Ever plot CML thickness and notice your piping or equipment growing over time? 

This “reverse corrosion” phenomenon sweeping inspection departments can defile your data and give false impressions of real metal loss and corrosion rates. 

How common are growth readings for CMLs? VERY. In fact, during Jeff Goldstein, P.E.’s latest webinar “Rethinking Piping Inspection Locations” he discussed a site with 12,000 of 40,000 CML readings showing growth over the last CML reading. That is 30% of readings showing a growth rate instead of an expected metal loss/corrosion rate. What you decide to do with these growth readings will have a big impact on maintenance and inspection plans moving forward. 

Maybe there was an undocumented piping or equipment replacement. 

Maybe the measuring device was not properly calibrated when the operator took CML readings. 

Maybe the CML readings were entered into software or spreadsheets incorrectly. 

Maybe the CML readings were taken in the wrong location.

Please spend some time trying to understand what is causing growth readings and act accordingly.

Have you started using the Appendix 46 solution for external pressure calculations?

I encourage you to become familiar with the Appendix 46 external pressure solution as the more accurate calculation method provides benefits over the Division 1 solution. Try creating a spreadsheet for the Appendix 46 external pressure method and compare it to the Division 1 solution over a wide range of diameters and materials. This helped build my confidence in the method and will build your confidence too. 

Take a look at this simple 8’ inner diameter carbon steel vessel at 425°F where the Appendix 46 solution passes the vacuum case and the Division 1 solution fails the vacuum case.

Can you use internal nozzle area to pass UG-37 area-of-replacement requirements?

Yes, you can! 

Any area within the limits of reinforcement counts as area towards the UG-37 area-of-replacement requirements, including internal nozzle area. Adding an internal nozzle projection is a way some designers meet the area-of-replacement rules when they need a little extra reinforcement. 

However, make sure that the process department for your Owner/Operator is okay with using internal nozzle projection in your designs before going forward. Some specification sheets allow internal projection for inlet nozzles but not for outlet nozzles. Others do not allow internal projection at all!

Ever see ASME B16.5 blind flanges as reducers?

Just throw any pipe in a standard ASME B16.5 blind and you’re good-to-go, right? 

No. That “standard” ASME B16.5 blind flange may not be “standard” once you decide to start drilling holes in it. Many times ASME B16.5 blind flanges used as reducers include pipes larger than allowed in B16.5 without a reinforcement calculation. Please make sure your facility has proper documentation for ASME B16.5 blind flanges used as reducers.

Are you using the Division 2 nozzle rules yet?

I know a lot of Owner/Operators are afraid of using the Division 2 rules because they are weary of using higher allowable stresses. Did you know that you can apply the lower Division 1 allowables with the Division 2 rules by using Appendix 46 in Division 1 designs? This means engineers like me can still save big money where the Division 2 rules offer a clear advantage. This is especially evident in nozzle design where the Bildy rules often avoid extraneous nozzle reinforcement compared to the Division 1 area-of-replacement rules.

Do you verify thicknesses before receiving your equipment?

The design calculations use NPS 8 Sch 120. 

The fabrication drawings use NPS 8 Sch 120. 

You received NPS 8 Sch 80 (XS). 

So what happened? A drawing was misinterpreted and a fabrication mistake was made. Fabrication mistakes happen. Don’t compound those mistakes by not finding them until after installation. Using a 3rd party to verify that your as-built thicknesses meet or exceed your design thicknesses is a quick and important step you should implement before receiving your equipment to help improve your bottom line.

I like to use 3rd parties for thickness verification because they are excellent at communicating between fabricators and Owner/Operators during stressful projects. I trust Turnaround EPC for my 3rd party as-built thickness verification because of their excellent communication, fast service, and attention to detail.

Do you have an issue with leaky gaskets?

A lot of people message me complaining about leaky flange gaskets and wondering what kind of gaskets I recommend to help. The fact of the matter is that gasket quality and performance has never been better. I have found that many times the real problem is that the gasket was not considered during the design phase. Gaskets are an important part of flange design. Treat them that way. 

Have you performed ASME PCC-1 Appendix O for your flanges? 

An approach in ASME PCC-1 Appendix O can help you make sure your gaskets are being seated correctly. The gasket needs to be below the maximum permissible gasket stress, above the minimum seating gasket stress, and above the minimum gasket operating stress. In this example the minimum gasket operating stress is not being maintained at the operating condition. 

Make sure you perform ASME PCC-1 next time you open up a pair of flanges.

Do you consider longitudinal stresses when you check for brittle fracture?

I’ve had a lot of people asking me to check their hydrotest temperatures. A design where a governing longitudinal stress case was not considered in the Minimum Design Metal Temperature (MDMT) stood out to me. 

Did you know that longitudinal stresses can govern the MDMT even if circumferential stresses govern the Maximum Allowable Working Pressure (MAWP)? Not checking the longitudinal stresses can lead to devastating results as the brittle mode of failure is swift and catastrophic. Always remember to consider longitudinal stresses in your brittle fracture check.

Did you know you can design safer and save money using ASME Division 2 rules instead of Division 1? 

There are many ways to take advantage of the higher allowables and more accurate numerical methods in Division 2. One such method to consider is using the Bildy Rules instead of Area of Replacement for nozzle design. 

Take a look at this simple nozzle internal pressure case:

Design Pressure = 290 psi (2000 kPa)

Design Temperature = 475°F (246°C)

Nozzle Material = SA-106 C Smls Pipe

Nozzle Size = NPS 12 Sch 140

The Division 1 vessel does not pass the design condition of 290 psi (2000 kPa) @475°F (246°C).

The Division 2 vessel passes the design condition of 290 psi (2000 kPa) @475°F (246°C).

Are you ready for AI? Your data is not.

One of the big themes of Reuters Downstream 2025 was the use of AI in maintenance and inspection. I found the concept of using AI for predictive analytics in maintenance incredibly intriguing. However, let’s not pretend that your important data is in a format that AI can actually use. You need to digitize your data.

Are your U-Forms sitting in a filing cabinet somewhere?

How about the historical measured thickness data from internal inspections? Are those values just floating around in PDFs?

What about the interesting information found during external inspections? Repairs? Rerates? Derates? FFS assessments? Wouldn’t you want all of that data digitized in a way that makes it easy for AI to pull from?

We all know that AI is the future but AI is only as good as the data you are feeding it. Consider yourself behind if the items listed above are not in a usable, digital format.

Can you save money using ASME Division 2 rules instead of Division 1? 

There are many ways to take advantage of the higher allowables and more accurate numerical methods in Division 2. 

Take a look at this simple external pressure (vacuum) case:

Outside Diameter, Do = 60.75” (1543 mm)

Thickness, t = 0.375” (9.53 mm)

Unsupported Length, L = 264” (6706 mm)

The Division 1 vessel does not pass the vacuum condition of 14.7 psi (101 kPa).

The Division 2 vessel passes the vacuum condition of 14.7 psi (101 kPa).

Have you seen a tower fall during Post Weld Heat Treatment (PWHT)? 

As crazy as it sounds I’ve seen this occur when PWHT is applied around the circumferential seams of tall towers in the field. The solution is often even crazier as Owner/Operators prefer expensive materials analysis and complex FEA calculations over simple explanations. The fact is the tower behaved exactly as expected as it was never designed to take such loadings at temperature. 

So what is actually happening?

1 - The Wind case was not properly considered by the design engineer

2 - The Wind case in the Empty condition was not calculated

3 - The much lower allowable stress in the material at temperature was not considered

4 - Wind loads during PWHT create a large moment that the material was not designed to resist 

5 - DOWN IT GOES!!!

Do you use calculation exemptions on important process nozzles?

Yes, we all know nozzles under a certain size are exempted from the Division 1 nozzle area-of-replacement rules. That does not mean that you should use this exemption every time you can. Think very carefully about what each nozzle is used for and what the consequences of failure may be. 

You may want to double-check exempted nozzles against another set of nozzle rules before finalizing your design. Are you around 90% of area-of-replacement? Fine. Are you around 50% of area-of-replacement like my example? You may want to add a little thickness to the nozzle. This is a bad place to try to save money on your design. 

What if local corrosion is higher than the design corrosion?

There is no nozzle thickness to compensate for corrosion. Through wall corrosion is common.

What if nozzle loadings are higher than anticipated?

There is no extra thickness to handle pressure plus unexpected loadings. 

What if an upset condition increases the process velocity?

There is no extra thickness to handle process erosion.

Why would anyone de-rate a pressure vessel or heat exchanger? What is the point of formally accepting less out of your equipment than it was designed for?

The simple answer is to reduce the operational risk at your facility by ensuring that pressurizing to the original Maximum Allowable Working Pressure (MAWP) is never attempted. The most common scenario I observe is when a design with no corrosion allowance does not align with measured in-service corrosion.

Many times pressure vessels and heat exchangers operate much lower than their ratings so de-rating causes no operational disruptions. Is there any equipment in your facility that could use a de-rate?

Like to use ASME B16.9 Elbows as main pressure containing components? Please remember to treat them as seamless pipe and to calculate the required thickness by outer diameter. This means using Appendix 1-1 instead of UG-27 for internal pressure in Division 1.

I see off-the-shelf ASME standard flanges and pipe fittings often used as main pressure containing components in pressure vessels and heat exchangers. I’ve noticed the biggest variance in approach when I investigate ASME B16.9 Elbow calculations. Some design engineers believe that ASME B16.9 Elbows do not require any calculations at all. Others calculate ASME B16.9 elbows using UG-27. Very few use Appendix 1-1. 

Notice that Appendix 1-1 may require extra thickness compared to UG-27. Please refer to UG-44 for more information regarding the calculation of standard flanges and pipe fittings.

I think it’s time we talk about “RT 4” vessels and heat exchangers. I see them everywhere in corrosive processes and highly consequential service and I really wish I didn’t. 

So what is “RT 4”?  Well “RT 4” simply means that “RT 1”, “RT 2”, or “RT 3” do not apply. It does not mean anything else specifically about what kind of weld inspection was applied to the vessel. 

How much radiography or other weld inspection is applied to a vessel marked “RT 4”? Some? None? I don’t know, but probably not a lot. I personally assume that no weld inspection was performed unless proven otherwise. This is where I have a problem with “RT 4” vessels and exchangers. I don’t feel comfortable placing vessels and heat exchangers in service without having some sort of assurance regarding the quality of the welds.   When I uncover the root cause of unplanned shutdowns too many times these “RT 4” vessels are to blame.

I consider the cost of maintenance and inspection in my designs and I do not believe any money is saved in the long run by operating “RT 4” vessels.

Let’s talk about my favorite product form “Welded and Seamless Pipe”. 

Is it Welded? Is it Seamless? Is it Both? Wait, that’s impossible…

Well then what is a “Welded and Seamless Pipe”? I have no idea. What I do know is that there is a big difference between how welded pipe and seamless pipe are handled when it comes to ASME Code calculations. You need to know exactly which version of the material you have before performing calculations or there is a good chance you will be making a big mistake. 

Let’s take a look at SA-312 TP316L, which includes the seamless and welded pipe product form in the same specification. 

Seamless: No need for a longitudinal weld joint efficiency factor in the allowable stress. Double-check to make sure the allowable stress reflects this. 

Welded Pipe Made without Filler: The longitudinal weld joint efficiency factor needs to be included in the allowable. Notice the allowable is (15,300 x 0.85 = 13,005). There is no need to count this twice and multiply by 0.85 again in the joint efficiency value.

Welded Pipe Made with Filler: The joint efficiency factor was not included in the allowable so it needs to be taken into consideration in the joint efficiency value itself. Keep in mind that the joint efficiency value may be lower than 0.85 depending on the radiography performed.

Are your bolts designed to handle the uplift from moments caused by external forces like wind and seismic loads? 

I commonly see equipment where only the operating case is checked and I’ve even seen cases where the bolt uplift calculation is ignored altogether. Don’t be fooled into ignoring this simple calculation as catastrophic bolt failure can cause severe damage to your equipment.

The empty condition often governs because process weight helps to prevent uplift. Please calculate all design conditions to determine the governing case.

Are you getting the credit you deserve?

The ASME Code allows you to take mechanical strength credit for cladding in design calculations. Most design engineers do not take credit for cladding and end up purchasing a thicker shell than necessary. 

Take a look at this simple internal pressure with cladding example:

Design P = 350 psi (2413 kPa)

Shell Material = SA-516 70

Cladding Material = SA-240 316L

MAWP from shell component: 343 psi (2365 kPa)

MAWP from cladding component: 40 psi (276 kPa)

MAWP from shell plus cladding: 383 psi (2641 kPa)

Using only the shell component for mechanical strength fails at the Design Pressure of 350 psi (2413 kPa).

Using the shell plus cladding for mechanical strength passes at the Design Pressure of 350 psi (2413 kPa).

Can your saddle bolt holes handle the effects of thermal expansion in your equipment?

Bent saddles and sheared bolts from undersized slotted bolt holes pose a major operational risk that is often overlooked. It is much easier to design for thermal expansion before fabrication than it is to try to fix an inadequate slotted bolt hole length while in service. There is a simple calculation to determine if your slotted bolt hole length is enough to handle thermal expansion at design temperature. 

Please review PIP VEFV1100 for more information regarding the design of supports and other attachments.