Pardon me while I take cover under my desk, there may be incoming from across the street.


Sure it is, 14th post down in that thread, "As the steel industry has modernized, the newer techniques (electric furnaces and the like) have resulted in a much cleaner product than what was available in the “steam era”. Ironically, the dirtier the steel, the more corrosion resistant it is. I asked one of the metallurgists where we could go to get some good old dirty open hearth or Bessemer steel, and he thought Russia or China, but advised me “not to go there”."

As he explained it to us at the meeting, the inclusions and impurities act as speed bumps to the corrosion which cruises unhindered through clean steel.

At the risk of completely hi-jacking this thread, here is some more information regarding the questions about steel and corrosion resistance. The terms "dirty" and "clean" for steel is really a poor use of terminology. Pure steel, which is the only truly "clean" steel, is a binary alloy of iron and carbon. However, no such product actually exists for a myriad of reasons. So, in effect, all steel is "dirty" for one reason or another. Using the term "dirty" in this case should not be taken to mean steel with manufacturing defects like inclusions or laminations, but merely chemically speaking. The dirt in plain carbon steel is actually a number of different elements present in the final product of the steel manufacturing process and in varying amounts. Some of the dirt is put there intentionally; some remains in the final product because the cost to extract it would outweigh the benefit gained from its removal. Each element present in steel has an effect on the physical properties of the metal corresponding to its proportion. Some of these properties are desirable; others are detrimental depending on the intended use of the product. For example, manganese, which was the principal oxygen killing element added to older steels, increases hardness and wear resistance in steel, but reduces machinability and ductility.

Why does "dirty" old steel often resist corrosion better than new steel? Simply, it contains more of the dirt that resists corrosion, namely the elements nickel, chromium and especially copper. Older steel making processes were conducive to this because they were made from a smaller percentage of scrap steel which already had a larger proportion of impurities removed during the initial smelting and also because steel makers did not have the tools at their disposal to make more pure steels. The net result of this is that these steels contained larger residual amounts of alloying agents which impart their anti-corrosive properties while not being too great to alter the other properties of the metal to a degree that they would not answer for service in boilers.

To back this up, not only does this hold true in theory, empirical data gathered through chemical analysis of steel samples taken from boilers showing more or less corrosion tend to support this idea. It isn't as simple as dirty vs. clean, but a matter of what kind of dirt is present and to what amount.

Now that we understand the clean versus dirty terms better, I have to ask, can we mix a better steel for our boilers?

Assuming clean sheet for now, making an all new boiler from all new steel. Can we by careful metaluurgy mix a steel that is both strong and flexible enough to handle the stresses of a boiler, but is also more corrosion resistent and tolerent of marginal water chemistry?

I know you can get almost anything if you have the money, but just watch dirty jobs and you'll see mixing additives into batches of steel isn't that complex. And I suspect we're not the only ones looking for the right steel for our needs. I know the Navy has a huge variety of steels for all sorts of applications.

Also, making sure we don't cause problems due to galvanic corrosion, would the steel in the tubes and the inner sheets of the firebox be best if we used a certain steel that had better heat transfer, and the outer walls and boiler sheel something that has less?

For that matter, is steel really our best choice? Take a tour of any beer brewery and you see stainless steels are used in all sorts of "hot" though unfired pressure vessels. Is there a grade of stainless that can be used? I know copper has been used, though if I read things right, it might not be allowed in the US?


Greg

Yes, but you have to order a mill run (five to ten tons) of each different thickness and shape of steel in the boiler. By the time you are done, you will have accumulated enough steel to build a dozen boilers.

Brendan Zeigler, well put on the metallurgical matter!

Kelly, in keeping with the idea that german tube material is better suited for greater longevity due to it's make up.... what about the idea of importing german plate steel? Has this been looked into?

Cheers, Jason

Not by me. US made steel is easier to return when the mill test reports don’t match what was delivered.

Let's start with the simplest part of the calculation process, the minimum required thickness of the shell. Assume a boiler shell has three courses in it plus the wrapper: the barrel course nearest the smoke box is 36" outside diameter (OD); the next course tapers from 36" OD to 42" OD; the next course runs from the taper course to the wrapper, is 42" OD and has the steam dome attached to it; finally the wrapper. We must also assume some design conditions. The design pressure is 200psi. Design pressure should not be confused with the maximum allowable working pressure (MAWP). The MAWP may be the same as the design pressure, but some customers want to know to the pound, how high they can operate their boiler. In some cases, customer specifications require certain parts of the vessel be overdesigned preventing those components from being the limiting factor when calculating the MAWP. Nozzles and domes are good examples of components customers may want overdesigned. The design temperature dictates the allowable stress in psi a material can be exposed to and comes from tables in ASME Section II "Materials". Since 200psi saturated steam is 388° F, we'll use 400° F for our design. When we get to the firebox we’ll use a much higher design temperature.

The largest and weakest part of the barrel is the 42" OD course. To calculate the minimum required thickness, ASME Section I, paragraph PG-27.2.2 gives the formula: thickness=Pressure x outside Diameter / (2 x allowable Stress x joint Efficiency) + (2 x y [a temperature coefficient from PG-27.4, Note 6] x Pressure). Add the result to the desired corrosion allowance and that is the minimum required thickness of the plate. For our example we will use SA-516-70 steel plate. The material is plentiful and is used by many manufacturers of fire tube boilers. Applying the formula: (200psi x 42")/(2 x 20,000psi x 1) + (2 x .4 x 200psi) = .209". The efficiency used in this calculation is 1. Longitudinal butt welds in boilers are required to be examined 100% by radiography so the Code allows an efficiency of 1 for these types of joints. Joint efficiencies drop when using rivets for the long joint in a shell course, but that's a discussion for a different thread.

Corrosion allowance is the amount of extra thickness a designer adds to a component to allow for corrosion to take place without reducing the structure's ability to withstand the pressure. Our customer has a good program of boiler washing and water chemistry management so we'll use .125" in our example bringing the minimum thickness for the 42" shell course to .334”. This number does not take into account the material required to adequately reinforce the dome attachment which may require additional shell thickness. The minimum thickness also does not account for loads placed on the course from hanging an air compressor or running board brackets or any number of things that get attached to steam locomotive boilers.

Next post I'll jump into flat plates supported by staybolts, exciting stuff!

And this is my point regarding the boiler being a structural member of the machine. I can see altering design for ergonomics of maintenance such as washout plug location and quantity etc.. Using more modern locomotive boiler construction techniques, from the late 40's/early 50's. And, perhaps adapting national standards such as fillet welded bolts.
Speaking of which, am I correct in my thought that the FRA could not deny the use of fillet welded bolts since there are 6 locomotives in this country currently equipped? That is assuming the proper process is used for installation backed up by documentation for the procedure from the country of origin....say germany or china?
Part 230 does after all state in part 230.29(b)(1) "All defects disclosed by inspection shall be repaired in accordance with accepted industry standards---which may include established railroad practices, or nbic or api established standards" It does not specify the country of which "industry" or "railroad practices" are to be drawn from.
Of course I am not advocating using fillet welded bolts for repairs lest they are used on an ENTIRE sheet. Rather, I am looking at new construction and entire box replacements such as the 844 or 3713. The FRA did, prior to the issuing of the new regs allow the use of mudring pins on a US built engine. The group tried riveting but did not have a large enough gun....after several failed attempts they went to the FRA and got the pins approved. A precedent has been set, no?

Cheers, Jason

Paul,

I am new to railroading, and even newer to boiler design: Can you expand a little bit on the difference between MAWP and design pressure?

Thanks,

--Steve

I suspect this to be true, but perhaps someone with more experience can comment. As of the 2010 Edition of the ASME Code, fillet welded staybolts are not allowed, only threaded and full penetration welded stays are permitted for new construction.


Since you brought up the FRA, it is worth noting that the allowable stress for materials in Section II of the ASME code is based on a design margin of 3.5, not 4 as the FRA requires. In order to design an FRA compliant boiler you can use the same formula as noted in my earlier post, but use the allowable stress from the 1998 Edition of Section II which is based on a design margin of 4.

If you design a boiler shell 30” in outside diameter with a design pressure of 150psi and design temperature of 400° F and a fully radiographed long weld, the Code minimum required thickness is .112” for SA-516-70 plate. If you add 3/16” to the thickness for corrosion allowance you get a required thickness of .300”. Chances are a manufacturer will not want to purchase a mill run of .300” thick plate so the manufacturer uses 3/8” thick plate, giving the shell an extra .075” of thickness. Using the formula for pressure from PG-27.2.2 the 3/8” thick material is good for 250.6psi. There may be many other factors that limit MAWP, but this is a simple example of how the MAWP can be different from the design pressure.

Since you brought up the FRA, it is worth noting that the allowable stress for materials in Section II of the ASME code is based on a design margin of 3.5, not 4 as the FRA requires. In order to design an FRA compliant boiler you can use the same formula as noted in my earlier post, but use the allowable stress from the 1998 Edition of Section II which is based on a design margin of 4.[/quote]

This sounds correct. However, if one uses the compendium calculations (those intended for locomotives) and refers only to the ASME practices for methods and procedures by which to join components together then you will end up with a compliant vessel. I point this out because amongst many in the industry there seems to be confusion regarding FRA versus ASME.
Further, if one goes off original design, save possibly modernizing a firebox from 5/16" sheet to 3/8", you will find that you have a boiler that is in GREAT excess of ASME minimums. That is to say, if you assume original sheet thicknesses and then run the calculations for MAWP based on those numbers(rather than from the other direction of from minimums) then everything will be in excess compliance. Again, save certain belpaire and santa fe 2900 class wrappers, not much in the way of corrosion allowance.
To illustrate, the 401 boiler (as I have been informed) was "designed" in a very simple manner. The original boiler was blue printed in autocad, at the seams the sheets were "dragged" into alignment and the resulting boiler had the engineering work done to determine MAWP. This was done versus designing up from minimums and starting from scratch.

Cheers, Jason

Greetings:
A couple of comments:
First, the FRA doesn't specify how staybolts are to be attached (threaded, fillet weld, etc.), they only care about the load the bolt must carry (and that it doesn't leak).
Second, I believe that there are only four steam locomotives in the US with fillet welded staybolts: the JS and three QJs. The SYs both have full penetration welded staybolts (and other items) because they were built to be equivelent to the ASME standards (Section 1- Power Boilers) at that time.
J.David

Stay bolts and the area of a piece of plate supported by a staybolt has been a topic of sometimes controversial discussion around the industry as of late. My intention here is to cover the basic concepts and address the ASME Code requirements for staybolts attached by a full penetration weld.

A water wall in a simple locomotive boiler is the water space between two plates with the pressure on the water side trying to push the plates apart. The job of the staybolt and its close cousin the brace, is to resist the force of the steam pressure trying to push the two plates away from each other. Three things need to be considered when designing sections of a boiler supported by a staybolt: 1) The material specification, grade and thickness of the plate being supported; 2) How far apart the staybolts are or "pitch" of the staybolts; 3) The load on the staybolt. For the purpose of this example we will consider a firebox side sheet that is flat, 3/8” thick, SA-516-70 steel plate. The design pressure and temperature for the example is 160psi at 700° F.

A note on allowable stress: The allowable stress, as I have mentioned in a previous post, is a fraction of the minimum specified tensile strength of a material. SA-516-70 plate has a minimum tensile strength of 70,000psi, abbreviated as 70ksi. In actuality most SA-516-70 will have a much higher minimum tensile strength, but the allowable stress found in the tables of ASME Section II, Part D are based on the minimum tensile strength of the material so that’s what I’ll use in my example. The stress tables in the 2010 edition of the Code reflect a design margin of 3.5. This means the allowable stress of SA-516-70 plate at room temperature is 20ksi. If we were deigning a boiler for use on an FRA regulated railroad we would need to use a design margin of 4, making the allowable stress at room temperature 17.5ksi. There was a change in design margin from 4 to 3.5 in the 2001 edition of the Code. When designing for a Jurisdiction that requires a design margin of 4, the stress tables from the 1998 edition can be used. Temperature has an effect on the tensile strength of steel. As the temperature rises, the strength of the plate lowers. These changes in the physical properties of the steel at different temperatures are taken into account in the stress tables. The allowable stress of SA-516-70 plate at 700° F from Table 1A of ASME Section II, Part D, 2010 edition is 18.1ksi.

The formula for minimum thickness of plates supported by stays comes from paragraph PG-46 of ASME Section I:

t=p√(P/SC)

Where t = minimum thickness, p = distance between the centers of the staybolts or pitch, P = design pressure, S = the allowable stress of the plate, C = a constant. For 3/8” thick plate C = 2.1. Using the formula:

t=4√(160/((18100)(2.1)))

Or t = .26” since we want to use 3/8” thick plate we will also have .115” of corrosion allowance. If our customer wants a full 1/8” of corrosion allowance then a thicker plate would need to be used or the pitch of the staybolts could be reduced until the minimum thickness plus the required corrosion allowance was less than or equal to 3/8”.

In order to know how large the diameter of a staybolt must be the load on the staybolt must be known. The load on a staybolt is addressed in PFT-26. Essentially, the load is calculated by determining the area of the plate supported by a stay minus the area occupied by the stay. Using a 4” X 4” stay pitch and a 1” diameter staybolt the area supported by the staybolt is:

(4 X 4) - .785 = 15.215 square inches

The load on the stay is therefore:

15.215 X 160 = 2434.4lbs.

The diameter of the staybolt is determined using the rules in PG-49. To get the required cross sectional area of the staybolt divide the load on the staybolt (2434.4lbs) by the allowable stress from Section II, Part D for the staybolt material then multiply the results by 1.1. The allowable stress for an SA-36 round bar at 400° F is 16.6ksi. Therefore:

(2434.4/16600) X 1.1 = .161 square inches

Welded in staybolts do not require telltale holes per the Code. I will use a 3/16” diameter telltale hole in our example. The cross sectional area of a 1” diameter staybolt is .785 square inches. The area of a 3/16” tell tale hole is .028 square inches making the net cross sectional area of a 1” diameter staybolt with a 3/16” diameter telltale hole .757 square inches, significantly larger than the required area of .161 square inches.

The FRA requires a maximum staybolt stress of 7500psi. If we take the load on the staybolt and divide it by the net cross sectional area of the staybolt we get the stress on the staybolt:

2434.4 / .757 = 3216psi

Next: Stayed curved surfaces!

Paul,

I just found this discussion and have been glad to see some examples of design parameters with the explanations.

However, I am hoping you and the other posters will continue the thread

Please give us the lesson on STAYED CURVED SURFACES.

i and others i suspect also want to see how you do the dome opening and angular stays/braces.

SteamerDave