ASME Methods Blog

Joint Efficiency

Joint efficiency is a factor required in all head and shell calculations that accounts for how closely a finished weld joint approximates the quality of the seamless parent material. Without further inspection it is assumed the welded joint is weaker than the material around it due to potential defects such as porosity, slag inclusions, and others. Shell thickness and therefore weld quantity is increased to account for this reduction in strength. Code welders following a qualified weld procedure are tested to weld a finished joint that maintains 100% of the parent material strength, but without further testing the allowed strength of a production joint is reduced to 70%.

For some design conditions, such as lethal service, the Code requires the designer to specify full radiography. However, when not required, the designer can specify optional radiographic examination to increase joint efficiency and reduce the required thickness of shells and heads. The designer weighs the material and welding costs against inspection costs to determine which course is best suited for the application. 

The figures below show the ASME VIII-1 joint efficiency values based on Type 1 joints (butt joints fully welded from both sides or equivalent) and degree of radiographic examination. The information is generated using the radiography logic diagrams and samples from Part 7 of PTB-4-2013 ASME Section VIII – Division 1 Example Problem Manual – the PTB-4 ‘E7.1’ through ‘E7.4’ example numbers are indicated where applicable.

No Radiography
ASME PTB-4 Ref. No.: None
	E = 0.70
	E = 0.85

Figure 1. Sample vessel illustrating joint locations and efficiency for No Radiography

Visual examination with no radiography is the simplest inspection option. All shell joints (A and B) have an efficiency of 0.70.

The seamless head efficiency is reduced from 1.00 to 0.85 since the shell circumferential seam it intersects is not inspected per code rule UW-12(d). This is shown as the “imaginary” seam H in the figure.

RT-4 Option 1
ASME PTB-4 Ref. No.: None
	E = 0.85
	E = 0.70
	E = 0.85

Figure 2. Sample vessel illustrating joint locations for RT-4 that will improve the shell long seam joint efficiency.

Since circumferential stress governs cylindrical shell design, performing spot radiography on long seams is the easiest way to improve joint efficiency and thus reduce shell thickness.

When specified, spot radiography requires one examination for every 50 feet of the same type of weld, with the provision that each welder’s work is represented. One spot could cover all of the Type 1 joints in this vessel if their total length adds up to less than 50 ft. This increases the long seam efficiency from 0.70 to 0.85 and reduces the cylindrical shell thickness at minimal cost.

The head imaginary joint efficiency remains at 0.85 due to UW-12(d).

RT-3
ASME PTB-4 Ref. No.: E7.3
	E = 0.85
	E = 0.85

Figure 3. Sample vessel illustrating joint locations for RT-3 that will yield the same results as RT-4 Option 1.

RT-3 increases the inspection requirements to spot radiography on both the long and circumferential seams of a vessel. There is no value added for the spot radiography of the circumferential joints since the long seam joint efficiency governs the design and RT-4 Option 1 already increased the long seam efficiency to 0.85.

The head imaginary joint efficiency remains at 0.85 due to UW-12(d).

RT-2
ASME PTB-4 Ref. No.: E7.2
	E = 1.00
	E = 0.70
	E = 1.00

Figure 4. Sample vessel illustrating joint locations for RT-2 that will improve the shell long seam and head joint efficiency relative to RT-4 Option 1 and RT-3.

RT-2 is often used to reduce the thickness of a seamless, non-hemispherical head by improving the head joint efficiency – all long seams must be fully examined to take advantage of this option. 

For the first time rule UW-12(d) is met and the shell long and imaginary head seam efficiencies are 1.00.

RT-4 Option 2
ASME PTB-4 Ref. No.: E7.4
	E = 1.00
	E = 0.85
	E = 1.00

Figure 5. Sample vessel illustrating joint locations for RT-4 Option 2 that will improve the shell circumferential seam joint efficiency relative to RT-2.

RT-4 Option 2 is similar to RT-2, but uses additional spot radiography to improve the circumferential joint efficiency of the shell. This option costs more than RT-2 and yields the same component thicknesses – circumferential seams do not govern the design of cylindrical shells.

Again rule UW-12(d) is met and the shell long and imaginary head seam efficiencies are 1.00.

RT-1
ASME PTB-4 Ref. No.: E7.1
	E = 1.00
	E = 1.00

Figure 6. Sample vessel illustrating full radiography of all seams.

As shown, RT-1 requires all seams to be examined for their full length and yields E = 1.00 for all joints. RT-1 inspection is required for lethal service.

Table 1. Summary of joint efficiencies for Type 1 joints on shells and seamless heads.

Dual Certified Vessels for Low Temperature Service

PVE-5920, Last Updated: Apr 12, 2018, By: LRB

Background:

Many older in-service vessels built from SA-212 coarse grained carbon steel do not meet current code requirements for low temperature service. The service history on these vessels is good, resulting in an attitude that the current code rules are too restrictive. However, there is an alternate way to understand these vessels: many are in service conditions like propane storage where the pressure at lower temperatures can never get very high. Many of these vessels can never experience situations of combined high pressure and low temperature.

Propane Pressure Temperature Curve:

A propane storage vessel typically has a pressure rating of 250 psi. In Canada it can also have -50°F MDMT (Minimum Design Metal Temperature) depending on the service location. What it does not experience is 250 psi at -50°F. The pressure temperature curve for Propane is shown Fig-1 below. Some data points have been highlighted on the propane P-T curve. At -50°F, a common design temperature for exposed locations in Canada, the pressure in a propane storage vessel is a small vacuum of -2 psig (or 13 psia, 2 psi below atmospheric pressure) – if the valve is open, air will enter the tank instead of propane exiting. At -20°F, a common minimum temperature for older design codes, the pressure is still only 11 psi. At 250 psi, a common propane design pressure, the temperature is 127°F. Once the contents of the vessel reach this temperature, a 250 psig relief valve will open allowing the release of some gas reducing the temperature of the remaining contents through boiling heat transfer.

Fig 1 – Propane pressure vs temperature curve.  Data source

A Typical Vessel Design:

We designed a sample vessel with a minimum wall thickness to just pass 250 psi service. In different calculation runs, the material category was changed from Curve A to B to C to D.  A crude generalization is that Curve A represents coarse grained materials with poor low temperature impact (toughness) properties like some SA-212 materials. Curve D materials have the best properties obtained through methods like normalization or quench and tempering. Curve B and C are intermediate materials.

The program reported the allowed minimum temperature for the design as the pressure was changed (Fig 2). For example, the coarse grained Group A material had a minimum temperature of 50°F for the full 250 psi service. In comparison, the curve D material was good all the way down to -35°F.  But for all material groups, by lowering the design pressure, lower minimum temperatures resulted.

Fig 2 – Sample vessel P-T ratings for four different curves of steel used in pressure vessel design

Although the strength dropped as the temperature dropped, in all cases the vessel pressure rating exceeded the propane P/T curve. Fig 2 shows that all could be used safely, however, curve A in this case could not reach all the way to -50°F. The Curve A material can be seen to have the poorest pressure rating at low temperature, and the Curve D the best. When a vessel is designed for new construction it is possible to combine the selection of material with the appropriate testing to obtain MDMT of -50°F, even at full pressure. However, for this sample vessel, no material combination provides a full pressure rating at -50°F.

Continuing this sample assuming Curve B material: ASME has additional rules like UG-20(f) that allow higher pressures to be used for some materials down to -20°F service. With UG-20(f) applied, the Material pressure temperature curve for a Curve-B vessel now looks like:

Fig 3 – Just looking at curve B material. UG-20(f) allows the full pressure to be used down to -20°F. A second rating of 115 psi covers the range from -50 to -20°F.

The design based on curve B material has been dual rated: 250 psig at down to -20°F and 115 psig for down to -50°F (Fig 3). This is recorded according to the rules of UG-116(a(5) footnote 37 which does not restrict the number of minimum temperature pressure combinations used. One full calculation set is required for each P-T combination. Two full calculation sets are required for this vessel: the first is calculated at a pressure of 250 psig and shows a minimum temperature of -20°F; The second is calculated at 115 psig and shows a minimum temperature of -50°F.

As an alternative to using these curves, it is possible to impact test materials and welds. In general it does not pay to be optimistic about coarse grained materials and welds passing impact tests prior to seeing the actual test results.

More on SA-212 material:

Source: www.krrao.com (web site no longer exists)

ASTM SA-212-39 (S-55) was put into Section II in the 1940 edition of the Code. There were two grades in S-55: A and B, each with two different minimum tensile strength requirements controlled by carbon content.

In 1952… it was required in SA-212 that plates intended for low-temperature service must meet the impact requirements in SA-300. SA-212 could be purchased to a fine-grain-melting practice-and subsequently normalized and tempered-for low temperature service, or purchased to a coarse grain-melting practice; the single specification permitted the manufacture of both plate grades. The SA-212 Specification continued up to 1962 as the carbon steel plate material of choice for low-temperature service for boiler drums and pressure vessels.

[In the] 1968 edition of Section II … the SA-212 Specification was deleted …it was replaced with two specifications. The SA-212 steel plate melted to coarse-grain practice was replaced with SA-515 (Specification for Pressure Vessel Plates, Carbon Steel, for Intermediate and Higher Temperature Service) and the SA-212 steel plate melted to finegrain practice was replaced with SA-516 (Specification for Pressure Vessel Plates, Carbon Steel, for Moderate and Lower Temperature Service). These two specifications-along with SA-299 (Specification for Pressure Vessel Plates, Carbon Steel, Manganese-Silicon), which has slightly higher room-temperature strength-were first published in the 1949 Edition of Section II. They continue to be used today as the carbon steel plate materials of choice for boiler and pressure-vessel applications.

SA-212 in older vessels being recalculated could be either coarse or fine grain. Either Curve A or a better curve. Proving that it was built to fine grain practice and impact tested to SA-300 could be difficult for an old vessel. Sometimes it is only possible to assume that it was made to Curve A. There are also some concerns that special care is required for hydrotesting coarse grained pressure vessels. National Board (search page for SA-212)

When a used vessel moves to a new location in Canada a new CRN registration number is usually required. The CRN calculations are based on the inspected wall thickness. Three possible calculation methods are used: 1) calculations to be based on current code rules (see note from ABSA below); 2) Calculations based on code rules at time of construction or; 3) Calculations based on both the time of construction and the current code rules, the most conservative to be applied. Some of our customers place reserve bids on used vessels and do not complete the transactions until the CRN has been obtained. It is important to get this sorted out before moving the vessel! More info from ABSA:

1.Q2. Is it permissible to bring into and operate a used pressure vessel that was manufactured of SA-212 Grade B steel? The vessel was not impact tested when it was manufactured.

1.R2. A used pressure vessel made of SA-212 Grade B steel may be brought into and registered for operation in provided that its proposed design conditions meet the intent of the current ASME Pressure Vessel Code. Since the current Code requires a minimum design metal temperature (MDMT) for a pressure vessel, such an MDMT must be established for the used vessel using the current Code methodology. SA-212-B material would be considered a Curve A material for the purposes of Code paragraph UCS-66. Therefore, a maximum allowable working pressure (MAWP) that supports the MDMT without impact testing would have to be established. It is assumed that it is not feasible to impact test all the shell and head plates and weld joints to support an MDMT lower than that without impact testing.

Unlisted Materials

File: File:PVE-4245, Last Updated: Aug 20 2014, By: LRB

Background

To register fittings, design validation based on code calculations, finite element analysis or proof testing is required. When a design is based on code listed materials, the code of construction provides allowable operating stress levels. If the design of the pressure containing item is simple, the regular code rules can be used and will supply a pass/fail judgement. If no code rules exist for a complex or unusual shape, Finite Element Analysis (FEA) can provide the stresses which can be compared with the listed allowables for a pass/fail judgement.

An alternate to FEA is to proof test the item at stress levels far above operating. The items actual and guaranteed minimum tensile strengths are required for the proof test. The formula used is from VIII-1 UG-101(m):

Eqn-1

Where B is the burst test pressure and P the allowed operating pressure. The burst test has to be at least 4 times the operating pressure. E is the welding efficiency if the item is welded – typically between 0.7 and 1.0. Two more pieces of information are required – Su – the specified minimum tensile strength of the material and Suavg – the tensile test results from the item under test. Typical proof test pressures are 5-6x operating pressure, a requirement in many cases more conservative than regular code calculations or FEA.

For code listed materials, all of the required information is available for either calculations/FEA or for burst testing.

Unlisted Materials

Codes B31.1 and B31.3 are useful for registering fittings because they allow unlisted materials to be adopted and because they provide fewer restrictive design rules. However, if your design looks like a small vessel you might have to use VIII-1.

An unlisted material made to a specification can be adopted if the material’s guaranteed minimum yield and tensile strength are available at the operating temperature. The code adopted strength is based on a formula using these two inputs resulting in allowable design strength. Or the minimum tensile strength can be used in the proof test. Using this process, almost any IID listed material can be adopted for use in B31.1 or B31.3.

Eqn-2

This is a typical formula for adopting unlisted material in B31.3. Sy and St are the materials guaranteed minimum strength. More complex methods are used at higher temperatures where the materials creep properties need to be taken into account. Lack of availability of elevated temperature material properties can severely limit the adoptability of unlisted materials. Caution: see Unlisted Material Registration Problems below.

Unlisted Materials With No Specified Strength

Many fittings materials are not code listed and have no guaranteed minimum tensile or yield strength information. Two common examples: SAE1010 carbon steel and B85 A380 die cast aluminium. Both are made to chemical only specifications and both are used in fittings.

To use either of these materials in Canadian registered fittings, the purchaser has to agree with the mill/foundry what minimum tensile and yield strength level is acceptable. A specification referenced or written into each material batch purchase order is required. Chosen strength levels are obviously important. Set too high and excessive batches will be rejected upon physical testing. Set to low and the parts will not pass code calculations. Also note that the ratio between the actual and minimum tensile strength impacts the required burst test pressure. The lower the minimum specified strength, the higher the required proof test. A sample purchase order or a copy of the specification would be required with the CRN application. Caution: see Unlisted Material Registration Problems below. Setting appropriate guaranteed minimum stress levels commonly causes confusion, an example follows.

Example: A manufacturer who is investigating a new unlisted material gets some pull test results. 4 tensile test results at ambient [ksi] 47, 46, 44, 48. 4 yield stress results [ksi] 25, 26, 23, 28. The results are at ambient only, and the product will only be used at ambient so elevated temperature testing is not required. What should the guaranteed minimum yield and tensile be? Each material batch will be tested, so setting the specified minimum too high will result in batches being rejected. For example, a specified minimum tensile of 45 ksi would cause the 3rd specimen to be rejected. Some number around 40 ksi tensile and 20 ksi yield might be reasonable as is shown in this graph.

Fig 1 – the unknown materials test results after specified minimum tensile and yield strengths are chosen.

What happens if the guaranteed minimum is set too low? If the product is to be burst test, per Eqn-1, the required burst test is increased by the ratio of Su/Sur, where Su is the specified minimum burst test, and Sur the test results from the item under test. If samples 1, 2 and 3 are taken from the test object, Sur = average(47,46,44) = 45.6. If the specified minimum is 40 ksi, then the burst test ratio is 4 x 45.6/40 or 4.56x. However if the specified minimum was set too low, to 20 ksi, then the ratio would be 4 x 45.6/20 or 9.12x.

If the product will not be used at ambient, then elevated materials properties are required. For CRN applications, temperatures above 100ºF are considered elevated (source unknown). Additional elevated temperature material testing is required to cover the design conditions.

The manufacturer needs to document the minimum specified properties and other characteristics of the unlisted material with no specified strength per B31.3:

B31.3 2010 323.1.2 Unlisted Materials. Unlisted materials may be used provided they conform to a published specification covering chemistry, physical and mechanical properties, method and process of manufacture, heat treatment, and quality control, and otherwise meet the requirements of this Code. See also ASME BPV Code Section II, Part D, Appendix 5. Allowable stresses shall be determined in accordance with the applicable allowable stress basis of this Code or a more conservative basis.

Alberta requires that this document be published on the manufacturers web site available for unrestricted access.

Unknown Materials

If all of the above fails, many Canadian reviewers will allow a fitting to be registered with “unknown” materials if it can be proof tested to 10x operating pressure (no tensile strength testing required, no guaranteed minimum specification provided). This category includes many plastics that are not covered by the piping codes, glass, ceramics and steels that cannot be adopted by the above methods.

However, testing to 10x operating pressure is a severe test not possible with many products.

Stainless Stress Values

Selection of the Correct IID Listed Stainless Steel Stress Values

File: PVE-8296, Last Updated: Apr 8 2015, LB

This is an extract from one of our in house training topics, we thought our customers would find it useful.

Lots of IID Entries

The IID book provides many rows of stress properties for the 304 grade of stainless steel pipe. These 12 line items cover SA-312 304 stainless steel in welded, seamless, low, high and regular carbon and high and low stress. Quite a lot – here are the 3 variables covering the 12 lines.

1. Seamless or welded
2. Low or high strength
3. The amount of carbon in the grade: L, regular or H.

This is further confused by Dual Certified stainless (304/304L). Other variations like N, LN also exist with different chemistry not discussed here.

For a simple 304 grade stainless pipe, the different combinations here provide for 12 different IID table 1A lines with different allowed stress/temperature curves. You want only one.

12 lines of IID material properties for 304 stainless steel pipe.

The table notes are required to choose the correct line. For Section VIII-1 use, if we do not care how the material properties are calculated, the following notes can be ignored:

G21 and W13 apply to section I only, W12 is for Section III use

T4, T6, T7 and T8 explain how the high temperature properties were calculated

Removing the notes that are not helpful for this application

These are the useful notes:

G3 These stress values include a joint efficiency factor of 0.85.

G5 Due to the relatively low yield strength of these materials, these higher stress values were established at temperatures where the short-time tensile properties govern to permit the use of these alloys where slightly greater deformation Is acceptable. The stress values in this range exceed 66% but do not exceed 90% of the yield strength at temperature. Use of these stresses may result in dimensional changes due to permanent strain. These stress values are not recommended for the flanges of gasketed joints or other applications where slight amounts of distortion can cause leakage or malfunction. For Section Ill applications, Table Y -2 lists multiplying factors that, when applied to the yield strength values shown in Table Y -1, will give allowable stress values that will result in lower levels of permanent strain.

G12 At temperatures above 1000°F, these stress values apply only when the carbon is 0.04% or higher on heat analysis.

G24 A factor of 0.85 has been applied in arriving at the maximum allowable stress values in tension for this material. Divide tabulated values by 0.85 for maximum allowable longitudinal tensile stress.

W14 These S values do not include a weld factor. For Section VIII, Division 1 and Section XII applications using welds made without filler metal, the tabulated tensile stress values shall be multiplied by 0.85. For welds made with filler metal, consult UW-12 for Section VIII, Division 1, or TW-130.4 for Section XII, as applicable.

Welding

Note G3 or G24 is applied to all welded pipes. The longitudinal efficiency for an ERW (Electric Resistance Welded or fillerless welded) pipe is set at 0.85. Note W14 indicates that a weld efficiency of 0.85 has not been included but should be included if the product is ERW.

As a suggestion, use the Smls. & wld. pipe grade only for seamless product, and Wld. pipe for welded grade. The additional 0.85 efficiency will not need to be applied resulting in less confusing calculation sets.

Even if an ERW efficiency factor of 0.85 has been applied, additional reduction in efficiency might be required according to the rules of UW-12(d). For example, pipe caps welded on the end of a vessel made out of ERW pipe with no radiography will require an efficiency of 0.85 to be applied to the pipe long seam, this is in addition to the 0.85 already taken off for the ERW efficiency. This is easier to do if the welded material line has been chosen from table IID.  Here is an interpretation.

Standard Designation: BPV Section VIII Div 1
Subject Description: Section VIII, Division 1, UW-12(c)
Date Issued: 02/18/1988
Record Number: BC88-043
Historical Interpretation numbers : VIII-1-86-218
Question(s) and Reply(ies): Question: Is it the intent of the new stress multiplier rules that, for a vessel consisting of an ERW pipe shell with seamless ellipsoidal or torispherical dished heads and no radiography of the Category B seams, the stress values from Table UCS-23 for ERW pipe be multiplied by E = 0.85 for calculations involving circumferential stress in the shell?
Reply: Yes.

Many drawings do not specify if the product form is seamless or welded. If this cannot be clarified on the drawing then the lower strength value (welded) must be assumed.

High Strength or Low Strength

Half of the listings have note G5, indicating that the strength level of the material is set above the customary 66% yield limit. The use of these values is not recommended for flanges, but not prohibited. Our experience indicates that ASME VIII-1 Appendix 2 and Y flange designs are highly conservative. It is our policy to use high strength materials for these applications except when registering in the province of Alberta which has a requirement to use the low strength values, or when the customer prohibits it.

The difference between high strength and low strength values for stainless steel at temperature

The Chemistry of the Grade (L, Regular or H)

Usually it is the job of the customer to specify the correct grade of stainless to use, but sometimes we have to ask for changes based on the operating temperature. The regular grade (TP304) has a maximum carbon content of 0.08%. TP304L has a maximum content of 0.035% and TP304H ranges form 0.04-0.010%

Carbon content ranges of plain, L and H grade 304 stainless steel

304L is not listed for applications above 1200°F. Note G12 prohibits the use of regular grade 304 above 1000°F unless the carbon content is above 0.04%. (There is no explanation for the difference between 1200°F for the 304L grade and 1000°F for 304 without the extra carbon.) 304H always has 0.04% or greater carbon so no note is required for high temperature use.

Dual Certified

Dual certified stainless is produced by mills to meet the requirements of both 304 and 304L. From the above graph, the carbon content must be below 0.035%, and from the top table, the higher 75,000 tensile and 30,000 yield must be met. This is not a challenge for modern production methods.

Sometimes drawings have a confusing 304/304L designation on the bill of materials. Does this mean materials that meet both specifications (dual certified) which would allow higher strength levels to be used in the calculations, or is the customer giving themselves a choice between using the stronger or the weaker material, in which case the weaker would have to be calculated. Usually the first interpretation is correct, but where this cannot be known, the weaker material must be calculated.

Conclusion

With these 3 variables understood, the most appropriate IID listed line can be selected. Other product forms like plate are simpler because the welded or seamless variable does not exist, but the method is the same.

Lethal Service – Quick Guide

File: PVE-3856, Last updated June 6 2017 LRB / CT

Introduction

Requirements for lethal service are scattered through VIII-1, the code cases and the interpretations. ASME VIII-1 section UW-2 has most of the requirements for lethal service. Two definitions of lethal service are provided in UW-2. The second definition is also duplicated in the end notes:

[A] vessel is to contain fluids of such a nature that a very small amount mixed or unmixed with air is dangerous to life when inhaled[.]

By “lethal substances” are meant poisonous gases or liquids of such a nature that a very small amount of the gas or of the vapour of the liquid mixed or unmixed with air is dangerous to-life when inhaled. For purposes of this Division, this class includes substances of this nature which are stored under pressure or may generate a pressure if stored in a closed vessel.

It is up to the user to determine if a service is lethal. We at Pressure Vessel Engineering do not determine if the vessel service is lethal (but we might have opinions based on previous jobs).

ASME VIII-1 Lethal Service Code References

UW-2 Service Restrictions is a very important section for lethal service vessels and must be read in its entirety. A few brief points from UW-2:

• UW-2(a) and UW-11(a)(1) – All butt welds shall be 100% radiographed
• UW-2(a) – ERW pipe (like some grades of SA-53) is not permitted but interpretation VIII-1-01-118 says it is acceptable if the long seam is fully radiographed
• UW-2(a) – Post weld heat treatment is required for CS and Low Alloy
• UW-2(a)(1)(a) – Category A welds shall be type 1 only (butt welded with no permanent backing strip)
• UW-2(a)(1)(b&c) &Interpretation VIII-1 92-211 – Category B & C welds shall be type 1 or 2 only (butt welded). No slip on flanges! No Figure UW-13.2 Flange or Head to Shell attachments
• Interpretation VIII-I-98-23 – Category D welds (typically nozzles) shall be full penetration.
• UW-2(a)(1)(c) – Category C joints for stub ends have a long list of requirements
• UW-2(a)(2 and 3) – Heat exchangers have a long list of requirements
• Read all of UW-2 for more restrictions…

When it has been determined that a vessel is in Lethal Service, some of the following code sections will apply. Other code sections might also be required. (List last updated June 16 2015 based on the ASME 2013 VIII-1 code edition.)

• U-2(a)(2) – The user determines if the service is lethal.
• UG-16(5)(a) – Air cooled and cooling tower heat exchanger tube walls to be 1/16″ min.
• UG-24(6) – Casting RT requirements.
• UG-25(e) – No telltale holes.
• UG-99(g),(k) – Hydro test visual leak inspection cannot be waived. Do not paint or line prior to the hydro test.
• UG-100(d)(4) – Pneumatic tests cannot be used for lethal service vessels, but also see code cases 2046-2, 2055-2, 2407 regarding pneumatic instead of hydrostatic testing.  
• UG-100(e)(3) – Do not paint or line prior to the pneumatic test.
• UG-116(c) – “L” stamping must be added to the nameplate.
• UG-120(d)(1) – “lethal service” is added to the data report
• UW-2 – Service Restrictions – main source of info on Lethal Service (see lots more from UW-2 below)
• UW-11(a)(1) – All shell and head butt welds to be 100% RT
• Fig UW-13.5 – One possible lap joint stub end configuration – see also interpretation BC-79-680 / VIII-80-111
• UB-3 – Brazed vessels shall not be used
• UCS-6 – Do not use SA-36, 38W or SA-283
• UCS-79 – PWHT and extreme fiber elongation – read for rules when heat treatment is required
• UCI-2, UCD-2 Cast iron and ductile cast iron vessels are not allowed.
• UIG-2(c) – Metal parts for graphite vessels are to be designed to UIG requirements
• UIG-23(b) – Factor of safety for lethal service
• UIG-60 – Lethal service
• UIG-99 – Lethal service pressure tests
• ULW-1 & ULW-26(b)(4) – Layered vessels lethal restrictions apply to the inner shell and heads only
• UHX-19.1(b) – Heat exchanger markings
• Appendix 2-5(d) & 2-6 – Maximum Flange Bolt Spacing
• Appendix 2-14(a) – Mandatory flange flexibility
• Appendix 7-1, 7-5 – Steel casting examination for 100% quality factor
• Appendix 9-8 – if the jacket does not carry lethal substances, lethal service restrictions do not apply to it
• Appendix 17-2(a) – Dimple jackets will not contain lethal substances
• Appendix 35-3(c) & 35-7(d) – Mass production rules
• Appendix W Table W-3 – Filing out the U forms
• Appendix KK (note 12) – Filling out User Design Requirements forms
• Appendix NN Table NN-6-5 – User or designated again determines if the contents are lethal
• End notes 65 – Definition of lethal service

Weld restrictions by category

Special lap joint stub end for lethal service – see also Fig UW-13.3 and Fig 2-4(5) and (6)

All partial penetration nozzles are prohibited

All figure UW-13.2 corner welds are prohibited

No slip on flanges

Code Cases

The following Code Cases refer to lethal service for VIII-1 vessels (2013 edition):

• 1750-20(g) – A126 not to be used for PRVs, or rupture disks.
• 2249 – Furnace Brazing
• 2318 – slip on flanges in lethal service
• 2321-1 PWHT of P4 and P5A materials
• 2324 – Ultrasonic inspection
• 2334 – single fillet lap joint tubesheet to shell connection
• 2346-1 & 2537 – Alternative rules for SE and F&D head to shell attachment
• 2369 – Testing of covers separate from vessels
• 2377 – RT of SA-612 steel
• 2421 – Single fillet joints in Heat exchanger tubes
• 2437 – Diffusion bonded heat exchangers
• 2527 – Pneumatic testing of UM vessels
• 2537 – alternative F&D and SE head to shell attachment is not for lethal service
• 2621 – Diffusion bonding
• 2751 – alternative spherical head to shell attachment is not for lethal service

Interpretations

With interpretations now available on line at cstools.asme.org/Interpretation/SearchInterpretation.cfm it is now easier to find relevant and up to date interpretations. A search for “BPV Section VIII Div 1” keyword: “lethal” returns many records. The comments on the most interesting lethal service interpretations are listed below. Refer to the actual interpretations for the full text.

Inspection Openings – Quick Guide

File: PVE-3856, Last updated June 6 2017 LRB / CT

Introduction

Requirements for inspection openings are based on UG-46 with additional code passages, code cases and interpretations.

ASME VIII-1 Inspection Opening References

UG-46 should be read in its entirety – a few brief points:

• UG-46(a) – Vessels with compressed air  or with corrosion, or with erosion, or with mechanical abrasion need inspection openings.  Exemptions:
  • Air dried to atmospheric dew point -50°F or lower, 
  • Shell side of fixed tube heat exchangers
  • Non-corrosive service
• UG-46(b) – use of telltale holes instead of inspection openings
• UG-46(c,d&e) – opening requirements and exemptions based on the diameter of the vessel
• UG-46(f&g) – what qualifies as an inspection opening based on size of the vessel and the opening size
• UG-46(h,i&j) – design of the opening

Other VIII-1 code passages related to inspection openings:

• UG-32(o) and UG-33(j) – use the rules of UG-36 through UG-46 when designing openings in heads
• UG-42(e) – threaded inspection openings size vs threaded nozzle size
• UG-45 – exemption to mininum nozzle thickness requirements for inspection openings
• UCI-36(a) and UCD-36(a) – UG-46 applies to inspection openings in cast vessels with thickness restrictions
• UCL-25(b) – telltale holes instead of inspection openings in clad vessels
• ULW-16(a) and ULW-18(a&c) – inspection openings in layered vessels
• App 9-4(b) – inspection openings in jackets
• M-2(b) – inspection openings shall be accessible

Code Cases

Code Case 2634 refers to exemptions to inspection openings:

• 2634 – Exemptions to inspection opening requirements when the vessel is contained in a sealed container

Interpretations

Some interpretations from cstools.asme.org/Interpretation/SearchInterpretation.cfm.  Refer to the actual interpretations for the full text.  Caution- old interpretations remain even after they have been superseded by code changes that render them obsolete.

Use of Type (3) welds in ASME Pressure Vessel Design

File: PVE-6022, Last Updated: March 23, 2012, By: LRB

Conclusion: Table UW-12 provides 8 types of welds with appropriate efficiencies to use when differing levels of radiography is applied. Confusion exists when trying to determine if a single welded circ weld in a small vessel is type 1 or type 3. Our answer based on experience and code interpretations is that it is a type (1) weld as long as the back side can be inspected; otherwise it becomes a type (3).

Definition of Type (1):

Joint Description: Butt joints as attained by double-welding or by other means which will obtain the same quality of deposited weld metal on the inside and outside weld surfaces to agree with the requirements of UW-35. Welds using metal backing strips which remain in place are excluded.

Efficiency: is 1, 0.85 or 0.7 depending upon the degree of radiography applied

A long seam welded from one side with no backing strip – always Type 1

Definition of Type (3):

Joint Description: Single-welded butt joint with- out use of backing strip

Restrictions: Circumferential butt joints only, not over 5/8″ thick and not over 24 inch outside diameter

Efficiency: is 0.60 only for no radiography

A circ seam on a small vessel welded from one side only with no backing strip – is it Type (1) or Type (3) ?

The confusion always comes from determining what is the equivalent of double welding? Can a joint be welded from one side only and still be considered to be the equivalent of double welding? ASME has a few of interpretations that directly address this issue:

Interpretation:	VIII-1-83-220
Subject:	Section VIII, Division 1; UW-12
Date Issued:	February 22, 1984
File:	BC83-557
Question (1):	For vessels of small diameter, not accessible for welding from the inside, as well as for vessels of large diameter where welding from the inside is possible, it is proposed to weld both the longitudinal and circumferential seams with single side full penetration welds. TIG and SMAW or TIC; and SAW processes with argon backing for the root run will be used. May these be considered to be Type 1 joints in Table UW-12 of Section VIII, Division 1?
Reply (1):	Yes.
Question (2):	Will the degree of examination affect the determination of the type of joints?
Reply (2):	No.
Interpretation:	VIII-1-83-291
Subject:	Section VIII, Division 1; UW-12
Date Issued:	June 29, 1984
File:	BC84-191
Question:	For vessels of small diameter, not accessible for welding from inside, as well as for vessels of large diameters where welding from inside is possible, it is proposed to weld both longitudinal and circumferential seams with single side full penetration welds. GTAW, GMAW, SMAW, and SAW processes with fiberglass tape backing for the root run will be used. May these be considered to be Type No. (1) joints as described in Table UW-12 of Section VIII, Division 1?
Reply:	Yes.
Interpretation:	VIII-1-83-267
Subject:	Section VIII, Division 1; Table UW-12, Joint Types
Date Issued:	May 31, 1984
File:	BC84-090
Question:	A circumferential joint of greater than 24 in. 0. D. is made with a single-welded full penetration butt weld. Is this a Type No. (1) joint as given in Table UW-12 of Section VIII, Division 1?
Reply:	Yes, provided the requirements in UW-35 and UW-37(d) are met.

So it is permissible to consider single sided welds as type 1. Is it also permissible to consider them as Type (3)?

Interpretation:	VIII-1-92-138
Subject:	Section VIII, Division 1 (1992 Edition, 1992 Addenda); Table UW-12
Date Issued:	March 17, 1993
File:	BC93-110
Questions:	Under the limitation requirements for Type No. (3) joints given in Table UW-12 of Section VIII, Division 1, may the following be single welded and still be in compliance using only circumferential butt joints: 1) Weld a vessel that has a 3/4 in. thick wall and is 20 in. in diameter? 2) Weld a vessel that has a 3/8 in. thick wall and is 30 in. in diameter?
Replies:	1) No. 2) No.

This interpretation prohibits the use of Type (3) joints regardless of the vessel size (over or under 24 inch in diameter). If Type (3) is not allowed, all that is left is Type (1).

Our experience indicates that the use of Type (1) joints is acceptable as long as the back side of the weld can be visually inspected after welding. We have been asked to use Type (3) occasionally – primarily when the back sides of welds cannot be inspected. About once every few years we will be asked to change from a Type (1) to Type (3) weld for other reasons – which we will make as requested. The change from Type (3) to Type (1) usually does not affect the design of a vessel as the long seam efficiency normally governs the design thickness.

Loads on Flanges- The ASME Way

File: N/A, Last Updated: Sept. 9, 2008, LB

Summary

ASME VIII-1 Appendix 2 provides a method of sizing flanges. The calculations use three loads – HT, HG & HD and two operating conditions – seating and operating. What are these loads, how are they calculated, and where are they applied to the flange?

In the article below, a sample flange is calculated using ASME Appendix 2 methods and by finite element analysis (FEA) to illustrate the application of the loads and show the resulting stresses.

Downloads

• Loads on Flanges Article

Vessel with a Large Opening

Date: June 16, 2010, By: LB

A Collection of Unusual Design Features

This sample vessel is not modeled after a real vessel. It is a collection of difficult design features and seldom used code requirements: large nozzles, swing bolt covers, cone discontinuities, use of bar stock for nozzles (code case 2148) and the use of sanitary ferrules in vessels. Refer to the calculation sets for more details.

Unusual Flanges

The swing bolt cover is analyzed as an appendix 2 flange. The bolt circle is outside of the flange, which is a length of increased wall thickness pipe. The bolt loads try to twist the pipe inside out – the Appendix 2 calculations check for this. Additional calculations are run for the attachment lug weld and shear pin stress. Flange C is also a custom flange calculated to Appendix 2.

Ferrules

Ferrules and sanitary connections can not be calculated by the rules provided by Appendix 24 which requires metal to metal contact outside the gasket. Typically, ferrules need to be proof tested, or calculated by Finite Element Analysis. For this vessel, the manufacturer of the 2″ ferrule has provided a CRN number covering the design. The CRN implies that either proof testing or some other calculation method was used to prove the design. The 8″ ferrule does not have a CRN, here a proof test is specified.

Large Nozzles

The side nozzle C is checked against the rules of Appendix 1-7 because it is larger than 1/2 of the vessel body diameter. The rules of Appendix 1-7 can be confusing and difficult to interpret, so in this case, all the conditions regarding moments of inertia and area replacement have been applied.

Bar Stock

The rules that allow the use of bar stock in pressure vessel bodies or nozzles have been changing. At the time this sample was made, bar stock could only be used with code case 2148. This code case has since been annuled and later reinstated as 2148-1. See UG-14 and App 2-2(d) for up to date information on the use of bar stock in pressure vessels.

Example calculation sets using Advanced Pressure Vessel (APV, renamed to DesignCalcs) and PVEng Spreadsheets can be downloaded from the links below.

Downloads

• APV Calculations
• Spreadsheet Calculations
• Drawing

Nozzle Rules

Origins of the ASME area replacement rules?

File: PVE-2461, Last Updated: Oct. 11, 2007, By: LB

Origins:

In the book “The new Science of Strong Materials or Why You Don’t Fall Through the Floor” by J. E. Gordon, 1968, Penguin Books, there is a section quoted below (Page 60) dealing with tubular shapes like railway bridges and ships.

A ship is a long tube closed at both ends which happens to be afloat but is not otherwise structurally very different from Stephenson’s Menai bridge. [topic of the authors previous paragraph] The support which the water gives to the hull does not necessarily coincide with the weights of engines, cargo and fuel which are put into the ship and so there is a tendency for the hull to bend. It ought to be impossible to break a ship, floating alongside a quay, by careless and uneven loading of the holds and tanks, but this has happened often enough and will probably happen again. In dry-dock ships are supported with care upon keel-blocks arranged to give even support but there is not much even support at sea where a ship may be picked up by rude waves at each end, leaving her heavy middle unsustained, or else exposing a naked forefoot and propeller at the same moment. As ships tended to get longer and more lightly built, the Admiralty decided to make some practical experiments on the strength of ships. In 1903 a destroyer, H.M.S. Wolf, was specially prepared for the purpose. The ship was put into dry-dock and the water was pumped out while she was supported, in succession, amidships and at the ends. The stresses in various parts of the hull were measured with strain-gauges, which are sensitive means of measuring changes of length, and therefore of strain, in a material. The ship was then taken to sea to look for bad weather. It does not require very much imagination to visualize the observers, struggling with seasickness and with the old-fashioned temperamental strain-gauges, wedged into Plutonic compartments in the bottom of the ship, which was put through a sea which was described in the official report as ‘rough and especially steep with much force and vigor’. Her captain seems to have given the Wolf as bad a time as he could manage but, whatever they did, no stress greater than about 12,000 p.s.i. or 80 MN/m2 could be found in the ship’s hull.

As the tensile strength of the’ steel used in ships was about 60,000 p.s.i. or 400 MN/m2, and no stress anywhere near this figure could be measured, either at sea or during the bending trials in dry-dock, not only the Admiralty Constructors but Naval Architects in general concluded that the methods of calculating the strength of ships by simple beam theory, which had become standardized, were satisfactory and ensured an ample margin of safety. Sometimes nobody is quite as blind as the expert. Ships continued to break from time to time. A 300-foot (90 meters) ore-carrying steamer, for instance, broke in two and sank in a storm on one of the Great Lakes of America. The maximum calculated stress under the probable conditions was not more than a third of the breaking stress of the ship’s material. Even when major disasters did not actually happen, cracks appeared around hatchways and other openings in the hull and decks.* These openings are of course the key to the problem. Stephenson’s tubular bridge was eminently satisfactory because it is a continuous shell with no holes in it except the rivet holes. Ships have hatchways and all sorts of other openings. Naval Architects are not especially stupid and they made due allowance for the material which was cut away at the openings, increasing the calculated stresses around the holes pro rata. Professor Inglis, in a famous paper in 1913, showed however that ‘pro rata’ was not good enough and he introduced the concept of ‘stress-concentration’ which, as we shall see (Chapter 4), is of vital importance both in calculating the strength of structures and in understanding materials.

What Inglis was saying was that if we remove, say, a third of the cross-section of a member by cutting a hole in it then the stress at the edge of the hole is not 3/2 (or 1.5) of the average but it may, locally, be many times as high. The amount by which the stress is raised above the average by the hole – the stress-concentration factor – depends both upon the shape of the hole and upon the material, being worst for sharp re-entrants and for brittle materials. This conclusion, which Inglis arrived at by mathematical analysis, was regarded with the usual lack of respect by that curiously impractical tribe who call themselves ‘practical men’. This was largely because mild steel is, of all materials, perhaps the least susceptible to the effects of stress concentrations though it is by no means impervious (Plate 3). It is significant that, in the Wolf experiments, none of the strain gauges seems to have been put close to the edge of any important opening in the hull.

Is this really the origin of the Area replacement (or pro-rata) rules that we use in pressure vessels? Is it just a set of rules that failed when applied to ships but have been successfully applied to pressure vessels? Yes these pro rata rules are still in use in the ASME pressure vessel and piping codes. Basically, when we remove some material from a vessel in the form of a nozzle opening, we look for an equal amount of extra material to replace it, in both the surrounding shell, and in the nozzle pipe.

Picture that the stresses in vessels were too complicated to accurately understand at the time, so a rule like area replacement was adopted from another field like naval architecture. Or it is independently developed by the original pressure vessel designers (our rules UG-36 to 43). The designs work well most of the time but occasionally a pressure vessel blows up (this is after all an experience based code). With more experience, more restrictions like appendix 1-7 are added and our vessels fail less often. Are designers ignoring the intent of the code but purely following the rules – the more specific restrictions on the geometry are added – (like UW-14 to 16). And so it goes – we are still changing the pressure vessel code today. At no point is the problem fully understood but pressure vessels gradually get more reliable. We can expect more changes in the future…

Warning:

The ASME nozzle area replacement rules cannot be taken on their own. There are a large number of code sections that need to be considered on each nozzle – UG-36 to 43, App 1-7, UW-14 to 16, UG-45 and others. The rules explicitly only apply to circular, obround and elliptical openings – for the last two, the length cannot be more than twice the width. In practice, these limits are commonly violated.

The amount by which the stress is raised above the average by the hole – the stress-concentration factor – depends both upon the shape of the hole and upon the material, being worst for sharp re-entrants and for brittle materials [J.E. Gordon, copied from the above quote].

Brittle materials are not much of an issue with modern pressure vessels, however the shape of opening issues still remains. Also, we do not distinguish between nozzles attached to high stress areas of vessels like knuckles on heads, or lower stress areas like straight shells. 

Nozzle F Factor

Using the ASME VIII-1 Nozzle F Factor (UG-37)

File: PVE-3783 Jan. 24, 2013, LRB

A pressure vessel with two identical elliptical nozzles, but oriented in different directions

This is a FEA plot of a pressure vessel with two identical elliptical nozzles, but oriented in different directions. ASME says that the two nozzles have different stresses around them, as the FEA results confirm. A cylindrical shell circ stress is 2x the longitudinal stress. The nozzle that cuts more material in the long direction has higher stresses.

• The default F factor is 1.0 – this effect can be ignored if desired.
• F Factor can reduce the required amount of area replacement to 1/2 in certain directions – this allows less conservative nozzle designs if the non-round nozzle is oriented favorably.
• F Factors other than 1.0 can only be used for integral (full penetration welded, no re-pad) nozzles.
• The nozzle will need to be calculated twice – once in the longitudinal direction at F = 1.0 and once in the circ direction at F=0.5. Different d values will be used for the different directions.

An example follows:

F correction factor

F = correction factor that compensates for the variation in internal pressure stresses on different planes with respect to the axis of a vessel. A value of 1.00 shall be used for all configurations except that Figure UG-37 may be used for integrally reinforced openings in cylindrical shells and cones. [See UW-16(c)(1).]

ASME figure UG-37. At an angle of 0°, the maximum circ stress exists, F = 1.0. At 90°, the maximum longitudinal stress exist, which is half the circ stress. F = 0.5

Companion Sample Problem and Calculation Set

The enclosed example shows an elliptical manway nozzle that takes advantage of the F factor to get a higher pressure rating than otherwise possible.

Calculation Set

Making Repads Using Offset Surfaces

PVE-7253, Last Updated Nov 27 2015, By: Cameron Moore, Jordan Winger and Laurence Brundrett

Here is another repad modelling method, this one for more difficult designs where the two methods presented below will not work. The repad can be built onto complex surfaces. Developed by Jordan Winger of Hawk Ridge Systems, shown here with a few minor changes.

The finished product – a branch connection covering a swept feature with more than one radius, welds are not shown.

The start of the development: A flanged and dished head has a straight flange, a knuckle radius and crown radius. Normally when swept 3 different surfaces are created. These 3 surfaces lines are combined to create 1 fit spline.

This is important – once revolved, the fit spline creates a solid head with only one outer and one inner surface. This allows the repad to extend beyond the limit of the crown onto the knuckle and if required onto the shell as well.

A nozzle is extruded onto the head, it is not merged. No hole is cut yet.

A circle radius of the desired repad size provides the profile, and the end of the nozzle provides the path for a sweep. The radius of the profile provides a distance to the limit of the repad, not the true developed width of the repad which is larger due to the curvature.

The top surface of the head is duplicated.

A line at the intersection of the swept shape and the duplicated top surface defines the limits of the repad.

The duplicated top surface is trimmed to the intersection line.

Thicknened…

…inside cut and combined for the finish!

Making Repads Using the Flex Command

PVE-7253, Last Updated Nov 26 2015, By: Cameron Moore and Laurence Brundrett

Here is another method of making repads, this time using the Flex command. When this method works, it is simpler than using surface commands. This method developed by Cameron Moore here at PVEng is particularly useful because it creates a properly contoured repad that can be used for a Drawing or for Finite Element Analysis.

The finished product – a branch connection complete with repad – welds are not shown.

The start of the development: two solid parts – 1) the header and 2) OD and thickness of the final repad. The Repad is drawn on a plane tangent to the OD of the header.

This is the magic – the repad is flexed to match the OD of the header. The repad edge is perpendicular to the Header the whole way around.

The branch is extruded from a new plane at the end of its extent, down to the repad. The extrusion command merges the header, repad and branch.

Finally the branch hole is cut finishing the model. Welds can be added – there are no gaps for easier FEA analysis.

Making Size on Size Branch Connections with Repads

PVE-7253, Last Updated: July 29, 2013, By: Cameron Moore and Laurence Brundrett

Modeling size on size connections with repads can be difficult. This method developed by Cameron Moore here at PVEng is particularly useful because it creates a properly contoured repad that can be used for a Drawing or for Finite Element Analysis.

The finished product – a size on size branch connection complete with repad – welds are not shown

The start of the development: two surface parts – 1) the OD of the branch and 2) the OD of the Repad on the Header are developed.

This is what makes this method work so well – the repad outer surface is trimmed to the OD of the branch pipe…

… and extended an equal amount all the way around. Note that the repad is bent under itself as it goes around the branch pipe. A projected cut from the branch onto of the OD of the repad would not work.

Thickening the repad converts it from a surface to a solid. Note that the edges are perpendicular – a feature that cannot be obtained by using projected cuts.

A solid header has been created and the surface branch has been replaced by solid pipe. Here shown in section view.

An opening in the header for the branch completes the model unless repad and header fillet welds are also required…

… addition of the welds.

The solid model made this way is without gaps making for easy FEA analysis!

Spreader Bar Sample

File: PVE-3857, Last Updated: Feb. 1, 2013, By:CBM

This spreader bar design is to the new ASME BTH-1 “Design of Below-the-Hook Lifting Devices”. Labelling is per ASME B30.20. We can provide P. Eng. sealed drawings and calculation sets like the ones shown below.

Downloads

• Calculations
• Drawing