EN 13103:2009

2003-01.Slovenski inštitut za standardizacijo. Razmnoževanje celote ali delov tega standarda ni dovoljeno.Bahnanwendungen - Radsätze und Drehgestelle - Laufradsatzwellen - Konstruktions- und BerechnungsrichtlinieApplications ferroviaires - Essieux montés et bogies - Essieux-axes porteurs - Méthode de conceptionRailway applications - Wheelsets and bogies - Non-powered axles - Design method45.040Materiali in deli za železniško tehnikoMaterials and components for railway engineeringICS:Ta slovenski standard je istoveten z:EN 13103:2009SIST EN 13103:2009en,fr01-junij-2009SIST EN 13103:2009SLOVENSKI
STANDARDSIST EN 13103:20041DGRPHãþD

EUROPEAN STANDARDNORME EUROPÉENNEEUROPÄISCHE NORMEN 13103March 2009ICS 45.040Supersedes EN 13103:2001
English VersionRailway applications - Wheelsets and bogies - Non-poweredaxles - Design methodApplications ferroviaires - Essieux montés et bogies -Essieux-axes porteurs
- Méthode de conceptionBahnanwendungen - Radsätze und Drehgestelle -Laufradsatzwellen - Konstruktions- undBerechnungsrichtlinieThis European Standard was approved by CEN on 26 December 2008.CEN members are bound to comply with the CEN/CENELEC Internal Regulations which stipulate the conditions for giving this EuropeanStandard the status of a national standard without any alteration. Up-to-date lists and bibliographical references concerning such nationalstandards may be obtained on application to the CEN Management Centre or to any CEN member.This European Standard exists in three official versions (English, French, German). A version in any other language made by translationunder the responsibility of a CEN member into its own language and notified to the CEN Management Centre has the same status as theofficial versions.CEN members are the national standards bodies of Austria, Belgium, Bulgaria, Cyprus, Czech Republic, Denmark, Estonia, Finland,France, Germany, Greece, Hungary, Iceland, Ireland, Italy, Latvia, Lithuania, Luxembourg, Malta, Netherlands, Norway, Poland, Portugal,Romania, Slovakia, Slovenia, Spain, Sweden, Switzerland and United Kingdom.EUROPEAN COMMITTEE FOR STANDARDIZATIONCOMITÉ EUROPÉEN DE NORMALISATIONEUROPÄISCHES KOMITEE FÜR NORMUNGManagement Centre:
Avenue Marnix 17,
B-1000 Brussels© 2009 CENAll rights of exploitation in any form and by any means reservedworldwide for CEN national Members.Ref. No. EN 13103:2009: ESIST EN 13103:2009

Steel grade EA1N.27 7.3
Steel grades other than EA1N.28 Annex A (informative)
Model of axle calculation sheet.31 Annex B (informative)
Procedure for the calculation of the load coefficient for tilting vehicles.32 Annex C (informative)
Values of forces to take into consideration for wheelsets for reduced gauge track (metric or close to a metre).34 Annex D (normative)
Method for determination of full-scale fatigue limits for new materials.35 D.1 Scope.35 D.2 General requirements for the test pieces.35 D.3 General requirements for test apparatus.35 D.4 Axle body fatigue limit ("F1").36 D.4.1 Geometry.36 D.4.2 Verification of the applied stress.36 D.4.3 End of test criterion.37 D.4.4 Détermination of the fatigue limit.37 D.5 Axle bore fatigue limit ("F2").38 D.5.1 Geometry.38 D.5.2 Verification of the applied stress.38 D.5.3 End of test criterion.38 D.5.4 Determination of the fatigue limit.38 D.6 Wheel seat fatigue limit ("F3 et F4").39 D.6.1 Geometry.39 SIST EN 13103:2009

Relationship between this
European
Standard and the Essential Requirements of Council Directive 96/48/EC amended by Directive 2004/50/EC.42 Annex ZB (informative)
Relationship between this
European
Standard and the Essential Requirements
of Directive 2001/16/EC of the European Parliament and of the Council of 19 March 2001 on the interoperability of the trans-European conventional rail system amended by Directive 2004/50/EC of 29 April 2004.43 Bibliography.45
Introduction Railway axles were among the first train components to give rise to fatigue problems.
Many years ago, specific methods were developed in order to design these axles. They were based on a feedback process from the service behaviour of axles combined with the examination of failures and on fatigue tests conducted in the laboratory, so as to characterize and optimize the design and materials used for axles. A European working group under the aegis of UIC1 started to harmonize these methods at the beginning of the 1970s. This led to an ORE2 document applicable to the design of trailer stock axles, subsequently incorpo-rated into national standards (French, German, Italian) and consequently converted into a UIC leaflet. The bibliography lists the relevant documents used for reference purposes. The method described therein is largely based on conventional loadings and applies the beam theory for the stress calculation. The shape and stress recommendations are derived from laboratory tests and the outcome is validated by many years of operations on the various railway systems. This standard is based largely on this method which has been improved and its scope enlarged.
1 UIC : Union Internationale des Chemins de fer. 2 ORE: Office de Recherches et d'Essais de l'UIC. SIST EN 13103:2009

1 Scope This standard: 1) defines the forces and moments to be taken into account with reference to masses and braking conditions; 2) gives the stress calculation method for axles with outside axle journals; 3) specifies the maximum permissible stresses to be assumed in calculations for steel grade EA1N defined in
EN 13261; 4) describes the method for determination of the maximum permissible stresses for other steel grades; 5) determines the diameters for the various sections of the axle and recommends the preferred shapes and transitions to ensure adequate service performance.
This standard is applicable to: 6) solid and hollow axles of railway rolling stock used for the transportation of passengers and freight; 7) axles defined in EN 13261; 8) all gauges3. This standard is applicable to axles fitted to rolling stock intended to run under normal European conditions. Before using this standard, if there is any doubt as to whether the railway operating conditions are normal, it is necessary to determine whether an additional design factor has to be applied to the maximum permissible stresses. The calculation of wheelsets for special applications (e.g. tamping/lining/levelling machines) may be made according to this standard only for the load cases of free-running and running in train formation. This standard does not apply to workload cases. They are calculated separately. For light rail and tramway applications, other standards or documents agreed between the customer and supplier may be applied. Non-powered axles of motor bogies and locomotives are analysed according to the requirements of
EN 13104. 2 Normative references The following referenced documents are indispensable for the application of this document. For dated references, only the edition cited applies. For undated references, the latest edition of the referenced document (including any amendments) applies. EN 13260:2009, Railway applications – Wheelsets and bogies – Wheelsets – Product requirements
EN 13261:2009, Railway applications – Wheelsets and bogies – Axles – Product requirements
3 If the gauge is not standard, certain formulae need to be adapted. SIST EN 13103:2009

rail 2)(21gmm+ 0P N Vertical static force per journal when the wheelset is loaded symmetrically 21gm 1P N Vertical force on the more heavily-loaded journal 2P N Vertical force on the less heavily-loaded journal 'P N Proportion of P braked by any mechanical braking system
1Y N Wheel/rail horizontal force perpendicular to the rail on the side of the more heavily- loaded journal 2Y N Wheel/rail horizontal force perpendicular to the rail on the side of the less heavily-loaded journal H N Force balancing the forces 1Y and 2Y 1Q N Vertical reaction on the wheel situated on the side of the more heavily-loaded journal 2Q N Vertical reaction on the wheel situated on the side of the less heavily-loaded journal iF N Forces exerted by the masses of the unsprung elements situated between the two wheels (brake disc(s) etc.) fF N Maximum force input of the brake shoes of the same shoeholder on one wheel or interface force of the pads on one disc xM Nmm Bending moment due to the masses in motion 'xM , 'zM Nmm
Bending moments due to braking 'yM Nmm
Torsional moment due to braking MX, MZ Nmm
Sum of bending moments MY Nmm
Sum of torsional moments MR Nmm
Resultant moment b2 Mm Distance between vertical force input points on axle journals s2 Mm Distance between wheel rolling circles SIST EN 13103:2009

Average friction coefficient between the wheel and the brake shoe or between the brake pads and the disc s N/mm2 Stress calculated in one section K
Fatigue stress concentration factor R mm Nominal radius of the rolling circle of a wheel bR mm Brake radius d mm Diameter for one section of the axle 'd mm Bore diameter of a hollow axle D mm Diameter used for determining K r mm Radius of transition fillet or groove used to determine K S
Security coefficient G
Centre of gravity fLR N/mm2 Fatigue limit under rotating bending up to 107 cycles for smooth test pieces fER N/mm2 Fatigue limit under rotating bending up to 107 cycles for notched test pieces qa m/s2 Unbalanced transverse acceleration qf
Thrust factor 4 General The major phases for the design of an axle are: a) definition of the forces to be taken into account and calculation of the moments on the various sections of the axle; b) selection of the diameters of the axle body and journals and - on the basis of these diameters - calculation of the diameters for the other parts of the axle; c) the options taken are verified in the following manner: ¾ stress calculation for each section; ¾ comparison of these stresses with the maximum permissible stresses. The maximum permissible stresses are mainly defined by: ¾ the steel grade; ¾ whether the axle is solid or hollow. SIST EN 13103:2009

Figure 1 Unless otherwise defined by the customer, the masses )(21mm+ to be taken into account for the main types of rolling stock are defined in Table 2. For particular applications, e.g. suburban vehicles, other definitions for masses are necessary, in accordance with the specific operating requirements.
Table 2 Type of rolling stock units Mass )(21mm+ Freight wagons For the axle considered, proportion of the wagon mass under maximum permissible loading in service Coaches including accommodation for passengers, luggage or post
Mass in service + 1,2 ´ payload,
1 – Main line vehiclesa "mass in service" is defined as: the vehicle mass without passengers, tanks full (of water, sand, fuel, etc.); "payload" is defined as the mass of a passenger estimated at 80 kg, including hand luggage; ¾ 1 passenger per seat; ¾ 2 passengers per m² in corridors and vestibules; ¾ 2 passengers per attendant compartment; ¾ 300 kg per m2 in luggage compartments.
2 – Suburban vehiclesa b Mass in service + 1,2 ´ payload, "mass in service" is defined as the vehicle mass without passengers, tanks full (of water, sand, fuel, etc.); "payload" is defined as the mass of a passenger, which is estimated at 70 kg (little or no luggage); ¾ 1 passenger per seat; ¾ 3 passengers per m² in corridor areas; ¾ 4 or 5 passengers per m² in vestibule areas b; ¾ 300 kg per m2 in luggage compartments. a The payloads to be taken into account to determine the mass of the mainline and suburban vehicles broadly reflect the normal operating conditions of the member railways of the International Union of Railways (UIC). If the operating conditions differ significantly, these masses may be modified, for example, by increasing or decreasing the number of passengers per m² in corridors and vestibules. b These vehicles are sometimes associated with classes of passenger travel, i.e. 1st or 2nd class. The bending moment xM in any section is calculated from forces 1P, 2P, 1Q, 2Q, 1Y, 2Y and iF as shown in Figure 2. It represents the most adverse condition for the axle, i.e.: 1) asymmetric distribution of forces; 2) the direction of the forces iF due to the masses of the unsprung components selected in such a manner that their effect on bending is added to that due to the vertical forces; 3) the value of the forces iFresults from multiplying the mass of each unsprung component by 1 g.
Key G – centre of gravity of vehicle Figure 2 Table 3 shows the values of the forces calculated from1m. The formulae coefficient values are applicable to standard gauge axles and classical suspension. For very
different gauges, metric gauge for example, or a new system of suspension, tilting system for example, other values shall be considered (see Annexes B and C). SIST EN 13103:2009

Table 3 All axles except guiding axlea gmbhP111)/075,0625,0(+= gmbhP112)/075,0625,0(-= gmY1130,0= gmY1215,0= gmYYH12115,0=-= Guiding axlea gmbhP111)/0875,0625,0(+= gmbhP112)/0875,0625,0(-= gmY1135,0= gmY12175,0= gmYYH121175,0=-= For all axles
()()()()[]iiiysFRYYsbPsbPsQ-S--+--+=22121211
()()()[]iiiyFRYYsbPsbPsQS-----+=2112221 a The guiding axle is the axle of the first (i.e. leading) bogie of a coach used at the head of a reversible trainset. If an axle can be used in both positions (guiding or non-guiding), it is to be considered as a guiding axle.
Between loading plane and running surface yPMx1=
Between running surfaces Mx = P1y – Q1 (y – b + s) + Y1R – SiFi (y – b + s – yi)
iF: force(s) on the left of the section considered
General outline of xM
variations
a For a non-symmetric axle, the calculations shall be carried out after applying the load alternately to the two journals to determine the worst case.
5.3 Effects due to braking Braking generates moments that can be represented by three components:'xM,'yM,'zM (see Figure 3).
Figure 3 1) the bending component 'xM is due to the vertical forces parallel to the z axis; 2) the bending component 'zM is due to the horizontal forces parallel to the x axis; 3) the torsional component 'yM is directed along the axle centreline (y axis); it is due to the forces applied tangentially to the wheels. The components 'xM , 'yM and 'zM are shown in Table 5 for each method of braking. If several methods of braking are superimposed, the values corresponding to each method shall be added.
NOTE If other methods of braking are used, the forces and moments to be taken into account can be obtained on the basis of the same principles as those shown in Table 5. Special attention should be paid to the calculation of the 'xM component, which is to be added directly to the xM component representing masses in motion. 5.4 Effects due to curving and wheel geometry For an unbraked wheelset, the torsional moment 'yM is equal to 0,2 PR to account for possible differences in wheel diameters and the effect of passing through curves.
For a braked wheelset, these effects are included in the effects due to braking.
5.5 Calculation of the resultant moment In every section, the maximum stresses are calculated from the resultant moment MR(see the following note), which is equal to:
222MZMYMXMR++= where MX, MYand MZare the sums of the various components due to masses in motion and braking:
+='xxMMMX
='yMMY
='zMMZ
NOTE At a point on the outer surface of a solid cylinder (also in the case of a hollow one) with d as diameter, the components MX, MYand MZgenerate: SIST EN 13103:2009

32232dMZMXnps+= The value of the shear stress is the following (torsion of beams with a circular section):
316dMYtps= As a result, the two principal stresses 1s and 2s are obtained as:
24221tnnssss++=24222tnnssss+-= Since the normal stress has a much higher absolute value (10 to 20 times) than the shear stress, the diameter of the largest Mohr's circle is selected (21ss- in this case) as a check of the value assumed ford.
22232t221MYMZMXd324++p=s+s=s-s=s As a result, the definition of a resultant moment is:
222MZMYMXMR++= SIST EN 13103:2009

M’x = 0,3Ff G y M’x = 0,3Ff G (b – s) M’x = Ff G y M’x = Ff G (b – s)
a b
a b b b M’x
M’z = Ff (0,3 + G )y M’z = Ff (0,3 + G )(b – s) M’z = Ff (1 + G )y M’z = Ff (1 + G )(b – s)
a a
M’z
M’y
M’y = 0
M’y = 0,3P’R c d
M’y = 0
M’y = 0,3P’R c d
Two brake discs mounted on the axle Two brake discs attached to the wheel hubf Components M’x, M’z, M’y Between loading plane and running surface Between running surfaces and disc Between discs Between loading plane and running surface Between running surfaces
M’x = Ff G y M’x = Ff G (b – s + yi) M’x = Ff G y M’x = Ff G (b – s + yi)
b b b b M’x
M’z = Ff GyRRb M’z = Ff G RRb (b – s) M’z = Ff GyRRb M’z = Ff G (b – s) RRb
b b b b M’z
M’y M’y = 0 RPMy''3,0= d e
M’y = 0,3 P’R M’y = 0 M’y = 0,3 P’R
d, e
d, e SIST EN 13103:2009
M’x = Ff G ()ybysbi2-+
M’x = Ff G ()()ybbysbi-+-22 M’x = b21Ff G y (b+s-yi) M’x = Ff G()()ybbysbi-+-22
b B b b M’x
Between loading planes and running surface Between running surfaces
M’z = 21Ff GyRRb M’z =21Ff GRRb(b – s) M’z = 21Ff GyRRb M’z = 21Ff G (b – s) RRb M’z
Between loading planes and running surface Between running surfaces
M’y M’y = 0 M’y = 0,3 P’R M’y = 0 M’y = 0,3 P’R
d, e
d, e SIST EN 13103:2009
)(324'4dddMRK-´´´=ps ¾ in the bore:
)(324'4'dddMRK-´´´=ps
Figure 4a)
Figure 4b)
4 In the case of a conical wheel seat, the stress is calculated for the section where the resultant moment is the highest and the diameter of this section is taken to be equal to the lower diameter of the wheel seat. 5 Kis a fatigue stress concentration factor (i.e. it takes account of the geometry and properties of the materials). SIST EN 13103:2009

In a cylindrical part situated on the surface of a solid or hollow axle and in the bore of a hollow axle, the fatigue stress concentration factor Kis equal to 1. However, each change in section produces a stress increment, the
maximum value of which can be found:
at the bottom of a transition between two adjacent cylindrical parts with different diameters; ¾ at the groove bottom. NOTE When the transition comprises several radii, it is recommended that the critical section should not be at the intersection of two radii.
If this situation occurs, it is necessary to calculate the stress level at each intersection of the transition radii. The fatigue stress concentration factor Kto calculate this increment is shown in the nomograms in Figure 5 (transition between two cylindrical parts) and in Figure 6 (groove bottom). It is obtained from two ratios: dr and dD where:
r is the transition fillet radius; d is the diameter of the cylindrical part in which the stress concentration is calculated; D is the diameter of the other cylindrical part. SIST EN 13103:2009

Figure 5 — Fatigue stress concentration factor K as a function of D/d and r/d (at the bottom of the transition between two cylindrical parts) When a wheel or a disc has an interference fit on a seat, Dis to be assumed to be equal to the diameter of the hub (see Figure 5). For a collar or deflector, Dis assumed to be equal to the diameter of the bearing seat, since the interference fit of these parts is quite small.
Figure 6 — Fatigue stress concentration factor K as a function of D/d and r/d (groove bottom)
When a wheel, a brake disc, a gearwheel or a bearing has an interference fit on a seat, Dis to be assumed to be equal to the diameter of the hub or the bearing ring (see Figures 7a, 7b and 7c). For a collar or deflector or cross-bar, Dis assumed to be equal to the diameter of the bearing seat, since the interference fit of these parts is very small. SIST EN 13103:200
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