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Marine Structures 15 (2002) 113

Comparative fatigue strength assessment of a

structural detail in a containership using various

approaches of classification societies

W. Frickea,

*, W. Cuib

, H. Kierkegaardc

, D. Kihld

, M. Kovale

,T. Mikkolaf, G. Parmentierg, M. Toyosadah, J.-H. Yooni

aTU Hamburg-Harburg, AB 3-06, L .aammersieth 90, D-22305 Hamburg, GermanybSchool for Naval Architecture & Ocean Engineering, Shanghai Jiao Tong University, Shanghai 200030,

ChinacOdense Steel Shipyard Ltd., P.O. Box 176, DK-5100 Odense C., Denmark

dNaval Surface Warfare Center, Carderock Division, West Bethesda, MD 20817, USAeKvaerner Krylov Maritime Ltd., 196158 Moskovskoye shossee, 44 St. Petersburg, Russia

fVTT Manufacturing Technology, P.O. Box 1705, 02044 VTT, FinlandgBureau Veritas, 17 Bis Plau des la Defense 2, Paris la Defense 92077 Cedex, France

hDepartment of Marine Engineering, Kyushu University, 6-10-1 Hakozaki, Higashi-ku, Fukuoka 812-8581,

JapaniSamsung Heavy Industries, 530 Jangyung-ri, Shinhyun-up, Koje-City, Kyungnam 656-710, South Korea

Received 18 October 2000; received in revised form 23 March 2001; accepted 13 May 2001

Abstract

A comparative study on fatigue strength assessment procedures used by the classification

societies has been performed by Committee III.2, Fatigue and Fracture, of the International

Ship and Offshore Structures Congress (ISSC2000). A pad detail on the longitudinal coaming

of a Panamax container vessel was selected as an example. This detail was chosen because ofthe well-defined loading due to hull girder bending. Large differences in predicted fatigue lives

were found, ranging from 1.8 to 20.7 years. The spreading of results is attributable to

assumptions regarding loads, local stress determination and SN curve. For comparison, a

direct calculation of loads using the spectral method was performed. Also this calculation

showed a relatively short fatigue life of 5.3 years, although the structural detail is considered

not to be prone to fatigue failures. r 2001 Elsevier Science Ltd. All rights reserved.

Keywords: Fatigue strength assessment; Welded detail; Classification society; Container ship; Fatigue life

*Corresponding author. Tel.: 49-40-428-32-3148; fax: 49-40-428-32-3337.

E-mail address: [email protected] (W. Fricke).

0951-8339/02/$ - see front matter r 2001 Elsevier Science Ltd. All rights reserved.

PII: S 0 9 5 1 - 8 3 3 9 ( 0 1 ) 0 0 0 1 6 - 8


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1. Introduction

During the past decades, a great number of fatigue failures occurred in ship

structures, particularly in areas made of higher tensile steel. In the early 1990s,following the occurrence of cracks in side shell longitudinals of very large crude

carriers (VLCCs) after only a few years in service [1], all major classification societies

issued or revised rules and guidelines for explicit fatigue analyses. Failures in side

shell longitudinals were generally used for calibrating the load assumptions and

fatigue damage calculations.

It is quite natural that several investigations in the following years were focussed

on the complex load process at the ships sides (e.g., [24]) and on the fatigue

resistance of the affected structural details (e.g., [5]). Additional studies

were performed to review various procedures of the classification societies for

fatigue strength assessment, which showed large differences in the assumptions and

results [6].

In order to further investigate the different procedures for fatigue strength

assessment of ship structural details, the members of ISSC Committee III.2, Fatigue

and Fracture, decided to perform another comparative study in which the

procedures of the different classification societies were applied by the individual

participants. It was decided, however, not to investigate structural details at the

ships side again for following reasons:

* Side shell longitudinals are not the only ship details prone to fatigue. Longitudinalmembers in the upper and lower flanges of the hull girder require at least the same

attention given side shell longitudinals in view of the cyclic vertical and horizontal

wave bending moments. Recent casualties of ships breaking apart (MSC Carla,

Flare and Erika) underline the possible consequences of such failures,

although the contribution of fatigue to the cases mentioned still needs to be

clarified.* The load process at the ships sides is very complex, being characterized by

uncertainties in the combination of local and global hull girder loads and in non-

linear effects particularly related to local pressure loads. These uncertainties

consequently affect the assumptions contained in the different fatigue assessment

procedures and, hence, the spread of results.

Therefore, a detail in the upper longitudinal structural members was selected. The

loading for this detail is relatively well defined in the rules of the classification

societies. With respect to criticality, a detail in a relatively large container vessel has

been considered, where the use of higher-tensile steel and the presence of high tensile

stillwater stresses increase the failure probability. Frequent recommendations to pay

attention to outfitting details at the coaming of container vessels indicate that fatigue

cracks have occurred there.

W. Fricke et al. / Marine Structures 15 (2002) 1132


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2. Selected example

The example ship chosen is a 3800 TEU Panamax container vessel. The principal

particulars of the vessel are given as follows:

Length between perpendiculars, Lpp 242.000 m

Rule length, L=0.97LWL 234.740 m

Breadth moulded, B 32.250 m

Depth moulded, D 21.300 m

Draught moulded, T 14.000 m

Block coefficient, Cb 0.670

Maximum service speed, v 24.000 kn

Web frame spacing 3.144 m

Max still water bending moment, hogging 210000 t m

A welded pad detail on top of the longitudinal hatch coaming bar, where the hatch

covers are supported for vertical loads, was selected for the analysis. The

longitudinal elements in the midship section and the pad detail are given in Fig. 1.

The vessel has higher-tensile steel, HTS 36, in the upper part, noted AH and EH in

Fig. 1, with the rest of the ship made of mild steel, noted A. The pad detail for the

comparative study is located at mid-ship, midway between the fore and aft end

transverse coaming. The effects of torsional loading and stresses due to warping

distortion are therefore disregarded. The shape of the pad detail was designed toreduce the notch effect in the longitudinal direction. The size of welding is specified

by a throat thickness of 10.0 mm. Cracks are expected to initiate at the weld toes on

the top of the coaming plate.

The midship section properties of the hull grider were calculated and given as:

Moment of inertia about the vertical axis 659.765 m4

Moment of inertia about the horizontal axis 258.740 m4

Height of neutral axis above base line 8.496 m

Section modulus at coaming (rule z=22.781 m) 18.114 m3

3. Applied rules and guidelines

The fatigue analyses were performed according to current rules and guidelines

from the following classification societies:

* American Bureau of Shipping (ABS)* Bureau Veritas/Registro Italiano Navale (BV/RINA)* Det Norske Veritas (DNV)* Germanischer Lloyd (GL)* Korean Register (KR)

W. Fricke et al. / Marine Structures 15 (2002) 113 3


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Fig. 1. Midship section with longitudinal elements for a Panamax container vessel. Details of the pad

plate on the longitudinal hatch coaming bar are also seen.



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* Lloyds Register of Shipping (LR)* Nippon Kaiji Kyokai (NK)* Russian Register of Shipping (RS)

Direct calculations using the spectral method and statistical approach were also

performed.

The SN approach was generally applied, mostly in a simplified way, by

prescribing the shape of the stress histogram together with the number of load cycles.

The stresses used were either nominal stresses in the top of coaming, hot-spot stresses

at the critical weld toe, or else notch stresses. The hot-spot stress concentration

factor (SCF) was either calculated using a local finite element model or estimated on

the basis of tabulated SCFs, while the notch factor for the weld toe was taken from

the rules or guidelines, where applicable. Special aspects of the different rules and

guidelines are given in the following.

3.1. American Bureau of Shipping (ABS)

ABS supports both the nominal and the hot-spot stress approaches in their

simplified fatigue strength assessment method [7]. The fatigue stress ranges are

assumed to follow a Weibull probability distribution. For the present pad detail, the

design stress range is the same as that specified by IACS UR S11 [8]. The Weibull

shape parameter is a function of both ship length and location, and under some

assumptions, is estimated to be 0.81. The effect of mean stress has been ignored. DEn

SN curves were used to describe the fatigue strength of the details. In order to

account for corrosion, a net ship concept was used together with a stress reduction

factor of 0.95. For more detailed description using ABS approach to analyze this pad

detail, please refer to [9].

3.2. Bureau Veritas/Registro Italiano Navale (BV/RINA)

By the BV rules [10], which are comparable to RINA rules, two load cases should

be taken into account: half-time in head sea conditions and half-time in oblique sea

conditions. The stresses are based on the rule bending moment specified by IACSUR S11 [8]. The stresses and SN curve are based on the notch stress approach. The

hot-spot stress concentration factor was obtained from a finite element calculation,

while the notch stress concentration factor was obtained from a table depicting the

type of weld and quality of welding. The calculation included the stress component

from the IACS head sea vertical bending moment, transferred to a probability of

105 using a Weibull distribution with a shape parameter of 0.943 combined with

a horizontal bending moment. Another combination of vertical and horizontal

bending moments provided the oblique seas component. A reduction of life for

thickness above 16 mm is included as well as a reduction of stress amplitude when the

notch stress amplitude corresponding to a probability of 10

5 is above the yieldstress. The part of stress above the yield stress was weighted with a factor of 0.6

(compressive stress factor).



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3.3. Det Norske Veritas (DNV)

The procedure from DNV [11] for fatigue strength assessment is based on the hot-

spot stress approach. The stress concentration factor was obtained from a table ofstandard details which contains doubler plates. However, the stress range was

multiplied by a stress concentration factor for the weld (notch stress) before entering

the SN curve. The calculations only include the stress component from the vertical

bending moment, per IACS UR S11 [8], transferred to a probability of 104 by the

Weibull distribution with a shape parameter of 0.93. The stress range was further

reduced by a factor of 0.80 to account for worldwide operation. A reduction in life

for a plate thickness above 22 mm was included.

3.4. Germanischer Lloyd (GL)

The simplified approach of GL [12] assumes a long-term histogram of stress ranges

characterised (a) by a maximum stress range of 75% of the IACS UR S11 [8] design

value for vertical hull girder bending in order to account for the effect of varying

loading conditions and worldwide service, (b) by a total of 50 million load cycles,

and (c) by a Weibull shape parameter x 1: Both, the hot-spot as well as the

nominal stress approach are supported; the latter based on the recommendations of

the International Institute of Welding [13]. Here, a relatively conservative detail

category of FAT 50 is assumed, which may be substantially increased in case of a

reduced weld flank angle. Mean stress is taken into account by a correction factor.

3.5. Korean Register (KR)

The guidance for the fatigue strength assessment of ship structures of KR [14] is

based on the hot-spot stress approach. Since the mean stress effect is to be considered

in the case of compressive mean stress only, it was ignored in this calculation. The

thickness effect is to be considered for plate thicker than 22 mm. In the calculation of

the SCF, if the weld is not considered in the finite element model, the stresses at the

distance t/2 and 3t/2 from the weld toe are determined by the Lagrange interpolation

using the element stresses in the region. The hot-spot stress was then obtained bylinear extrapolation to the weld toe using the stresses determined at distances of t/2

and 3t/2 from the weld toe.

3.6. Lloyds Register of Shipping (LR)

From Lloyds Register, the ShipRight Fatigue Design Assessment Level 3

procedure [15] was applied. FDA Level 3 is a spectral approach where, in this case, a

scaled hull form and weight distribution from a similar ship was used for generation

of the vertical bending moment response amplitude operator from 2D strip theory. A

typical container trade pattern was assumed and a 20-year simulation period with27% non-sailing days was chosen. The fatigue stress was obtained from a detailed

FE-model composed of 4-noded shell elements, representing the geometric stress



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(hot-spot stress) with some embedded notch stress effects. The analysis was

performed in-house by LR.

3.7. Nippon Kaiji Kyokai (NK)

NK [16] guidance supports the hot-spot stress approach. The stresses are normally

based on direct calculation using a spectral method and statistical approach. The

hot-spot stress is determined with the aid of a zooming FE-model with element sizes

of approximately t t in the critical area. Several standard SN curves for each

welded joint were prepared and the effects of mean stress with residual stress were

considered. In the present case, the stress range was based on 100% of the rule

bending moment according to IACS UR S11 [8] since direct calculation using the

spectral method was not carried out. The SN curve for non-load carrying type fillet

welded joints in NK guidance was used.

3.8. Russian Register of Shipping (RS)

RS [17] supports both the nominal stress approach for standard details and the

hot-spot stress approach for other details, including the present pad detail. For this

calculation, only the stress component from the vertical bending moment was

included and transferred to a probability level of exceedance 103 by a Weibull

distribution with a shape parameter of 0.88. Mean stress was taken into account by a

correction factor. Corrosion effect may be considered.

4. Calculated results

The calculated results have been derived independently by the responsible

participants and, in most cases, in co-operation with the respective society. The

details of the applied approaches are very different, ranging from the nominal stress

approach to spectral load analysis and local FE analysis of the patch detail.

However, the collected results are assessed by comparing the loading, the local stress,

and the fatigue life. The results are presented in Table 1.The third column in Table 1 (after the classification society and type of approach)

contains information on the loading, i.e., the stress histogram as explained in the left-

hand diagram of Fig. 2. The highest stress range Dsmax was generally based on

nominal stresses in order to allow a better comparison. Although the loading was

governed by vertical hull girder bending only, significant differences can be observed

in Dsmax assumed, ranging from 199 to 319 MPa.

Also, the Weibull shape parameter, as well as the number of stress cycles, vary to a

certain extent. It is interesting to note that the Weibull shape parameter apparently

decreases with increasing maximum stress range, indicating some calibration behind

the load assumptions.The stress concentration factors in the centre column are partly derived from finite

element analyses with models as shown in Fig. 3. The hot-spot SCF derived varies



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Table 1

Comparison of results from comparative study

Rules and

guidelines

Type of stress

approach

Histogram of nominal stress ranges SCF

(hot spot/weld)

Design S

Dsmax (MPa) nmax x sma(MPa) DsR (MP

ABS [7] Nominal 318.7 5107 0.81 F F/F 49.9

Hot spot 318.7 5107 0.81 F 1.736/F 80.3

BV/RINA [10] Notch 278b 2.8107 0.943 F 1.63/1.84 142.6

136c 2.8107 0.943 F

DNV [11] Notch 233.0 6.65107 0.93 F 1.47/1.5 142.2

GL [12,18] Nominal 209.2 5107 1.0 104.9 F/F 50

Hot spot 209.2 5107 1.0 104.9 1.9/F 110f

KR [14] Hot spot 278.8 5.64107

0.943 1.66/F 91.3 LR [15] Hot spotg 210e 5.7107 F F 1.81g 124g

NK [16] Hot spot 281.5 108 1.0 108.9 2.15/F 95

RS [17] Hot spot 199.0 5107 0.88 113.8 1.80/F 100

aMean stress only given if it affects life.bPart from head seas.cPart from oblique seas.dEstimated from usage factor.eRead from 108 probability of exceedence.fIncluding 10% increase for exact stress analysis.gHot spot with some embedded notch stress effects.h

Life 13.2 years for corrosion wastage by 0.5% of section modulus every year.


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considerably from 1.47 to 2.15. Table 2 gives some additional information about thefinite element calculations, i.e., element types and sizes used and the stress evaluation

technique applied. The assumed notch factors vary as well.

The SN data in Table 1 are further explained in the right-hand diagram of Fig. 2.

The reference stress range DsR is related to the type of stress mentioned in the second

column. Again, the numbers vary considerably, while the knuckle points of the SN

curves and the slope exponents show less scatter.

As a consequence, the resulting fatigue lives in the last column differ by a factor of

more than ten, i.e. from 1.8 to 20.7 years. The low fatigue lives are somewhat

surprising because the opinion of some experienced designers was that the detail

analyzed is very good and should not show any problems in service. Also, the serviceof container vessels, with even worse details for more than 20 years, has

demonstrated a better fatigue life than predicted here.

Fig. 2. Definitions for Table 1.

Fig. 3. Typical FE-model for calculation of hot-spot stress concentration factor.



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Table 2

Calculation of hot-spot SCFs

Guidelines Element at hot-spot Stress Evaluation

Elem.

type

Size

long./transv./vert.

Weld

modeling

Stress

type

Extrapolation to

surface hot-s

ABS Solid20 30 3030mm3 Triang. v. Mises Linear Linea

BV/RINA Solid8 30 30 10mm3 No Principal Linear Linea

GL [18] Solid20 30 30 30mm3 Prism. Perpend. Linear LineaKR [14] Solid8 30 30 30mm3 No Principal Linear Linea

LR [15] Shell 30 30mm2 Shella Normalb Linear None

NK [16] Solid8 30 30 15mm3 Prism. Principal Linear Linea

RS [17] Solid20 10 10 7.5 mm3 Prism. Perpend. Nonlinear Linea

a45 mm thick elements.bStress component normal to crack plane within 7401 from principal crack plane.


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5. Direct calculation using a spectral method and statistical approach

The fatigue life of the coaming pad detail was further predicted using a linear

spectral method and long-term wave statistics. The analysis is briefly described in thefollowing. A detailed description is given in [19]. The following assumptions were

made:

* hydrodynamic calculations based on strip theory for 50 wave frequencies between

0.2 and 2.5 rad s1

* ISSC wave spectrum and cosine-squared spreading function* Long-term wave data from Global Wave Statistics [20] from relevant North

Atlantic areas* duration at sea 85% of 20 years

* speed profile (24 kn up to Hs=5 m, 20 kn up to 7 m, 15 kn up to 9 m, 10 kn up to11 m and 0 kn above 11 m)

The resulting long-term distribution of nominal stress ranges is characterized by a

Weibull distribution with the following data:

Ds=274.1 MPa

n =5.37 107

x =1.0205

The first two values agree quite well with the assumptions of IACS R 56 [21],whereas the shape parameter x of the Weibull distribution obtained by a linear least-

squares procedure is much higher than that proposed by IACS. Assuming the

rainflow correction factor in [22] as well as FAT 50 together with the associated

design SN curve according to IIW [13], a fatigue life of 5.3 years is obtained.

The effect of several input parameters on fatigue life was investigated in a

parametric study, which is summarized in Table 3. None of these parameters, i.e. sea

area, ships speed, rainflow correction and SN curve with or without change in

Table 3

Results of direct calculations with parameter variation

Sea area Speed Knuckled

SN curve

Rainflow

correction

Dsmax (MPa) nmax (107) x Fatigue life

(yr)

N. Atlantic Profile Yes Yes 274.1 5.37 1.0205 5.3

N. Atlantic Profile No No 274.1 5.37 1.0205 4.0

N. Atlantic Profile No Yes 274.1 5.37 1.0205 4.7

N. Atlantic Profile Yes No 274.1 5.37 1.0205 4.5

Northern N. A. Profile No Yes 290.8 5.18 1.0677 2.9

Pacific Profile No Yes 255.0 5.55 0.9991 5.3

N.A.+Pacific No Yes 267.6 5.46 1.0040 4.6

N. Atlantic Zero No Yes 274.6 5.62 0.9663 4.8N. Atlantic 24 kn No Yes 298.3 5.37 0.9893 3.5

N.A.+Pacific Profile Yes No 267.6 5.46 1.0040 6.2



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slope, had a large effect on the calculated fatigue life, ranging from 2.9 to 6.2 years.

The only major effect found was when the maximum significant wave height was

reduced, i.e., the wave climate became more moderate.

6. Conclusions

Approaches for fatigue strength assessment have been developed during the recent

years and implemented in several classification societies rules and guidelines. These

have been applied, within a comparative study, to the fatigue assessment of a pad

detail on the coaming of a Panamax container ship. Although the loading is quite

well defined in the upper part of the hull girder, the predicted lives vary considerably;

between 1.8 and 20.7 years. It has been shown that the high scatter is attributable todifferent assumptions regarding load effects as well as local stress analyses and SN

curves. In most cases, a simplified approach has been applied using standardized

stress histograms (Weibull distributions) together with the SN approach. The type

of stress varies between nominal, hot spot and notch stress, each coupled with certain

uncertainties regarding the detail classification or determination of stress concentra-

tion factors.

In addition, a direct calculation of the loading has been performed. It showed a

relatively short fatigue life of 5.3 years, assuming the North Atlantic wave climate

and a certain speed profile. In a parametric study several assumptions and input

parameters have been modified, resulting in calculated fatigue lives, ranging from 2.9to 6.2 years, which is in the lower end of the results based on the rules and guidelines

of the classification societies. In conclusion, the situation appears to be

unsatisfactory, particularly in view of the large scatter of result, but also because

of the fact that designers regard the detail as unproblematic with respect to fatigue.

This indicates that several procedures including the direct load calculation yield

overly conservative results.

Acknowledgements

The comparative study was performed as part of the work of Committee III.2,

Fatigue and Fracture, of the International Ship and Offshore Structures Congress

(ISSC) during the working period 19972000. The authors wish to thank those

members of the classification societies, who have actively contributed to the

calculations, as well as the remaining members of the Committee, who have

contributed valuable comments to the study.

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