DX OAHU design recommendations

[MAragorn]

I am becoming more and more grateful for the years I was able to do design with big built up chilled water cooling and steam heat air systems. Doing good design using "less sophisticated" equipment is HARD!

I have been involved recently in one way or another with about 25 - 100% OA, DX, electric heat Roof top units. I am becoming painfully aware of my own lack of experience with this type of equipment.

This is an example:

Gulf Coast, 93F db / 80F wb entering summer conditions.
20F db entering winter conditions
75F db / 50% RH summer, no RH control winter

Presume no energy recovery for this discussion.

System flow is 4000 cfm, 100% exhausted, thus 100% OA rooftop unit.

After learning a bit over the last few months, and also doing a search on here, this is what I think I might recommend:

Multiple compressors (how many would be optimum?)
Separate refrigerant circuits
Entertwined refrigerant circuits on the coils (is this better than 2 separate full face coils?)
Hot gas bypass
Multiple stage heating (in this case I am using 7 stages)

I would appreciate comments on what I am thinking wrong and what I should do additionally.

Thanks

 

[Yorkman]

I going with the thought that this will be a package unit and you are selecting options. I'm not sure the hot gas is necessary if your not planning on running the supply fan with VFD. Your talking probably a 10 ton unit+/- so two compressors max.
On the topic of evaporators you may be limited on the options offered by the manufacterer. Typically you will see a coil configuration called; horizontal split, face split, os sometimes these are called stacked coils. These do have the disadvantage of re-evaporating condensate on the inactive coil section, and bypassing this to the air conditioned space. Although, while the bottom coil is active some form of humidity control can be achived by the bypassed air.
The next type of evap. is the row split, vertical split, or sometimes called a series flow. These are seldom offered as standard equipment. It eliminates air bypassing and minimizes re-evaporation of condensate. But close attention is required for accurate sizing and selection.
Finally the interlaced coil or intertwined coils. These use the whole face area and depth of the coil. Because airflow is constant across this coil, modulating refrigerant flow can result in increased coil surface temperatures, which may require additional compressor protection. Example suction pressure regulators or compressor multiplexing.
I'm sure other members of this forum may have additional ideas and information, and I open the floor to them as it were.
You might also check ot the FAQ on Psychrometrics it may help make the unit sizing and selection a little easier. Goodluck.

I'm not a real engineer, but I play one on T.V.
A.J. Gest, York Int./JCI

 

[lilliput1]

You will have to recirculate some of the supply air direct to the return. This will make the DX coil entering air the mix of the 100% OA and the recirculated air. Your AHU will have to handle the sum of the 100% OA supplied to the bldg. + the recirculated air CFM.

 

[MAragorn]

Well I can't do that.

It has to be once thru air.

 

[imok2]

Hot gas has some major drawbacks that if your not familier with then you should read this:

http://www.trane.com/commercial/library/vol32_2/ADM_APN007_EN_0503.pdf

I think 2-6 ton variable speed scrool compressors maybe a better choice. At the out side conditions you gave for summer the sensible heat factor would be about 55%, a big latent load.

 

[quark]

I can't presume a 100% OA unit, with the wastage we see, unless there is a strong reason. In any case, I do consider energy recovery.

The OA condition of 93F DB and 80F WB has an enthalpy of 43.6 btu/lb and room condition of 75F DB and 50% RH (62.5 WB) has an enthalpy of 28.2 btu/lb.

The load excluding the room heat gain (in summer) is 4.5x4000x(43.6-28.2) = 277200 btu/hr

Considering 80% recirculation (in normal case) or 80% heat recovery, you will be loosing about 277200*0.8 = 221760 btu/hr or 18.48 TR.

Note that when you recirculate the 80% air or go for a heat recovery, you need not cool 80% of air from ambient conditions to the room condition.

In winter condition, you have to consider a reasonable RH in the area, say 50% again.

The enthalpy of OA at 20F and 50% (I took a simple case as you didn't mention winter outdoor RH) is about 5 btu/lb and that of room condition is 28.2 btu/lb. The specific humidity difference is about 0.0083lb/lb (i.e 0.0093-0.001)

So, the heat load excluding room loss is 4.5x4000x(28.2-5) = 417600 btu/hr

plus heat required for evaporation of water i.e
4000cfmx0.075lb/cu.ftx0.0083lb/lbx1000btu/lbx60min/hr = 149400 btu/hr

Which comes out to be 567000 btu/hr or 166 kW.
I would prefer 8 x 20kW heaters. I don't understand what you mean by 7 stage heating.

Room heat gain and loss should be considered in all these calculations.

what is the reason you want to go for hot gas bypass?

 

[MAragorn]

At THIS point, this is an academic exercise. I asked to NOT consider Heat recovery so as to isolate the issues at hand.

I have done heat recovery, and often, and in another topic I intend to bring it up.

In this case, as I expected, it masks the original discussion.

 

[lilliput1]

You will have to recirculate air around the unit. DX units operate at about 300 to 400 CFM per ton. With 100% OA you will find out that your 100% OA CFM is less than the minimum CFM for the correct AHU. Because you will be recirculating the dehumidifies air direct back to the ducted return, not going to the space, you will not be violating the once through requirement. You will have icing at the coil if your CFM is less than the rated minimum.

 

[AbbyNormal]

First off it sounds like you are looking at the design dry bulb with coincidental wet bulb for your design condition, where you could have times of lower dry bulb but higher dewpoint. Something to handle air with a dewpoint of 75F could become a humidity pump when the ambient dewpoint rises to say 78F.

You need to know the sensible and latent load of the space as you are going to be displacing humidity rather than pulling it out of room air.

Perhaps a supply with a dewpoint of 54.5F would suffice, and pulling down 4000 CFM of your stated ambient condition down to a dewpoint of 54.5F would need 30 tons.

Perhaps you have a somewhat higher dewpoint and will need a little bit lower entering dewpoint that could push the load to say 32 tons or more.


You would need a deep row coil perhaps 8 rows deep with a low face velocity.

You could go with 4 @ 7.5 ton stages then a final two ton stage with the two ton circuit having its own evaporator and condenser coil in the airstream.

The 2 ton evaporator and condenser would pull the dewpoint down and provide some reheat in case continual 4000 CFM make up at the low temperature could over cool the space.

Most likely be some total static involved here with the deep row coil for the four larger stages, plus the pressure drop of the two ton evaporator and then the condenser so most likely at least two degrees of reheat from the fan. With the fan reheat and the small two ton 2 dehu circuit on could be a supply dry bulb in mid 60F range.

I have about 3 similar systems in the Caribbean, and the dewpoints hover from 79 to 81 from June through to the end of November. I have hot gas bypass on the first stage of the larger compressors on all of these systems.

I looked at what Addison could do and then what Engineered Air could do, I ended up going with Engineered Air.

Systems have been in operation since 2002, and 2003. So far no complaints, they even survived a CAT 5 storm.