Pad-Reinforced Nozzles in Pressure Vessels

This article examines the application of reinforcing pads for strengthening nozzles in pressure vessels and the impact of the geometric gap that exists between the pad and the vessel shell

Many common forms of nozzle connections in technological applications are subjected to internal pressure and external loads. Pressure-vessel design engineers are often confronted with various considerations that they have to understand with regard to the nozzle design. Practice shows that the traditional design codes do not always provide adequate solutions for specific cases and do not sufficiently address relevant points that are important to ensure structural integrity, including the reinforcing pads that are often used for strengthening nozzles in pressure vessels that are subjected to internal pressure and piping reactions (Figure 1). Moreover, it cannot be confirmed that the resistance to gross plastic deformation (GPD) of a pad-reinforced shell is exactly equal to the resistance of a single-wall shell whose thickness is equal to the sum of the pad and shell. GPD under static loading is a fundamental failure mode that must be considered in the design by analysis (DBA) method for pressure vessels. In elastic DBA, GPD is prevented by limiting the primary stress in the vessel. In inelastic DBA, GPD is prevented by limiting the load applied to the vessel, restricting it to a fraction of the notional ductile collapse load of the vessel.

There are a number of parameters that are important to consider for pad-reinforced (repad) nozzles. This article discusses several aspects that design engineers must take seriously in order to ultimately achieve a prudent and safe nozzle design.

 

Characteristics of repads

Perfect contact between the vessel shell and the pad cannot be achieved for several reasons. First, this contact adversely affects the stress distribution in the vicinity of the nozzle intersection and results in a lower internal pressure capacity compared to a fully integral connection without a gap.

However, in most design codes, the gap between the pad and the shell, which prevents full contact between the two elements, is ignored, and the thickness of the pad plus the shell is considered integral, which raises some questions. Ref. 1 provides additional insight into the effect of a geometric gap between a cylindrical or spherical shell and a reinforcement pad.

The statement that a gap may (but not always) exist between the pad and the main vessel shell is in fact correct. This is due to the skill level of the sheet-metal worker who rolls and forms the material to make the pad.

All the nozzle-reinforcing calculations assume the pad, weld, shell and nozzle form an integrated assembly — this also accounts for internal pressure and for external piping loads.

However, there have been some occasional instances where poor fit does indeed play a part, especially if there is thermal expansion and no weep hole. A weep hole is a small hole drilled in the pad to allow any gap moisture to be released. If there is no weep hole, then the gas in the inner space can expand. In such cases, it is possible for the pad to “blow off” and crack the welds or the vessel itself.

 

Discussion

The region of intersection of a nozzle with a pressure vessel is generally one in which high stresses exist. Consequently, the designer of a pressure vessel must have means for accurately predicting the magnitudes and locations of maximum stresses in these regions.

According to Welding Research Council (WRC) Bulletins 107/297/368, which are frequently used in performing nozzle load analyses, if the width of the shell reinforcement (pad plate) is equal to or greater than the below expression, where R is the shell radius, T is the vessel thickness and d is the nozzle diameter:

 

1.65( RT) ^0.5 or d/2,

 

then it can be assumed that the thickness of the reinforced portion of the shell would determine the state of stress at the nozzle-to-shell junction. In such a case, the T value to be used in equations for calculating stresses should be T (the shell thickness plus the pad plate thickness). It is generally assumed that discontinuity stresses due to the penetration have been reduced to negligible levels at a distance of 1.65(RT)^0.5 or d/2, whichever is greater, from the nozzle neck. Stress attenuation is the key here.

Ref. 2 recommends that in finite element analysis (FEA) studies, the analysis not be based on the total pad plus vessel thickness, but on a thinner effective pad thickness that represents the bending stiffness of a separate pad and vessel shell, but is conservative for local membrane stress. When the local membrane stresses govern the maximum loads, an additional FEA evaluation should be performed using full thickness. The effective vessel shell plus pad combined thickness (t_e) is given by the following equation, where: T_v represents vessel thickness and t _p is the reinforcing pad thickness:

 

t_e = (T_v^2.5 + t_p^2.5) ^0.4

 

Ref. 3 states that if the reinforcing pad of adequate width has been applied, the total shell thickness can be assumed as the corroded shell thickness plus the pad thickness. Criteria for adequate pad width is as follows:

 

0.5 t_s ≤ t_p ≤ 1.5t_s

 

W ≥ r [(t/t_s) ^0.75 – 1]

 

W ≥ 0.611 (R t ) ^0.5

 

Where:

W = width of reinforcing pad (mm)

R = external radius of shell (mm)

r = external radius of nozzle (mm)

t_s = thickness of shell (mm)

t_p = thickness of reinforcing pad (mm)

t = combined thickness of shell and pad (mm)

 

Note that if the pad is not wide enough, then t should be replaced by a reference thickness t_r. To determine t_r, use the lesser of the following:

 

t_r = 2.678 W²/R

t_r = t_s[(r + W)/(r)]^4/3

 

Example

Consider an example of a cylindrical shell with an outside diameter of 1,600 mm and a wall thickness of 20 mm. A 16-in. NPS (NB 400) flush nozzle is fitted in the cylindrical shell, which is provided with a reinforcing pad with a thickness of 12 mm and a width of 80 mm. Table 1 outlines the relevant equations and results for this sample calculation.

Since the requirements of Equations (2) and (3) are not met, it is not permitted to take the total thickness of the shell plus reinforcing pad into account. This means that a substitute or equivalent thickness of 21.424 mm must be applied. It should be noted that this result differs little from the equivalent thickness calculated with Equation (1). The difference is approximately 3%.

Refs. 4 and 5, which use the so-called pressure-area method for the nozzle compensation calculation, which is common in European codes, introduces a reinforcement efficiency factor k of 0.75 for the reinforcement pad. This means that the cross-sectional area of the pad within its defined boundaries should be multiplied by this factor in order to compensate for the gap. It should be noted that the implementation of this k factor applies specifically to the design code mentioned here (from the Netherlands) and has not been adopted by the other European codes like EN 13445 (EU), PD 5500 (U.K.), AD 2000 (D) and CODAP (F).

Note that ASME BPVC Section VIII-Division 1 — paragraph UG-37 (g);(h), however, indicates that the area A5 being the area of ​​the reinforcing pad should be multiplied by 0.75 if some provisions are not met.

 

Findings

It can be concluded that views regarding the contact problem between the repad and shell — and the way in which this situation should be anticipated — differ considerably. A more in-depth investigation into this by means of a numerical analysis method (such as FEA) under different loading regimes (internal pressure and imposed pipe loads) and simulating the gap in combination with varying contact surfaces between the repad and shell could lead to workable solutions. However, modeling of pad-reinforced nozzles is quite complex due to the possible gap between the pad and shell, or to some extent, the lack of contact between both elements [6]. Some computer programs are available to design engineers that offer solutions for this. However, it is preferable to develop a specific analytical approach. The vessel design engineer must be aware that ignoring the effect of the gap can result in a non-conservative prediction of the load capacity. ■

 

References

1. Stikvoort, W., Effectiveness of reinforcement plates pertaining to pressure equipment, American Journal of Engineering Research, Vol. 10, pp. 127–146, August 2021.

2. Koves, W., and others, Establishing Allowable Nozzle Loads, Proceedings of the ASME 2021 Pressure Vessels & Piping Division Conference, Baltimore, USA, July 2021.

3. Stikvoort, W., Piping Reactions on Pressure-Vessel Nozzles, Chem. Eng., pp. 51-53, July 1986.

4. Rules for Pressure Vessels — Chapter D 0501, Openings in a curved wall: Strength reduction coefficient, reinforcement, Issue 04-11, Sdu Publishers, The Hague, The Netherlands.

5. Schwaigerer, S., Mühlenbeck, G., “Festigkeitsberechnung im Dampfkessel – Behälter – und Rohrleitungsbau,” 5 ^th Edition, Springer Verlag Berlin Heidelberg, 2012.

6. Chen, H., Chao, Y., Contact Between Vessel Shell and Welded Pad in Nozzle Reinforcement, Journal of Pressure Vessel Technol., pp. 364–372, Nov. 1993.

 

Acknowledgement

All figures provided by author unless otherwise noted

 

 

Author

Walther Stikvoort (Email: [email protected]) is a renowned authority in the field of mechanical and structural integrity of static pressure equipment. He has more than 50 years of experience in pressure vessel and piping design and has developed numerous technical standards and practices to improve the asset integrity of leading operating companies. He is the author of numerous peer-reviewed international journal articles in the field of mechanical and structural integrity. During his career, he was regularly active in developing and teaching courses and training to mechanical engineers in his area of expertise and he was a member of various expertise committees. He is currently active as a consultant on static pressure equipment integrity serving the engineering community on request.

 

Further reading

1. Benefiting from Nozzle Flexibility in Piping Design, Chem. Eng., August 2024, pp. 33–40.

2. Bolt-Load Considerations Associated with ‘Hot Bolting,’ Chem. Eng., August 2024, pp. 31–44.

3. Storage Tanks: Snapshots of Failures, Damages and Inspections, Chem. Eng., December 2019, pp. 34–37.

4. Pressure-Vessel Quality Control Requirements, Chem. Eng., January 2014, pp. 28–35.

5. Decoding Pressure Vessel Design, Chem. Eng., June 2010, pp. 28–35.

6. Piping Design, Part 2 — Flanges, Chem. Eng., March 2007, pp. 56–61.