Enclosed vessel, chiller sizing problem

[ginsoakedboy]

I am performing a preliminary thermal design for a test vessel.

The test vessel is required to have (sea)water maintained at 35°F. The test pieces will be suspended in this vessel to study the effect of long term exposure on coatings. The test pieces will be maintained at an elevated temperature ~300°F. Maintaining the seawater around this test piece at 35°F will give rise to a heat load of approximately 50,000 W.

There will be chiller-coiled tubing system with 30% glycol solution as the working fluid to take the heat away from the test vessel.

I have written down the equations for natural convection heat transfer from the test piece to the seawater. And, also calculated the OHTC (U) for the chiller tubing immersed in the seawater.

Assumed inlet temperature of chilled fluid to be 25°F and the exit temperature to be 50°F.

Also, the heat balance for the chiller fluid is set up.

My problem is that I cannot determine a good way to tie up the heat load to the heat being transferred away by the chilled fluid. The interaction between the ambient seawater and the chiller tubing is very complex involving natural convection, tubing geometry, and variable temperature differences.

This case is not a textbook heat exchanger so I cannot plug the traditional heat exchanger equations. Additionally, the seawater will be pressurized to around 8000 psi, so the chiller tubing cannot exceed 1" OD in size because of collapse loading concerns. This will constrain the maximum flow rate through the chiller tubing.

Can anyone guide me with the way to proceed? I need to determine amount of heat chiller-tubing system will take away and if I can even maintain the temperature of seawater at 35°F.

 

[BrianCNTFM]

Skywalker, some questions/statements to get things rolling and promote discussion.
What is ambient air temperature of facility?
Are there any physicals limits on test facility.
What is allowable range on seawater temperatures?
If we assume bulk seawater of 35 then need to reduce tmperature of outlet coolant to something more meaningful to establish LMTD. That will allow establishment of coolant flowrate. You can make some approximations on heat transfer though conductivity through pipe will likely dominate. It maybe that you will need relatively large mass of seawater and credit can be taken for natural convection from test vessel containing it. If you have very small mass of SW then it will approach source temp quickly with conduction and be outside spec. This will be an iterative process. I am unsure what difference thre will be with pressure stated but already have an idea that we are talking large volumes to accomodate bulk seawater temps based on previous experience.


300 F-
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-
-
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T sw inner-
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- T sw bulk 35F
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- T sw outer

Regards Brian C

 

[ginsoakedboy]

- Ambient air temperature of the facility could go as high as 100F on a hot sunny day. But, the test vessel will be comprehensively insulated to minimize the heat gain from the pressure vessel walls.

- I don't quite understand what you mean by physical limits, but there are none that I am aware of.

- The pressurized seawater temperatures are to be maintained within the 35°F to 40°F.

- I have modified my calcs to 22°F as coolant entry temp and 34°F as the coolant exit temp this leads to LMTD = 4.7°F.

- Speaking in the heat exchanger jargon, "shell side" convective resistance dominates so that changing the flow rates has a very small effect on the heat taken out by the coolant.

- Bringing the SW up to test temperature is relatively quick.

- But, the steady state test is carried out for extended periods of time where there will be a constant heat addition of 50,000 W from the coated pipe (uninsulated). The coolant needs to remove this heat while flowing inside a 0.7" bore tubing.

 

[chicopee]

The exit temperature of chiller fluid of 50 dF does not make sense, was it then modified down to 34 dF?

 

[ginsoakedboy]

chicopee,
This was indeed modified down to 34°F.