Nozzles on Loads and Pressure

8/2/2019 Nozzles on Loads and Pressure

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ELSEVIER

1111.1. Pres. V cs . ,\ l'll) illg 72 (1097) I-IXID 1097 Elsevier Sc ience Limited. All rights reservedPrinted In Northern Ireland

()JOX-O 161/97/$17.00II:S0308-016I(97)OOOI4-8

Nozzles-on external loads and internalpressureC. J. Dekker

Continental Engineering B.v .. loan Muyskenweg 22, 1096 Cl Amsterdam, The Netherlands&

H. J. 80SDynafiow Engineering B.v .. Tint/aun 73, 2719 A H Zoetermeer, The Netherlands

(Received 30 January 1997: accepted II February 1997)

Close comparison of local load stress calculation methods reveals considerabledifferences. To investigate we performed many finite element analyses ofnozzles on cylinders concentrating not just on the shell stresses but also on thestresses in the nozzel wall. Local load stresses were sometimes found to bemuch higher in the nozzle than in the shell. This led us to formulate a'modified improved shrink ring method' and to devise multiplication (contour-)charts for deriving local load nozzle stresses from local load shell stresses.Being important for a proper nozzle assessment, pressure induced stresseswere investigated too. This resulted in non-dimensional parameter graphs todetermine pressure induced stresses at nozzles. 1997 Elsevier Science Ltd.

NOMENCLATUREF ,. Radial thrust load on nozzleA( Circumferential moment load on nozzle

(au t-of-plane bending)M, Longitudinal moment load on nozzle

(in-plane bending)P Internal pressurer Mean radius of nozzler o Outside radius of nozzle, r o =r + ~tR Mean radius of vesselsRo Outside radius of vessel, R =R +!TSCF Stress concentration factorI Wall thickness of nozzleT Wall thickness of vesself 3 Relative nozzle size with respect to vessel

size WRC definition: f 3 =0875 X r o / R'}' Relative thinness of vessel, WRC defini-

tion: '}'=R/T(WRC stands for WRC Bulletin 107, seereferences)Stress or stress intensity

1 INTRODUCTIONIn the assessment of nozzles the weakening effect ofnozz.le openings on vessels with respect to internal

pressure is emphasized. This is mainly due to theprominence that design codes place on this aspect. Butreactions from connected piping may give rise to highstresses too and these stresses are in addition to thepressure induced stresses.The stresses due to external loads can be calculated

by various analytical methods, e.g. WRC-107,IAppendix G of BS 55002 and Wordsworth,' but theirresults may differ up to a factor of 2. For a thoroughcomparison of these methods see Dekker."To resolve the question which local load calculation

method gives reliable stresses we made numerousfinite clement analyses of radially placed nozzles oncylinders. Being important for a proper nozzleassessment the pressure-induced stresses were in-vestigated too.

2 ASSESSMENT OF NOZZLES

2.1 General

The raison d'etre for vessels are nozzles, aftcrall whoneeds vessels sealing forever their contents (nuclear


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2 C. 1. Dekker, H. 1. Boswaste industry?), and these nozzles experience twosimultaneously occurring loadings:-pressure-external loads due to piping reactions.

A proper stress assessment of a nozzle requires thesuperposition of the stress systems from both loadings.Then the maximum stress intensities for the variousstress categories are to be determined and comparedwith their specific stress limits. One is referred to e.g.Article 4.1 'Mandatory Design Based on StressAnalysis' in Appendix 4 of AS ME VIII, div. 2,5BS 55002 Appendix A 'Recommendations for designwhere loadings and components are not covered bysection three' or sheet D 1200 'General StrengthAssessment by Analysis' from the Dutch 'Rules forPressure Vessels'."The various stresses occurring at nozzle/vesseljunctions are to be categorized as follows:-membrane stresses (locall) due to internalpressure are primary stresses.

-the bending stresses due to internal pressure arcsecondary stresses.

-membrane stresses (local) due to external loadsbelong to the local primary membrane stresscategory whether the origin of the externalloading is mechanical or thermal. Though in thelatter case it has all the characteristics of asecondary stress.-bending stresses due to external loads belongalways to the secondary stress category.

-stress increments due to concentration in thetransition of vessel to nozzle proper belong to thepeak stress category and need to he consideredonly when a fatigue evaluation is required. Theamount of stress increment is very muchdependent on the actual weld geometry.

The stress intensity limit for primary stresses due topressure will be deemed to he satisfied by judiciouslyfollowing the applicable design code for theconsidered vessel. Assuming that the bending stressesdue to pressure and due to external loads to be atleast as large as the membrane stresses, then one needonly check the stress intensity due to the sum ofprimary stresses and secondary stresses. When thislatter stress requirement is fulfilled then the otherstress requirement (primary membrane stress) isfulfilled too. Note that the sum of primary stresses andsecondary stresses criterion is to prevent low cyclefatigue.If indeed a large number of load cycles occur

then additionally a fatigue assessment is to beperformed.

2.2 Practical implementation

Preferably one should superimpose the stress systemsdue to internal pressure and due to the variousexternal load components and only then determinethe maximum stress intensity. However, such anapproach is feasible only when using FEM-programs(finite Element Method). Having no access to suchprograms (or the timel) for such an analysis, then onehas to resort to more conventional methods.To calculate the external load's stress intensity one

could use for example WRC-l071 or the methodsgiven in G.2.2 and G.2.3 of BS 55002 (in Appendix G).Both WRC-107 as well as Appendix G of BS 5500distinguish between bending stresses and membranestresses and these stresses are given in 4 differentpoints, i.e. the two crown points and the two saddlepoints.However, the stress distribution due to internalpressure at nozzle/vessel junctions is quite another

matter. The only method! known to the authors isEnquiry Case 5500/19 from BS 5500 and this methodgives only the maximum stress intensity due tointernal pressure. No distinction in either stress type(membrane vs bending) or in position can be made.Remarkable is that BS 5500 restricts in clause

A.3.3.2 the sum of the stress intensity due to pressure(calculated in accordance with Enquiry Case 5500/19)and the stress intensity due to external loads(calculated in accordance with G .2.2 and G .2.3) to225 X f with f heing the basic design stress (note that225 X f corresponds with 15 times the yield stress).Normally the stress intensity of primary and secondarystresses is limited by 3 X f (or 2 X yield stress). Incontrast to BS 5500 the bulletin WRC-I07 doesneither mention any stress criterion nor provide anymethod to calculate the stress intensity due topressurc.t

2.3 Improved shrink ring method

Predating the WRC-l07 Bulletin and the BS 5500methods is the shrink ring method first published bythe MW Kellogg Company in their publication'Design of Piping Systems'." The major advantages are

t Admittedly, the Dutch 'Rules for Pressure Vessels" givenin sheet D 1141-Appendix 1 is an approximate method forthe maximum stress intensity due to pressure hut themethod's applicability is restr icted.t Some people advocate to cnter the longitudinal pressureforce in the cylindrical shell forming the nozzle (i.e. J[( rl) - I? Xp) as the radial nozzle load in a WRC-J 07calculation. The resulting stress intensity from such acalculation is thought to represent thc pressure inducedstresses.


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External loads and internal pressure in nozzlesits simplicity of use and its quickness for assessmentcalculations. There is no need to extract data fromnumerous graphs which would lead inevitably tointerpolation.However, comparing the original shrink ring

method with the 'WRC-lOT calculated stress inten-sities show that the stress intensity could be severelyunderestimated. In 'External Loads on Nozzles" animproved shrink ring method is proposed togetherwith a stress limit which takes into account theinternal pressure stresses.Improved shrink ring method:

Y R 7 T(T due to thrust =45 -- x thurstZnr; T

Y R 7 Tif due to MI =1'5--0 -x MIrrroT( ro ) Y R 7 Tif due to A(= 1 + 105,~ --2 - x M eYR. T rrroT

with (T (thrust) + if(M 1 ) + (T(M,J:s I :]Application range is for r o /R : 06 and IO:s (R/T):s100 .

In 'Proper Interface Design for Pressure Vessels'?the improved shrink ring method is combined with theso called load fraction rule. The resulting formulationgives the piping/vessel designer an even better insightinto the relative severity of each loading component:

F ;, act M1 , act }\{, act----+ + :s 1F ;, max M1 , max M e, maxwith:

the actual thrust load on the nozzle.the maximum allowable thrust load ascalculated by means of the improvedshrink ring method and in the absence ofany other external load component.the actual longitudinal moment loading onthe nozzle.

M , max the maximum allowable longitudinal mo-ment as calculated by means of theimproved shrink ring method and in theabsence of any other external loadcomponent.

M e, act the actual circumferential moment loadingon the nozzle.

M e, max the maximum allowable circumferentialmoment as calculated by means of theimproved shrink ring method and in theabsence of any other external loadcomponent.

F r o actF " max

33 NUMERICAL VERIFICATION OFEXTERNAL LOADS

3.1 Introductory notes

As mentioned in the Introduction the widely acceptedexternal load calculation methods do result in quitelarge differences, see Dekker." Trying to establish themore reliable method, we embarked upon 'finiteelement method' analyses. For these analyses we usedthe program FE-Pipe,t a dedicated finite elementprogram for nozzles.The external load on the nozzles always consisted of

a single load component, i.e. either a radial thrustload, a circumferential moment or a longitudinalmoment. The vessel, or better said the cylinder onwhich the nozzles are placed, are either moderatelythick or relatively thin as in the experience of theauthors thick-walled vessels never pose problems withregard to local load stresses. The y-pararneter beingthe ratio of the average cylinder radius and the wallthickness, is taken as 25 for moderately thick-walledvessels and as 50for thin-walled vessels.3.2 Results of cases analyzed

The nozzle geometries analyzed are listed with theirresulting stresses in tables, see the Appendix. For bothy-values a table is given for each specific I/T ratio.With four different I/T values (i.e. 0'5, 0'75, 10 andI '5, respectively) this amounts to eight tables.

In these tables the maximum stress intensity in thecylinder proper due to the three different single loadcomponents, arc given together with the load causingthese stress intensities.

In addition, the so called 'back-calculated' SCFs arelisted which were derived from the followingequations:

Y R 7 TS.1. due to thrust =SCF x -- x F ;2rrril TYRiTS.1. due to MI =SCF X --2 - X MIrrroTYRiTS.l. due to M; =SCF x --) - x M errroT

t This is a proprietary program of the Paulin ResearchGroup, Texas. It is a dedicated FE-program for pipeconfigurations, i.e. for many often occuring problems,so-called mesh-generating templates are provided. For radialnozzles on cylinders the user can suffice with entering themajor geometric data like radii and wall thicknesses ofcylinder and nozzle, respectively and the programautomatically generates the finite element mesh.


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4 C. 1. Dekker, H. 1. BosTo compare with the 'back-calculated' SCFs are theSCFs according to the improved shrink ring method.These latter SCFs arc either 45 (for thrust), 15 (forlongitudinal moments) or

( 1 + H ) 5 ~ )for circumferential moments, see Section 2.3Last entries in the tables are the stress intensities inthe branch (or nozzle) itself and these S.l.s areexpressed as a percentage of the corresponding S.1. inthe vessel.Obviously, the SCFs are functions of the specific

nozzle/vessel geometry characterized by the non-dimensional geometry parameters f 3 and y and the 1 / Tratio. This enables the graphic display of the variousSCFs as functions of the f3-paramcter:-The back-calculated SCFs from the finite clementanalyses are plotted as distinct points. To showmore markedly the I/T dependency, curves aredrawn through 'equal r/?" points.

-The (continuous) SCF-curvcs as derived from theWRC-J 07 bulletin, see Dekker.~

-The SCF functions from the improved shrink rinkmethod and the modified improved shrink ring

8,0

.for VT = 0.50""or VT = 0.75.'or VT = 1.00""orVT= 1.50

7.0

'MIST"6,0

5,0

4.0

3,0WRClor

2.0u,oCI l FE-Pipe results1,0

0,0 o 0,7,2 0,3----p0,1 0,4 0,5 0,6fig. 1. Stress intensity due to thrust for 'Y =25.

8,0

.'or VT = 0.50-e-tor VT = 0.75.forVT= 1.00""orVT= 1.50

7,0

6,0

5,0

4,0

3,0

2,0u,oCI l

[ ' 00,0

FE-Pipe results

0,2 0,3----p 0,70,1 0,5 0,6,4fig. 2. Stress intensity due to thrust for 'Y =50.

method are identified by 'improved' and 'mist'respectively. For the modified improved shrink ringmethod sec Section 5.1.

Six separate graphs are given for thrust, longitudinalmoment and circumferential moment respectively atboth considered v-values (Figs 1-6).3.3 Stress intensities in nozzle necksStudying the tables with numerical results one learnsthat the stress intensity in the nozzle neck proper maydiffer considerably from the stress intensity in the shellat externally loaded nozzles. For nozzle configurationswith t /T = 10 the stress intensities are about thesame, but for t/T < 10 the stress intensity in thenozzle neck becomes larger than the correspondingstress intensity in the shell. For t/T > HI the reverseis true: the nozzle's S.1. becomes less than thecorresponding shell's S.l.The reason for this is of course, that the

discontinuity bending moments in the nozzle neck andin the shell must be necessarily equal at the junction.Note that this moment is not constant along the lengthof the junction! The resulting bending stress from thismoment is inversely proportional to the square of the


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External loads and internal pressure in nozzles8,0

7,0

6,0

5,0

4,0

3,0

2,0u,U(f)

11,0

0,0

FE-Pipe results

WRC10Y"

.. for tIT = 0.50-e-tor tIT = 0.75.for tIT = 1.00. .. for t IT = 1.50

0,2-----p0,1 0,3 0,4 0,5 0,6 0,7Fig. 3. Stress intensity due to circular moment for y =25.

local thickness (i.e. (T =6M /(2) and hence unequalthicknesses of the nozzle and of the shell cause theoverall stress intensities of the nozzle neck and of theshell to differ quite substantially. Although not inproportion to the square of the t / T ratio, as not allconstituent stress components of the stress intensityexhibit this 'square t /T ' behaviour.In order to quantify this potential large stress

intensity raising effect due to unequal nozzle/shellthicknesses, we prepared contour charts of this factoras functions of the t / T ratio and the f3-parameter(Figs. 7-8). As these factors differ only slightly for thethree loading types, i.e. radial thrust, longitudinalmoment and circumferential moment, the contourcharts give the average value of these three factors.Hence the contour charts, given for y =25 and y =50,respectively, are applicable to all three loading types.Having established in one way or another the stress

intensity in the shell at the nozzle/shell junction of anexternally loaded nozzle, the stress intensity in thenozzle wall proper can be obtained by multiplying theformer with the factor read from these contour charts.Where necessary one could interpolate or even

8,0

7,0

6,0

5,0

4,0

3,0

2,0u,U(f)

11,0

WRClOY"

0,0

FE-Pipe results.. for tIT = 0.50-+-for tIT = 0.75.for tIT = 1.00. .. for tIT = 1.50

0,2----p 0,3 0,5 0,6 0,7,4 0,1Fig. 4. Stress intensity due to circular moment for y =SO.

extrapolate in a linear way between y = 25 and y = 50for other v-values.

4 NUMERICAL VERIFICATION OFINTERNAL PRESSURE STRESS4.1 GeneralFor a proper assessment of the stress intensity at anozzle/vessel junction the external load on the nozzleis important, but so is the internal pressure. Afterall,the stress intensity due to primary and secondarystresses from both loadings together is limited to 3 X f(or twice the yield stress). Being able to assess thestress intensity due to external loads is not enough. Itis vital to know too the stress intensity due to internalpressure.Analytical methods to calculate the stress intensity

due to internal pressure are 'Enquiry Case No.5500/19' from BS 55002 and the 'approximate method'from the Dutch 'Rules for Pressure Vessels',il sheetD 1141-Appendix I. However, the method 'EnquiryCase No. 5500/19' is apparently regarded with some


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6 C. 1. Dekker, H. 1. BosB, O

FE-Pipe results

lortIT = 0.50+Ior tIT = 0.75.'or tIT = 1.00""or tIT = 1.50

__ 'IMPROVED' + 'MIST'/" -. . .. . .'WRC107'

7,0

6,0

5,0

4,0

3,0

u ..oC IJ2,0

1,0

0,00,2----po 0,1 0,3 0,6 0,7,4 0,5

Fig, 5, Stress intensity due to longitudinal moment fory =25.

suspicion because, as already mentioned in Section2.2, BS 5500 limits then the total stress intensity to225 XI instead of 3


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where

External loads and internal pressure in nozzles

0,7

< 1.0

0,6

0,5

0,4

0,3

0,2pI0'

0,00,2 0,3 0,4 0,5 0,6 0,7 0,8 0,9----_1fT 1,1 1,2

Fig. 7. Factor for nozzle stress, y =25.

is a function depending on non-dimensional para-meters. The approximate method from the 'Rules' is:

where z is the strength reduction coefficient of thenozzle opening to be determined in accordance withchapter 00501 of the 'Rules'.The expression between square brackets can be

interpreted as a stress concentration factor for thestress intensity due to internal pressure at nozzles.

p.RS.1. pressure =SCFprcsslilcXTTo illustrate the differences between these variousdetermined pressure SCFs. let us consider a vessel,0.0. =2020 mrn and T =20 mm (i.e. 'Y =50), and anozzle with 0.0. nozzle =6R57mm and t =20 mm(i.e. f 3 =03). Assuming the nozzle material to be

7

0,7

0,6

0,5

0,4

0,3

0,2pI 0,'

0,00,2 0,3 0,4 0,5 0,6 0,7 0,6 0,9---_1fT

< 1.0

1,1 1,2

Fig. 8. Factor for nozzle stress, y =50.

equal in strength to the vessel material, then one finds:

-Using the graphs as presented here:SCFpresslire=474

-Based upon enquiry case no. 5500/19:SCF pressure=3AO-Approximate method from the 'Rules':SCFprcssurc=408

-Nozzle's longitudinal pressure force as thrust load ina WRC-I07 calculation and transforming the foundstress intensity into an SCF: SCFprC'ssurc=491

5 DISCUSSION AND CONCLUSIONS

5.1 MIST

Though we think that the performance of theimproved shrink ring method is fair, certainly whenone includes from the contour charts as given in 3.3,the factor for the stress intensity raising effect thatthin nozzle walls have, it may underestimatenevertheless the stress intensity in the shell wall by


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C. 1. Dekke r , H. .I. Bos10,0 10,0

9,0 9,0

8,0 8,0

7,0 7,0

6,0 6,0

5,0 5,0

4,0 4,0

u, u,o 3,0 o 3,0e n e nI 2,0 I 2,01,0 1,00,05 0,1 0,15 0,2 0,25 0,3 0,35 0,4 0,45 0,5 0,05 0,1 0,15 0,2 0,25 0,3 0,35 0,4 0,45 0,5P P

Fig. 9. Maximum stress intensity due to pressure forR/T =20.

about 25%. In case one prefers to use a simple localload stress method which is conservative over itsentire range, then we recommend using the ModifiedImproved Shrink ring method (or MIST for short),which was devised just for that purpose:

VRft(J' due to thrust =60 X -- X thrust2 nr o T

VRfta due to MI = 15 x--o-x MInroT

with a (thrust) + (r(M1 ) + lr(MJ oS 1 xl/factorand factor obtained from the contour charts as givenin 3.3, but not less than 10.Range of application of this MIST approach is

10 ~ y( = R/T) ~ 100 with ro /R ~ 08 (or f 3 ~ 0,7).

Fig. 10. Maximum stress intensity due to pressure forR/ T =30 .

5.2 Proposed design and assessment method fornozzlesNozzles which will experience external loads fromconnecting (process) piping should of course conformto the applicable design code.Next, with the internal pressure SCF-graphs one

should establish that the S.1. due to internal pressureis not more than 2 X f (twice the design stress) and soleaving a stress margin of I X I for external loadstresses. Where necessary, nozzle and/or vesselthicknesses are to be increased to achieve this.Then once the external load on the nozzle is known

(by means of a pipe stress analysis or otherwise) thestress intensity due to this load is to be assessed bymeans of a suitable method, e.g. improved shrink ringmethod or the MIST-approach. Hereby one shouldtake into account the sometimes higher stress levels inthe nozzle necks, i.e. to include the multiplicationfactor from the contour charts. This external load'sS.1. is not to exceed 1 X f as this is the margin left bysetting the pressure's S.1. at 2 X f.RemarkIt is possible to reserve a larger stress range than just1 X f for the external load's S.I., but then one has to


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External/oads and internal pressure in nozzles

10,0

9,0

8,0

7,0

6,0

5,0

4,0

IL& l 3,0

2,0

1,00,05 0,35 0,4 0,45 0,5,1 0,15 0,2----p 0,25 0,3


lower simultaneously the stress range for the internalpressure's S.I., as the limit for the combined S.I. isfixed at 3 X [. In our experience the suggested 2 X f forthe internal pressure is not a bad choice as the S.l.s ofrealistic nozzle loads seldom exceed 1 X f In this waythe vessel design can be finalized before the pipestress analysis reveals the actual external load on thenozzle. In the rare cases that the nozzle load provestoo large then pipe lay-out changes have to solve theproblem.5.3 Conclusions-The local load stresses in cylindrical shells at nozzleswith external loads are dependent too on the ratioof nozzle wall thickness and shell wall thickness.The thicker the nozzle wall is in comparison withthe shell wall, the more the nozzle behaves like arigid insert and, consequently, locally steeperdeformation gradients in the shell wall occur andhence the larger the stresses in the shell wall are.This effect is reflected neither in WRC-l07 nor inthe improved shrink ring method. The sometimesup to 25% underestimation of stresses by the

9

10,0

9,0

8,0

7,0

6,0

5,0

4,0

1,00 ,_05_ 0_,1_ 0_,1.5 P 0,2 0,25 0,3 0,35 0,4 0,45 0,5


improved shrink ring method made us formulate themodified improved shrink ring method.

-The stress intensity from local load stresses due toexternally loaded nozzles is sometimes larger in thewall of the nozzle than in the shell's wall. This is thecase when the thickness of the nozzle wall is lessthan the thickness of the shell. This stress raisingeffect is more or less the same for thrusts,longitudinal moments and circumferential moments.To quantify this effect we devised contour charts forthis magnification factor which apply equally to allthree load types. Note that these contour charts areindependent of the used local load stress calculationmethod and could be used, if so desired, inconjunction with e.g. the 'Appendix G' method ofBS 5500. Needless to say that we recommend the fareasier (modified) improved shrink ring method.

-The bending stresses are always larger than the(local) membrane stresses for both internal pressureas well as external loads. Though the membranestresses are not included separately in the tableswith FE_Pipe results we do confirm the correctnessof this assumption made in Section 2.1 for the wholerange of nozzles considered here.


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10 C. .I. Dekker, H. .I. Bos-The stress intensity due to internal pressure is largerthan predicted by analytical methods like EnquiryCase No. 5500/19 of BS 5500. Here too themaximum stress intensity in the nozzle wall issometimes higher than in the shell wall, this is thecase when the thickness of the nozzle wall is lessthan the thickness of the shell. As in general theinternal pressure induced S.I. takes up the largerpart of the allowable stress range of 3 :], it isimperative to usc accurate pressure S.l.s. For thispurpose we made graphs of the (maximum) SCFdue to internal pressure for 'Y =20, 30, 40 and 50,respectively.

-Following our proposed design/assessment methodfor nozzles, i.e. using the here-given pressure graphsand contour charts in conjunction with either theimproved shrink ring method or the MIST-approach, will result in soundly designed nozzles.Admittedly, our proposed design method issometimes a bit conservative in that it adds upalgebraically the various load component's maxi-mum stress intensities which do not necessarily occurat the same position, but the resulting nozzle designwill be safe from low cycle fatigue. In the exceptionalcase that you are designing a pressure vessel from goldthen use FE-techniques in order to save as muchmaterial as possible, but otherwise use the relativelysimple and quick design method for nozzles asproposed by us.

-As shown, the stresses at nozzle/vessel junctionsmay be much higher than predicted by 'conven-tional' methods. Especially due to the stress raisingeffect of a nozzle with a wall much thinner than thevessel wall. This applies both to pressure inducedstresses as well as to extenal load induced stresses.One might wonder why in the daily practice of plantoperating this has never been noticed throughnozzle failures due to overstressing. Well, in ouropinion the following explanation deals with thatobjection (plus operating conditions being lesssevere than the design conditions and the yieldstresses of materials often heing higher than theirspecified minimum yield stresses).In case such a vessel with a thin walled nozzle does

not experience many pressure-cum-load cycles, thenfailure need not occur as the failure mode is low cyclefatigue for too large 'primary-and-secondary' stressintensities. But if enough cycles occur to result infailure of the nozzle then the true nature of the failureis often not recognized. At the point of failure (orvery ncar to it) will also be the weld between nozzle

and vessel. The plant manager will call in a weldspecialist and that is quite understandable! However,as minor defects and blemishes are always present in aweld, the weld specialist is indeed able to pinpointsuch a weld irregularity and blame it for having causedthe failure. After welding a new nozzle in the vessel,the vessel will operate well for many years to come asthe plant's operators are experienced by now. Thesmooth way of operating the plant results in few(start-stop) load cycles and low cycle fatigue does notget a second chance.

I t is our opll1lOn that many nozzle failuresmasquerade as weld defects while in reality the nozzledesign (or better said: the nozzle design method) is toblame. One would be wise to also consult a stressspecialist at nozzle failures to establish the true causeof the failure!

ACKNOWLEDGEMENTMuch of the here-presented material is from studiesundertaken in cooperation with the authors' COI11-panies under commission by NAM (Assen, TheNetherlands). We would like to thank Mr W. 1.Stikvoort of NAM, business unit Groningen, forallowing us to use this material and for hisencouragement in preparing this paper.

REFERENCES1. Wichman, K. R., Hopper, A. G. and Mershon, J. L.,Local stresses in spherical and cylindrical shells due toexternal loadings. WRC Bulletin 107/ August 1965,

Revision March 1979.2. BS 5500: 1991, Specification for Unfired Fusion WeldedPressure Vessels. British Standards Institution, London,

1991.3. Wordsworth, A.C., Stresses in cylindrical pressure vesselsdue to local loads. In Structural Integrity Assessment, ed.P. Standley. Elsevier Applied Science, London, 1992.

4. Dekker, C. J., Comparison of local load stress calculationmethods for nozzles on cylinders. lnt . .I . Pres. Yes. &Piping, 1994, 58,203-213.

5. ASME Boiler and Pressure Vessel Code, SectionVIII-Division 2, 1995 edition. The American Society ofMechanical Engineers, New York, 1995.6. Rules for Pressure Vessels, Issue 96-02. Published onbehalf of Stoomwezen B.V. by SOU Publishers, TheHague.7. The MW Kellogg Co., Design of Piping Systems. 2ndedition. John Wiley, New York, 1956.

8. Dekker, C. J., External loads on nozzles. lnt . .I . Pres. Yes.& Piping, 1993, 53, 335-350.

9. Stikvoort, W. J., Proper interface design for pressurevessels. Chemical Engineering, 1994, 133-134.

APPENDIX: TABLES WITH NUMERICAL RESULTS FROM FE-PIPE CALCULATIONSTables with the main results from our FE-Pipe analyses, for both external loads as well as internal pressure, arcincluded here for reference purposes.There are eight external load tables and four internal pressure tables,


11/18

External loads and internal pressure in nozzies 11Local load stresses, i.e. FE-pipe versus the improved shrink ring method

Gamma ratio: 2500 Mean shell radius: 40000 mrnt/T ratio: 050 Thickness of shell: 1600 mrnThickness of nozzle: 8.00 mrnIdentification Results from 'Ffi-pipc ' calculations SCF acc. Stress in branch

improvedLoad type Load S.1. SCF (back shrink S.1. Stress in(kN. kNm) (MPa) calcula ted) ring method (MPa) header (%)

beta 0100 Thrust 50 171 3143 4-500 211 1234r.o [mm] 45714 Mi.circ. 50 3997 1697 1600 5338 1336Mi.long. 50 3358 1411 15()() 4254 1267beta 0150 Thrust 50 135 3723 4500 187 1385r_o[mm] 68571 Mi circ, 50 2471 2336 1900 3606 1459M_long. 50 1681 1589 1500 2354 1400beta 0200 Thrust 50 103 3787 4500 176 1709r_o(mm] 91-429 Mi.circ, 50 1647 zzes 2200 2852 1732

Mi.long. 50 949 1595 1500 1428 1505beta 0250 Thrust 50 110 3677 4500 145 1813I'_O [mm] 1142116 Mi.circ, 50 1134 2987 2500 20711 1832Mi.Iong. 50 617 1620 1500 979 1587beta 0350 Thrust 50 54 3-474 4500 116 2148r .. o [r nr n] 160000 Mi.circ. 50 622 3202 3100 1375 2211

Me.long. 50 303 1560 15()() 516 1703

Local load stresses, i.e, FE-pipe versus the improved shrink ring methodGamma ratio: 2500 Mean shell radius: 40(HlO mrnt/T ratio: 075 Thickness of shell: 16()0mmThickness of nozzle: 12.0 m rn

Identification Results from 'Ffi-pipe calculations SCF ace. Stress in branchimprovedLoad type Load S.1. SCF (back shrink S.1. Stress in(kN, kNm) (MPa) calculated) ring method (MPa) header (%)

beta 0250 Thrust 50 81 3723 4500 113 1395r.,o [mm] 114286 Mi.circ. 50 989 2597 2500 1452 14611Mi.long. 50 511 1342 1500 657 1286beta 0350 Thrust 50 59 3796 4500 85 1441r_o[mm] 160000 Mi.circ. 50 614 3160 3100 927 1510M_long. 50 254 1307 1500 349 137-4beta 0-450 Thrust 50 44 .l-640 4500 65 1477LO [mm] 205714 Mi.circ. 50 405 3-446 3700 623 1538Mi.long. 50 147 1251 15()() 213 1449beta O550 Thrust 50 34 3-438 4500 50 1471r_o[mm] 251429 Mi.circ. 50 276 3508 4300 4211 1551Mi.long. 50 93 1182 1500 138 148-4


12/18

12 C. 1. Dekker, H. 1. BosLocal load stresses, i.e. FE-pipe versus the improved shrink ring method

Gamma ratio: 25 - o o Mean shell radius: 40000 rnmIT ratio: 1.00 Thickness of shell: 1600 rnrnThickness of nozzle: 16.00 rnrn

Identification Results from 'Ffi-pipe ' calculations SCF acc. Stress in branchimprovedLoad type Load S.1. SCF (back shrink S.1. Stress in(kN, kNm) (MPa) calculated) ring method (MPa) header (%)

beta 0200 Thrust 50 110 4044 4500 97 882r..o [mm] 91429 Mi.circ, 50 1415 2378 2200 1309 925M_long. 50 801 1346 1500 651 813beta O250 Thrust 50 93 4274 4500 79 849LO [mm] 114286 M_circ. 50 1076 2826 2500 968 900Mz.long . 50 535 1405 1500 448 837beta O350 Thrust 50 67 4311 4500 60 896LO [mm] 16000D Mi.circ. 50 67l 3454 3100 628 936Mi.long. 50 267 1374 1'500 233 873beta 0450 Thrust 50 51 4219 4500 46 902LO [mm] 205714 Mi.circ. 50 454 3863 3700 426 938Mi.long, 50 154 1310 1500 144 935beta ()550 Thrust 50 38 3842 4500 34 895r.,o [mm] 251429 Mi.circ. 50 306 38S9 4,300 2S1 9J.SM_long. 50 9S 1246 I 500 93 949beta 0700 Thrust 50 28 3 603 4500 27 964LO [mm] 320000 Mi.circ, 50 190 3912 5200 IS3 963Mi.long. 50 63 1297 1500 63 1000

Local load stresses, i.e, FE-pipe versus the improved shrink ring methodGamma ratio: 25 00 Mean shell radius: 40000 mrnt l T ratio: 150 Thickness of shell: 16()()mm

Thickness of nozzle: 24.00 mmIdentification Results from 'Ffi-pipe' calculations SCF aec. Stress in branchimprovedLoad type Load S.1. SCF (back shrink S.1. Stress in(kN, kNm) (MPa) calcula ted) ring method (MPa) header(%)

beta O200 Thrust 50 117 4302 4500 54 462LO [mm] 91-429 Mi.circ. 50 1412 2373 2200 691 489Mi.long. 50 789 1326 I 500 356 451beta 0250 Thrust 50 98 4504 4500 45 459r.,o [mm] 114286 Mi.ci rc , 50 1085 2849 2500 520 479Mi.long. 50 533 IAOO 1500 250 469beta 0350 Thrust 50 72 4'632 4,500 34 472r_o[mm] 16(HlOO l'vLcirc. 50 707 3'639 3100 339 479l'vLlong. 50 274 1410 1500 133 485beta 0450 Thrust 50 54 4A67 4500 26 481r..o [mm] 205714 l'vLcirc. 50 475 4042 3700 232 488l'vLlong. 50 156 1327 I .5(lO 83 532beta 0550 Thrust 50 41 4145 4500 19 463r.,o [mm] 251429 Mi.circ. 50 326 4144 4300 157 482M_long. 50 103 1309 1500 55 534beta 0700 Thrust 50 33 4246 4,500 16 4S'5r_o[mm] 32(J.(lOO l'vLcirc. 50 221 4-5.50 5200 106 480Mi.long. 50 71 1462 1500 40 563


13/18

External loads and internal pressure in nozzles 13Local load stresses, i.e, FE-pipe versus the improved shrink ring method

Gamma ratio: 5000 Mean shell radius: HOO'OOmmur ratio: 050 Thickness of shell: 1600 mmThickness of nozzle: H.OOmmIdentification Results from 'FE-pipe' calculations SCF ace, Stress in branchimproved

Load type Load S.1. SCF (back shrink S.1. Stress in(kN, kNm) (MPa) calculated) ring method (MPa) header ( ' Y o )beta 0100 Thrust 50 142 3692 4500 231 1627r_o[mm] 91429 Mi.circ. 50 1654 1966 1849 2917 1764Mi.long. 50 1269 150H 1500 196H ISS,]beta 0150 Thrust 50 101 3939 4500 202 2000r.,o [mm] 137143 Mi.circ. 50 9HO 2621 2273 2014 2055M_long. 50 584 1562 1500 9H8 1692beta 0195 Thrust 50 76 3H53 4500 163 2145r..o [mm] 178286 Mi.circ. 50 620 2802 2655 1398 2255M_long. 50 337 1523 1500 620 IH40beta 0250 Thrust 50 56 3640 4500 139 2482r.,o [mm] 228571 Mi.circ. 50 410 3045 3121 1049 2559Mi.long. 50 215 1597 1500 380 1767beta O350 Thrust 50 40 3640 4500 101 2525r .i o jm m] 32000 Mi.circ. 50 230 3348 3970 647 2813M_long. 50 lOS 1529 1500 207 1971

Local load stresses, i.e. FE-pipe versus the improved shrink ring methodGamma ratio: 5000 Mean shell radius: HOOOOmmur ratio: 075 Thickness of shell: 1600mm

Thickness of nozzle: 12.00 mrnIdentification Results from 'FE-pipe' calculations SCF ace. Stress in branch

improvedLoad type Load S.I. SCF (back shrink S.1. Stress in(kN, kNm) (MPa) calculated) ring method (MPa) header (%)beta 0250 Thrust 50 70 4550 4500 101 1443r..o [mm] 22H571 Mi.circ. 50 46H 3476 3121 697 14H9Mi.long. 50 194 1-441 1500 257 1325beta 0350 Thrust 50 50 4550 4500 72 1440r .i o jm m] 320,000 Mi.circ. 50 286 4164 3970 427 1493Mi.long. 50 95 1383 1500 136 1432beta 0450 Thrust 50 36 4212 4500 52 144-4Lo[mm] 411-429 Mi.circ. 50 IH4 4428 4HIH 275 1495M_long. 50 57 1372 1500 80 14004beta 0550 Thrust 50 27 3861 4500 39 14404r.,o [mm] 502857 IvLcirc. 50 122 4386 5667 182 1492Mi.long. 50 37 1330 1500 53 1432


14/18

14 C. J. Dekker, H. J. BosLocal load stresses, i.e. FE-pipe versus the improved shrink ring method

Gamma ratio: 5000 Mean shell radius: ROOOOmmur ratio: 100 Thickness of shell: 1600 mmThickness of nozzle: 16.00 rnrnIdentification Results from 'Ffi-pipe ' calculations SCF acc. Stress in branchimproved

Load type Load S.1. SCF (back shrink S.I. Stress in(kN, kNm) (MPa) calculated) ring method (MPa) header (%)beta 0200 Thrust 50 99 5147 45()() R2 R28LO [mm] 182857 Mi.circ. 50 706 3356 2697 616 873Mi.long. 50 322 1531 1500 257 798beta 0250 Thrust 50 80 5199 4500 69 863r..o [mm] 228571 Mi.circ. 50 512 3803 3121 461 9(H)M_long. 50 203 1508 1500 173 852beta 0350 Thrust 50 56 5095 4500 49 875r.,o [mm] 320000 Mi.circ. 50 313 4557 3970 285 911M_long. 50 103 1500 1500 90 K74beta 0-450 Thrust 50 41 4796 4500 34 829LO [mm] 411-429 Mi.circ. 50 205 4934 4818 176 859M_long. 50 60 1-444 1500 53 883beta 0550 Thrust 50 31 4-433 4500 24 77-4LO [mm] 502857 Mi.circ. 50 138 4961 5667 III 80-4Mi.long. 50 38 1366 1500 34 895beta 0700 Thrust 50 22 4004 4500 18 818LO [mm] 640000 Mi.circ, 50 79 4600 6940 6K K6 1Mi.long. 50 25 1-456 1500 22 880

Local load stresses, i.e. FE-pipe versus the improved shrink ring methodGamma ratio: 5000 Mean shell radius: 80000 mmur ratio: 150 Thickness of shell: 1600 mrnThickness of nozzle: 24.00 rum

Identification Results from 'Ffi-pipe calculations SCF ace. Stress in branchimproved

Load type Load S.1. SCF (back shrink S.1. Stress in(kN,kNm) (MPa) calculated) ring method (MPa) header ( l r)beta 0200 Thrust 50 107 5563 4500 44 411r..o [mm] 182857 M_circ. 50 735 3-494 2697 322 438M_long. 50 327 1554 1500 140 42Rbeta 0250 Thrust 50 87 5654 4500 37 425r..o [mm] 228571 M_circ. 50 545 4048 3121 241 442

Mi.long. 50 212 1575 1500 93 439bcta 0350 Thrust 50 61 5550 4500 27 443r..o [mm] 320000 M_circ. 50 335 4877 3970 152 454M_long. 50 107 1558 1500 48 449beta 0-450 Thrust 50 45 5264 4500 18 40'()r..o [mm] 411-429 Mi.circ. 50 220 5295 4818 94 427M_long. 50 62 1-492 1500 31 500beta o 550 Thrust 50 34 4R61 450(J 13 382LO [mm] 502857 Mi.circ. 50 149 5357 5667 61 409Mi.long. 50 42 1510 1500 21 50 nbeta 0700 Thrust 50 25 4550 4500 10 400LO [mm] 64(JOOO Mi.circ. 50 92 5358 6940 38 413M_long. 50 27 1572 1500 13 481


15/18

External loads and internal pressure in nozzlesNozzles in cylindrical shells under internal pressure

R IT -parameter =20Internal pressure = 10 MPaAll dimensions in table below are in [mm] and all stresses are given in [MPa]

15

0.0.

Nozzle dimensions Vessel dimensions0.0.

Geometry parametersT sto it reta Stress in

header

FE-pipe resultsSCF backcalculated

MaximumstressStress in Percentage intensity

branch of stress as SCFin header

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16/18

16 C. .I. Dekker, H. .I. BosNozzles in cylindrical shells under internal pressure

RI T -parameter = 30Internal pressure = 10 MPaAU dimensions in table below are .in [mm] and all stresses are given in [MPal

O.D.

Nozzle dimensions Vessel dimensionsO.D.

Gcomctrv parametersT .uo I/Teta Stress inheader

Ff.-pipc resul tsSCF backcalculated

Stress inbranch

Maximumstress

Percentage intensityof stress as SCFin header

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17/18

External loads and internal pressure in nozzlesNozzles in cylindrical shells under internal pressure

R IT -parameter =40Internal pressure = 10MPaAll dimensions in table below are in [mm] and all stresses are given in [MPaJ

17

0.0.

Nozzle dimensions Vessel dimensions0.0.

Geometry parametersT to ureta Stress in

header

Fli-pipc resultsSCF backcalculated

MaximumstressStress in Percentage intensity

branch of stress as SCFin header

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18/18

18 C. 1. Dekker , H. 1_ 80SNozzles in cylindrical shells under internal pressure

R/T -parameter = 50Internal pressure =10 MPaAll dimensions in table below are in [mm] and all stresses are given in [MPa]

O _ D _Nozzle dimensions Vessel dimensions

O _ D _Geometry parameters

T sto itreta Stress inheaderFE-pipe results

SCF backcalculated

Maximumstress

Stress in Percentage intensitybranch of stress as SCF

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Documents

Nozzles on Loads and Pressure