Guidebook for the Design of ASME Section VIII, Pressure Vessels DOWNLOAD PDF Casti Guidebook to Asme Section IX: Welding Qualifications. ASME Section VIII provides requirements for the design, fabrication, testing, inspection, and certification of both fired and unfired pressure vessels. Established in as a non-profit educational and technical organization, The American Society of Mechanical Engineers. (ASME) is the.

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Section VIII, Divisions 1 or 2, nor minimum pressure limits for this Division. Rules pertaining to the use of the single ASME certification mark with the U3, UV3 and. Section V, Nondestructive Examination. BPVC-VIII-1 Section VIII, Rules for Construction of. Pressure Vessels, Division 1. BPVC-IX Section IX, Welding and. 1) ASME Section V (Nondestructive Exami- nation). 2) ASME Section IX (Welding and Brazing. Qualifications). lesforgesdessalles.info

Then the following parameters are obtained: Similarly, the rules in VIII-2 for calculating the allowable pressure on tbe convex side of hemispherical heads are given in Paragraph AD The rules in this section are limited to vessels of noncircular cross section with a straight longitudinal axis. Some of the examples in this book have changed completely and some remainunchanged. The rules for carbon steels are extensive and are discussed. For this head, the opening and its reinforcement shall be withiu a circle with a diameter ofO. Solution From Example 2.

Tabular values of the curves in these charts are also given in II-D, for example those shown in Table 2. See Table 2. Thus, the stress, B, from Fig. However, experiments performed subsequently to the publication of Eq.

A comparison of Eq. This fact should be considered when designing cylinders with large diameter to thickness ratios. The contents weigh kips. Deterntine the required thickness of the supporting skirt. Allowable tensile stress is 16 ksi. Use Fig. Also, the moment may have to be modified for shape and drag factors. The bending stress is obtained from the classical equation for the bending of beams: Thus, the selected thickness is adequate at the bottom of the skirt.

The axial stress due to dead load stays the same. However, these equations can be simplified for desigu purposes by plotting them so that the minimum buckling strain is expressed in terms of length, diameter, and thickness of the cylinder. These plots are utilized by the ASME as discussed next.

Accordingly, references in this section are made to paragraphs in Vill-l only. The designer calculates the known quantities LIDo and D,1t and then uses the figure to determine buckling strain, A.

To correlate buckling strain to allowable external pressure, the designer uses the stress-strain diagram of Fig. The allowable external pressure can then be determined from this B value, as explained below. Accordingly, the procedure in ASME Vill-I Paragraph UG for determining the allowable external pressure for cylinders with Dolt ratios equal to or greater than ten consists of the following steps: Assnme a value of t for the cyliuder.

Calculate the quantities LIDo and Dolt. A FIG. Use an External Pressure Chart such as Fig. These curves can also be used, conservatively, for cases where the pressure is on the sides only. These curves can also be used, conservatively, for cylinders with fixed rather than simply supported ends.

The effective length of a cylinder, L, needed to use Fig. Figure 2. The factor of safety for the allowable external pressure obtained by using Eq. The cylinder is subjected to an external pressure of 15 psi at "F? Then from Fig. The allowable pressure is obtained from Eq. Thus, from Eq.

Thus, the required thickness to be used is 3 I 8 in. Cylindrical Shells 45 T r: The value of S is taken as the smaller of two times the allowable tensile stress, or 0. The yield stress is obtained from the External Pressure Cbart of the material by using twice the B value obtained from the extreme right-hand side of the termination point of the appropriate temperature curve. The factor of safety in Eqs.

This gradual reduction in the factor of safety as the cylinder gets thicker is justified since buckling ceases to be a consideration and the factor of safety for external pressure is kept the same as that for internal pressure, which is 2 I 3 s; Example 2.

What is the maximum allowable jacket pressure? The allowable stress from tension is 17, psi. Cylindrical Shells 47 From Eq. Twice the allowable stress is 35, psi. Notice, however, that this pressure is greater than 0.

Such equations for external pressure are not in VUl-I yet. Larger values are not permitted presently by the ASME. One approximate equation Jawad, that is frequently used by designers for large Dol t ratios was developed by the U.

Many pressure vessels are subjected routinely to vacuum as well as axial loads from wind and dead load. Section VU1 does not give any method for calculating the allowable compressive stress due to combined effect of vacuum and axial loads. One such method is given by Bergman Bergman, It uses an equivalent external pressure to account for the axial compression effect on external pressure. Another method that is used to combine axial and external pressure is that of Gilbert Gilbert and Polani, This method uses an interactive equation sintilar to the one used for calculating the buckling of beam columns.

Solution From Example 2. Theu from Eq. The length can be reduced by providing stiffening rings at various intervals, as shown in Fig. A line of support is 1 a stiffening ring, 2 jacket bar, 3 circumferential line on a head at one, third the depth of the head, 4 cone-to-cylinder junction. To design the stiffening ring 1.

Assume first an area, A" of the stiffening ring and calculate the available moment of inertia. I, or Is. Use the appropriate External Pressure Chart and determine an A value. Solve Eq. The furnished moment of inertia must be greater than the required one.

The shell and ring material are SA ,C. External pressnre is 12 psi, and the design temperature is lOOoP. For the stiffeniug ring, try a 3 X 1 14 in. For ease of calculations, assume that the stiffening ring is not integral with the shell.

Hence, Eq. The moment of inertia of the bar is bd31 I I I l-r-r--r- I I 5' 0. However, before such a new ring is chosen, let us use the effective moment of inertia of the existing ring and shell and compare that to Eq. The welds must he able to support a radial pressure load from the shell of PL,.

This is based on the code assumption that the stiffening rings must support the total lateral load if the shell segments between the rings collapse. Also, the code requires that the welds support a shear load of 0. I 2in. Equation 2. When F exceeds 0. Let the pressure be a psi, b psi. L9,are used in nozzle connections, transition sections, and reducers. Determine the required thickness. Solntion From Eq. The stresses, away from discontinuities, in the shell due to internal pressure can be approximated by using the membrane theory of elliptical cylinders Flugge, Tlie hasic equation for hoop stress is expressed as 2.

This term was added by the ASME to take into consideration the nonlinearity in stress that develops in thick spherical shells. The designer should be aware that Eq. Large spherical shells for liquid storage usually have low internal pressure.

Thus, the governing thickness is controlled by the liquid weight rather than Eq. One method for determiniug the thickness in such spheres is given in API , In some instances, the outside radius of a shell is known rather than the inside radius.

In this case the governing equatiou is obtained from Eq. This is similar to the case for cylindrical shells discussed in 2. This is illustrated in Fig. For an applied internal pressure in compartment A, the hemispherical heads abc and def are subjected to concave pressure and Eqs.

Paragraph UGe32 f of Vlfl- I gives the rules for the design of hemispherical heads due to pressure on the concave side. Example 3. What is the required thickness of tbe hemispherical beads if the allowable stress is 20, psi?

The calculated thickness is less than 0. Solution The maximum pressure is obtained from Eq, 3. It has an outside diameter of 8 ft and an internal design pressure of psi at "F.

Calculate the required thickness of the hemispberical head if the allowable stress is 19, psi. The required thickness of the shell for this vessel is calculated in Example 2.

Attaching the head to the shell requires a transition with a 3: One method of attaching the head to the shell is shown in Fig.

Equations 3. Tbis is illustrated in Fig. For an applied internal pressure in compartment B, the hentispherical head def is subjected to convex pressure and Eqs. Solution From Example 3. We will use this thickness as our assumed t. And from Eq. The material is SA The thickness may have to be increased due to handling and fabrication requirements.

F is taken as positive when it is in tension and negative when it is in compression. When F is larger than 0. The rules for calculating the required thickness of hemispherical heads subjected to pressure on the concave side are given in Paragraph AD The rules are identical to those for spherical shells given by Eqs.

The procedure and the factors of safety for calculating the a1lowahle external pressure on spherical shells in VIll-2 are given in Paragraph AD The rules and factors of safety are identical to those given in VIII-!.

Similarly, the rules in VIII-2 for calculating the allowable pressure on tbe convex side of hemispherical heads are given in Paragraph AD Since this ratio is larger than 0. The shape can be approximated by a spberical radius of 0. The required thickness of 2: Equation 3. Nor does it have any rules for ellipsoidal heads when the ratio of PIS is large. Spherical Shells, Heads, and Transition Sections 67 3.

The required thickness is the greater of the two thicknesses determined from the steps below. Multiply the design pressnre on the convex side by the factor 1. Determine first the crown radius of the ellipsoidal head. Then use this value as an equivalent spherical radius to calculate a permissible external pressure in a manner similar to the procednre given for spherical shells in Section 3.

The head is made of low-carbon steel. TABLE 3. For Convex Pressure 1. First calculate the pressure and thickness. For external pressure, we determine K; from Table 3. This compressive stress, which is not considered by Eq.

This equation normally results in a thickuess that is greater than that calculated from Eqs. Paragraph UG e of VIII-l states that the maximum allowable stress used to calculate the required thickuess of torispherical heads cannot exceed 20 ksi, regardless of the strength of the material.

This requirement was added in the code to prevent the possibility of buckling of the heads as the thickuess is reduced due to the use of materials with higher strength. Find the pressure and the thickuess. The procedure utilizes a chart, Fig. The thickness for 2: Figure 3. Use Vlll-2 and then VlII-1 rules.

Spherical Shells, Heads, and Transition Sections 73 0. For 2: Using VIIl-l rules and Eq. The required thicknesses of the conical and knuckle regions are calculated in a different manner. In addition, conical sections without a knuckle that are attached to shells result in an unbalanced force at the junction that must be considered by the designer.

Vlll-l provides rules for the design of the junctions. These rules differ for internal and external pressure.

This type of construction will be discussed later in this section. After determining the thickness of the cone for internal pressure, the designer must evaluate the cone- to-shell junction, The cone-to-shell junction at the large end of the cone is in compression due to internal pressure, in most cases, The designer must check the junction for required reinforcement needed to contain the unbalanced forces in accordance with Paragraph of Appendix I of VIII-I.

Part of this area may be available at the junction as excess area. This excess area can be calculated from the equation 3. If this excess area is less than that calculated from Eq. The cone-to-shell junction at the small end of the cone is in tension due to internal pressure, in most cases. The designer must check the junction for required reinforcement in accordance with Paragraph of Appendix I of VIll Thus, the design of the cone as given by Eq, 3.

One design method uses the pressure-area procedure Zick and Germain, to obtain the required thickness. I I FIG. The flue angle is normally the same as the cone angle. Cone From Eq. Large Shell Again, using Eq. Large Cone-to-Shell lunction Assume that a reinforcing ring, if needed, is to he added to the shell. Spherical Shells, Heads, and Transition Sections 81 Next, we need to calculate the need for reinforcement in accordance with Table 3.

The amount of reinforcement is calculated from Eq. Small Cone-to-Shell Junction Assume that a reinforcing ring, if needed, is to be added to shell. The need for reinforcement is obtained from Table 3. This information is needed because the cone thickness used so faris based on the large diameter rather than on the small one.

Solution Small Shell From Eq. Large Shell From Eq. Flue From Eq. Due to external pressure, the cone-to-shell junction at the large end of the cone is tension, in most cases. The designer must check the junction for required reinforcement in accordance with Paragraph of Appendix 1 of VlII The required area is obtained from 3.

The area calculated from Eq. Some of this area may be available as excess area at the junction. This excess area can be calculated from the equation fl. In addition to having a sufficient reinforcement area, the cone-to-shell junction must have an adequate moment of inertia to resist external pressure forces when the junction is considered as a line of support.

The required moment of inertia is calculated as follows: Determine the quantity An from the equation 3. Enter the appropriate EPC and determine an A value. If B falls helow the left end of the temperature line, calculate A from the equation 3.

Calculate the moment of inertia from one of the following equations: The required momeut of inertia must be greater than the furnished one. The cone-to-shell junction at the small end of the cone due to external pressure is in compression, in most cases. The designer must check the junction for required reinforcement in accordance with Paragraph of Appendix I of Vlll-I. Some of this area may be available at the junction as excess area.

In addition to having a sufficient area. Determine the quantity A" from the equation 3. If B falls below the left end of the temperature line, calculate A from the equation 3. Calculate the moment of inertia from one of the following equations; 3. The required moment of inertia must be greater than the furuished one. When the cone is fianged and flued, then the required thickness of the cone is determined as before, except that Eq, 3.

Notice that the modulus of elasticity, shown in Fig. Large Cone-to-Shell function Assume that a reinforcing ring, if needed, is to be added to the shell. Hence, reinforcement is needed.

Next, determine the required moment of inertia at the largejunction needed for external pressure. Assume that a 2 in.

The available moment of inertia is obtained from Fig. Reinforcement is required in accordance with Eq. Next, determine the required moment of inertia at the small junction needed for external pressure. Spherical Shells, Heads, and Transition Sections 95 3.

These equations, which are for cylindrical shells, are used with the radius R perpendicnlar to the surface of the cone, as shown in Fig. VIIl-2 does not list any rules for toriconical heads.

The need for reinforcement at the large end of a cone-to-cylinder junction subjected to internal pressure is obtained from Fig. The figure is used to determine a maximum angle a when the PIS value is known. If the actnal angle a is less than that obtained from the figure, then no additional reinforcement is needed and the original thickness is adequate.

On the other hand, if the actnal angle a is larger than that obtained from the figure, then additional reinforcement is needed in accordance with Fig. The Q factor obtained from Fig. The Q factor in this case is multiplied by the required thickness of the shell at the small end of the cone. Figures 3. Allowable stress and pressure data is given in Table E3. Curva gowrned by maximum ,tr. Hence, reinforcement is required.

Curves governed by membrane stress intensitY ldue to average circumferential teneton stress end average radial compression stresst limited by 1. I"-- t-- 0. Circular plates are used for most applications; however, there are some applications where the flat plate is ohround, square, rectangular, or some other shape. When a flat plate or cover is used as the end closure or head of a pressure vessel, it may be an integral part of the vessel by virtue of having been formed with the cylindrical shell or welded to it or it may be a separate component that is attached by bolts or some quick-opening mechanism utilizing a gasketed joint attached to a companion flange on the end of the shell.

Flat plates and covers may contain no openings, a single opening, or multiple openings. To satisfy the loadings and allowable stresses, the plate may need to be of an increased thickness or it may reqnire reinforcement from attachments.

For the design of flat plates and covers which are attached hy bolting that causes an edge moment due to the gasket and bolt loading actiou, both gasket seating loads and operating loads shall be considered in a similar manner to that required for determining the acceptability of a bolted, flanged joint.

Since the loadings and dimensions required for analysis of bolted, flat plates and flanges in both VIIl-I and VIlI-2 are very similar, if not the same, they will be treated together. Spherically-dished covers are considered in section 4.

For further description of the terms given above, refer to Fig. Depending on the shape and welding details of the comer, a value of C is selected. A value for E, the butt- weld joint efficiency within the fiat plate, is required if the diameter of the head is sufficiently large that the head needs to be made of more than one piece. The value ofE depends on the degree of NDE' performed. It is not a weld efficiency of the head-to-shell comer joint! Example 4.

There are no butt welded joints within the head. There is a corrosion allowance, c.

The comer details conform to Fig. Solution First, the minimum required thickness of the cylindrical shell, t; must be calculated using Eq.

Flat Plates, Covers, and Flanges tttt.. C'" O. J o C"0. Reteiniog ,in. When pipe threads are C - 0. The plate is integrally welded into place. There is no corrosion allowance and no butt-welded joints in the plate. Solution I Using Eq. There is no corrosion allowance, and no butt-welded joints in the plate.

A plate that is 8 in. The bolted connection may consist of a flat plate or so-cal1ed blind flange, a loose-type flange, or an integral- type flange. An early method to analyze bolted flange connections with gaskets entirely within the circle enclosed by the bolt holes was developed by Taylor Forge in Waters, Determination of gasket requirements, bolt sizing, and bolt loading is the same for a bolted f1ate plate or blind flange, loose-type flange, and integral-type flange.

Loadings arc developed for gasket seating or bolt-up condition and for hydrostatic end load or operating condition. Guidance is given for the selection of the gasket and design factors, m, the gasket factor, and y, the gasket unit seating load, psi. Once the gasket material and sizing is determined, the bolt-up and operating loads are determined, bolting is selected, and the design bolt loading is calculated.

The selection of gasket type and material is set by the designer after considering the design specifications. The vaiues of m and y in the table are nonmandatory, and different values of m and y, wbich are either higher or lower, may be used if data are available to indicate acceptability.

After the gasket material and type are selected, the effective gasket width, b, is determined by the fol1owing procedure: The basic gasket seating width, b. When b. The required bolt load for operating condition, Wm1, is determined as follows: The bolt load used for the design of flanges, W, is theu determined from the following.

For operating condition, 4. In addition to safety, Eq. Where additional protection is desired or required by the design specifications, the following equation is used: Although not considered in VIII-I, it is prndent to design against crushout of the gasket by determining the minimum gasket width using the following: Staudard flanges are acceptable without further calculations for diameters aud pressure!

Flat Plates, Covers, and Flanges temperature ratings in the respective standards when of the types shown in Fig.

When there is no standard flange available, the minimum required thickness of tbe circular flat plate is calculated by using the following equation: Particular care should be taken when standard blind flanges are used so as to not exceed the size permitted by the Standard.

The reason for the factor of 0. For multiple openings in which no diameter is greater than half the plate diameter, no pair of openings has an average diameter greater than one-fourth the plate diameter, and the spacing between pairs of openings is no less than twice the averagediameter, Eq.

If the spacing is less than I V4 the average diameter, U-2 g applies. In all cases, the width of the ligament hetween the pair of openings shall be no less than one-fourth of the diameter of the smaller of the pair, and the edge ligament between an opening and the edge of the plate shall be no less than one-fourth of the diameter of that opening. In VIII-I, as an alternative to the rules for reinforcement of a single opening, the following procedure may be used to determine the minimum required thickness of a flat plate.

In Eqs. As an alternative to the rules for multiple openings, when the spacing for all pairs of openings is equal to or greater than twice the average diameter of that pair, the alternative rules for single openings may be used. For small openings which are located in the rim of the flat plate surrounding a large opening, as shown in Fig.

Using these loadings and moment arms, gasket seating moments and operating moments are determined and stresses are calculated and compared with allowable stress values.

The procedure for calculating stresses and acceptability of stresses is essentially the sarne for welding neck flanges as for loose, slip-on, or ring-type flanges.

For Code consideration, there are three types of flanges: Loose means that, for calculations, the flange ring provides the entire strength of the flange, even though the ring is attached to the vessel or pipe by threads or welds.

Integral means that, for calculations, the ring- and-shell or ring-and-pipe combination provides the strength of the flange, and the assumption is made that the connection between these parts has enough strength that they act together. Optional means that the connection is basically integral; however, it may be calculated as loose, which requires only one stress to be calculated. The difference in these various flanges is the line of load application and the magnitude of the loads.

However, the applied moments are determined in a similar manner. Either of these may have a hub, but the weld size and strength are not great enough to have the flange and vessel or pipe act together. See Appendix D for blank fill-in Sheet D. Integral-type flanges usually contain a tapered hub and flange ring which may be integrally formed or the hub and neck are welded together and act integrally so that each carries part of the loadings.

For gasket seating condition: I; Gasket is spiral-wound, fiber-filled, stainless steel, This flaoge has facing details, gasket size, and bolting that are the same as those given in Example 4. S, except that this is a welding neck flange instead of a ring flange. He -W.. Radial fIg. Since at gasket seating condition, the moment is smaller and tbe allowable stress is larger, only the operating condition is calculated. SH ,; 1. If an optimum minimum thickness of the flange is desired, calculations must be repeated with a smaller value of t until one of the calculated stresses or stress combinations is approximately equal to the allowable stress, even though other calculated stresses are less than the allowable stress for that calculated stress.

Included are: Flanges with other geometry and loading shall follow U-2 g. They are similar to standard flanges, except some of the loads on the flange ring cross section may he applied at different locations and in a reverse direction, possibly causing a reverse moment. Vlll-I has chosen to use the term aD to convert a standard flange to a reverse flange. The method of analysis for a reverse flange is similar to that used for an integral flat head with a large, single, circular.

For both analyses, a special limitation of the geometry is given. When K: For this reason, use of the analysis procedure should be limited to K: No Cor' 0i'U: S C-G - '5. N Z5R O. This flange bas facing details, gasket size, and bolting that are the same as those given in Example 4. Solution I The allowable tensile stress of the bolts from lI. Since the moment at operating condition is less than 0. SH'" 1. Although Fig. This type of gasket may be used with either a loose or an integral flange.

See Appendix D for blank fill- in Sheets D. Most often, a full-face gasket is used where the m aud y factors are relatively low, so that the bolt loading is kept within acceptable limits. A full-face gasket design generally results in the total moments from gasket seating aud from operation to be fairly low, and consequently, only a nominal flauge thickness is required.

However, bolt loads are usually higher. Ap- pendix Y of VIII-I contains rules for the design of a flat-face flange with metal-to-metal contact across the whole face or with a metal spacer added to the outer edge between pairs of flauges. Gasket loadings usually are small, as most gaskets are of the self-sealing type.

Inorder to make an analysis easier, assemblies are classified and individual flanges are categorized. Classification of Assemblies Class I: A pair of flanges which are identical except for the gasket groove Class 2: A pair of nonidentical flauges in which the inside diameter of the reducing flange exceeds half the bolt circle diameter Class 3: A flange combined with a flat head or a reducing flauge with au inside diameter that is small and does not exceed half the bolt circle diameter Categories of Flanges Category I: An integral flange or an optional flange calculated as an integral flange Category 2: A loose-type flange with a hub that is considered to add strength Category 3: The contact force, He, aud its moment arm, he, involve an interaction between the bolt elongation and the flange deflection and the moments from the bolt loading and pressure loading.

The equations given in VIII-l are approximate and may be conservative because they do not take into account the discontinuity condition which exists at the intersectiou of the ring and head aud that would distribute forces aud moments between the two parts relative to theirresistauce stiffness.

The Code procedures and equations assume the entire loadings at the intersection are taken by the bolting ring alone. Flat Plates, Covers, and Flanges 4. Typo te Integr. This is the type of cover shown in Fig.

The flange ring thickness is determined by combining the circumferential ring stress and the tangential bending stress, as follows: The term P is an absolute value for either internal or external pressure.

The value of M; is determined by combining the moments from bolt loading and gasket loading with the moment caused by the intemal pressure loading at the head-ta- ring intersection.

When M; is used in the equations, the absolute value is used. Determine the minimum required thickness of the head and flange ring.

There is no corrosion allowance, and no joint efficiency is required. The dish radius is 0. Flange Loads Same flange loads as in Example 4. The thickness of the head parallel to the flange ring face is: For operating condition: Since this is less than the 5. As with all calculated stress where a value of thickness, T, is assumed, a lesser value of required thickness may be determined by repeated assumptions of thickness and further calculations until the assumed thickness and the calculated thickness are the same.

Loads may be generated from both internal and external pressure and from applied external loadings.

An examination of the pressure boundary may indicate that extra material is needed near the opening to keep stresses from loadings at an acceptable level. This may be provided by increasing the wall thickness of the shell or nozzle or by adding a reinforcement plate around the opening. At some openings, there may be a nozzle to which is attached external piping generating external forces and moments from dead loads or thermal expansion. At other openings only a blind flange or flat cover with little or no available reinforcement may exist.

In designing openings, two types of stresses are important: Other loadings shall be considered separately. The first method, the reinforced opening or area replacement method, is used when that area which was to carry the primary membrane stress is missing due to tbe opening.

To replace this area, close-in substitute areas are called upon to carry the stress. The second method is called the ligament efficiency method.

This method examines the area of metal remaining between adjacent openings compared with the area of metal that was there before the openings existed. The primary membrane stress and shear stress are then examined for acceptability, Curves have been developed to simplify this examination. For single openings, only the reinforced opening method is used, while for multiple openiugs, either the reinforced opening method or the ligament efficiency method may be used.

Although the reinforced opening method and the ligament efficiency method are not developed on the same basis, VIII-I permits either one to be used. It is appropriate to use whichever method is more liberal, that is, the method giving the lower value for the increase in thickness.

Consequently, both methods may require examination. Article D-S of VIII-2 contains reinforced opening rnles for a satisfactory design for pressure loading only when a fatigue analysis is not required: It does not contain provisions for added loadings from piping or other loadings which may be imposed. When there is an opening through the shell, except for fiat heads, primary membrane stresses which develop from the pressure loading over the area formed by the opening diameter and the minimum required thickness are interrupted.

A substitute pathway is required. For flat heads, the situation is similar, except primary bending stresses are interrupted. The assumption is made that since primary bending stresses are maximum at the surfaces and zero at the centerline of the thickness while primary membrane stresses are uniform across the wall thickness, the replacement area for fiat heads needs to be only half the area required for cylindrical shells and formed heads.

Themethod presented for determining any needed reinforcement examines theregionaround theopening for available areas to carry the primary stress around the opening. Since stress is related to the load and cross- sectional area, areas: Placementandlocation of the replacementarea is important. The replacement area should be close to the opening; but care should be taken, if temperature is a consideration, not to generate an area of high thermal stresses.

If it is not too difficult to place some of the replacement area inside as well as outside of the vessel wall, try to place about two-thirds of the replacement area on the outside and one-third on the inside.

Although the procedure for evaluating stresses for external pressure is based on a buckling and stability analysis, the method for determining the reinforcement requirements for openings in shells under external pressure is very similar to that for shells under internal pressure, but with the following changes: This method deterntines the effectiveness of the material close to the opening for carrying loads.

The relationship of the nozzle wall thickness compared to the opening diameter or chord dimension, as appropriate, usually decides which of the two values controls. When this vertical limit was set by the ASME Code committee years ago, the assumption was made that rlt of 10 was appropriate. For application in various Code sections, the value was rounded to 2.

For VIII-! However,engineering experience and additional data have shown that these rules are satisfactory for most designs and so are still used. As mentioned previously, reinforced opening rules are for pressure loading only. Other loadings that need to be evaluated shall be considered separately using such methods as those given iu Welding Research Council Bulletin No. Openings 5.

Openings in cylindrical shells and formed heads are usually circular, elliptical. The latter shape is often developed for a nonradial nozzle opening. However, any other shape is also permitted, but there may be no method of analysis given in the Code. In shells 60 in. In shells over 60 in. When the size of the opening meets these limits, the rules given in UG through UG and in sections 5.

When the size of the opening exceeds these limits, the rules given in Appendix of VIII-l and in section 5. The total cross-sectional area of reinforcement required for any plane through the center of the opening is determined by: This correction is necessary to adjust for a minimum required thickness when the plane being examined is somewhere between the longitudinal plane and the circumferential plane. If the circumferential plane is being examined, t, is determined by 0.

Excess cross-sectional area of material within these limits is available for reinforcement. When the opening dimensions are within the limits given in section 5. When the opening is within the limits in section 5.

The nozzle is inserted through the head and attached by a fnll penetration weld. The inside diameter of the head skirt is The head material is SA-5l6 Gr. There is no corrosion allowance, and the weld joint! Nominal thickness used is 0. For this head, the opening and its reinforcement shall be withiu a circle with a diameter ofO. This radius is used in UG f to determine the t, for reinforcement calculations as: Nominal thickness used is 1. St or 2. This is less than the limit of Therefore, the provision to use the spherical head rule is valid.

According to UW b. Since this nozzle is similar to that detail, no calculations are required.

Example 5. The man way forging is inserted through the vessel wall and attached by a full penetration weld. The 12 in. The manway cover seals against the outside surface of tbe manway forging. The LD. The shell material is SA Gr. The design pressure is psi. There is no corrosion allowance. Since there are rio equations for determining the minimum required thickness of an elliptical shell in VIII-I, the rules of U-2 g are followed.

For an elliptical shell, equation 2. The maximum value of minimum required thickness is used for all planes as follows: NoJl1inal thickness used is 1. The opening has a 16 in. However, since AT is larger than Ar, the design is satisfactory and there is no need to consider the increased limit in our evaluation of this design.

According to UW b , weld strength arid load path calculations for pressure loading are not required for nozzles like the one shown in Fig. Since this nozzle is similar to that one, no calculations are required. According to UW b , weld strength and load path calculations for pressure loading are not required for nozzles of the type shown in Fig. Since this nozzle is similar to that one in detail, no calculations are required. The nozzle abuts the vessel wall and is attached by a fun penetration weld.

The J. When a corrosion allowance is considered, all calculations are based on the corrosion allowance being fully corroded away. Solution 1 The allowable tensile stress for SA Gr. B is UG c is: According to UW b , weld strength and load path calculations for pressure loading are not reqnired for nozzles of the type shown in Fig.

Therefore, for such openings, the following special reqnirements shall be met in addition to those listed in sections 5. The horizontal limits are the larger of: The vertical limits are exactly the same as those given in section 5.

In VIII-I, a method is provided for examining membrane and bending stresses in radial nozzle openings when the nozzle radius I shell radius 0; 0. Other alternatives given in Code Case may also be used. This method is as follows: Membrane stress, Sm, and bending stress, S; for either Case A for a nozzle with a reinforcing pad or Case B for a nozzle with integral reinforcement are calculated and compared to the allowable stress value.

I The membrane stress using the limits given in Fig. Case A: Reinforcement limits and spacing are based upon the damping length of a beam on an elastic foundation using the actual dimensions. In lieu of using these rules, the rules of Appendix 4 and Appendix 5 of vm-2 may be used. For other planes through the shell, the value of F from Fig. For h, L, x, and other dimensions, see Fig. The sum of the diameters is S; 0. T - Alternate nouHt to pipo transrtion 1. The limits given below also apply: For an opening with a shape and size not within these limits, design-by-analysis shall be used.

The total cross sectional area of reinforcement required for any plane through the center of the opening is determined by: For the Nozzle Shown in Fig. The perpendicular limits are the greater of 1 0. Openings In all cases, t,: Metai included in meeting the reinforcing area requirements shail be located in the zone boundary shown in Fig.

The nozzle is iuserted through the head and attached by a full penetration weld. The head materiai is SA Gr. The design pressure is psi, and the design temperature is 'F. Lc or Ln through the thickness. C it is Since the nozzle material is stronger than the head material, no adjustment is required. The manway forging is inserted through the vessel wall and attached by a full penetration weld.

The I. The design pressure is psi, and the design temperature is SOoop. An adjustment of t: Since there are no eqnations for determining the minimum required thickness of an elliptical shell in VIII-2, oue ueeds to be located. For an elliptical shell, Eq. The maximum value of the minimum required thickness is used for all planes being examined as follows: Use 8. Openings b Determine the limit perpendicular to the shell surface.

Section VIII-l permits two methods for calculating the replacement metal removed at openings. They are the reinforced opening method and the ligament efficiency method. The ligament efficiency method considers the metal area removed from the pressure boundary and the metal remaining between the two or more openings in the pressure boundary.

No metal is cousidered from any uozzle attached at the opening. The ligament efficiency curves apply only to the cylindrical shell of a pressure vessel where the circumfer- ential stress has twice the intensity of the longitudinal stress.

Once this was established, Rankine's Ellipse of Stress was used to determine the tension stress and the shear stress on any diagonal ligament plane. Using these data for tension and shear, curves were developed with respect to the longitudinal plane circumferential stress for various values of e, p' d, and t.

Equations were developed to make the determina- tion of the ligament efficiency an easy task. These equatious and a plot of the curves are given in Figs. Calculations should be made using both the reinforced opening method and the ligament efficiency method: Openings Example 5.

The openings are 2. Shell material is SA Gr. The openings are not located in or near any butt weldedjoint. Solution 1 The allowable tensile stress for both SA Gr. C at 'F is Since the reinforcement area available comes only from the shell, the shell thickness will have to be increased. A trial thickness will be assumed and verified. Therefore, the reinforcement limits overlap, and the rules of UG apply.

Those limits state that no reinforcement area shall be used more than once. Therefore, the limits overlap, and the rules of UG apply. Therefore, most, if not all, of the reinforcement area will come from the shell.

Solution 1 Determine the minimum longitudinal ligament efficiency or equivalent longitudinal efficiency and compare it with the longitudinal butt joint efficiency.

The lesser efficiency is used to calculate the minimum required thickness of the shell. However, there are some design rules for components with special geometries and configurations which require additional consideration. ForVIII-2,procedures aregiven for design by analysis of componentswith special geometries.

Consequently, very few rules for design of special geometries are given. This chapter contains guidance for the design of some of those special geometries in VITI-I as follows: Stays are used in pressure vessels to carry part or all of the pressure loading when it is desirable and possible to reduce the span and! Opposite stayed surfaces are "tied" together by staybolts, tubes, or baffles, which carry pressure loading as tension memhers.

Depending on the numher of ties, the thickness of braced and stayed surfaces may he less than wheu the surfaces are not stayed, hecause the loading is now resisted by both bending moments and bending strength and hy tensile strength of the stays. If washers are used, they shall be at least half as thick as the plate being stayed. Welded stays, shown in Fig. Welded stays may be used in con- struction of a dimpled or embossed assembly where a dimpled or embossed plate is welded to another dimpled or embossed plate or to a plain plate and the following rules from Appendix 17 of VIll-1 are met: Views Total views.

Actions Shares. Embeds 0 No embeds. No notes for slide. Officer Inspection Attock Refinery Ltd. ASME Section VIII It provides requirements applicable to the design, fabrication, inspection, testing, and certification of pressure vessels operating at either internal or external pressures.

This code was established in Sec. VIII, Div. Classes of Material - Special Constructions Materials: NDE 5. PWHT Supplementary 6. Lamination or Ultrasonic Testing e.

Ovality Formula: D - Minimum I. Otherwise will be a corner joint Continued Continued UG a E. Nozzle Reinforcement Pad attachment. UG b E. Insulation Ring Profile, Defect e. Punching Thickness limitation: V When required acceptance criteria: Spot RT Spot UG a 1 , UG a 2.

UG a 2. UG h Continued Foot Note 34 of Ed. UG k UG a 1 2 Not readily dried, that are to be used in services where traces of the testing liquid cannot be tolerated. Thereafter, the test pressure shall be increased in steps of approximately one-tenth of the test pressure until the required test pressure has been reached.