I've been trying to find the IWC pressure drop across the evaporator on a Climatemaster 27 3-ton horizontal unit, but I am unable to find anything in the manufacturer's literature. Is there anyone that knows that this value is? I'm trying to design the duct work right now, and I would much rather have the true value to plug into the Excel sheet. Thanks!

I don't know the particulars of the specific unit you cite, and I'm not motivated enough to surf the number of tube rows or fin pitch of that coil, but off the top of my head I would expect a wet coil to impose something in the range of 0.10 - 0.16 inWC. "Wet" refers to cooling mode nearly everywhere other than in the desert southwest - airflow must get past staggered rows of tubes, fins and also the beads of water constantly forming, falling, growing and draining away since indoor air dewpoint is below evaporator temperature.

Don't forget to allow for the filter ( you can cut that pressure drop by specifying a deeper filter cabinet providing more surface area).

Don't forget the heat strips - lower kW aux heat kits impose less pressure drop.

I think a CM 27 is a 2 stage system, and if you size it right it'll be in low stage 80-90% of the time, with much lower pressure drops. ECM blowers provide a bit of extra oomph to push past pressure drops, albeit at a cost in blower Watts and noise. If that condition occurs only a small fraction of total operating hours (high stage) it may be acceptable.

All of the above may be academic since air handler performance data is normally specified in TESP (Total External Static Pressure) IOW, available pressure drop is net of strips and wet coil, so no need worrying about the particulars of your particular coil.

Finally, consider somewhat oversizing ductwork if space permits AND ductwork is mostly or wholly located within conditioned or indirectly conditioned space. Just as ECM blowers are able to overcome somewhat crappy duct work via extra speed and power, they also reward us for good ductwork by slowing down - less noise, fewer Watts. That's another reason a large filter cabinet pays off.

Engineer, thanks for the response--you are a true asset to this community.

You are correct in your assumption that the unit multistage. Since the units only come in nominal size, and my load calculations netted a 2.5-ton unit, I ended up going with the next largest size, the 3-ton. My hope is that it will run in the first stage for nearly all of the heating/cooling demands and only use the second stage on those extremely hot/cold days. When designing for a multistage unit, do you design for the highest velocity (my current approach), the lower velocity, or somewhere in between? My concern with designing for the higher velocities is that there would not be enough throw from the registers to mix the air in the room.

When I was doing the friction rate calculations, I ended up using 0.15 for the coil. I also added for the filter (0.16), supply outlet (0.01), return grill (0.01), and a balancing damper (0.03).

Playing with the Excel spreadsheet, I couldn't help but notice by changing the friction rate by 0.02 I could reduce the size of my main supply and return--is it better to error on the larger side as you mentioned above? In other words, is it more efficient to use a large duct than the next step down? Unfortunately 90% of my ducting is located in an unconditioned attic--which goes against your recommendation of enlarging the duct work if there is ample space (there is plenty of that!). I have posted my Excel sheet here if you would be so kind as to double check my numbers.

Here is a link to the Excel sheet. Thanks for any additional input you can provide.

Manual D Calculations

If your load calc is spot on then you are essentially buying and operating a very expensive (but very efficient) single stage unit, because the second stage will need to operate about 3 hours per day about 3 days per year.

Oddly enough, your situation is quite similar to my own home. It loads out to about 30k, summer and winter, and has a WF Envision 038. We do manual setbacks, so to prevent unnecessary second stage operation during setback recovery I put a lockout switch on the stage 2 (Y2) lead to the unit. I enable Y2 only when we return from a long weekend or vacation away from home in January or July and when I am not willing to wait for low stage operation to pull house up / down to comfortable temps after a multiday shutdown.

Hourmeters in place since startup during spring 2008 now show almost 10,000 hours in Y1 and 48 hours in Y2. That begs the question of whether I'd have been better off with an 026 (next size down). Perhaps that would have been the better choice, and it would have simplified minimum airflow distribution issues affecting our 4 zone duct system, but an 026 would need to be in less efficient Y2 mode far more often.

The availability of reserve capacity is reassuring, and monthly average heating / cooling costs of $30 for a 3400 SF 3 story house with 48 windows and 5 residents (4 of whom pay no attention to energy costs) suggest I may have stumbled upon the right answer.

Your design friction rate of 0.06 mirrors my own standard practice for foamed homes (all ducts within thermal and pressure envelope) In other words, little or no heat gain / loss penalty for moderate oversizing.

That said, bumping the friction rate to 0.08 in exchange for dropping duct diamters is likely a prudent tradeoff.

Excercise particular attention to duct sealing, especially on the return side.

Thanks again for such a great reply!

If I'm reading this correctly, since my ducts are not in the environmental envelope, that going with the smaller duct sizing that is calculated when using a 00.8 friction rate would be more efficient than with using the duct sizes given for a 0.06 friction rate?

The thermostat that I am installing has the ability to measure the delta between the supply and return air temperatures to determine (I believe) how, and when, the second stage is called. I'm also using an outside thermostat that I will use to try and minimize any secondary stage heating unless it is below a certain outside temperature.

Thanks again for the input!

Maybe, maybe not. The only way to be sure is to model it using load calculation software. My Man J s/w calculates both sensible and latent duct losses and gains based on where the ducts are located, the level of insulation, how much they leak (actual or projected) and their total square footage, supply and return. I use a ductblaster kit to inject a bit more reality into the numbers to stave off the usual GIGO (garbage in, garbage out) problems with modeling software.

Do I always drill down that deep while loading a house? No, but I look at what the package defaults return and apply a bit of Kentucky windage when warranted. For deep energy retrofits I advocate spray foam so as to pull the attic ductwork into the pressure and thermal envelope. I invest time and effort into proving the foam job using smoke and blow and then quit worrying about the ductwork since it is then in a mild dry sealed environment and any leaks / losses are returned to the home albeit indirectly.

Foam typically cuts required tonnage and airflow by 1/4 or more, so previously marginal ductwork suddenly becomes right or even oversized, and I can have the guys install a few $2 manual plate dampers to divert air from underblown rooms into previously starved rooms, and call it a day.

Curt,

I guess I would ask the question why you think 1st stage operation is more efficient than 2nd stage operation?

Part load power, capacity and efficiency tables published by manufacturers, my own experience, data from TED energy monitors on customer systems.

Well,

my experience is the opposite, namely that I monitor no difference between 1st and 2nd stage in the COPs of the systems. Keep in mind that the 1st stage operation is measured by the manufacturers with an EWT of 9 degrees warmer (in heating mode) or 9 degrees colder (in cooling mode than 2nd stage). That difference in EWT during lab testing conditions itself will result in a 15% performance difference between 1st and 2nd stage. However, in an installed system, the EWT does not change when a heat pump goes into 2nd stage, thus there is really no difference in performance. The first which showed this in a study

(http://www.builditsolar.com/Projects/SpaceHeating/InField%20PerformanceTestingofGSHP_updated%2011_11_2010.pdf)

showed actually a better system COP in 2nd stage than in 1st stage. As bad as the system designs were in this study, but the data monitoring is consistent. I see the same in our systems.

A WF 038 operating with 4 GPM of 70*F water posts COP of 5.74 and EER 25.7 at part load

At full load COP drops to 4.9 and EER to 18.2 with 5 GPM of 70*F water

That's from WF tables

Posted By engineer on 08 Sep 2012 04:04 PM
A WF 038 operating with 4 GPM of 70*F water posts COP of 5.74 and EER 25.7 at part load

At full load COP drops to 4.9 and EER to 18.2 with 5 GPM of 70*F water

That's from WF tables

This is my observation too.  An almost one-third EER degradation going from 1st to 2nd.

Some Dallas area geo installers design exclusively to use 1st stage only to stay with max efficiency for EER (COP is of no concern due to load being so imbalanced in favor of cooling vs heating).

Best regards,

Bill

Mine jumps from 5.7 to 10 amps in 2nd stage, which is why I interlock Y2 unless really needed.

You are correct, the tables reveal similar things for Climatemaster, however, cfm and waterflow are slightly different between 1st and 2nd stage, but not different enough to explain the non-linear performance drop they specify in 2nd stage. Also pumping power per BTU delivered is less, since the gpm does not change with a constant speed pump, but again, not enough to make up the for the different specs in the tables. Makes me scratch my head why the whole system COP I usually monitor for my systems does not differ between 1st and 2nd stage....???

You are correct, the tables reveal similar things for Climatemaster, however, cfm and waterflow are slightly different between 1st and 2nd stage, but not different enough to explain the non-linear performance drop they specify in 2nd stage. Also pumping power per BTU delivered is less, since the gpm does not change with a constant speed pump, but again, not enough to make up the for the different specs in the tables. Makes me scratch my head why the whole system COP I usually monitor for my systems does not differ between 1st and 2nd stage....???

Your data may be better than the tables, which are probably modeled rather than measured. My own system draws about 250 Watts more than WF tables predict, despite the fact that in my case pumping power is zero