A BETTER WAY TO SUPPORT HORIZONTAL PRESSURE VESSELS SUBJECT TO THERMAL LOADING,

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1 THE AMERICAN SOCIETY OF MECHANICAL ENGINEERS 345 E. 47th St., New York, N.Y AA-57 The Society shall not be responsible for statements or opinions advanced inpapers or crticussion at meetings of the Society or of its Division; or Sections, or printed in its publications. Discussion is printed only if the paper is published in an ASME Journal. Authorization to photocopy 0 material for Internal. or personal use under circumstance not falling within the fair useprovisions otthe Copyright Act Is granted by ASME to libraries and other users registered with the Copyright Clearance Center (OCC) Transactional Reporting Service provided that the base fee of $0.30 per page is paid directly to the CCC, 27 Congress Street Salem MA Request; for special permission or bulk reproduction should be addressed to the ASMETechnical Publishing Department CopyrigM by ASME All Rights Reserved. Printed in U.S.A A BETTER WAY TO SUPPORT HORIZONTAL PRESSURE VESSELS SUBJECT TO THERMAL LOADING, Alwyn S Tooth, , Department of Mechanical Engineering, University of Strathclyde, Glasgow, G1 1XJ, UK. John ST Cheung, Heong W Ng, Lin S Ong School of Mechanical & Production Engineering, Nanyang Technological University, Singapore Chithranjan Nadarajah Exxon Engineering Asia Pacific, Singapore ABSTRACT When storing liquids at high temperature, in horizontal vessels, the current design methods aim to minimise the thermal stresses by introducing a sliding surface at the base of one of the twin saddle supports. However, regular site maintenance is required to ensure that adequate sliding is achieved. This may be difficult and costly to carry out. The aim of the present work, therefore, is to dispense with the sliding support and to provide saddle designs which although fixed to the platform, or foundation, do not result in the storage/pressure vessel being overstressed when thermal loading occurs. The paper provides general recommendations for the most appropriate caddie geometries, and details the way in which 'Design by Analysis' and 'Fatigue Life Assessments' may be carried out using the stresses which arise from these designs. NOTATION A length of vessel beyond saddle - overhang. I', breadth of the saddle top plate in axial direction. dp basic saddle width. E elastic modulus of the vessel and saddle material. height of saddle, measured from nadir of vessel to base plate (see Figs 1 &2). length of vessel between supports. mean radius of the vessel. wall thickness of vessel. thickness of saddle web and stiffeners. stress intensity (i.e. max principal stress difference) AT a temperature differential. extended width of saddle top plate. linear coefficient of thermal expansion of vessel and saddle material. INTRODUCTION Horizontal cylindrical storage/pressure vessels as used in the power, petroleum and other process industries are designed according to recognised Codes and Standards (for example; ASME, BS, Codap, etc) to withstand both the test and operating conditions. The common practice, in terms of support, is to provide two caddle-like supports symmetrically located along the length of the vessel. To avoid induced axial restraint stresses, as in the case of thermal loading, the codes recommend that one of the saddles be free to slide in the axial direction. This can be achieved in a number of different ways; by the use of foundation bolts positioned in slotted holes, by the use of low friction material (such as polytetrafluoroehtylene, PTFE) bearing pads bonded to the backing plate of the saddle base and the foundation plate, or by the introduction of a roller at the base of the support. A further recommendation is that the hot liquid storage vessel and the supports be fully insulated. Such a requirement is obviously necessary, to prevent heat loss which could be detrimental to the process, to prevent fire damage, and to protect personnel from inadvertent contact with the vessel or the support. Such insulation also avoids the high rate of temperature loss down the saddle itself which induces correspondingly high values of vessel stress. Presented at the ASME ASIA '97 Congress & Exhibition Singapore - September 30-October 2, 1997

2 In spite of the long standing practice of providing a sliding surface (or roller), mistakes can and do occur in practice. Vessels are found where the free sliding of the movable saddle cannot occur. This may be becmige slotted holes are inadvertently not provided, or the slotted holes that are provided are installed on the wrong side of the saddle base plate thus preventing any movement when thermal expansion takes place. Mother common experience, with the slotted hole type of installation, is that the nuts'are over-tightened and become rigidly fixed in position, the slotted hole is totally filled with concrete debris, or the sliding surface region of both the saddle and mounting are rigidly corroded together in installations mounted on 'off-shore' platforms. In view of this there are advantages from the operational, maintenance and a safety point of view, to dispense with the sliding support and install the vessels with both carldles fixed at their bases. The bases may be bolted to the foundation mounting in the 'on-shore' ground foundations, or welded to 'off-shore' platform beams or the deck of a ship. Having taken this step, the element of uncertainty is removed from the design approach, but one is now faced with the requirement to design the caddies and vessels to carry the total value of the thermal stresses which arise from the restraint of the cneklle feet. This paper briefly reports the findings of an extensive investigation, both experimental and theoretical, which provides the designer with the necessary tools to carry out this process. THE STORAGE OF HOT FLUID When a horizontal cylindrical vessel is used to store hot fluid and installed so that the saddles are fixed to the deck beams of a platform or the foundation, the thermal expansion of the vessel is restrained. If the hot liquid only occupies the lower part of the vessel then nonuniform heating of the vessel will occur. A further complexity could occur if during filling, the hot fluid impinges rapidly on the surface of the vessel surface in a local region. In this case a transient analysis of a local hot spot may be required to analyse the problem. To avoid the complexity of these cases the following assumptions are made:- (a) (b) (c) the hot fluid is inserted slowly so that the problem may be considered as 'steady state', the whole vessel is heated uniformly by hot liquid, to a temperature AT above ambient, the vessel and the supports are in the first instance fully insulated to avoid heat loss from the vessel wall and the saddles; this also serves the purpose of protecting plant personnel and reducing acoustic noise. The Radial Expansion of the Vessel During storage of the hot fluid the vessel expands both longitudinally and radially. The radial expansion occurs over the whole vessel. It is, however, restrained locally in the region of the support due to the wrap around effect of the caddie supports, using a saddle of at least 1200 embracing angle. The thermal stresses occur because there is a temperature gradient from ambient at the saddle base, to the vessel temperature at the uppermost point of contact, known as the horn of the saddle. This restraint is expected to cause local bending mainly in the circumferential direction at the uppermost region of the saddle top plate, similar to that caused by internal pressure. The Longitudinal Expansion of the Vessel The longitudinal expansion of the vessel is also locally restrained, in this case by the non-sliding saddle support. Because of this, a horizontal force and fixing moments are induced at the base of the support by the deck beams of the platform. If the fixing is considered to be rigid then the value of this force and moment will be such that the horizontal displacement and rotation will both be zero at the base fixture. In practice the platform deck will contain some flexibility so that these forces will be somewhat less than those in the rigid case. However, from a design point of view it is considered that the full force and moment constraint loadings should be assumed such that the base displacements are zero. When the system was analysed it was found that the longitudinal restraint resulted in higher vessel stresses than those due to the radial restraint. In fact, when an ideal sliding support is introduced at the base of one of the saddles the thermal axial (longitudinal) and circumferential stresses were substantially reduced. This fact does, of course, validate the existing recommendations of the code for those cases where an ideal support can be guaranteed. TYPICAL SADDLE DESIGNS Three widely used caddie designs were investigated in a range of preliminary studies. The details are given by Cheung,et.al.(1995). The result of these studies indicated that for the smaller vessels, up to and including 1200 mm diameter, a design based on that given in the pressure Vessel Design Handbook by Megyesy (1992) is appropriate. However, modifications to this design are proposed, as a result of these studies, to provide greater flexibility at the caddie horn and over the saddle width. The design proposed is shown in Figure 1. For larger diameter vessels a design based on that given in a dimensional British Standard (BS 5276:1983) 2

3 is proposed. Again modifications to this design are suggested, in which the width, w, is optimised to increase, the flexibility across the width. A typical design is shown in Figure 2 Megyesy, does not provide much radial flexibility in the horn region. As a result of the current studies it is proposed that this dimension be increased to correspond to an angle of 12 on both sides and an extended width, w of 25.4 nun on each side, [Fig 1]. It is also noted that for vessels up to 1200 mm diameter a stiffening rib is not provided in the lower vessel region. The saddle design for the larger diameter vessels [Fig 2] is based on that in BS 5276 (1983). This design includes an element of radial flexibility in the horn region, since the central web is not stiffened by, an end plate, as in the case of the Megyesy saddle. The width of the saddle top plate was, however, increased so that the ratio of the basic taddie width, dp and the extended width, W Was maintained at The value of this ratio was found to provide optimum flexibility and correspondingly reduced thermal axial and circumferential stresses in the vessel. ANALYTICAL PROCEDURES - NUMERICAL & EXPERIMENTAL FIG 1 DESIGN BASED ON MEGYESY (1992) FIG 2 DESIGN BASED ON BS 5276 (1983) The saddle design by Megyesy (1992) is a relatively simple design and is useful for the support of smaller diameter vessels. This saddle, as originally detailed by 6 The Finite Element Method The vessels were analysed using the Finite Element method with 20-noded brick elements (ANSYS software). A temperature differential of 100 C was achieved by making the internal vessel surface temperature 100 C, while the base of the saddle was assumed to be at a temperature of 0 C. In the first instance, the outside surfaces of the vessel and saddle were assumed to be "perfectly" insulated so that there was no heat loss to the surroundings. The heat was therefore transferred by conduction from the vessel to the base support. A small displacement linear elastic analysis was carried out in two phases. In the first instance the heat transfer problem was addressed to determine the temperature distributions in the vessel and saddle. These values were used in the second phase to determine the thermal stresses. Further details of this work is given by Cheung,et.al.(1995), by Tootketal. (1996) and by Ong,et.al.(1997). Experimental Studies Experimental studies have been conducted on small cylindrical storage vessels (1035 nun long and 228 mm inside diameter and 2.1 nun wall thickness). These were supported on both the Megyesy and BS type saddles, which incorporated the proposed modifications, referred to above. They were extensively strain gauged and subjected to two types of tests to verify the validity of the finite element numerical analysis. The axial restraint was investigated using an isothermal saddle base push-pull displacement test. The combined radial and axial 3

4 restraint was explored by progressively heating the vessels with hot liquid. The overall conclusion of these studies was that the finite element studies gave reasonable predictions of the strains in the experimental vessels and could be used with some confidence to carry out further parametric studies. The full details of this work is given by Ng, et.al. (1997). PARAMETRIC INVESTIGATIONS Using the finite element method described above, an extensive number of finite element runs were carried out on the two designs of saddles shown above. In this investigation the saddle designs were progressively modified to examine all the geometric features of the saddles, together with a complete range of vessel geometries. The scope of this is given by Ong etal.(1997). In each FE run the maximum stress intensity, the absolute maximum circumferential stress and the absolute maximum axial stress were determine for each geometric configuration. This work is an extension of the work reported by Tooth, et.al. (1996). In these studies the radial stresses, i.e. normal to the wall of the vessel, were found to be small compared to the circumferential and axial stresses. In view of this and since the aim of the work was to provide stress values for fatigue assessment, they were not considered further. Using a least square curve fitting procedure, parametric equations for the above maximum stresses have been established. In this work a power law relationship has been used in which dimensional groups of the leading vessel and support parameters are given. The details of this work are given by Ong,etal (1997). After a study of 15 power series expressions for their quality of fit, error estimates, and consistency the best non-dimensional expression for the stress intensity which occurs in the Megyesy design, was considered to be:- 511 EatIT - i7 a [bp l[h c [ [ts f [ v rmp rm tei 1t e I (1) where the constant a and the indices b to g are given (for three values of Alrm ) by- AIrm a b c d Alr, e f g Similar values were obtained for the absolute maximum circumferential and axial stresses based upon the Megyesy and the British Standard saddles, suitably modified as above. To further improve the accuracy of fit, 'trend line' equations were obtained into which the values obtained from the parametric equation can be substituted. Full details of the method and the appropriate equations are given by Ong, et.al. (1997). GENERAL CONCLUSIONS OF PARAMETRIC SURVEY From the 'parametric' and 'trend line' equations it is possible to draw a number of general conclusions which, if used, will reduce the value of the thermal stress produced in the vessel. 1. Tall saddles should be used since they introduce axial flexibility. However, in such cases care should be exercised to design the saddle webs and stiffeners to avoid panel buckling. 2. Narrow width saddles reduce the radial displacement at the ends of the saddle plate, and should be used where possible. The saddle top plate (or 'wear plate') should be extended, by modest amounts, both in the circumferential direction at the horn, but also across the width, beyond the basic saddle width, dp. In extending the top plate the requirements of a saddle of overall width outlined above in 2, should be recognised. Such a procedure introduces flexibility and thereby reduces saddle reactive interface forces and thermal stresses. 4. The distance between the two fixed caddie supports should be reduced, as far as possible, thereby reducing the value of the axial thermal expansion to be restrained. 5. If possible the saddle embracing angle should not exceed 120 for hot liquid storage vessels, since using the smaller angle of support provides flexibility to rotational movement. 4

5 INSULATION OF THE VESSEL AND SADDLE In the work reported above, it was assumed that both the vessel and the saddles were fully insulated and did not loose heat to the surrounding atmosphere (that is, adiabatic conditions). That is, the heat capacity of the hot liquid was transferred to the base of the enriches by means of conduction only. During discussions with an industrial support group, set up by NTU, Singapore to monitor the project, it was pointed out that it was not normal practice to fully insulate the support regions. The insulation was restricted to the vessel itself where a heat loss could be detrimental to the process. The supports were, however, protected with a fireproofing concrete material to help preserve the structural integrity of the steel in the event of a fire local to the vessel. Such protection has the additional advantage of reducing the risk of injury, by personnel inadvertently coming into contact with the supports and thereby sustaining a burn. It was considered that the fireproofing material used to protect the saddle, would be less effective in providing good insulation than the conventional insulation material, with the result that convection and radiation could occur from the outer saddle surface. Analysis of this case results in a temperature profile down the saddle, and particularly in the immediate region of the saddle/vessel junction, that is more detrimental to the occurrence of thermal stresses than the temperature profile derived under adiabatic conditions. The influence of such changes in the heat transfer behaviour was examined in detail for a range of vessels supported on the Megyesy saddle. In this work the thermal conductivity for the steel was retained at k, = 45 W/m. K. An overall heat transfer coefficient, 14, was determined assuming a convection heat transfer coefficient and a radiation heat transfer coefficient, and using a thermal conductivity for the fireproofing concrete of 0.1 W/m. K. For the values considered, the value for he was found to be 1.83 W/m 2 K. When this value was used in the analysis for a saddle of h p, Ir = 2 the stress intensity values were increased from that assuming pure conduction by 6.6% to 16.7% as the ratio of L.A. was reduced from 12.0 to 4.0. The detailed results of this work is given by Tooth, et. al (1997), from which stress values for other heat transfer coefficients may be obtained. It is sufficient to comment here that it is essential that some form of good insulation be provided in the saddle/vessel region. If there was a total absence of fireproofing concrete altogether, either by design or by damage, then the outer surface of the steel saddle would be subject to a high value of heat transfer coefficient estimated to be W/m 2. K. This value should be compared with the value of 1.83 W/m 2. K. obtained when the fireproofing material was present. This high value would result in unduly high values of the thermal stress which could cause the onset of failure. THE DESIGN ASSESSMENT OF THE STRESSES From the parametric survey conducted using the FEA and briefly reported above, the maximum stress intensity, the absolute maximum circumferential and the absolute maximum axial stress, were obtained. In the case of the axial and circumferential stresses no distinction is made between tensile or compressive stresses, since both are damaging and likely to cause propagation of a fatigue crack, due to the presence of tensile residual stress. It is considered that two possible modes of failure are likely to occur in these vessels. The one is due to failure by ratcheting and the other failure by fatigue. These are discussed below Failure by ratchetinq - design for shakedown If the vessel is subject to cyclic temperature loading and unloading, which in the present case could occur during the frequent filling and emptying of the hot liquid storage vessel, this would cause a heating and cooling cycle, resulting in thermal stresses of a cyclic nature being set up in the vessel. It was found that regions of high stress occurred close to the profile of the saddle top plate and the vessel. If the cyclic effects cause large strains and plastic action occurs in these high stress regions during each progressive cycle, then damage could occur to the vessel. Failure, in this case, should be distinguished from the possibility of low-cycle fatigue in the regions of peak stress. We are here concerned with an overall structural behaviour due to cycles of thermal loading. In general, for cyclic loading the vessel is designed for a shakedown condition in order to avoid ratcheting, which can cause incremental collapse. Shakedown is the condition that after the first cycle of load, the component behaviour is purely elastic. Some plastic behaviour does take place in the first cycle but not in the second or subsequent cycles. If shakedown is not achieved then in each cycle there is additional plastic strain accumulated - this behaviour is called ratcheting, which causes incremental collapse, and should clearly be avoided in design. Ratcheting is avoided in the codes [Annex A of BS 5500: (1997) and ASME Section VIII Div 2 (1995)] in a rather simplistic way in which the secondary stress is limited to twice the yield stress or three times the design stress. The thermal loading of the type considered in this project is assumed to produce a 'General Thermal Stress'. 5

6 It is defined as a secondary stress and therefore, to assess the acceptability of the stress level it is necessary to find the total maximum stress intensity from all the loading which occurs on the vessel. In general this will arise from the stresses caused by the following; the vessel and liquid contents loading, the internal pressure stress, and the thermal stress. The value of this stress intensity [as given by equations such as eqt (1)] must not exceed a value of 3 x design stress at the design temperature. In general, the stressing from the thermal loading is the dominant part of this total stress intensity, and therefore, it is of value to determine its magnitude as accurately as passible. Failure by fatigue loading. The maximum stresses which occur in the storage vessel are found to occur in a region close to the saddle support/vessel welds. If cracks occur in the vessel, they invariably occur in the region of the welds and progress into, or around, the vessel from the highly stressed weld region. Although these vessels may not be subjected to a large number of cycles, it is important that a fatigue life assessment be made in view of the fact that they are storing liquids at a high temperature which, if released due to vessel failure, are highly dangerous to personnel on the platform or process plant. Two methods of fatigue assessment are available, they are as follows. That followed by ASME (1995) uses the value of the Stress Intensity Range, Sri), obtained from equations such as eq. (1) and from this the 'alternating stress intensity', Sidt, which is one half the value of Sr,., is determined. The alternating stress intensity, S i t, is used with the appropriate curves for the material and temperature to obtain the fatigue life. The second method for fatigue life assessment, is that contained in BS 5500 (1997), which is similar to that to. be included in the new European Pressure Standard. This method recognises that the fatigue life of welded joints are dominated by fatigue crack propagfirion. The fact that the crack initiation period has already taken place in the creation of the welded joint is reflected in the method outlined. This approach is thus quite different from that used in the ASME (1995) procedure outlined above. In the BS approach the type of weld is classified, in association with the direction of the applied loading, and from the appropriate 'fatigue design S - N curves' the life assessment can be made. In the case of the saddle support the weld is classified as `G' which is a rather low class of weld detail. In BS 5500 (1997) a power series equation has been fitted to the fatigue curves which enables the life assessment to be carried out with accuracy. In this treatment the component stresses are used, rather than the stress intensity; that is, the circumferential and axial stresses which would be normal to the weld directions. In the present case both the magnitudes of the absolute maximum circumferential and axial stresses can be determined from the parametric equations. In this case all the welds around the saddle have been classified as class 'G', [from BS 5500 (1997)1 so the assessment can be carried out simply using the absolute maximum stress of the two stresses (circumferential or axial). In doing this it is noted that in these saddles the maximum stresses invariably occur on the side of the saddle in the region of the weld running in the circumferential direction. In such a case the maximum axial stress would be the most appropriate stress to use in the assessment. Nevertheless, to be absolutely sure it is proposed that the absolute maximum stress be used, whatever the direction. In point of fact the axial and circumferential stresses are of a similar magnitude anyway. occasionally the maximum stress is found to occur on the inside surface of the vessel immediately adjacent to the weld. Despite this, to build in a further measure of safety, all the maximum stresses should be used as if they did in fact occur in the most critical orientation to the welds. These vessels are invariably subjected to a repeated filling and emptying routine. In many cases the vessel may only be partially emptied prior to refilling and thus the full range of pressure and temperature may not be realised. However, in order to simplify the computer model, and to provide a worst case scenario, it is often assumed at the design stage that full cycles of temperature OCCUr. CONCLUDING COMMENTS The paper provides the background to the use of the non-sliding support in vessels where thermal expansion is known to occur. The results from the FEA have been provided by Ong, et.al. (1997) in the form of parametric equations to enable maximum stress values to be obtained. These stresses are used to design the vessel to operate in the 'shakedown' range and to carry out a fatigue life assessment. The general conclusions with regard to saddle design should be helpful in providing a saddle of near optimum design. The fact that these installations can be analysed with some certainty gives additional motivation for using the non-sliding saddle, thus providing a better way to support these vessels. It is anticipated that the general conclusions which are drawn from the parametric survey on both designs of saddles, could also be applied to other designs of saddle 6

7 in general use in the industry. It should be appreciated, however, that when large diameter horizontal vessels are used for storing high temperature liquid under high internal. pressure, cognisance should be taken of the requirements of the self weight and internal pressure on saddle design. In such cases a compromise design is required. Tooth,A S., Cheung, J S T., Ng H W & Ong,L S. (1997) "Analysis and design of horizontal pressure vessels with non-sliding saddle supports" Final Report to The European Commission, Contract No, Cl 1* CT , March ACKNOWLEDGEMENTS This research has been funded by a grant from the Commission of the European Union for joint work by the University of Strathclyde, Glasgow, UK and Nanyang Technological University, Singapore. The use of ANSYS software through an educational license from Swanson Analysis is also acknowledged. REFERENCES ASME, (1995) "Boiler and Pressure Vessel Code" Section VIII Div 2, American Society of Mechanical Engineers, New York BS 5276: (1983), "Pressure Vessel Details (Dimensions), Pan 2. Specification for saddle supports for horizontal cylindrical pressure vessels" British Standards Institution, London. BS 5500: (1997), "Unfired Fusion Welded Pressure Vessels" British Standards Institution, London. Cheung, J ST., Tooth, A S., Nadarajah, C., Ong, L S. & Ng, H W (1995) "Horizontal pressure vessels on fixed saddle supports under thermal expansion loading - a study of 3 different saddle designs" Int. Conf. on Mechanics of Solids and material Engineering, Vol (C), pp , 5-7 June 1995, Singapore, organised by School of Mech. & Prod. Eng.; Nanyang Technological University, Singapore. Megyesy, E F (1992), "Pressure vessel handbook" 9th edition, published by Inc. Tulsa, OK, USA. Ng, H W., Tooth, A S., Cheung, J ST 8c Ong, L S ' (1997) " Experiments and FEA on horizontal vessels under thermal expansion", ASME ASIA '97, 30 Sept to 2 October 1997, Singapore. Ong, L S., Cheung, J S T., Ng, H W. & Tooth, A S (1997) "Parametric equations for maximum stresses in cylindrical vessels subjected to thermal expansion loading", ASME ASIA '97, 30 Sept to 2 October 1997, Singapore. Tooth, A S., Cheung, J ST. Nadarajah, C., Ong, L S & Ng, H W. (1996) 'The support of horizontal vessels containing high temperature fluids - a design study" Proc. of the Eighth Intern. Conf. on Pressure Vessel Tech. (ICPVT-8) Montreal, Quebec, Canada, Vol 2 'Design and Analysis' pp , July 1996, ASME, New York, NY