CFM vs Capacity 1

[MechGuy22]

Realistic hypothetical: Lets say you are designing an office with DX split systems and the load calc comes to:

35,700 sensible btu/h
3,500 latent btu/h
39,200 Total.

This could probably be covered nicely by a nominal 3.5 ton system pretty nicely. Now lets say the cooling cfm required is 1595 based on a 20 degree TD. Would you size your unit based on the capacities, or select a unit that can be the CFM.

I struggle with this a lot with offices because of the high sensible heat ratios where I will have a 0.95 + sensible heat ratio. Any opinions?

 

[willard3]

Whether you size by tonnage or CFM will depend upon the air velocity in the space you're trying to cool.

 

[MechGuy22]

Im not sure I follow that one. You mean the air velocity in the space, not the duct work correct?

 

[urgross]

I usually select the compressor based on thermal load, and select the fan motor based on fan curve. Fan curve selection is based on system static pressure to be overcome to meet flow. The static would be based on duct velocity. You have a 3 TR load based on sensible load (1.08*20*1595=34452)), what is the duct loss at the duct velocity selected (are you using some guidance such as an 0.08ft/100 ft from the SMACNA chart?)

 

[MechGuy22]

One question, what is a "3 TR"? 3 ton?

I understand what you are saying, but if I selected a compressor based on the load, and a fan based on the CFM, the unit would have to produce well over 400 cfm/ton which isn't possible with DX equipment usually.

In my case, lets say I selected a 3.5 ton compressor to meet the capacity. Then I selected a fan that could provide the 1595 CFM. That would equate to 456 CFM/ton. The unit probably would not be able to do any de-humidification at all. It doesen't really have any latent involved here but even so.

In my opinion, its better just to go to a 4-ton system.

 

[urgross]

First, you gotta do what you gotta do, you own the design, not the talk room.

Not sure how dehumidication is an issue if you have no latent load.

The CFM/TON (TR-tons refrigeration) depended greatly upon the outside air portion and internal load. As you seem to have neglible internal and outside air latent loading-otherwise you woud not have 95% sensible load.

At roughly tons and 1600 CFM, won't you have 400 CFM/TN? No outside air, minimum internal load, what difference does a rule-of-thumb make? For the climate where I'm at, high humidity, applications for 100% OA often come out to 190 CFM/TN.

 

[MechGuy22]

Also, I typically use 0.1/100' for a smaller system, then 0.08/100' for a larger system. If I verify the ESP and it is too low or high, then I go back and do a little re-sizing.

I know some people like to try and make the static exactly equal to the CFM they are trying to achieve based on the fan curve but this usually isn't reasonable in my experience. I always figured the higher the static, the more money it costs to run.

 

[MechGuy22]

I am located in southern Florida so humidity is a huge issue. I usually won't use a outside air unit unless the designs requires it such as in a assembly occupancy or a dense hair salon.

I would say using the rule of thumb 400 CFM/ ton is just that, a rule of thumb but I wouldn't want to exceed that by too much.

 

[willard3]

Whether you use 500 cfm/ton or 250 cfm/ton depends upon the sensible heat ratio of the load.

 

[MTES93]

Sounds like your climate is similar to mine. We don't give the family cat a nice static shock by shuffling across the carpet in our socks and touching it. We could light the little fur ball on fire..

For packaged equipment (roof tops) we will size on the sensible capacity based on our loads. Look closely at the sensible capacity of equipment you are looking at. Coil bypass factors and related come into play in addition to the zone SHR. A 4 ton roof top unit will most likely not give you 4-tons of sensible cooling as they are designed approximately around the 400 CFM/ton rule (meaning coil bypass factors and some assumed zone SHR?? Not sure what they do... I used to). Thus/ with packaged equipment you almost always end up w/ an over-sized jobber.

On split systems I have successfully designed larger indoor coils (for the sensible load) and smaller ACCU's (condensing units). However check closely w/ the mfgr when doing this.

So on your hypothetical system I would be looking at a 4-ton coil (1,600 CFM) and a 3.5 ton ACCU.

 

[urgross]

If you are starting out with a 0.95 SHR, then first question may be, are you using the dry bulb with mean coincident wet bulb, or WB with MCDB, or dew point with MCDB, for your load calc's. I would've guessed that you are in a dry climate.

 

[MechGuy22]

Im not exactly sure how my program is calculating it. I wonder where I can go to find out which method it uses... This is an office with no internal latent gains except people. Also, the OSA requirement is very small.

There is virtually no infiltration in this building (2nd floor of a almost new 3 story building)and the outside air requirement is almost negligible so I don't really think the climate affects me too much with respect to dry vs humid.

 

[willard3]

If you have people in internal spaces, you are required to provide outside air which, in Fl, is pretty moist.

 

[MechGuy22]

Yea, I have per ashrae 62.1 - 2004 but for this unit, it is a very small amount. There are only about 10 people that will occupy the space at a maximum. Ashrae 62.1-2004 says 5 CFM/ person + .06 the floor area. This equates to about 80 CFM of OSA. In this space I have large western windows which is where all the sensible is coming from.

 

[willard3]

If you have large windows, you will also have infiltration adding more moist air.

 

[urgross]

If you have politicians, you will have more moist air as well.

 

[MechGuy22]

Agree to disagree. Infiltration is accounted for in my load calc but it is very little. A commercial building is required to have ventilation running continuously during occupied hours. That means that the building is positively pressurized already. This building is pretty air tight and the windows are heavy, impact windows. Not to mention, by code, they have to be sealed to leak less than 0.3 cfm/ ft^2.

 

[MechEngNCPE]

The sensible load is what gives you your cfm.

35,700 btu/h = CFM*1.085* delta T

delta T = Tmix-Tsupply

You know your delta T because you know your mixed air temperature and your discharge air temperature. Solving it for a twenty degree delta T you get 1645 CFM.

Use your psych chart to determine what supply air discharge temp you need to achieve a certain relative humidity in the space. Remember to use the slope at the 95 sensible ratio line.