An alternative way to support horizontal pressure vessels subject tothermal loading

Alwyn S. Tootha,*, John S.T. Cheungb, Heong W. Ngb, Lin S. Ongb, Chithranjan NadarajahcaDepartment of Mechanical Engineering, University of Strathclyde, Glasgow G1 1XJ, UK

bSchool of Mechanical and Production Engineering, Nanyang Technological University, Singapore 639798, SingaporecExxon Engineering Asia Pacific, Singapore 048693, Singapore

Received 27 May 1998; accepted 9 June 1998

Abstract

When storing liquids at high temperature in horizontal vessels, the current design methods aim to minimise the thermal stresses byintroducing a sliding surface at the base of one of the twin saddle supports. However, regular site maintenance is required to ensure thatadequate sliding is achieved. This may be difficult and costly to carry out. The aim of the present work, therefore, is to dispense with thesliding support and to provide saddle designs which, although fixed to the platform or foundation, do not result in the storage/pressure vesselbeing overstressed when thermal loading occurs. This paper provides general recommendations for the most appropriate saddle geometries,and details the way in which design-by-analysis and fatigue-life- assessments may be carried out using the stresses that arise from thesedesigns. q 1998 Elsevier Science Ltd. All rights reserved

Keywords: Pressure vessels; Storage vessels; Supports; Thermal loading

1. Nomenclature

A length of vessel beyond saddle overhangbp breadth of the saddle top plate in the axial directiondp basic saddle widthE elastic modulus of the vessel and saddle materialh e overall heat transfer coefficienthp height of the saddle, measured from nadir of the vessel to

saddle base platek s thermal conductivity of steelL s length of the vessel between supportsrm mean radius of the vesselt c wall thickness of the vesselt s thickness of the saddle web and stiffenersS I stress intensity (i.e. maximum principal stress difference)DT temperature differentialw extended width of the saddle top plate

a linear coefficient of thermal expansion of the vessel and saddlematerial

2. Introduction

Horizontal cylindrical storage/pressure vessels as used inthe power, petroleum and other process industries are

designed according to recognised codes and standards (forexample; the ASME, BS, CODAP, etc. pressure vessel stan-dards) to withstand both the test and operating conditions.The common practice, in terms of support, is to provide twosaddle-like supports symmetrically located along the lengthof the vessel. To avoid induced axial restraint stresses, as inthe case of thermal loading, the codes recommend that oneof the saddles be free to slide in the axial direction. This canbe achieved in a number of different ways: by the use offoundation bolts positioned in slotted holes, by the use oflow-friction material (such as polytetrafluoroethylene(PTFE)) bearing pads bonded to the backing plate of thesaddle base and the foundation plate, or by the introductionof a roller at the base of the support.

A further recommendation is that the hot liquid storagevessel and the supports be fully insulated. Such a require-ment is obviously necessary to prevent heat loss whichcould be detrimental to the process, to prevent fire damagethereby inducing structural weakness in the vessel and thesupport, and to protect personnel from inadvertent contactwith the vessel or the support. Such insulation also avoids ahigh rate of temperature loss down the saddle itself whichinduces correspondingly high values of vessel stress.

In spite of the long-standing practice of providing a slid-ing surface (or roller), mistakes can and do occur in practice.

0308-0161/98/$ - see front matter q 1998 Elsevier Science Ltd. All rights reservedPII: S0308-0161(98)00065-9

* Corresponding author. Tel.: +44-141-552-4400; Fax: +44-141-552-5105

International Journal of Pressure Vessels and Piping 75 (1998) 617623

IPVP 1855

Vessels are occasionally found where free sliding of themovable saddle cannot occur. This may be because slottedholes are inadvertently not provided, or the slotted holes thatare provided are installed on the wrong side of the saddlebase plate, thus preventing any movement when thermalexpansion takes place. Another common experience withthe slotted hole type of installation is that the nuts areover-tightened and become rigidly fixed in position, or alter-natively the slotted hole is totally filled with concrete debrisor the sliding surface region of both the saddle and mountingare rigidly corroded together.

In view of this, there are advantages from the operational,maintenance and safety point of view in dispensing entirelywith the sliding support and installing the vessels with bothsaddles fixed at their bases. The bases may be bolted to thefoundation mounting in on-shore ground foundations, orwelded to off-shore platform beams or the deck of a ship.Having taken this step, the element of uncertainty isremoved from the design approach but one is now facedwith the requirement to design the saddles and vessels tocarry the total value of the thermal stresses that arise fromthe restraint of the saddle feet. This paper briefly reports thefindings of an extensive investigation, both experimentaland theoretical, which provides the designer with the neces-sary tools to carry out this process.

3. The storage of hot fluid in horizontal vessels

When a horizontal cylindrical vessel is used to store hotfluid and is installed so that the saddles are fixed to the deckbeams of a platform or the foundation, the thermal expan-sion of the vessel is restrained. If the hot liquid onlyoccupies the lower part of the vessel then non-uniformheating of the vessel will occur. A further complexitycould occur if during filling the hot fluid impinges rapidlyon the surface of the vessel in a local region. In this case, atransient analysis of a local hot spot may be required toanalyse the problem. To avoid the complexity of thesecases, the following assumptions are made.

1. The hot fluid is inserted slowly so that the problem maybe considered as steady state.

2. The whole vessel is heated uniformly by hot liquid to atemperature DT above ambient.

3. The vessel and the supports are fully insulated to avoidheat loss from the vessel wall and the saddles; this alsoserves the purpose of protecting plant personnel andreducing acoustic noise.

3.1. The radial expansion of the vessel

During storage of the hot fluid, the vessel expands bothlongitudinally and radially. The radial expansion occursover the whole vessel. It is, however, restrained locally inthe region of the support due to the wrap-around effect of

the saddle, where an embracing angle of at least 1208 isused. The thermal stresses occur because there is a tempera-ture gradient from ambient at the saddle base to the vesseltemperature at the uppermost point of contact, known as thehorn of the saddle. This radial restraint causes local bendingmainly in the circumferential direction at the uppermostregion of the saddle top plate, similar to that caused byinternal pressure.

3.2. The longitudinal expansion of the vessel

The longitudinal expansion of the vessel is also locallyrestrained, in this case by the non-sliding saddle support.Because of this, a horizontal force and fixing moment areinduced at the base of the support by the deck beams of theplatform. If the fixing is considered to be totally rigid thenthe value of this force and moment will be such that thehorizontal displacement and rotation will both be zero at thebase fixture. In practice, the platform deck will contain someflexibility so that these forces will be somewhat less thanthose in the rigid case. However, from a design point ofview it is considered that the full force and moment con-straint loadings should be assumed such that the base dis-placements are zero.

When the system was analysed, it was found that thelongitudinal restraint resulted in higher vessel stressesthan those due to the radial restraint. When an ideal slidingsupport was introduced at the base of one of the saddles, thethermal axial (longitudinal) and circumferential stresseswere substantially reduced. This fact does, of course, vali-date the existing recommendations of the code for thosecases where an ideal support can be guaranteed.

4. Typical saddle designs

Three widely used saddle designs were investigated in arange of preliminary studies. The details, given by Cheunget al. [1], indicate that for the smaller vessels, up to andincluding 1200 mm in diameter, a design based on thatgiven in the pressure vessel design handbook by Megyesy[2] is appropriate. However, certain modifications to thisdesign are proposed to provide greater flexibility at the sad-dle horn and over the saddle width. The proposed design isshown in Fig. 1.

For larger diameter vessels, from 1200 to 3000 mm insize, a design on based that given in a dimensional BritishStandard (BS 5276:1983 [3]) is proposed. Again, modifi-cations to this design are suggested in which the width, w,is optimised to increase the flexibility across it. A typicaldesign is shown in Fig. 2.

The saddle design by Megyesy [2] is a relatively simpleone and is useful for the support of smaller diameter vessels.In the original design (see Ref. [2]) this saddle does notprovide much radial flexibility in the horn region. As a resultof the current studies, it is proposed that this dimension be

618 A.S. Tooth et al./International Journal of Pressure Vessels and Piping 75 (1998) 617623

increased to correspond to an angle of 128 on both sides andan extended width, w, of 25.4 mm again on each side; seeFig. 1. It is also noted in Ref. [2] that for vessels up to1200 mm in diameter, a stiffening rib is not provided inthe lower vessel region.

The saddle design proposed for the larger diametervessels, shown in Fig. 2, is based on that in BS 5276 [3].This design includes an element of radial flexibility in the

horn region, since the central web is not stiffened by an endplate as in the case of the Megyesy saddle. The width of thesaddle top plate is, however, increased so that the ratio of thebasic saddle width, dp, and the extended width, w, is main-tained at 6.25 over the range of vessel diameters 12003000 mm. The value of this ratio was found to provideoptimum flexibility and correspondingly reduced thermalaxial and circumferential stresses in the vessel. For this, it

Fig. 1. Saddle design based on Megyesy handbook [2].

Fig. 2. Saddle design based on British Standard BS 5276:1983 [3].

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was found that the maximum thermal stress in the vessel didnot occur at the horn of the saddle and, therefore, was inde-pendent of the angular extension. This value was thus fixedas the average of those contained in BS 5276 [3], namely 78,as shown in Fig. 2.

5. Analytical proceduresnumerical and experimental

5.1. The finite element method

The vessels were analysed using the finite elementmethod with 20-noded brick elements (ANSYS software).A temperature differential of 1008C was achieved bymaking the internal vessel surface temperature 1008Cwhile the base of the saddle was assumed to be at a tem-perature of 08C. In the first instance, the outside surfaces ofthe vessel and saddle were assumed to be perfectlyinsulated so that there was no heat loss to the surroundings.The heat was therefore transferred by conduction from thevessel to the base support. Later work, reported below inSection 8, examines the influence of heat transfer by radia-tion and convection.

A small displacement linear elastic analysis was carriedout in two phases. In the first instance, the heat transferproblem was addressed to determine the temperature distri-butions in the vessel and saddle. These values were used inthe second phase to determine the thermal stresses. Furtherdetails of this work are given by Cheung et al. [1], Toothet al. [4] and Ong et al. [5].

5.2. Experimental studies

Experimental studies have been conducted on smallcylindrical storage vessels (1035 mm long, 228 mm ininside diameter, and a 2.1 mm wall thickness). These weresupported on both the Megyesy and BS type saddles, whichincorporated the proposed modifications, referred to above.They were extensively strain gauged and subjected to twotypes of tests to verify the validity of the finite elementnumerical analysis. The axial restraint was investigatedusing an isothermal saddle base pushpull displacementtest. The combined radial and axial restraint was exploredby progressively heating the vessels with hot liquid. Theoverall conclusion of these studies was that the finiteelement analysis gave reasonable predictions of the strainsin the experimental vessels and could be used with some

confidence to carry out further parametric studies. Fulldetails of this work are given by Ng et al. [6]

6. Parametric investigations

Using the finite element analysis described above, anextensive number of finite element runs were carried outon the modified designs of two saddles shown in Fig. 1and Fig. 2. In this investigation, all the geometric featuresof the saddles together with a complete range of vesselgeometries were examined. For example, in the Megyesyand British Standard saddles, 900 and 729 different FEAruns were carried out respectively. Details of this aregiven by Ong et al. [5]. In each FE run, the maximum stressintensity, the absolute maximum circumferential stress andthe absolute maximum axial stress were determined for eachgeometric configuration. These studies are an extension ofthe work reported by Tooth et al. [4].

In these analyses, the radial stresses, i.e. normal to thewall of the vessel, were found to be small compared to thecircumferential and axial stresses. In view of this and sincethe aim of the work was to provide stress values for fatigueassessment, they were not considered further.

Using a least-squares curve fitting procedure, parametricequations for the above maximum stresses have been estab-lished. In this work a power law relationship has been usedin which dimensional groups of the leading vessel and sup-port parameters are given. The details of this work are givenby Ong et al. [5]. After a study of 15 power series expres-sions for their quality of fit, error estimates and consistency,the best non-dimensional expression for the maximum valueof the stress intensity that occurs in the Megyesy design wasconsidered to be

SIEaDT

abprmtc

p !b hprm

c

Lsrm

d

Armtc

p !e

tstc

f

wrmtc

p !g 1where the constant a and the indices bg are given (for threevalues of A/rm) in Table 1.

Similar values were obtained for the absolute maximumcircumferential and axial stresses based upon the Megyesyand the British Standard saddles, suitably modified asabove.

Table 1Constant a and indices bg in Eq. (1)A/rm a b c d e f g0.5 0.05091 1.42627 0.80956 0.76971 1.05420 0.43872 0.526121.0 0.08209 1.19043 0.69126 0.82142 0.88672 0.55078 0.404302.0 0.14128 1.46387 0.61416 0.86332 1.166211 0.74121 0.56641

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To further improve the accuracy of fit, trend lineequations were obtained into which the values obtainedfrom the parametric equation can be substituted. Full detailsof the method and the appropriate equations are given byOng et al. [5]. A more sophisticated curve fitting procedurefor these results has been obtained by Fok and Tooth [7], inwhich some 54 coefficients were derived to provide a closedform equation over the whole geometric field covered by theFEA results. The intention is to make these results availableon the Internet; University of Strathclyde, Department ofMechanical Engineering, website.

7. General conclusions from the parametric survey

From the parametric and trend line equations, it ispossible to draw a number of general conclusions which,if they are incorporated, will reduce the value of the thermalstress produced in the vessel.

1. Tall saddles should be used since they introduce axialflexibility. However, in such cases care should be exer-cised to design the saddle webs and stiffeners to avoidpanel buckling.

2. The basic saddle width dp should be as narrow as possi-ble in order to reduce the radial displacement (and thusthe stresses) which occurs at the sides of the saddle platewhen the restraining moment loading is applied to thesaddle base.

3. In contrast to conclusion 2, the saddle top plate (or wearplate) should be extended beyond the basic width, dp,both in the circumferential direction at the horn and alsoacross the width to introduce flexibility and therebyreduce saddle reactive interface and thus thermalstresses.

4. The distance between the two fixed saddle supportsshould be reduced as far as possible, thereby reducingthe value of the axial thermal expansion to be restrained.However, this must be consistent with acceptable valuesof the membrane longitudinal stresses at the mid-spanand saddle support profiles when the vessel is used forstoring liquid.

5. Ideally, the saddle embracing angle should not exceed1208, since using the smaller angle of support providesflexibility to rotational movement.

8. Insulation of the vessel and saddle

In the work reported above, it was assumed that both thevessel and the saddles were fully insulated and did not loseheat to the surrounding atmosphere. That is, the heat capa-city of the hot liquid was transferred to the base of thesaddles by means of conduction only.

During discussions with an industrial support group, setup in Singapore by NTU to monitor the project, it waspointed out that it was not normal practice to fully insulate

the support regions. The insulation was restricted to thevessel itself, where heat loss could be detrimental to theprocess. The supports were, however, protected with a fire-proofing concrete material to help preserve the structuralintegrity of the steel in the event of a fire local to the vessel.Such protection has the additional advantage of reducing therisk of injury by personnel inadvertently coming into con-tact with the supports and thereby sustaining a burn.

It was considered that the fireproofing material used toprotect the saddle would be less effective in providing goodinsulation than conventional insulation material, with theresult that convection and radiation could occur from theouter saddle surface. Analysis of this case results in a tem-perature profile down the saddle and particularly in theimmediate region of the saddle/vessel junction, which ismore detrimental to the occurrence of thermal stressesthan the temperature profile derived under adiabatic condi-tions. The influence of such changes in the heat transferbehaviour was examined in detail in Ref. [8] for a rangeof vessels supported on the Megyesy saddle.

In these heat transfer studies, the thermal conductivity forthe steel was retained at ks 45 W m 1 K 1. Further, anoverall heat transfer coefficient, he, was determined assum-ing a convection heat transfer coefficient, a radiation heattransfer coefficient, and using a thermal conductivity for thefireproofing concrete of 0.1 W m1 K1. For the values con-sidered, the magnitude of he was found to be1.83 W m2 K1, as opposed to zero for the pure conductioncase. When this value was used in the analysis for a saddleof hp=rm 2 the stress intensity values were increased fromthat assuming pure conduction, by values ranging from 6.6to 16.7% as the ratio of Ls/rm was reduced from 12.0 to 4.0.Detailed results from this work are given by Tooth et al. [8],from which stress values for a range of heat transfercoefficients may be obtained for the full range of vesselgeometries.

It is sufficient to comment here that some form of goodinsulation is essential in the saddle/vessel region. If therewere a total absence of fireproofing concrete altogether,either by design or by damage, then the outer surface ofthe steel saddle would be subject to a high value of theheat transfer coefficient, estimated to be42.89 W m2 K1, compared to 1.83 W m2 K1, whenthe fireproofing material is present. This high value of hemay result in thermal stresses which could cause the onsetof structural failure of the support.

9. The design assessment of the stresses

From the parametric survey conducted using the FEA andbriefly reported above, the maximum stress intensity, theabsolute maximum circumferential and the absolute maxi-mum axial stress were obtained. In the case of the axial andcircumferential stresses, no distinction is made between ten-sile or compressive stresses since, due to the presence of

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tensile residual stress induced by welding, the maximumstress range will be in the tensile field and thus likely tocause the propagation of a fatigue crack.

It is considered that two possible modes of failure arelikely to occur in these vessels. One is due to failure byratcheting and the other failure by fatigue. These are dis-cussed below.

9.1. Failure by ratchetingdesign for shakedown

If the vessel is subject to cyclic temperature loading andunloading, which in the present case could occur during thefrequent filling and emptying of the hot liquid storagevessel, this would cause a heating and cooling cycle, result-ing in thermal stresses of a cyclic nature being set up in thevessel. It was found that regions of high stress occurredclose to the profile of the saddle top plate and the vessel.If the cyclic effects cause large strains and plastic actionoccurs in these high stress regions during each progressivecycle, then damage could occur to the vessel. Failure in thiscase should be distinguished from the possibility of low-cycle fatigue in the regions of peak stress. We are hereconcerned with the overall structural behaviour due tocycles of thermal loading.

In general, for cyclic loading the vessel is designed for ashakedown condition in order to avoid ratcheting, which cancause incremental collapse. Shakedown is said to occurwhen, after the first cycle of load, the component behaviouris purely elastic. Some plastic behaviour may take place inthe first cycle but not in the second or subsequent cycles. Ifshakedown is not achieved, then in each cycle additionalplastic strain is accumulated. This behaviour is called ratch-eting and causes incremental collapse; it should clearly beavoided in the design.

Ratcheting is avoided in the codes (Annex A of BS 5500[9] and ASME section VIII division 2 [10]) in a rathersimplistic way by limiting the maximum value of the pri-mary plus secondary stress intensities to twice the yieldstress or three times the design stress, at the design tempera-ture. The thermal stress of the type considered in this pro-gramme of work is defined as a secondary stress, themaximum value of which can be obtained from the para-metric Eq. (1) given in Section 6. In general, the stressingfrom the thermal loading is the dominant part of this totalstress intensity, and therefore it is important to determine itsmagnitude as accurately as possible.

9.2. Failure by fatigue loading

The maximum stresses which occur in the storage vesselare found to occur in a region close to the saddle support/vessel welds. If cracks occur in the vessel, they invariablyoccur in the region of the welds and progress into or aroundthe vessel from the highly stressed weld region. Althoughthese vessels may not be subjected to a large number ofcycles, it is important that a fatigue life assessment be

made in view of the fact that they are storing liquids at ahigh temperature which, if released due to vessel failure, arehighly dangerous to personnel on the platform or processplant.

Two methods of fatigue assessment are available, givenas follows.

The procedure followed by ASME [10] uses the value ofthe alternating stress intensity, Salt, with the appropriatecurves for the material and temperature to obtain the fatiguelife. The stress intensity range, which is twice Salt, may beobtained from Eq. (1).

The second method is that contained in BS 5500 [9],which is similar to that to be included in the new Europeanpressure vessel standard. This method recognises that thefatigue life of welded joints is dominated by fatigue crackpropagation. The fact that the crack initiation period hasalready occurred in the creation of the welded joint isreflected in the method outlined. This approach is thusquite different from that used in the ASME [10] procedureoutlined above, and will be given here in some detail.

In the BS approach, the type of weld is classified in asso-ciation with the direction of the applied loading, and fromthe appropriate fatigue design SN curves the life assess-ment can be made. In the case of the saddle support, theweld is classified as G, which is a rather low class of welddetail. In BS 5500 [9], a power series equation has beenfitted to the fatigue curves which enables the life assessmentto be carried out with accuracy. In this treatment, the com-ponent stresses are used rather than the stress intensity; thatis, the circumferential and axial stresses which would benormal to the weld directions associated with the saddle.

In the present case, both the magnitudes of the absolutemaximum circumferential and axial stresses can be deter-mined from the parametric equations given by Ong et al. [5].In this case, all the welds around the saddle have beenclassified as G (from BS 5500 [9]), so the assessmentcan be carried out by simply using the absolute maximumstress of the two stresses (circumferential or axial).

In doing this, it is noted that for these saddles the maxi-mum stresses invariably occur on the side of the saddle inthe region of the weld running in the circumferential direc-tion. In such a case, the maximum axial stress would be themost appropriate stress to use in the assessment. Neverthe-less, to be absolutely safe it is proposed that the absolutemaximum stress be used, whatever the direction. In point offact, the axial and circumferential stresses are of similarmagnitude anyway.

Occasionally, the maximum stress is found to occur onthe inside surface of the vessel immediately adjacent to theweld. Despite this, and to build in a further measure ofsafety, all the maximum stresses should be used as if theydid in fact occur in the most critical orientation to the welds.

These vessels are invariably subjected to a repeated fillingand emptying routine. In many cases, the vessel may only bepartially emptied prior to refilling and thus the full range ofpressures and temperatures may not be realised. However,

622 A.S. Tooth et al./International Journal of Pressure Vessels and Piping 75 (1998) 617623

in order to simplify the computer model, and to provide aworst case scenario, it may be assumed at the design stagethat full cycles of temperature occur. Later refinement of theassessment considering the actual pressure and temperaturerange may be worthwhile to obtain a more exact lifeassessment.

10. Concluding comments

The paper provides the background to the use of the non-sliding support in vessels where thermal expansion is knownto occur. The stress results from the FEA have been pro-vided by Ong et al. [5] in the form of parametric equations toenable maximum stress values to be obtained. Thesestresses are used to design the vessel for operation in theshakedown range and to carry out a fatigue life assess-ment. The conclusions from the parametric survey, givenin Section 7, with regard to saddle design should be helpfulin providing an arrangement where the vessel stresses arereduced to a minimum. The fact that these installations canbe analysed with some certainty gives additional motivationfor using the non-sliding saddle, thus providing an alter-native way to support these vessels. It is shown thatvessel stresses can be reduced markedly by the use ofgood insulation material both round the vessel and thesaddle supports. If damage occurs to the insulation roundthe supports it is essential that it be replaced as soon aspossible.

It is appreciated that the parametric equations providedhere and by Ong et al. [5] only give the stress values for thetwo recommended saddle designs. However, it is antici-pated that the general conclusions, given in Section 7,which are drawn from the parametric survey on both designsof saddles, could also be applied to other designs of saddlein general use in the industry. This should provide a firstbasic design in which the overall saddle dimensions areestablished. If further assessment is required and the saddlesare of a different design to those given in this paper, then anFEA will be necessary to establish accurate stress values. Itshould also be appreciated that when large- diameter hori-zontal vessels are used for storing a high-temperature liquidunder high internal pressure, cognisance should be taken ofthe requirements of the self weight and internal pressure on

saddle design. In such cases, a compromise design may berequired in order to obtain an optimum saddle design.

Acknowledgements

This research has been funded by a grant from theCommission of the European Union for joint work by theUniversity of Strathclyde, Glasgow, UK and Nanyang Tech-nological University, Singapore. The use of ANSYS soft-ware through an educational license from Swanson Analysisis also acknowledged.

References

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[8] Tooth AS, Cheung JST, Ng HW, ONg LS. Analysis and design ofhorizontal pressure vessels with non-sliding saddle supports. Finalreport to the European Commission, contract no. CI 1*CT92 0067,March 1997.

[9] BS 5500. Unfired fusion welded pressure vessels. London: BritishStandards Institution, 1997.

[10] Boiler and pressure vessel code, section VIII Div. 2. New York:American Society of Mechanical Engineers, 1995.

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