillations of the load can be troublesome in some cases.

8-2


8.2. TOLERANCES OF CRANES AND TRACKS GENERAL The use of the Design Rules presupposes that the tolerances specified hereafter for cranes and tracks shall be maintained. These tolerances apply unless other conditions have been agreed with the user, and take no account of elastic deformations during the operation. The elastic deformations have to be taken into consideration if required. The specified tolerances are valid for overhead travelling cranes, gantry cranes and jib cranes, but not for railway cranes. For cranes which have been erected for temporary use only, e.g. building cranes, these rules are only partially valid, in other cases they are to be used judiciously.

8.2.1. MEASURING PROCEDURE when using measuring tapes, calibrated steel measuring types are to be used. The rules for the use of these measuring types are to be observed. The readings obtained are to be corrected for the sag of the tape measure as well as for the divergence of the ambient temperature from the standard temperature. All measurements on one and the same crane have to be made with the same tape and the same tension force.

8.2.2. MANUFACTURING TOLERANCES FOR CRANES 8.2.2.1. Span The greatest divergence ∆s of the crane span’s from the drawing dimension must not exceed the following values : for s ≤ 15 m : ∆s = ± 2 mm mm ( max. ± 15 mm ) for s > 15 m : ∆s = ± [ 2 + 0,15 . ( s-15 ) ] (s is to be expressed in m)

Figure 8.2.2.1.

8-3


8.2.2.2. Crane girders camber Crane girders, freely supported at their ends, must have no sag, even if the drawing does not prescribe a camber. This means that the track of the trolley with unloaded crane (without trolley) must have no deviation downward from the horizontal. This requirement only applies to cranes with a span longer than 20 m. 8.2.2.3. Inclination of the wheel In cases where the top of the rail is flat, the inclination of the wheel axis from the horizontal, for the unladen crane, must be between + 0,2 % and -0,05 %.

Figure 8.2.2.3. By unladen crane is meant the crane bridge without trolley, freely supported on the end carriages. 8.2.2.4. Trolley rail center distance The trolley rail centre distance must not differ from the nominal dimension s by more than ± 3 mm.

Figure 8.2.2.4.

8-4


8.2.2.5. Difference in height of two opposite points In a plane perpendicular to the travel direction of the trolley, the difference in height of two opposite points of the trolley track shall not exceed 0,15 % of the trolley rail centre distance, with a maximum of 10 mm .

Figure 8.2.2.5. 8.2.2.6. Bearing surface Trolley rails shall be laid in such a way that the running surface is horizontal and that the greatest unevenness of the bearing surface is no more than ± 3 mm for rail centres up to 3 m and no more than ± 0,1 % of the trolley wheel centre distance if it exceeds 3 m .

Figure 8.2.2.6.

8-5


8.2.2.7. Vertical axis of the trolley The vertical axis of the trolley rail must not diverge from the vertical axis of the rail girder web by more than half the thickness of the rail girder web .

Figure 8.2.2.7 8.2.2.8. Trolley rails linearity The axes of the trolley rails must not diverge from their theoretical axis by more than ± 1,0 mm in a rail length of 2 m. There should be no misalignments at rail joints.

Center axis of the trolley rails

Figure 8.2.2.8.

8-6


8.2.2.9. Axes of the wheel bores The axes of the wheel bores must not have an angular deviation greater than ± 0,04 % from its theoretical axis, in the horizontal plane . Theoretical position of all wheels

Figure 8.2.2.9. 8.2.2.10. Axle bores of opposite wheels The axle bores of wheels opposite to each other at each side of the track, and if wheels are mounted in bogies the axes of the bogie pins of the unwarped trolley and crane bridge shall have an alignment divergence in the vertical plane, less than 0,15 %, maximum 2 mm of the wheel centre distance .

Figure 8.2.2.10.

8-7


8.2.2.11. Bushed wheels The centre planes of wheels rolling on a common rail must not diverge more than ± 1 mm from the rail axis . Center axis of the rail head

Center of wheel

Center of wheel

Figure 8.2.2.11. For bushed wheels the above tolerances apply with the wheel in a central position between the contact surfaces at either side of the wheel. 8.2.2.12. Guide rollers If horizontal guide rollers are used, the centre of the distance between guide rollers at one corner must not deviate more than ± 1 mm from the axis of the rail . Center of distance between guide rollers Distance between guide rollers Center axis of the rail head

Figure 8.2.2.12. 8.2.2.13. Bushed wheels diameter The diameter tolerance of the wheels should correspond to the ISO tolerance classification h9. If runner wheel speeds are synchronized by an "electrical shaft", tighter tolerances may be required. These will have to be determined from case to case. These tolerances will apply also to non driven wheels, as the wheels must be interchangeable.

8-8


8.2.3. TOLERANCES FOR CRANE TRACKS The tolerances specified below apply to new crane tracks. If in the course of use, these tolerances are exceeded by 20 %, the track must be realigned. If the travelling behaviour is noticeably deteriorating, it may be necessary to realign the track, even if the tolerance excess has not reached 20 %. 1 - The greatest divergence ∆s from the spans s is : for s ≤ 15 m : ∆s = ± 3 mm for s > 15 m : ∆s = ± [ 3 + 0,25 . ( s-15 ) ] mm (s is to be expressed in m) (see figure 8.2.2.1.).

(max. ± 25 mm)

If horizontal guide rollers are provided on one rail only, the tolerances for the other rail only may be increased to three times the above values, but must not exceed 25 mm. 2 - It is assumed that with the trolley positioned in the centre of the span the deflection of both rail tracks is approximately equal. 3 - The greatest permissible tolerance of the upper edge of the rail is ± 10 mm from the theoretical height position. The theoretical height is either the horizontal position, or if applicable, the theoretical camber curve. The height position of the two rails may show a divergence of 10 mm. The curvature in a longitudinal direction may, at each point of a measured length of 2 m, not exceed a rise of ± 2 mm. 4 - The inclination of the rail rolling surface must not exceed the following values as compared with the theoretical position :

Longitudinally : 0,3 % Laterally : 0,3 %

Figure 8.2.3.a.

8-9


5 - The maximum permissible lateral deviation of each rail in a horizontal plane is ± 10 mm. The curvature in the longitudinal axis at any point shall not exceed ± 1 mm in a length of 2 m . Random sampling

Figure 8.2.3.b. For cranes guided on both sides by horizontal rollers, the above values are valid also for the rail surfaces of the horizontal rollers. For cranes guided on one rail only, the requirement for the straightness of the non guiding rail can be lowered, in agreement with manufacturer. 6 - No account has to be taken of misalignment at the rail joints. It is recommended that welded rail joints are used.

8 - 10


.

APPENDIX A - 8.1.3. - TEST LOADS FOR CRANES IN SOME EUROPEAN COUNTRIES Country AUSTRIA BELGIUM

Dynamic tests 125 % up to 25 t 110 % over 25 t Up to 20 t 125 % From 20 to 50 t +5t Over 50 t 110 %

Static tests

25 t 25 to 50 t 50 t

140 % 10 t 120 %

SWITZERLAND GERMANY

Pk = 1,25 . P (H1 and H2) Pg = 1,33 . P (H1 and H2) Pk = 1,25 . P Pk = 1,25 . P1 + 0,25 . P0 FRANCE 120 % (excluding builder's tower cranes and some dismountable appliances : 110 %) GREAT BRITAIN 125 % of the SWL ITALY NETHERLANDS NORWAY

SWEDEN

128 % self propelled cranes 120 % tower cranes 110 % other lifting appliances Up to 20 t 125 % From 20 to 50 t +5t Up to 20 t 125 % From 20 to 50 t +5t Over 50 t 110 % or FEM Up to 5 t 125 % From 5 to 20 t 120 % From 20 to 50 t 115 % Over 50 t 110 %

Comments

Pg = 1,50 . P (H3 and H4) Pg = 1,33 . P - 1,4 . P

According to DIN 15030 DIN 15018 part 1 DIN 15019 part 1 DIN 15030

150 % (excluding builder's tower cranes and some dismountable appliances : 133 %)

During 15 min. Not compulsory FEM

8 - 11



F.E.M.

SECTION I

3rd EDITION REVISED


1.001

1998.10.01


SUPPLEMENTS AND COMMENTS TO BOOKLETS 1 to 8

The total 3rd Edition revised comprises booklets 1 to 5 and 7 to 9Copyright by FEM Section I Also available in French and German


Booklet 9

SUPPLEMENTS AND COMMENTS TO BOOKLETS 1 TO 8

9.1. PREFACE..............................................................................................................................................3 9.2. PRESENTATION..................................................................................................................................3 9.3. VALUES OF THE DYNAMIC COEFFICIENT ψ (2.2.2.1.1.) ..............................................................4 9.4. FORCE DUE TO HORIZONTAL MOTIONS SH (2.2.3.)....................................................................6 9.4.1. Transverse action due to rolling action........................................................................................ 6 9.4.1.1. 9.4.1.2. 9.4.1.3. 9.4.1.4. 9.4.1.5.

Model of appliance................................................................................................................. 6 Relationship between tangential forces and displacements .................................................... 8 Forces due to skewing ........................................................................................................... 9 Tangential forces, Fx and Fy ................................................................................................ 10 Skewing angle α .................................................................................................................. 11

9.4.2. Buffer effects on the structure ................................................................................................. 11

9.5. WIND ACTION (2.2.4.1.)...................................................................................................................12 9.6. QUALITY OF STEEL (3.1.3.) ............................................................................................................13 9.7. STRUCTURAL MEMBERS OTHER THAN JOINTS - PERMISSIBLE STRESSES (3.2.2.1.)....18 9.8. JOINTS MADE WITH TENSION BOLTS WITH CONTROLLED TIGHTENING (3.2.2.2.1.).......20 9.9. CHECKING MEMBERS SUBJECT TO CRIPPLING (3.3)..............................................................20 9.10. CHECKING MEMBERS SUBJECT TO BUCKLING (3.4).............................................................21 9.11. CASE OF STRUCTURES SUBJECTED TO SIGNIFICANT DEFORMATION (3.5.)...................22 9.11.1. Non-proportional effect on the structure due to the forces .................................................... 22 9.11.2. Non linear structures and favourable effects of own weight................................................. 24

9.12. CHOICE OF RAIL WHEELS (4.2.4.)..............................................................................................26 9.13. DESIGN OF GEARS (4.2.5.) ...........................................................................................................27

9 -1


9.14. DETERMINATION OF PERMISSIBLE STRESSES IN MECHANISM COMPONENTS SUBJECTED TO FATIGUE ( 2.1.4.3., 4.1.3.5., 4.1.3.6., 4.1.3.7.) .......................................................27 9.14.1. Introduction .............................................................................................................................. 27 9.14.2. Partial modifications of booklet 2 and 4.................................................................................... 28 9.14.3. Checking for fatigue of a mechanical component - example................................................... 34

9.15. STABILITY AND SAFETY AGAINST MOVEMENT BY WIND (Booklet 6 : deleted)................36 9.15.1. Scope....................................................................................................................................... 36 9.15.2. Stability - Calculations.............................................................................................................. 36 9.15.2.1. 9.15.2.2. 9.15.2.3. 9.15.2.4. 9.15.2.5. 9.15.2.6. 9.15.2.7.

Stability .............................................................................................................................. 36 Calculations ....................................................................................................................... 36 travel effect........................................................................................................................ 36 site effect........................................................................................................................... 36 attachments effect ............................................................................................................. 36 collision effect .................................................................................................................... 36 For tower cranes ................................................................................................................ 37

9.15.3. Backward stability in service conditions ................................................................................. 38 9.15.4. Application of wind loads ........................................................................................................ 38 9.15.4.1. On-service.......................................................................................................................... 38 9.15.4.2. Out-of-service..................................................................................................................... 38

9.15.5. Crane base .............................................................................................................................. 38 9.15.6. Temporary additional stability devices..................................................................................... 39 9.15.7. Deformation.............................................................................................................................. 39 9.15.8. Resistance to drifting caused by wind.................................................................................... 39

9.16. TESTS (8.1.) ....................................................................................................................................41 9.17. TORANCES OF CRANES AND TRACKS (8.2.) ...........................................................................42

9 -2


9.1. PREFACE The Rules for the Design of Hoisting Appliances established by the Technical Committee of Section I of the Fédération Européenne de la manutention (F.E.M), which were published in the form of 8 booklets, have been increasing widely used in many countries all over the world. However, these rules were elaborated at the beginning of the eighties and must advance to keep pace with the improving state of knowledge and the increasingly efficient conception tool being used. The need for a revision is based on several observations : •

The harmonized standard EN 13001 from the work of CEN/TC147/WG2 will be applied progressively only at the beginning of the 21st century.

Thus, it is not desirable to await this date and continue to refer to the FEM rules, certain parts of which became obsolete. •

FEM 1.001 consolidates a great deal of experience and serves manufacturers and customers often as a basis for calculations.

•

Development of the rules should make future application of the harmonized standards easier, based on methods with limit states, among others.

The text below shall be considered as a supplement to those texts which are the subject of booklets 2, 3, 4 and 8. Booklet 6 is deleted.

9.2. PRESENTATION At the begining of each clause, there is a reference to the clause coming from booklet 2, 3, 4, 6 or 8 that the new text may replace. Example : Clause 2.2.2.1.1.of booklet 2 may be replaced by the following text:

9 -3


9.3. VALUES OF THE DYNAMIC COEFFICIENT ψ (2.2.2.1.1.) Clause 2.2.2.1.1.of booklet 2 may be replaced by the following text: For the coefficient ψ given by the clause 2.2.2.1.1 of booklet 2, we can take the value φ2 given by the following text : In the case of hoisting an unrestrained grounded load, the dynamic effects of transferring the load from the ground to the crane shall be taken into account by multiplying the gravitational force due to the mass of the hoist load by a factor φ2 (see figure F.9.3). The mass of the hoist load includes the masses of the payload, lifting attachments and a portion of the suspended hoist ropes or chains, etc.

Figure F.9.3 - Factor φ2 The factor φ2 shall be obtained as follows : φ2 = φ 2min + β2νh φ 2min and β2 are given in table T.9.3.a for the appropriate hoisting class. For the purpose of this standard, cranes are assigned to hoisting classes HC1 to HC4 according to their dynamic characteristics. (The selection of hoisting classes depends on the particular types of cranes and is dealt with in the European Standards for specific crane types.) Equally, values for φ2 can be determined by experiments or by analysis without reference to hoisting class. νh is the steady hoisting speed, related to the lifting attachment. Values of νh are given in table T.9.3.b.

9 -4


Table T.9.3.a - Values of β2 and φ 2min Hoisting class of the appliance

β2 s/m

φ 2 min.

HC1

0,17

1,05

HC2

0,34

1,10

HC3

0,51

1,15

HC4

0,68

1,20

Table T.9.3.b - Values of νh for estimation of φ2 Load combination

Type of hoist drive and its operation method of operation HD1

HD2

HD3

HD4

HD5

Case I, Case II

νh,max

νh,CS

νh,CS

0,5 . νh,max

νh. = 0

Case III

-

νh,max

-

νh,max

0,5 . νh,max

Where : HD1

hoist drive cannot be operated with creep speed;

HD2

a steady creep speed for the hoist drive can be selected by the crane driver;

HD3

hoist drive control system ensures the use of a steady creep speed until the load is lifted from the ground;

HD4

a stepless variable speed control can be operated by the crane driver;

HD5

after prestressing the hoist medium, a stepless variable speed control is provided by the drive control system independent of the crane driver;

νh,max.

is the maximum steady hoisting speed;

νh,CS

is the steady hoisting creep speed.

9 -5


9.4. FORCE DUE TO HORIZONTAL MOTIONS SH (2.2.3.) Clause 2.2.3.of booklet 2 may be replaced by the following text: 9.4.1. Transverse action due to rolling action Example of a method for analysing forces due to skewing. 9.4.1.1. Model of appliance

To enable an estimation to be made of the tangential forces between wheels and rails and also of the forces between the acting guide means that are caused by skewing of the lifting appliance, a simple travel-mechanics model is necessary. The lifting appliance is considered to be travelling at a constant speed without anti-skewing control. The model consists of n pairs of wheels in line, of which p pairs are coupled. An individual (i) pair of wheels can be defined, either as coupled (C) mechanically or electrically, or mounted independently (l) of each other. The latter condition is also valid in the case of independent single drives. The wheels are arranged in ideal geometric positions in a rigid crane structure which is travelling on a rigid track. Differences in wheel diameters are neglected in this model. They are either fixed (F) or movable (M) in respect of lateral movement. The lateral degree of freedom can, for example, be provided by a hinged leg. The different combinations of transversally in-line wheel pairs that are possible are shown in figure F.9.4.a. Coupled (C) Fixed/Fixed (F/F)

Independent (I)

IFF

CFF Fixed/Movable (F/M)

IFM

CFM Figure F.9.4.a - Different combinations of wheel pairs 9 -6


In figure F.9.4.b the positions of the wheel pairs relative to the position of the guide means in front of the travelling crane are defined by the distance di. NOTE: Where flanged wheels are used instead of an external guide means, d1 = 0.

9 -7


It is assumed that the gravitational forces due to the masses of the loaded appliance (mg) are acting at a distance µl from rail 1 and are distributed equally to the n wheels at each side of the crane runway. Sl ac k of the guide

Width of t he rail head T ravell ing di rect ion Wheel pair 1 Wheel pair 2

Wheel pair i Wheel pair n

Span Rail 1

Rai l 2

Figure F.9.4.b - Positions of wheel pairs 9.4.1.2. Relationship between tangential forces and displacements

It is at first necessary to assume a relationship between the tangential forces and the corresponding displacements occurring between wheel and rail. Since the wheel has to transfer drive moments (My) to the rail and its movement is restricted by the system (crane and runway), it slides in the longitudinal and lateral directions [u(ux,uy)]; corresponding tangential forces (Fx, Fy) react on the crane (see figure F.9.4.c). Sliding distance

Geometry Rolling distance

Forces

Figure F.9.4.c - Tangential forces and displacements

9 -8


In general, a relationship exists between the sliding distances (ux,uy), the free-rolling distance rψ, the wheel load Fz and the tangential forces (Fx, Fy), as follows : Fx = fx(sx, sy, pc, surface conditions) ⋅ Fz Fz

Fy = fy(sx, sy, pc, surface conditions) ⋅

The friction coefficients of the rolling wheel (fx, fy) depend on the slip, i.e. the relation between slide and free-rolling distances (sx = ux / rψ, sy = uY / rψ), on the contact pressure between wheel and rail (pc) and on the surface conditions of the rail. To simplify the calculation, the following empirical relationships may be used :  fx = 0,3⋅ 1 − 

e

(−250. s x ) , for sx Û 0,015

 fy = 0,3⋅ 1 −  



(

−250. s

e

) , for s

y

 

y

Û 0,015

9.4.1.3. Forces due to skewing

The crane model is assumed to be travelling in steady motion and to have skewed to an angle α, as shown in figure F.9.4.d. The appliance may be guided horizontally by external means or by wheel flanges. Direction of motion

Direction of rail

Lateral slip Guide means

Rail 1

Rail 2

Instantaneous Slide pole

Figure F.9.4.d - Forces acting on crane in skewed position

9 -9


A guide force Fy is in balance with the tangential wheel forces Fx1i, Fy1i, Fx2i, Fy2i, which are caused by rotation of the appliance about the instantaneous slide pole. With the maximum lateral slip sy = α at the guide means and a linear distribution of the lateral slip syi between the guide means and the instantaneous slide pole, the corresponding skewing forces can be calculated as follows : a) Distance between instantaneous slide pole and guide means h For systems F/F, h = (pµµ'l2 + Σd2i) / Σdi For systems F/M, h = (pµl2 + Σd2i) / Σdi where : p

is the number of pairs of coupled wheels;

µ

is the distance of the instantaneous slide pole from rail 1;

µ'

is the distance of the instantaneous slide pole from rail 2;

l

is the span of the appliance;

di

is the distance of wheel pair i from the close-fitting guide means. b) guide force Fy

Fy = ν f mg where : ν = 1 - Σdi /nh,

for systems F/F

= µ' (1 - Σdi /nh), -250 α

f = 0,3·(1 - e

)

for systems F/M where α < 0,015 rad;

mg

is the gravitational force due to the mass of the loaded appliance;

n

is the number of wheels at each side of the crane runway.

9.4.1.4. Tangential forces, Fx and Fy

Fx1i = ξ1i fmg Fx2i = ξ2i fmg FY1i = ν1i fmg FY2i = ν2i fmg Where : f and mg are as given in clause 9.4.1.3.b) ξ1i, ξ2i, ν1i and ν2i are as given in table T.9.4. 9 -10


Table T.9.4 - Values of ξ1i, ξ2i, ν1i and ν2i

Combinations of wheel pairs (see figure F.9.4.a)

ξ1i = ξ2i

CFF

µµ'l/nh

IFF

0

CFM

µµ'l/nh

IFM

0

ν1i

ν2i

µ d (1 − i ) n h µ' di (1− ) n h 0

9.4.1.5. Skewing angle α

The skewing angle α , which should not exceed 0,015 radians, shall be chosen taking into account the space between the guide means and the rail as well as reasonable dimensional variation and wear of the appliance wheels and the rails as follows : α = αg + αw + αt Where : αg = sg/wb

is the part of the skewing angle due to the slack of the guide;

sg

is the slack of the guide;

wb

is the distance between the guide means;

αw = 0,1 (b/wb)

is the part of the skewing angle due to wear;

b

is the width of the rail head;

αt = 0,001 rad

is the part of the skewing angle due to tolerances.

9.4.2. Buffer effects on the structure In clause 2.2.3.4.1 replace 0,7 m/s with 0,4 m/s

9 -11


9.5. WIND ACTION (2.2.4.1.) Clause 2.2.4.1.of booklet 2 may be replaced by the following text: Other recommendations or work results can also be used provided that the same level of safety is obtained

9 -12


9.6. QUALITY OF STEEL (3.1.3.) Clause 3.1.3 of booklet 3 may be replaced by the following text: The properties of the steel grades frequently used are provided in the following standards : EN 10025

Hot-rolled products of non-alloy structural steels. Technical delivery conditions;

EN 10113-1 Hot-rolled products in weldable fine grain structural steels Part 1: General delivery conditions; EN 10137-1 Plates and wide flats made of high yield strength structural steels in the quenched and tempered or precipitation hardened conditions - Part 1: General delivery conditions; EN 10149-1 Hot-rolled flat products made of high yield strength steels for cold forming - Part 1: General delivery conditions; EN 10210-1 Hot finished structural hollow sections of non-alloy and fine grain structural steels Part 1: Technical delivery requirements; EN 10219-1 Cold formed welded structural hollow sections of non-alloy and fine grain steels Part 1: Technical delivery requirements. The quality of steels in these design rules refers to the property of the steel to exhibit ductile behaviour at determined temperatures. The steels are divided into four quality groups. The group in which the steel is classified, is obtained from its notch ductility in a given test and at a given temperature. Tables T.9.6.a, T.9.6.b, T.9.6.c and T.9.6.d comprises the notch ductility values and test temperatures for the four quality groups. The indicated notch ductilities are minimum values, being the mean values from three tests, longitudinal test pieces are used. The notch ductility shall be determined in accordance with V-notch impact tests according to the European Standard EN 10045-1. Steels of different quality groups can be welded together. Tc

is the test temperature for the V-notch impact test.

T

is the temperature at the erection site of the crane.

Tc and T

are not directly comparable as the V-notch impact test imposes a more unfavourable condition than the loading on the crane in or out of service.

9 -13


Table T.9.6.a - Quality groups

Quality

Impact energy

Test temperature

group

according to EN 10045-1 J

Tc °C

1

2

3

-

27

27

-

+ 20

±0

27 40 4

- 20 27 40 27 40 27

Steels, corresponding to the quality group Designation of steels According to According to According old standard EN 10027-1 & to ECISS IC 10 EN 10027-2 ST 37-2 S235JR 1.0037 ST 44-2 S275JR 1.0044 50 B S355JR 1.0045 Fe 360-B Fe 430-B Fe 510-B R St 37-2 St 44-2 E 24(A37)-2 E 28 - 2 E 36 (A52)-2 40 B 43 B

S235JR S275JR S355JR S235JRG2 S275JR S235JR S275JR S355JR S235JRG2 S275JR

1.0037 1.0044 1.0045 1.0038 1.0044 1.0037 1.0044 1.0045 1.0038 1.0044

Fe 360-C Fe 430-C Fe 510-C St 37-3U St 44-3U ST 52-3U E 24 (A37)-3 E 28 - 3 E 36 (A52)-3 40 C 43 C 50 C Fe 360-D1 Fe 360-D2 Fe 430-D1 Fe 430-D2 Fe 510-D1 Fe 510-D2 Fe 510-DD1 Fe 510-DD2

S235JO S275JO S355JO S235JO S275JO S355JO S235JO S275JO S355JO S235JO S275JO S355JO S235J2G3 S235J2G4 S275J2G3 S275J2G4 S355J2G3 S355J2G4 S355K2G3 S355K2G4

1.0114 1.0143 1.0553 1.0114 1.0143 1.0553 1.0114 1.0143 1.0553 1.0114 1.0143 1.0553 1.0116 1.0117 1.0144 1.0145 1.0570 1.0577 1.0595 1.0596

St 37-3N St 44-3N St 52-3N E 24 (A37)-4 E 28 - 4 E 36 (A52)-4

S235J2G3 S235J2G4 S275J2G3 S275J2G4 S355J2G3 S355J2G4 S355K2G3 S355K2G4 S235J2G3 S275J2G3 S355K2G3

1.0116 1.0117 1.0144 1.0145 1.0570 1.0577 1.0595 1.0596 1.0116 1.0144 1.0595

40 D 43 D 50 D St 52-3N

S235J2G3 S275J2G3 S355J2G3 S355J2H

1.0116 1.0144 1.0570 1.0576

50D

S355J2H

1.0576

9 -14

Old

New

standard

standard

DIN 17100 BS 4360 (1972) EN 10025 (1990) DIN 17100

NF A 35-501 BS 4360 (1972) EN 10025 (1990) DIN 17100

NF A 35-501

BS 4360 (1972)

EN 10025 (1993)

EN 10025 (1990)

DIN 17100

NF A 35-501

BS 4360 (1972) DIN 17100

EN 10210-1

BS 4360 (1972)

(1994)


Table T.9.6.b - Quality groups Quality

group

Impact energy

Test temperature

according to EN 10 045-1

Tc °C

J

4

40 47 40 47 40 47 40 47 40 47 40 47 40 47 40 47 40 47 40 47 40 47 40 47 40 47 40 47 40 47 40 47

- 20

Steels, corresponding to the quality group Designation of steels According to According to According old standard EN 10027-1 & to ECISS IC 10 EN 10027-2 S275N S275NL

1.0490 1.0491

E 355 R E 355 FP

S355N S355NL

1.0545 1.0546

E 420 R E 420 FR

S420N S420NL

1.8902 1.8912

E 460 R E 460 FP

S460N S460NL

1.8901 1.8903

StE285 TStE285

S275N S275NL

1.0490 1.0491

StE355 TStE355

S355N S355NL

1.0545 1.0546

StE420 TStE420

S420N S420NL

1.8902 1.8912

StE460 TStE460

S460N S460NL

1.8901 1.8903

40EE

S275N S275NL

1.0490 1.0491

50EE

S355N S355NL

1.0545 1.0546

S420N S420NL

1.8902 1.8912

S460N S460NL

1.8901 1.8903

S275M S275ML

1.8818 1.8819

S355M S355ML

1.8823 1.8834

S420M S420ML

1.8825 1.8836

S460M S460ML

1.8827 1.8838

55EE

9 -15

Old

New

standard

standard

NF A 36-201 (1984)

EN 10113-2 (1993)

DIN 17102 (1983)

EN 10113-2 (1993)

(United Kingdom)

EN 10113-2 (1993)

EN 10 113-3 (1993)


Table T.9.6.c - Quality groups Quality

group

Impact energy

Test temperature

according to EN 10045-1

Tc °C

J

4

30 40 50 30 40 50 30 40 50 30 40 50 30 40 50 30 40 50 30 40 30 40 50 30 40 50 30 40 50 30 40 50 30 40 50 30 40 50 30 40 40 50 40 50 40 50 40 50

Steels, corresponding to the quality group Designation of steels According to According to According old standard EN 10027-1 & to ECISS IC 10 EN 10027-2 S 460 T

S 500 T

S 550 T

S 620 T

S 690 T

S 960 T

-20

TStE 460 V StE 500 V TStE 500 V EStE 500 V StE 550 V TStE 550 V EStE 550 V StE 620 V TStE 620 V EStE 620 V StE 690 V TStE 690 V EStE 690 V TStE 890 V EStE 890 V TStE 960 V

S460Q S460QL S460QL1 S500Q S500QL S500QL1 S550Q S550QL S550QL1 S620Q S620QL S620QL1 S690Q S690QL S690QL1 S890Q S890QL S890QL1 S960Q S960QL S460Q S460QL S460QL1 S500Q S500QL S500QL1 S550Q S550QL S550QL1 S620Q S620QL S620QL1 S690Q S690QL S690QL1 S890Q S890QL S890QL1 S960Q S960QL S500A S500AL S550A S550AL S620A S620AL S690A S690AL

9 -16

1.8908 1.8906 1.8916 1.8924 1.8909 1.8984 1.8904 1.8926 1.8986 1.8914 1.8927 1.8987 1.8931 1.8928 1.8988 1.8940 1.8983 1.8925 1.8941 1.8933 1.8908 1.8906 1.8916 1.8924 1.8909 1.8984 1.8904 1.8926 1.8986 1.8914 1.8927 1.8987 1.8931 1.8928 1.8988 1.8940 1.8983 1.8925 1.8941 1.8933 1.8980 1.8990 1.8991 1.8992 1.8993 1.8994 1.8995 1.8996

Old

New

norme

norme

NFA 36-204 (1992)

EN 10137-2 (1995)

(Allemagne)

EN 10137-2 (1995)

EN 10137-3 (1995)


Table T.9.6.d - Quality groups

ality

oup

Impact energy

Test temperature

according to EN 10 045-1

Tc °C

J

Steels, corresponding to the quality group Designation of steels According to old According to According standard EN 10027-1 & to ECISS IC 10 EN 10027-2 E 315 D

S315MC

1.0972

E 355 D

S355MC

1.0976

E 420 D

S420MC

1.0980

S460MC

1.0982

S500MC

1.0984

S550MC

1.0986

S600MC

1.8969

S650MC

1.8976

S700MC

1.8974

E 560 D

E 690 D QStE 300 TM

4

40

- 20

Old

New

standard

standard

NF A 36-231 (1992)

S315MC

1.0972

QStE 360 TM

S355MC

1.0976

QStE 420 TM

S420MC

1.0980

QStE 460 TM

S460MC

1.0982

QStE 500 TM

S500MC

1.0984

QStE 550 TM

S550MC

1.0986

QStE 600 TM

S600MC

1.8969

QStE 650 TM

S650MC

1.8976

QStE 690 TM

S700MC

1.8974

S315MC

1.0972

46F40

S355MC

1.0976

50F45

S420MC

1.0980

S460MC

1.0982

(United

S500MC

1.0984

Kingdom)

S550MC

1.0986

S600MC

1.8969

43F35

60F55

SEW 092

EN 10149-2 (1995)

S650MC

1.8976

75F70

S700MC

1.8974

QStE 260 N

S260NC

1.0971

QStE 300 N

S315NC

1.0973

QStE 360 N

S355NC

1.0977

QStE 420 N

S420NC

1.0981

EN 10149-3

S260NC

1.0971

(1995)

S315NC

1.0973

(United

S355NC

1.0977

Kingdom)

S420NC

1.0981

40/30 43/35

9 -17

SEW 92-75


9.7. STRUCTURAL MEMBERS OTHER THAN JOINTS - PERMISSIBLE STRESSES (3.2.2.1.) Clause 3.2.1.1.of booklet 3 may be replaced by the following text: Table T.9.7.a - Values for fy, fu, and σa for non alloy, fine grain structural steels and such in the quenched and tempered conditions Standard

Steel

Thickness t mm

Yield stress fy N/mm2

Ultimate stress fu N/mm2

S235 (Fe360)

≤ 16 ≤ 40 ≤ 100 ≤ 150 ≤ 200 ≤ 250 ≤ 16 ≤ 40 ≤ 63 ≤ 80 ≤ 100 ≤ 150 ≤ 200 ≤ 250 ≤ 16 ≤ 40 ≤ 63 ≤ 80 ≤ 100 ≤ 150 ≤ 200 ≤ 250 ≤ 16 ≤ 40 ≤ 63 ≤ 80 ≤ 100 ≤ 50 ≤ 100 ≤ 150 ≤ 50 ≤ 100 ≤ 150 ≤ 50 ≤ 100 ≤ 50

235 225 215 195 185 175 275 265 255 245 235 225 215 205 355 345 335 325 315 295 285 275 460 440 430 410 400 460 440 400 690 650 630 890 830 960

340 340 340 340 320 320 410 410 410 410 410 400 380 380 490 490 490 490 490 470 450 450 550 550 550 550 550 550 550 500 770 760 710 940 880 980

EN 10025

S275 (Fe440)

EN 10025 and EN 10113

S355 (Fe510) S355N and S355NL steels up to t ≤150

EN 10113

S460

EN 10137

S460

S690

S890 S960

9 -18

Permissible stresses: σa Case I Case II Case III 2 2 N/mm N/mm N/mm2 157 150 143 130 123 117 183 177 170 163 157 150 143 137 237 230 223 217 210 197 190 183 307 293 287 273 267 307 293 267 460 433 420 593 553 640

177 169 162 147 139 132 207 199 192 184 177 169 162 154 267 259 252 244 237 222 214 207 346 331 323 308 301 346 331 301 519 489 474 669 624 722

214 205 195 177 168 159 250 241 232 223 214 205 195 186 323 314 305 295 286 268 259 250 418 400 391 373 364 418 400 364 627 591 573 809 755 873


NOTE 1: The yield stress fy and the permissible stress σa of the hot finished structural hollow sections according to EN 10210-1 comply with those in Table T.9.7, t ≤ 65 mm, for grades 235 to 460. NOTE 2: The yield stress fy and the permissible stress σa of the cold formed welded structural hollow sections according to EN 10219-1 comply with those in Table T.9.7, t ≤ 40 mm, for grades 235 to 460. Table T.9.7.b - Values for fy, fu, and σa for high yield strength steels for cold forming and hollow sections Standard

EN 10149

Steel

S315 S355 S420 S460 S500 S550 S600 S650 S700

EN 10219-1

S420MH and MLH

Thickness t mm

all t

≤ > ≤ >

8 8 8 8

≤ 16 ≤ 40

Yield stress fy N/mm2 315 355 420 460 500 550 600 650 630 700 680

Ultimate stress fu N/mm2 390 430 480 520 550 600 650 700 700 750 750

420 400

500 500

9 -19

Permissible stresses: σa Case I Case II Case III N/mm2 N/mm2 N/mm2 210 237 286 237 267 323 280 316 382 307 346 418 333 376 455 367 414 500 400 451 545 433 489 591 420 474 573 467 526 636 453 511 618 280 267

315 300

382 363


9.8. JOINTS MADE WITH TENSION BOLTS WITH CONTROLLED TIGHTENING (3.2.2.2.1.) Clause 3.2.2.2.1.of booklet 3 may be replaced by the following text: For the calculation developed in clause 3.2.2.2.1, other recommendations or standards (for example : VDI 2230, FDE 25030, or the work of CEN/TC 185/WG 7) can be used. However, different methods may not be mixed. Tests (for example : extensometric) can complete and/or replace the calculations.

9.9. CHECKING MEMBERS SUBJECT TO CRIPPLING (3.3) Clause 3.3.of booklet 3 may be replaced by the following text: The method presented in ENV 1993-1 :1992 Eurocode 3 : Design of steel structures Part 1.1 may be used.

9 -20


9.10. CHECKING MEMBERS SUBJECT TO BUCKLING (3.4) Clause 3.4 of booklet 3 may be replaced by the following text: In determining the buckling safety coefficients, stated below, it was considered that flat plates under compressive stresses equally distributed over the plate width, are exposed to a greater danger of buckling than plates under stresses changing from compression to tension over the plate width. In consequence, safety against buckling was made dependent on the ratio ψ of stresses at the plate edges (appendix A-3.4. of booklet 3). It shall be verified that the calculated stress is not higher than the critical buckling stress divided by the following coefficients ηv given by table T.9.10: Table T.9.10 Case

Buckling safety ηv

Buckling of plane members

I II III

Buckling of curved members : Circular cylinders (e.g. tubes)

I II III

1,70 + 0,175 (ψ - 1) 1,50 + 0,125 (ψ - 1) 1,35 + 0,075 (ψ - 1) 1,70 1,50 1,35

The edge-stresses ratio ψ varies between + 1 and - 1. Appendix A.3.4.of booklet 3 gives the procedure for determining the critical buckling stress. Checking members subjected to buckling can be carried out according to other recommendations , for example ENV 1993-1. ENV 1993-1 is based on limit state analysis : partial safety factors γF and γM are used.

9 -21


9.11. CASE OF STRUCTURES SUBJECTED TO SIGNIFICANT DEFORMATION (3.5.) Clause 3.5.of booklet 3 may be replaced by the following text: 9.11.1. Non-proportional effect on the structure due to the forces In this case the stresses in the members cannot be proportional to the forces which cause them due to the deformation of the structure as a result of the application of these forces. This is the case, for example, with the stresses produced in the column of a crane (illustrated in figure F.9.11.a) where it is clear that the moment in the column is not proportional to the forces applied because of deformations that increase their moment arm. Figure F.9.11.a

In this case the calculation can be carried out : - either by using the limit states method; - or by using the method described in clause 3.5 of booklet 3.

Limit states method The figure F.9.11.b shows the limit states method :

Figure F.9.11.b - Typical flow chart of the limit state method fi

is the load i on the element or component;

Fj

is the load combination j from loads fi, multiplied by partial load coefficients and risk coefficient, when applicable;

Sk

are the load effects in section k of members or supporting parts, such as inner forces and moments, resulting from load combination Fj;

σ1l

are the stresses in the particular element l as a result of load effects Sk;

σ2l

are the stresses in the particular element l arising from local effects;

σl

is the resulting design stress in the particular element l;

R

is the specified strength or characteristic resistance of the material, particular element or connection, such as the stress corresponding to the yield point, limit of elastic stability or fatigue strength (limit states); 9 -22


lim σ is the limit design stress; γP

are the partial load coefficients applied to individual loads according to the load combination under consideration;

γn

is the risk coefficient, where applicable;

γm

is the resistance coefficient. NOTE 1: Instead of a comparison of stresses, as mentioned above, a comparison of forces, moments, deflections, ets. may be made. NOTE 2: A general description of the limit state, method of design is given in ISO 2394 : 1986, General principle on reliability for structures.

The individual specific fi loads are calculated according to the data in booklet 2. They are multiplied by the appropriate partial coefficients of load. Then they are combined according to the combinations given in clause 2.3 of booklet 2. Table T.9.11 - Partial coefficients γ p Loads

Clause

Loading condition (see 2.3) Case I 1,22

1,16

1,10

Favourable effect - estimated weight

0,90

0,95

1,00

Favourable effect - measured weight

1,00

1,00

1,00

2.2.2.1

1,34

1,22

Acceleration from drives

2.2.3

1,34

1,22

1,10

Effects of climate

2.2.4

1,16

1,10

Dead loads

unfavourable effect

2.2.1

Case II Case III

Working loads

with γm = 1,10 9 -23


For checking erection or dismanling, the case III is appropriate. NOTE: The values given in table 10 of the document pr EN 13001-2 may be accepted. 9.11.2. Non linear structures and favourable effects of own weight Clause 3.5 of Booklet 3 describes a corrective method for proof of competence calculations in cases of structures subjected to significant deformation. However, the significant deformations are not the only cases where the designer shall consider the use of similar correction. By combining clauses 2.3.1 and 3.2.1 or 3.4 the conditions for calculated stress in load combination Case I can be expressed as follows: σ {γc (SG + ψ SL + SH)} ≤ σcr / ν where σcr is the yield stress, crippling stress or buckling stress, whichever is the most critical one, ν

is the relevant coefficient νE or νV.

The above formula can also be presented as: νσ{γc (SG + ψSL + SH)} ≤ σcr for any structural system. If the structural system is behaving nearly linearly, the above formula can be modified as follows: σ = kg ν γc SG + kL ν γc ψ SL + kh ν γc SH ≤ σcr where coefficients kg, kL, and kh represent the linear relationships between the load effects (SG, SL, SH) and the calculated stress. Those coefficients depend on the configuration and type of loading of the crane.

9 -24


A simplified example: A simply supported beam,with span l and section modulus W, loaded by its own weight mg (SG ) and in the middle by force F (produced by the load effects SL) . The bending stress is calculated by the formula:

σ=

ν ⋅ γ c ⋅ mg ⋅ l ν ⋅ γ c ⋅ ψ ⋅ F ⋅ l + 8 ⋅W 4⋅W

where it is seen that k g =

l 8⋅W

and

kL =

l 4⋅W

In order to check the most critical effect of all the loads for a particular design detail, it is evident that the signs for the variable loads shall be selected so that they lead to the maximum combined stress (if such a combination is physically possible). Furthermore, the loads are multiplied by coefficients taking into account dynamic effects and an adequate margin for failure. However, in a case where the dead weight SG decreases the absolute value of the stress due to the variable loads (SG having an opposite, favourable stress effect) the multiplication of the dead weight by its coefficients would lead to an situation where the actual margin for the critical stress might be dangerously reduced. To maintain the intended margin for failure, the calculation method described in clause 3.5 shall be applied in the following cases: 1. - When the dead weight has an effect in the opposite direction to the effect of the variable loads, i.e. the dead weight has a balancing effect. Examples: towers and lower structures of slewing jib cranes and tower cranes. 2. - Especially for structures where the dead weight has an effect in the opposite direction to the effect of the variable loads and the tension and compression forces are carried by different structural elements. Example 1: Tie-downs for storm anchoring take the difference between upwards and downwards loads while the wheels carry all the downwards loads in the case of opposite wind action. Example 2: Bogie pins and the ties of pin joints (such as end cups). The pins carry all the compression while the ties carry the difference between upwards and downwards loads. 3. - Pre-stressed structures. Example: Bolted flange joints. In this case the variable loads shall be multiplied by the coefficient ν, but the pre-stress loads shall be taken as the lowest estimated nominal values. The coefficient for the dead weight shall be selected between 1 and ν depending on whether its effect is favourable or unfavourable for the bolts.

9 -25


9.12. CHOICE OF RAIL WHEELS (4.2.4.) Clause 4.2.4.of booklet 4 may be replaced by the following text: The method proposed in clause 4.2.4.1 can be used with the values of PL and c2 given below tables T.9.12.a and T.9.12.b: Table T.9.12.a - Values of PL Ultimate strength of metal used for rail wheel

PL

Minimum strength for the rail

N/mm_

N/mm2

N/mm2

fu > 500

5,00

350

fu > 600

5,60

350

fu > 700

6,50

510

fu > 800

7,20

510

fu > 900

7,80

600

fu > 1000

8,50

700

Table T.9.12.b - Values of c2 Group classification of mechanism

c2

M1 and M2

1,25

M3 and M4

1,12

M5

1,00

M6

0,90

M7 and M8

0,80

The hardening of the running surface at the depth of 0,01D may be taken into account when selecting the value of PL When using the tables above, it is not necessary to consider the 5 last paragraphs of clause 4.2.4.1.3 in booklet 4. 9 -26


9.13. DESIGN OF GEARS (4.2.5.) Clause 4.2.5.of booklet 4 may be replaced by the following text: Standards or calculation methods such as follows may be used, for example: • NF E 23015, Henriot method; • DIN 3990; • ISO 6336. For the design calculation for gears, the coefficient γm is not cumulative with the service factor (ka). However, it shall be at least equal to γm.

9.14. DETERMINATION OF PERMISSIBLE STRESSES IN MECHANISM COMPONENTS SUBJECTED TO FATIGUE ( 2.1.4.3., 4.1.3.5., 4.1.3.6., 4.1.3.7.) Clauses 2.1.4.3, 4.1.3.5, 4.1.3.6, 4.1.3.7 and appendix A 4.1.3 of booklets 2 and 4 may be replaced by the following text: 9.14.1. Introduction The calculation methods for determining the fatigue strength of mechanism components are similar in the documents of section I (FEM 1.001 edition 1987) and section II (FEM 2.131 and FEM 2.132 edition 1992) In the above-mentioned editions, the Wöhler curve of a component includes a second slope 6 (factor c') for the number of cycles n greater than 2.10 :

Figure F.9.14.a

9 -27


The presence of this second slope results in the determination of very low values for fatigue strength for very heigh number of cycles n, and consequently the safety level is too high. The text below proposes in particular the cancelling of the second slope of the Wöhler curve is presented below. 9.14.2. Partial modifications of booklet 2 and 4 BOOKLET 2 MODIFIED NOTE: The modifications are written in bold face type. 2.1.4.3

STRESS SPECTRUM

....In many applications the function f(x) may be approximated by a function consisting of a certain number r of steps, comprising respectively n1, n2,... nr stress cycles ; the stress σ may be considered as practically constant and equal to σi during ni cycles. If n represents the total number of cycles and σmax the greatest of the stresses σ1, σ2, ..., σr there exists a relation : r

n1 + n2 + ... + nr =

∑ ni = n

σ1 > σ2 > ... > σr

et

i =1

we obtain an approximated form :  σ  ksp =  1   σ max 

c

n1  σ  +  2  n  σ max 

c

n2  σ  + ..... +  r  n  σ max 

c

nr = n

c

 σ  n ∑  σmaxi  ni i= 1 r

the summation is truncated for the first ni ≥ 2.106. This ni is taken as nr and replaced with nr = 2.106 cycles. Depending on its stress spectrum, a component is placed in one of the spectrum classes P1, P2, P3, P4, defined in table T.2.1.4.3.

9 -28


BOOKLET 4 MODIFIED NOTE: The modifications are written in bold face type. 4.1.3.5

WÖHLER CURVE

In this context, the Wöhler curve, shows the number of stress cycles n which can be withstood before fatigue failure as a function of the maximum stress σ (or τ), when all stress cycles present the same amplitude and the same ratio k between extreme values. With regard to this WÖHLER curve, the following hypotheses are made respectively - for n = 8.103 : - for

σ = σR

or

τ=

σR 3

8·103 ≤ n ≤ 2·106, the area of limited endurance, the function is represented by a straight line TD in a reference comprising two logarithmic scale axes (figure 4.1.3.5 modified). The slope of the WÖHLER curve, in the interval considered, is characterized by the factor : log 2 ⋅ 106 − log 8 ⋅ 103 log 2 ⋅ 106 − log 8 ⋅ 103 c = tan ϕ = or c = tan ϕ = σ log σ R − log σ d log R − log τ d 3

- for

n > 2.106 :

σ = σd

or

τ =τd

Figure 4.1.3.5 modified The spectrum factor ksp of the component is determined by means of the above mentioned value of c.

9 -29


4.1.3.6. FATIGUE STRENGTH OF A MECHANICAL COMPONENT The fatigue strength σk or τk of a given mechanical component is determined by the following expressions respectively :  8− j σk =  2 c  ⋅ σd or   where j is the component's group number.

τk

 8− j =  2 c  ⋅ τ d  

The group classification of components, on the basis of their total number of cycles n and their spectrum factor ksp, as well as the critical fatigue stresses associated with each group, are illustrated in figure 4.1.3.6 modified where σjk represents the stress applying to group Ej. For the critical shear stresses, the letter σ must be replaced with τ.

Figure 4.1.3.6 modified Concerning the relation between spectrum classes P1 to P4 and the spectrum factor ksp see table T.2.1.4.3 in booklet 2.

9 -30


Comment : The above fatigue strengths are based on the component’s group number so those values are discontinuous. The above formulas may be usefully replaced with the following: σk =

σd  n  k sp ⋅   nd 

1 c

=

σd n    k sp ⋅   2 ⋅ 106 

1 c

or τ k =

See example in 9.14.3.

9 -31

τd 1  c

 n  k sp ⋅   nd 

=

τd n    k sp ⋅   2 ⋅ 106 

1 c


BOOKLET 4 MODIFIED NOTE: The modifications are written in bold face type. 4.1.3.7 PERMISSIBLE STRESSES AND CALCULATIONS The permissible stresses σaf and τaf are obtained by dividing the stresses σk and τk, defined in 4.1.3.6., respectively by a safety factor of νk . 1

νk = 3,2c

Taking :

σaf and τaf will be obtained by the relations :

σaf =

and it is verified that :

σk νk

τaf =

τk νk

σ ≤ σaf

τ ≤ τ af

with : σ maximum calculated normal stress amplitude, τ maximum calculated shear stress amplitude. In the case of components acted upon simultaneously by normal stresses and shear stresses with different ratios κ between extreme stresses, the following condition must be satisfied : 2 2 2  σ xσ y   σy   σx   τ 11 ,     −  +   ≤ 2   +     σ kx   τk  νk  σky   σkx ⋅ σky  in which : σx, σy

= maximum normal stresses in the directions x and y respectively,

τ

= maximum shear stress,

σkx, σky

= fatigue strength for normal stresses, in the directions x and y respectively,

τk = shear fatigue strength. If it is not possible to determine the most unfavourable case of the foregoing relation from the corresponding stresses σx, σy and τ, calculations must be performed separately for the loads σx max, σy max and τmax and the most unfavourable corresponding stresses. It should be noted that the checks described above do not guarantee safety against brittle fracture. Such safety can be ensured only by a suitable choice of material quality. 9 -32


BOOKLET 4 MODIFIED A- 4.1.3. - DETERMINATION OF PERMISSIBLE STRESSES IN MECHANISM COMPONENTS SUBJECTED TO FATIGUE NOTE: The modifications are written in bold face type. The endurance limit for a polished specimen is a laboratory value, which is practically never attained in parts actually used. Numerous factors - shape, size, surface condition (machining quality) and possible corrosion - induce discontinuities resulting in "notch effects", which decrease the permissible stresses in the part, when these stresses are calculated by conventional elementary methods for the strength of materials. These factors are taken into account by coefficients, called ks, kd, ku, kc, respectively all greater than or equal to unity.The endurance limit fora polished specimen isdivided by the product of theses coefficients. Designers are advised against using a skin factor taking into account the influence of surface treatments. ... The calculation for permissible stresses for fatigue can also be constructed with the stress gradient method (or Siebel method), which takes into account plastic adaptation with notch root .. This method is used in the following documents : - "Handbuch für Werkstoffprüfung", E. SIEBEL, Berlin 1958, - "Calcul des pièces à la fatigue - Méthode du gradient", A. BRAND, CETIM 1980 - FKM Forschungskuratorium Maschinenbau e. V. (Hrsg.) : Festigkeitsnachweis. Vorhaben Nr. 154, FKM-Heft 183-1Frankfurt 1994, - E DIN 743 : Tragfähigkeit von Wellen und Achsen. Teile 1-4, Beuth-Verlag, Berlin, April 1996

9 -33


9.14.3. Checking for fatigue of a mechanical component - example We consider a shaft whose initial stress spectrum is as given in table T.9.14: Table T.9.14 Level

σi (N/mm_)

σ i/σ max

ni (real)

ni (effective)

1

200

1

10 000

10 000

2

160

0,8

50 000

50 000

3

125

0,625

200 000

200 000

4

90

0,45

1 500 000

1 500 000

5

80

0,4

5 000 000

2 000 000

6

71

0,355

20 000 000

0

7

63

0,315

50 000 000

0 n = ∑ni = 3 760 000

According to booklet 2

Classification of the component

n = 3,76.106 It belongs to the class of utilization B8 (clause 2.1.4.2). c = 3 (slope of the Wöhler curve for the component) We calculate the spectrum factor ksp (clause 2.1.4.3): k sp = 1 ⋅

10 4 3,76 ⋅ 10 6

+ 0,83 ⋅

5 ⋅ 104 3,76 ⋅ 106

+ 0,6253 ⋅

2 ⋅ 105 3,76 ⋅ 106

+ 0,45 3 ⋅

1,5 ⋅ 106 3,76 ⋅ 106

+ 0,4 3 ⋅

2 ⋅ 106 3,76 ⋅ 106

= 0,0026 + 0,006809 + 0,012986 + 0,036353 + 0,034043 = 0,09285 It belongs to the spectrum class P1 and, consequently, to the component group E6 (clause 2.1.4.4).

9 -34


According to booklet 4

Checking for fatigue

The endurance limit of the component is : σd = 100 N/mm2 (clause 4.1.3.4). The fatigue strength of the shaft is (clause 4.1.3.6): σk = 2

 8− j    c 

. σd = 2

 8− 6   3 

.100 = 158 N / mm 2 1/c

The safety factor is (clause 4.1.3.7):

νk = 3.2

The permissible stress of the shaft is :

σaf =

= 3.21/3 = 1.473

σk 158 = = 1073 , N/mm2 νk 1473 ,

The maximum calculated stress is : σ = 200 N/mm2

σ = 200 N/mm2 > σaf = 107.3 N/mm2

The shaft is not acceptable for fatigue, because the max stress amplitude is higher than the permissible value. NOTE: If we use the fatigue 4.1.3.6, it becomes : σd = σk = 1/ c n    k sp ⋅   2 ⋅ 106 

strength formula proposed in the comment of the clause 100 3760000   0,09285 ⋅   2000000

1/ 3

2 = 178,9 N/mm

The permissible stress of the shaft is : σaf =

σk 178,9 = = 1215 , N/mm2 νk 1473 ,

The maximum calculated stress is : σ = 200 N/mm

σ = 200 N/mm2 > σaf = 121.5 N/mm2

The shaft is still not acceptable for fatigue, because the max stress amplitude is higher than the permissible value.

9 -35


9.15. STABILITY AND SAFETY AGAINST MOVEMENT BY WIND (Booklet 6 : deleted) The following text replaces booklet 6 9.15.1. Scope These requirements specifiy the conditions to be met when verifying, by calculation, the stability of cranes that are subject to tipping and drifting; it assumes that the cranes are standing on a firm, level supporting surface or track. NOTE: Where the crane is required to operate on an inclined surface, the manufacturer shall take the specified conditions into account. 9.15.2. Stability - Calculations 9.15.2.1. Stability

A crane is said to be stable when the algebraic sum of the stabilizing moments is greater than or equal to the sum of the overturning moments. 9.15.2.2. Calculations

Calculations shall be made to verify the stability of the crane by computing the sum of the overturning moments and the stabilizing moments using the loads multiplied by the load factor given in table T.9.15.a In all calculations, the position of the crane and its components, and the effect of all loads and forces, shall be considered in their least favourable combination, direction and effect. 9.15.2.3. travel effect

For cranes designed to travel with load, the forces induced by the maximum allowable vertical track variation as specified by the manufacturer shall be taken into account, in addition to other loads specified in condition II of table T.9.15.a. 9.15.2.4. site effect

Where required, excitation effects appropriate to the particular site or zone shall be considered as an additional loading condition. 9.15.2.5. attachments effect

In the calculations shown in table T.9.15.a, consideration shall be given to the loads induced by the weight of the crane and its components, including any lifting attachments which are a permanent part of the crane in its working condition. 9.15.2.6. collision effect

For the case of collision (e. g. buffer impact), the stability calculations shall be based on dynamic considerations.

9 -36


9.15.2.7. For tower cranes

For tower cranes the stability case according to table T.9.15.b shall be met. Table T.9.15.a. Condition

Loading

Load factor to be considered

Loads induced by the dead weight

1,0

Applied load

1,6P

Wind load

0

Inertia forces*

0


1,0

Applied load

1,35 P

Wind load

1,0 W1

Inertia forces

1,0 D

III. Backward Stability


1,0

(Sudden release of load)

Applied load

-0,2 P

Wind load

1,0 W1

Inertia forces

0


1,0

Applied load

1,0 P1

Wind load

1,2 W2

Inertia forces

0


1,0

Applied load

1,25 P2

Wind load

1,0 W3

Inertia forces

1,0 D

I. Basic Stability

II. Dynamic Stability

IV. Extreme Wind Loading

V. Stability During Erection or dismantling

Where :

D

are the inertia forces from drives

P

is the net load

P1

is the fixed load lifting attachment; out-of-service the fixed load lifting attachment shall be considered as part of the weight of the crane and its components

P2

is the weight of the part being installed/removed during erection or dismantling

W1 is the in-service wind effect W2 is the out-of-service wind effect - gusting effects are included W3 is the in-service wind effect W1 or the effect of the wind limit for erection work in accordance with the instruction handbook of the manufacturer

9 -37


Table T.9.15.b. Condition VI Stability During Erection or dismantling see figure F.9.15.

Loading



1,0

Horizontal applied load

0,10 P2

Vertical applied load

1,16 P2

Wind load

1,0 W3

Inertia forces

1,0 D

Figure F.9.15 - Example: application of a load P2 for fitting a jib 9.15.3. Backward stability in service conditions The backward stability is covered by condition III. 9.15.4. Application of wind loads 9.15.4.1. On-service

On-service wind forces shall always be applied in the least favourable direction. 9.15.4.2. Out-of-service

Out-of-service wind forces shall be applied in the least favourable direction for those cranes which are not free to rotate with the wind. For those cranes which are designed to rotate with the wind, the force shall be applied on the superstructure in the direction contemplated, and in the least favourable direction on the lower structure. 9.15.5. Crane base The crane manufacturer shall specify the forces imposed by the crane on the ground or supporting structure. The information given by the manufacturer should state all applicable conditions for which the forces have been stipulated (including out-of-service wind). Where the crane base provides all or part of the stability of the crane, the manufacturer shall specify the requirement applicable to the crane base.

9 -38


9.15.6. Temporary additional stability devices Tower cranes shall be stable in their operating configuration (condition I to IV in table T.9.15.a) without use of temporary additional devices. Temporary additional devices may be used to satisfy condition V in table T.9.15.a, erection or dismantling. Detachable ballast may be used to satisfy the case in condition IV of table T.9.15.a. However, this condition shall be met without this extra ballast, using a factor of 1,1 W2. 9.15.7. Deformation Where it can be shown that with the most unfavourable loads for the most destabilizing configuration, and consideration given to deformation (Second Order Theory), the effect on the overturning moment does not exceed 10%.Then stability calculations may be carried out ignoring deflections (First Order Theory) for ease of calculation. However, when this is the case, the overturning moments for each condition in table T.9.15.a shall be increased by the value obtained above according to the second order. 9.15.8. Resistance to drifting caused by wind The resistance to drifting caused by wind shall be proven by calculation for all cranes on rails operating in the open air under the conditions in table T.9.15.c. Table T.9.15.c - Drifting caused by wind Condition 1.- IN SERVICE

2.- OUT OF SERVICE

Loading



1,0

Applied load

1,35 P

Wind load

1,2 W1

Inertia forces

1,0 D


1,0

Applied load

1,0 P1

Wind load

1,2 W2

Inertia forces

0

9 -39


Where rail clamps or similar measures are necessary to avoid out-of-service drifting, the operator's manual shall advise that they shall be applied when the in-service wind limit has been reached. The resistance to travel due to friction and the coefficients of friction shown in table T.9.15.d shall apply. Table T.9.15.d - Resistance to travel and coefficients of friction Ratio : Resistance to travel / Radial load

Coefficient of friction between track and

Plain bearings

Anti friction bearings

the braked wheel

the rail clamp

0,02

0,005

0,14

0,25

NOTE: Higher coefficients of friction may be allowed for if it can be shown that these are

present in all surface conditions and qualities (e.g. oil, dirt, ice).

9 -40


9.16. TESTS (8.1.) Clause 8.1.of booklet 8 may be replaced by the following text: Prior to being in service, appliances must be tested under overload conditions, as follows : The cranes shall be tested dynamically, using the maximum nominal speed for each drive movement and overload that is not less than that obtained by multiplying the rated load by coefficient ρ given in table T.9.16: Table T.9.16 - Values of ρ dynamic test coefficient Last (t)

ψ ≤ 1,2

ψ ≤ 1,4

ψ > 1,4

≤ 30

1,2

1,25

1,3

≤ 100

1,15

1,2

1,25

< 100

1,10

1,15

1,2

NOTE: These values are not applicable for cranes equipped with powered series hoist mechanisms with a direct action lifting force limiter. In that case the values given in FEM 9.751 may be used. where ψ = dynamic coefficient according clause 9.3. For cranes this dynamic test also covers the requirements of static overload and stability testing.

9 -41


9.17. TORANCES OF CRANES AND TRACKS (8.2.) Clause 8.2.of booklet 8 may be replaced by the following text: The axes of the wheel bores shall not have an angular deviation greater than α from its theoretical axis, in the horizontal plane, see figure F.9.17 The theoretical axis is the arithmetic mean value of the direction angles of all wheel axes. The values for α are given in table T.9.17 below. Theoretical position of all wheels

Figure F.9.17 Table T.9.17 - Wheel direction deviation angle α/rad

Class of

Travel speed v (m/min)

mechanism

≤25

≤50

≤100

≤200

>200

M1

0.0012

0,0012

0,0012

0,0010

0,0008

M2

0,0012

0,0012

0,0010

0,0008

0,0007

M3

0,0012

0,0010

0,0008

0,0007

0,0006

M4

0,0010

0,0008

0,0007

0,0006

0,0005

M5

0,0008

0,0007

0,0006

0,0005

0,0004

M6

0,0007

0,0006

0,0005

0,0004

0,0004

M7

0,0006

0,0005

0,0004

0,0004

0,0004

M8

0,0005

0,0004

0,0004

0,0004

0,0004

NOTE: The angles α give approximately same amount of wear of the wheels and rails, when the wheels are designed according to 9.12. 9 -42

FEM 1.001 Rules for the Design of Hoisting Appliances

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