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BronYrAur (Mechanical) (OP)
24 May 06 17:45
It has been a little while since I studied my thermodymanics, so I need a little help here.  Reducing steam pressure through a PRV is an isenthalpic process (no enthalpy change), correct?  So If I have 100 psig (338 deg F) saturated steam and reduce it to 15 psi, what changes?  I assume that I will have superheated steam at 15 psi, right?  How can I determine what the temperature is?

Where I am going with this is that I currently have a steam to water heat exchanger being served by 15 psig steam.  However, this 15 psig steam is coming immediately out of a PRV that reduces it from 100 psig.  I want explore the possibility of just running a 15 psig main and not having to reduce the pressure, but a saturated 15 psig main is only about 250 deg F.  I assume that I have a higher temperature steam now, but how high?  Is that even inportant in the grand scheme of heat transfer, since I know that the latent heat is the driving force - at least I think it is.

Any thoughts?
Helpful Member!  UtilityLouie (Mechanical)
24 May 06 18:15
Typically you can find a steam property calculator that you can find the enthalpy of the steam into the PRV - then you enter your new pressure with the enthalpy of the entering steam.... and voila... you should have your steam temperature on the downstream side of the PRV.

A good steam state calculator is here....

and it's free...

Hope that helps....
katmar (Chemical)
25 May 06 4:29
You are correct about the temperatures. At 100 psig saturated the enthalpy of the steam is 1189.6 btu/lb. The enthalpy will remain the same as it expands to 15 psig, giving a theoretical temperature of 300 deg F, compared with saturated 15 psig steam temperature of 250 deg F.

But this increase in temperature may actually reduce your heat transfer, rather than increase it. A portion of your exchanger area will be used to transfer the sensible heat from this superheated steam, bringing its temperature down from 300 F to 250 F without any condensation. The heat transfer coefficient for sensible heat transfer is usually much less than it is for latent heat transfer (condensation). You may well find a 5 to 1 ratio of heat transfer coefficients.

So, depending on your temperature differences, you could end up with a situation where a significant portion of your area is consumed by sensible heat transfer, and you are actually getting only a small amount of heat transferred in that area.

Katmar Software
Engineering & Risk Analysis Software

Helpful Member!  Latexman (Chemical)
25 May 06 8:49

If BronYrAur's water at the steam inlet end of the exchanger is less than (or ~ equal to) the saturation temperature of 15 psig steam, condensing will start immediately.  There will not be a dry desuperheating section of the exchanger.  Instead, thr superheat will transfer into the condensate film itself.

Good luck,

Helpful Member!  TBP (Mechanical)
25 May 06 8:57
In order to calculate superheat, you'll need the dryness fraction of the steam upstream of the PRV. Very few plants have a means of determining this value. In any event, if you're dealing with "normal" (general industry) conditions - saturated steam under 300#, superheat after PRV stations is much more of a theoretical issue than a practical one. Typically, the steam isn't 100% dry to begin with, and the bit of superheat you get across a PRV provides nice dry steam downstream. Once, we actually installed a thermometer  immediately after a PRV station - 125# down to 10#. At high loads, there was some superheat, but at low loads, essentially nothing. A few feet down the line, any superheat was gone entirely.

If you're starting off with any amount of superheat ahead of the PRV station, that CAN be an issue. However, I've never seen an operational problem due to superheat in installations with pressures like you're dealing with. Botched condensate piping on the other hand ...

katmar (Chemical)
25 May 06 8:58

Whether you regard the heat being transferred from the superheated steam as being transferred to the condensate film or to the tube you are still looking at sensible heat transfer, which is slower than latent heat transfer.  Steam at 300 deg C will not condense at 15 psig until its temperature has been reduced to 250 deg C, and this is the slow step.

Katmar Software
Engineering & Risk Analysis Software

Latexman (Chemical)
25 May 06 9:09

The local heat transfer coefficient for desuperheating steam with a falling film of condensate is quite large.  There is essentially no resistance.  He can include the duty of desuperheating in the total heat duty, use the temperature difference between sat'd 15 psig steam and his water, and he'll be just fine.  I've sized dozens of condensers this way with satisfactory results.

Good luck,

katmar (Chemical)
25 May 06 9:26

I disagree that you can condense superheated steam with condensate that is at the saturation temperature, but it is actually irrelevant. What I was trying to get at was that BronYrAur would not improve his situation by using superheated steam.

And as TBP has pointed out, the practicality of the matter is that there would be essentially no difference.

Katmar Software
Engineering & Risk Analysis Software

Latexman (Chemical)
25 May 06 14:30

My point was that your image of condensing superheated steam is flawed.  As long as the water at the steam inlet end of the exchanger is less than the saturation temperature of 15 psig steam, condensing will start immediately and there will not be a dry desuperheating section as you alluded.  For somewhat similar reasons, that’s why a glass of iced tea sweats in < 100% relative humidity.

So even if BronYrAur has the superheated steam predicted by the textbook calculation, he will not have any heat transfer problems if he switches to 15 psi saturated steam as he is looking to do.  He’ll have the same *real* temperature difference he has now, but with less heat duty, i.e. no superheat to remove.

Good luck,

Helpful Member!  rmw (Mechanical)
25 May 06 21:10

If you are right, some major big name heat exchanger manufacturers are wasting lots of time and effort making their heat exchangers with DSH sections, a very expensive part of the Hx.  Could it be that they are all wrong and you are right.  I think they will be quite shocked.

I firmly disagree with you as well.  Katmar is right.  The heating steam does not begin to condense and give up its latent heat until it reaches saturation temperature via a sensible HT process regardless if this heat transfer is against a dry wall or a wet wall.

Some of the wet wall you envision, if it exists at all, will be revaporized in the process and HT surface will be required to recondense this vapor as well, robbing the HT of surface area.

And, you are incorrect on another point.  The one benefit of SH steam, if there is any, is that the delta T of the superheated portion of the steam is at a higher value than the saturated steam so at least a portion of the SH heat transfer is at a higher LMTD for what that little bit of driving force is worth.

An iced tea glass sweating on a less than 100% humidity day is not related to superheat.  It is related to partial pressure relationships of the vapor in the air surrounding the tea glass.

Latexman (Chemical)
26 May 06 0:28

Under conditions similar to those in this post and as I prefaced each time before, provided the condensing surface is cooler than the saturation temperature of the steam, I know I'm right.  I learned it from the guys in the world's largest chemical company that *tell*, not ask, the heat exchanger manufacturers what to build.

What's in it for these companies to make smaller heat exchangers for companies that have no heat exchanger expertise?  Now do you get it?

I repeat, provided the condensing surface is cooler than the saturation temperature of the steam, condensing begins immediately.  

What is the velocity of the superheated steam flowing at the boundary of the heat transfer area?  Zero, right?  Well, if it's velocity is zero, it's going to sit there, cool, and condense!  When the first bit condenses the volume of the steam collapses to the volume of condensate which decreases pressure and induces more steam into the heat transfer surface.  And on and on and on!

The superheat is removed by a mechanism of condensing and re-evaporation at the condensate/vapor interface.  For superheated steam (within limits at constant heat level), condensate loading is lower, the film thickness is less, and the condensing coefficient is slightly higher.  Desuperheating occurs at the expense of the temperature difference between the superheat temperature and the saturation temperature.

Good luck,

katmar (Chemical)
26 May 06 6:05

I like to believe that I am open-minded enough to know that learning something to be a truth while at university does not necessarily make it the truth. Some of the most fundamental truths I learned as a boy, I later found out were just mythology. So I have been doing some serious thinking about what you have experienced with superheated steam condensing. But we should also bear in mind that money and prestige also do not turn myths into thruths.

My own hands-on experience with direct contact (rain tray) condensers has proved to me that you are correct when you say that there is essentially no resistance to heat transfer when condensing superheated steam into cold condensate - even in the presence of non-condensibles. But these direct contact condensers are always fed with cold water that only achieves a 3 or 4 degree C approach to the saturation temperature.  

It is impossible to condense steam (superheated or not) into condensate at the saturation temperature. This is one truth that I (and hopefully you) do accept. So it means that the condensate film in your model has to be significantly sub-cooled.  In your model, if you use a higher HT coefficient for transfering the heat from the superheated steam then you must off-set this with a lower temperature difference between the condensate and the cooling water on the other side of the tube wall.  In the classical model of a tubular condenser the condensate is assumed to be at the saturation temperature (unless the condenser is specifically designed with a sub-cooling zone).

This balance between higher HTC and lower temperature difference in your model probably comes back to giving the same net effect as the classical model (which would have low HTC and higher temperature difference).  Or it could be that your experience with superheated steam is explained by TBP's very pertinent comment that in real life superheated steam from let down valves is not a problem because of the quality of the steam before being let down.

Whichever model we use, at least we agree that BronYrAur will not be any worse off if he feeds his exchanger with saturated 15 psig steam.

Katmar Software
Engineering & Risk Analysis Software

Latexman (Chemical)
26 May 06 7:32

Is it impossible to condense steam (superheated or not) into condensate *at* the saturation temperature?  On a macroscopic scale, yes, but at the vapor/liquid interface there is a dynamic equilibrium situation going on.  Water molecules are continuously going back and forth into the liquid phase (condensing) and vapor phase (evaporation) depending on what direction they are headed (chaos), what energy state they are at due to their past history of collisions and thermal treatment (thermodynamics), local pressure gradients (fluid mechanics), etc.  This is what “vapor pressure” is.

My experience is mainly with vertical, vapor-in-tube, downflow condensers where condensate subcooling is very effective due to falling film heat transfer.  The superheat removal mechanism of condensing and re-evaporation at the condensate/vapor interface maintains the condensing-surface temperature at essentially the same level as that obtained with saturated steam.  Of course, the temperatures decrease through the falling film to the tube wall.  My idea of a little subcooling, 5 to 20o C, may be significant subcooling to you.  We usually like some subcooling to minimize vent losses, especially when the condensate is an organic chemical.

We have operations where the quality of steam is not important, and TBP's description is alive and well.  We also have operations where the quality of steam is critical and we maintain significant superheat.

Yes, BronYrAur should be fine if he feeds his exchanger with saturated 15 psig steam.

Good luck,

BronYrAur (Mechanical) (OP)
26 May 06 8:32
Thanks everyone for your help.  I see my post led to some lively discussion.  I'll verify it with the heat exchanger manufacturer, but your answers have convinced me that switching to saturated steam will most likely not reduce my overall heat transfer capability.  I appreciate the time everyone took to reply.  Thanks again!
TBP (Mechanical)
26 May 06 9:19
The upside of using lower pressure steam in a HX, is that it has (marinally) more latent heat. The downside is that with LP steam is that everything - control valves, traps, piping, the HX itself - needs to be larger, in order to handle the required steam flow.
rmw (Mechanical)
26 May 06 9:33
I'm sorry, I just want to be on record as saying "I don't buy it".  But I'm not going to expend much more effort on this.

Latexman, you defeat your own arguement and make ours with your statement "It is going to sit there, cool, and condense...."  That makes our case.  Saturated steam doesn't 'sit there' it condenses when it gets there.  The sensible heat transfer mechanism that Katmar discussed is what is going on while the steam is "sitting there."  Meanwhile a much lower heat transfer rate is occuring while it is 'sitting there and cooling.'  Meanwhile, the overall duty of the Hx is being robbed by the amount of area where steam is just "sitting there."  Our case is made.  Thanks  

I do have one parting comment.  We do seem to agree on one thing.  And that is your comment that the condensing surface has to be cooler than the saturation temperature for condensing to occur.  Boy, we are making real progress.

TBP (Mechanical)
26 May 06 9:57
Is the argument in motion based on the differences between applications like surface condensers under steam turbines vs process HXs?

At the end of the day, in the real world, if you feed superheated steam to a process heat exchanger, it'll behave as if it's airbound until you get past the sensible heat that is superheat.
Latexman (Chemical)
26 May 06 14:57

I'm referring to my experience with vertical, vapor-in-tube, downflow condensers.

"If you feed superheated steam to a process heat exchanger, it'll behave as if it's airbound until you get past the sensible heat that is superheat" is exactly the argument.  RMW agrees with you.  My view is "provided the condensing surface is cooler than the saturation temperature of the steam, condensing begins immediately."

Gentlemen, I quote from Perry's 7th Edition on page 11-11, "If the vapor is superheated at the inlet, the vapor may first be desuperheated by sensible heat transfer from the vapor. This occurs if the surface temperature is above the saturation temperature, and a single-phase heat-transfer correlation is used. If the surface is below the saturation temperature, condensation will occur directly from the superheated vapor, and the effective coefficient is determined from the appropriate condensation correlation, using the saturation temperature in the LMTD."

It goes on to describe how to determine whether or not condensation will occur directly from the superheated vapor by calculating the surface temperature by assuming single-phase heat transfer.  Equation 11-26 is given as:

Tsurface = Tvapor ? U/h x (Tvapor ? Tcoolant)

h is the sensible heat-transfer coefficient for the vapor
U is calculated by using h
both are on the same area basis

If Tsurface > Tsaturation, no condensation occurs at that point and the heat flux is actually higher than if Tsurface ? Tsaturation and condensation did occur. It is generally conservative to design a pure-component desuperheatercondenser as if the entire heat load were transferred by condensation, using the saturation temperature in the LMTD.

This is exactly what I was trying to say.  I thought I was being clear, but . . . .

Good luck,

Latexman (Chemical)
26 May 06 15:08
Tsurface = Tvapor - U/h x (Tvapor - Tcoolant)

Good luck,

katmar (Chemical)
26 May 06 16:12


It is generally conservative to design a pure-component desuperheater/condenser as if the entire heat load were transferred by condensation, using the saturation temperature in the LMTD.

So the HX manufacturers who design for a dry desuperheating zone aren't such bad guys after all. They are actually trying to save us some money! winky smile

This has been (for me at least) an instructive discussion, plus the reference to Perry 7th Ed has allowed me to fix the typo in my old 5th edition. Thanks to Latexman for perservering with what he believed to be right in the face of all our arguments. One more "truth" becomes a myth.


Katmar Software
Engineering & Risk Analysis Software

rmw (Mechanical)
26 May 06 20:54
It will come as a shock to the manufacturers all over the world of high pressure feed water heaters, who design their DSH sections very carefully in order to balance high velocities against tube vibration and dry wall margins so that the steam leaves the DSH section with a modicum of SH so as not to cut the tubes in the high velocity DSH section.  The velocity is as high as it can be without causing vibrations so as to maximize the HTC in that zone so it can be DSH'd to get it to the condensing zone where the real HT is.  The tubing is below saturation temperature in this zone until the very end.  These Hx's typically have a 0ttd or a -1 ttd.

See or a good cutaway view here;proid=35

Now, I don't know where Perry got his information, nor where he is coming from, but on this one, he has quite a set of detractors.  

Harvey, maybe truth is just hard to find.  I think the myths are flowing freely.

katmar (Chemical)
27 May 06 2:45

The Perry formula surprised me too. However, what it means is not that there is always direct condensation but that direct condensation is a possibility under the right circumstances. I would not have expected the heat flux to be higher with superheated gas than with condensation unless there were extreme conditions. But I have not tried to calculate any HTC's to test this.

With the conditions described in you last post (high velocity and temperature) the tube surface temperature will approach the vapor temperature and there will NOT be direct condensation. It is therefore sensible (sorry for the pun) to design for dry gas at high velocity and take all the precautions you mention.

It is one thing to get sufficient area into the exchanger by making the conservative assumptions Latexman recommended, but it is entirely a different thing to design the exchanger for long and safe operation.  I would never design an exchanger of this nature myself, whereas I have designed plenty of simple saturated vapor condensers. I would leave this to the specialists - and preferably one who would give a guarantee.

Katmar Software
Engineering & Risk Analysis Software

Latexman (Chemical)
27 May 06 12:04
rmw and katmar,

What type of heat exchanger are ya'll speaking of for a desuperheatercondenser?

Good luck,

rmw (Mechanical)
28 May 06 0:12

The Perry formula given is essentially a statement of the skin temperature for the conditions specified.

First, let's us for the sake of this discussion say that the skin temperature of the HT surface we are taking about is always less than Tsat.  Yes, the skin temp in some of the DSH zones of the Hx's I mentioned are above Tsat for parts of the zone, but not for all of it.  But that is irrelevant.

Let's carefully look at what Perry is saying.  First he addresses sensible heat transfer with the tube surface being above Tsat which we are not dealing with.  Then he states that condensation will occur directly from the superheated vapor when the skin temperature is less than Tsat.  He is right in saying this.  BUT, HE DOES NOT STATE AT WHAT RATE THIS HEAT TRANSFER OCCURS.

I have done the calculations, and in English units, it is at a rate around a U value of 52 at reasonable velocities while the condensing U value for the same Hx after the steam is cooled to Tsat is in the 450 range.

These calculations were done (first manually and then with web tools) for a Roberts type rising film evaporator common to the sugar industry.  The steam velocities were nowhere near the range of the type of Hx's I gave links to above, (Latexman, they are called 3 zone heaters, each having a DSH zone, a condensing zone and a condensate sub-cooling zone) which even with the velocities that they get to (to the point of causing hydrodynamic whip-vibration of the tubing) only attain a "U" value for the DSH zone in the range of 160.

The Roberts evaporator studied obviously couldn't have that much velocity because once the steam reached saturation and became wet, it would cut the copper tubed calandria to ribbons.  That is the same reason the designers of the high pressure heaters linked to above are so concerned about maintaining a 5F dry wall margin over saturation as the steam exits the DSH zone, because they don't want the (steel and stainless steel) tubing cut to ribbons.  I have personally witnessed many heaters where conditions changed and the DSH zone was destroyed by the presence of moisture in the DSH zone.  Note for the record that the tcoolant in the DSH zone is below Tsat for the heater pressure and the tube skin temperature in the region of the steam exit from this zone is below tsat as well.  But I digress.

Now to Perry/Latexman.  They are both absolutely right.  The superheated steam will (eventually) condense, but the 64,000 dollar question is; at what rate?  Latexman hits the nail on the head when he uses the phraseology "sit there, cool and condense".  Saturated steam, on the other hand, as it enters the heater and contacts a cold tube, condenses immediately.  No "sitting there" involved.  

The "sitting there" Latexman describes is the blanketing effect that gives the "air-bound" symptom that TBP mentioned in his 26May post.  Oh yes, it is transferring heat while it is "sitting there" but at a much lower rate of heat transfer, sensible heat transfer rates (single phase as Perry puts it) instead of condensing (he might have said two phase for condensing since a change of state of the steam takes place).  

Since the goal of most Hx designers is to minimize the cost of the Hx, minimizing its surface area is one of the most cost effective ways of doing that.

Since the U values for condensing, in the 400-600 range, are significantly higher than the best DSH zones that money can buy, in the 150-160 range in heaters where velocity is the limiting factor, it wouldn't take a smart designer long to figure out that adding more surface to account for the sensible heat transfer necessary to reduce the steam to saturation temperature isn't the thing to do.

The study where I got the numbers I quoted above was done to justify the addition of desuperheaters to the turbine exhaust steam headers supplying an evaporation station at a sugar mill.  The calculations showed that easily 1/3 or more of the calandria was devoted strictly to desuperheating the incoming exhaust steam.  It was easy to see this in the sight glasses on the body above the top tube sheet.  The area of the top tubesheet nearest the steam inlet was virtually dead, very little of the percolation effect common to this type of rising film evaporator.  A little, but not much.

The backside, however, once the steam was cooled to saturation, was very lively, with sugar juice jetting up higher than you could see through the sight glass.  This was a newer evaporator with several good, clear sight glasses, and easy to make the observation.

After the desuperheaters were added, and the exhaust was brought down to Tsat, two things happened.  The overall evaporation rate for the first stage evaporators (Pre's they are called in our part of the world) increased dramatically, and the visual observation in the sight glass showed that the lively boiling was uniform across the entire surface of the top tubesheet.  Gone was the dead zone that was there during the previous grind.

Latexman, has (or had) in their 'tools' section the capability to model heat exchanger performance, and/or calculate HTC's.  I challenge you to go set up a couple of different scenarios.  One being pure condensation where your inlet conditions are right at Tsat, and another where your inlet conditions are above Tsat keeping steam side pressure and Tcoolant constant.

I think if you do it right, you will see something different from what you are stating in this thread.

I have been there, done that, and I think I know the result you will find.  BronYrAur, if you are still following this thread, I recommend the same for you.  Mr Perry, I recommend the same for you too.

I have enjoyed this discussion.  It has cleared out a lot of cobwebs.

25362 (Chemical)
28 May 06 1:18

For those interested in a different (heat flux) approach there is an article in the Dec. 29, 1980 ChE issue, titled Superheated vapor condensation in heat exchanger design by Foxall and Chappell.
katmar (Chemical)
28 May 06 6:06

Thank you for a very well reasoned post - you have obviously put a lot effort and time into this and I appreciate it. Your experience with the Roberts unit proves that the heat flux was lower in the dry zone. This was what I expected when I said earlier that the heat flux in a dry zone would only exceed that in the condensing zone under extreme conditions.

I dug out a copy of the Foxall and Chappell article referenced by 25362. Their examples do show higher fluxes in the dry zones, but this is only because of the temperatures they used. The heat transfer coefficients they calculate are much lower for the dry zone than the condensing zone, but because they selected coolant temperatures so close to the condensing temperature and because they had large amounts of superheat the temperature effects outweighed the HTC effect. In their steam condenser example they condensed superheated steam at atmospheric pressure with coolant at 200 deg F. With a condensing temperature of 212 F there is only 12 F of driving force in the wet zone, but in the dry zone the vapor temperature was 500 F giving an overall temp difference of 300 F.

Probably the most important lesson to take out of all of this is to remember how dangerous it is to make assumptions in engineering.  A properly formulated and calculated solution is the only way to be sure.


Katmar Software
Engineering & Risk Analysis Software

Latexman (Chemical)
28 May 06 12:29

The Babcock url wasn't working when I asked what type those units were.  It is working now, and I see they are horizontal, vapor-in-shell units.  The Robert’s rising film evaporator sounds like a vertical, downflow, vapor-in-shell unit to me.  In these cases, the vapor is usually steam.

I believe I have been clear in saying my experience and comments are on vertical, downflow, vapor-in-tube condensers.  In most of my experience, the vapor is usually an organic, but has been steam from time to time.

Condensing outside the tubes is very different than condensing inside the tubes.  In fact, condensing outside horizontal tubes is different from condensing outside vertical tubes.

Let’s list a few differences between the Robert’s vertical, downflow, steam-in-shell unit to a vertical, downflow, steam-in-tube condenser I am familiar with:
  • Inside the tubes, the condensate flows in layer form down the length of the tube.  The condensate film is “thinned out” by the flowing vapor which increases heat transfer.
  • Outside the tubes, the condensate is impeded by tube supports.  Most of the condensate falls off the edge of each baffle providing poor contact with the cooling surface where the larger h exists.
  • Outside the tubes, especially if crossflow baffles are used and in the high velocity section where the vapors enter, condensate is stripped from the tubes by the vapors flowing through the unit.  Again, this provides poor contact between the condensate and the cooling surface where the larger h exists.
  • The mechanical design issues are different too.  With outside the tube condensing, the shell could be several hundreds of degrees higher than the tubes where condensate exists and flow induced vibration is a mighty issue.  Not so for inside the tube condensing.
It appears to me we are not discussing an “apples-to-apples” comparison.  I know I’m right with my discussion on vertical, downflow, vapor-in-tube condensers, so I would recommend BronYrAur understand these differences when reading them and apply the correct one to his application

By the way, when I said "sitting there" my intent was not in describing a blanketing effect that gives the "air-bound" symptom that TBP mentioned in his 26May post.  I was referring to the “boundary layer” at the tube wall that exists during turbulent flow.  In this “laminar sublayer”, the velocity at the wall is zero.  This is the classical textbook model that is taught in every fluid flow course.  So, since the velocity at the wall is zero and if the tube wall is < Tsat, the steam will condense, even if the steam in the “turbulent core” is superheated.  This includes at the steam inlet.  In a vertical, downflow, steam-in-tube condenser once the film is formed, it stays formed and is not impeded by tube supports or baffles and is not stripped off the tube wall by vapor flow.  We may be describing the same thing, just using different words.

In the vertical, downflow, vapor-in-tube condensers I’ve designed IF condensing started at the beginning of the tubes and the tubeside fluid was superheated steam, the h would be in the 400-600 range you mentioned for condensing, or maybe even higher, not the 52 you said you calculated.

I also studied the “Process Tools” at and they do not currently have a tool that handles a condensing heat transfer coefficient.  They say it is “coming soon”.  When this tool arrives if it is as simplistic as their existing tools, I doubt it will have a flash calculation to determine when and where condensation occurs anyway, but I’m speculating.  Anyway, since their current tools do not handle condensing heat transfer, I must decline your challenge.

I do have a question I’d love the answer to though.  If you have “been there”, , and “done that”, what tool did you use exactly?  

Good luck,

rmw (Mechanical)
29 May 06 22:58

I'll give a brief answer because I am packing for a trip to South America that I have to leave early tomorrow AM on.

I kept the print outs of the results of the processassociates runs in a file, but due to a recent office move, they are still packed, if not in storage, so I can't put my hands right on them.  I do remember the condensing result, because it verified the HTC for condensing that I had seen used in vendor literature and on various websites but had not done for myself.

You have me curious, so I will check it when I can, but it won't be right away.

Both Yuba and TEI (division of BPI) have vertical heaters, but you are correct in surmising that they are steam outside the tube, and have the presence of the tube supports.  They don't use flow through except in the DSH zones.

I read your thorough and complete post through once, but want to re-read it before making any comment.  It was chock full of deep material that is going to take me a couple trips through at some later time when I can get my mind into it.

And, Harvey, yea, the "per degree F" component of the HTC formula can generate higher heat flux if the delta T is high enough, but in most cases that I deal with, the degree of SH is low enough that it is a penalty, not a benefit.  The goal is to get the HT by getting the latent heat, so the SH present has to be dealt with separately.

Good night all, more later.


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