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Impact of ventilation fresh air on AC loac calc

On Friday, a mechanical engineer I hired sent me a load calculation for an AC system for 1800 square feet of office space in the first floor of a two-story building. (I posted about this space about a month ago.) The office is for an auto repair garage that occupies the rest of the first floor. The office has one window-less masonry exterior wall. On the other side of two of the other three walls is unconditioned garage space. The last wall adjoins conditioned space of the adjacent building. Above the office is conditioned space on the second floor. Only 3 people in the office. One desktop computer. No plants. So the sensible heat load calculation for the walls and people came out to 12,350 BTUH.

The combined sensible and latent load is calculated at 60,800 BTUH. The biggest part of this is the sensible load and latent load of the 880 CFM fresh air ventilated from the outside: sensible load of 19,400 BTUH ((1.1)x880CFMx20dt) and latent load of 28,960 BTUH (880CFMx4840x.0068).

First Question: So the latent heat load is nearly 400 percent of the sensible heat load. Put another way, the latent load represents 80 percent of the total load. Does this seem “bout right” or way off?

Second Question: The 60,800 BTUH load seems to call for a 5 ton compressor. Could the compressor be sized just to deal with 30,000 BTUH of total sensible heat load? On days where the outside design temperature is exceeded, would the system focus first on maintaining the indoor temperature at 75 degrees and allow the indoor humidity to rise above 50 percent? If so, this would allow a 3 ton system.

Third Question: Product descriptions of Heat Recovery Ventilators (HRV) say that they can capture up to 70 percent of the sensible and latent energy of ventilated air as it is exhausted. For my situation, would this mean that an HRV could reduce the 19,400 BTUH of sensible load of the ventilated air by 70 percent down to 5,820 BTUH? Would it reduce the 28,900 BTUH of latent load of the ventilated air 70 percent down to 8,688? If this were so, the total load would appear to drop to 26,868 BTUH, allowing a 2 ton or 3 ton system, wouldn’t it?

I am trying very hard not to end of with an oversized system that cycles on and off alot without removing humidity. I am going to bring up these questions with the engineer when I talk to him on Monday. But I would really appreciate any reaction from people reading this beforehand. Thanks.

Tom

Comments

  • Brad White
    Brad White Member Posts: 2,393
    A second look

    Just another pair of eyes here with experience, obviously not direct knowledge of the details beyond what you have offered:

    If I had an 1800 SF office space, I would take the lighting load at about 1.3 Watts psf and if power means just one PC, add another 200-250 Watts. Call the total electrical contribution 2590 Watts or 8,840 BTUH. Put another way, total of power and lights would be 1.44 Watts/SF as a check figure.

    (In my designs I will often would take, absent detailed information, about 1.75 to 2.0 Watts total for power and lights in an office for a comparative range.) Add three people at say 230 BTUH sensible each and you have 9,530 BTUH sensible for internal gains.

    Bottom line? For internal gains, your engineer's total 12,350 number would allow three people and 1.97 Watts per SF so maybe a printer was allowed, a copier, small office refrigerator? Things I would expect to see beyond what you described. I do not think that 12,350 number supports any appreciable external wall/partition gains.

    External gains have a wide variable range both in mass, insulation and solar direction. The difference to your stated 12,350 room sensible load would allow us 2,820 BTUH to play with. Not much for walls and such. I see four times that.

    If you have, a guess, say 500 LF of exterior wall at an R of 10 (0.10 u) and sunlit at a surface delta-T of 40 degrees above indoor, that is 2,016 BTUH sensible. The two remaining partition walls (1,000 SF?) adjoining the garage at say a 15 degree difference, uninsulated concrete block at R 1.6, would be 9,600 BTUH -see how much that adds up? 11,000 BTUH external gains.

    So- the space sensible load as I see it would be 21,150 to 23,350 BTUH depending if you take my net number or your engineer's number as internal loads only.

    You can see that I suspect a mis-understanding on the internal versus external gains. The 12,350 has to be all internal by my figuring. At those internal gains plus wall/partition gains, I would expect to see an evaporator airflow of about 1,000 to 1,100 CFM or 0.60 CFM per SF. Pretty light but for a windowless space, not unreasonable. The space latent load would be just the people plus any leakage. If 1,000 BTUH latent, that would not surprise me.

    Without any ventilation load, the space gains would be about 2.1 tons but to get a stock unit selection, you would need about a 2.8 call it 3.0 nominal ton unit in order to meet your sensible load. This assumes a 0.75 sensible heat ratio,

    The outside air load: I do not know what the criteria is, but absent other information, I would definitely want to see the greater of:

    20 CFM per person rounded up to allow for visitors (easy to do, say 100 CFM for crying out loud..) or:

    Make-up air for toilet exhaust plus a margin for pressurization and/or:

    At least 100 to 150 CFM surplus pressurization per door between the garage and the office to help keep fumes out.

    I do not want to second-guess anyone here, just am not clear why 880 CFM of outside air is required, no judgment intended. May be a lot of toilets to make up, covering the garage crew also plus their transient ventilation. Let me take the 880 CFM at face value.

    That air has to be taken down to ambient conditions plus wrung out to dry it. Normally say it goes from 90 degrees and 74 Wet Bulb (your local conditions may vary), to 55 degrees saturated at least. That give me about 5 tons right there, without recovery.

    With energy recovery (not just heat but total enthalpy), you might expect 70% recovery meaning the difference between outdoor conditions and indoor conditions (90/74 vs. 75/62), not outdoor conditions and coil leaving air conditions (90/74 vs. 55 saturated).

    Say your leaving recovery outside air is 80/67 with an enthalpy of 31.52 and you have to cool this to 55 degrees saturated so as not to impose a space latent load. This gets you to just over 3 tons net ventilation load.

    Thus you may expect to save about 2.0 tons, said without an actual unit selection on my part. Call the ventilation load about 3.0 tons then, 2.0 tons of which is sensible heat (23,870 BTUH in this case).

    So.... your space net tonnage of 2.1 tons plus your ventilation load of about 3.0 tons gets you to 5.1 tons.

    Of that 5.1 tons, 23,350 is space sensible and 23,870 is ventilation sensible for a total sensible load of 47,220 BTUH or almost four tons sensible. If an 80% SHR unit is selected, you will wind up with a 5-ton unit. Your engineer did well, in my opinion, but you may need a breakdown to understand the parts better.

    Remember though, the sensible heat capacity is a function of the refrigerant and the evaporator coil together. A deeper coil means more latent heat removal and if the OA is sent through that evaporator, you will need a deeper coil maybe 6 rows.

    So, what do you think?
    "If you do not know the answer, say, "I do not know the answer", and you will be correct!"



    -Ernie White, my Dad
  • don_192
    don_192 Member Posts: 45
    I think Tom

    should send you the check.
  • Tom_99
    Tom_99 Member Posts: 9
    IMPACT OF VENTILATION FRESH AIR ON AC LOAD CALC.

    Whoa!, Brad. I am speechless. Except to say that I agree with Don. I need to send you a check. I will email you my contact information. Let’s talk tomorrow. I’m serious. If any Wall reader is a contractor in the New York City area (that is where the building is) with an understanding of the issues that Brad set out, please call me at 240-271-7457.

    Just a little background about how I got into this position of being such a hapless customer. My father has owned this building for 35 years and recently I have had to get more involved in his affairs. The auto repair shop tenant moved in about two years ago, and a little while ago I got a call that the 35 year old compressor and air handler had died. (BTW, the old system was a 7.5 ton York system designed by an architect ). I live out of the area, near Washington D.C., and have been managing this from afar.

    As far as my experience with HVAC matter, I know very little but luckily found an HVAC zen master in my area a few years ago who practices with the kind of competence displayed by Brad. He designed and installed a great radiant floor system in an addition and explained to me that the excessive humidity on the second floor of my house was due to having an oversized compressor in the upstairs system. The system is only 13 years old and my wife doesn’t see anything wrong with the climate while the high humidity drives me nuts (last fall we replaced old single pane windows with high efficiency double panes and this summer the humidity now reaches 74 degrees most days. Any suggestions on how to make a compressor only last 13 years please share.)

    So the evils of an oversized AC system is the last battle that I am now refighting with this replacement project. As you can tell, I have demonstrated the limits of how far you can go in learning about something as complex as load calculations by reading stuff on the internet. It was helpful, but Brad’s virtuoso display is a good reminder about the humility needed when tackling something so complex. In any case, when I heard that the old system was 7.5 tons for only 1,800 SF of office space, I thought that the old system had been oversized (using a 300 SF per ton ratio led to a 6 ton system). So I was suspicious of those who recommended replacing it with the same sized equipment. (Replacing it with the same sized system would be easier because it would allow the same duct system to be used.)

    For the benefit of those who have taken an interest in the issues Brad raised in his reply, here is an update.

    Space Gain

    Brad noted that the exterior and interior space gain of 12,350 BTUH seemed understated. In fact, the calculation was an exterior sensible heat gain of 11,010 (very close to Brad’s estimate of 11,000). The interior calculation, however, was limited to the 750 BTUH generated by the three people ( I mistakenly included the three people’s 600 BTUH of latent load). I need to follow up on getting the interior sensible load calculation to include the lights and power . I see why when these are added, the sensible load gain for the space would likely be 2.1 tons, rather than 1 ton.

    Ventilation

    Brad questioned how the 880 CFM of outside air was arrived at. That is something I want to go over tomorrow when I talk with my engineer. There are three bathrooms with exhaust vents (I think NYC requires 150 CFMs per bathroom, so that would be 450 CFMs there. There are three doors from the offices to the garage area, so that would be an additional 300 to 450 CFMs. Add 100 CFMs for the three people and the occasional visitor and its pretty close to the 880 CFM. There is also an exhaust fan dedicated to capturing tailpipe exhaust from each of the 4 or 5 bays in the auto repair shop. 35 years ago, the architects calculated the outside air ventilation at 1,000 CFM (Oddly, I found their drawings in my Dad’s desk drawer). But given the significant impact this number has on the overall load calculation, I definitely want to take a closer look.

    Outside Air Load Without Recovery

    I found very interesting Brad’s comment about it being about a 5 tons load to take the outside air down to ambient conditions. (Brad used a 90 degree outside air temperature. The design temperature for New York City may be a little higher. I think my engineer’s calculation is based on 95 ). 5 tons before any recovery is considered. This seems to suggest that the outside air sensible load combined with the 2.1 space load would give a total sensible load of 7.1 tons. Does that tend to validate the calculation 35 years ago by the architects who caused a 7.5 ton system to be installed? Is this a good lesson in the dangers of blindly relying on quick ratios like “300 SF per ton” without understanding the nuance of the science involved?

    Total Load with Recovery

    It was toward the last part of Brad’s post that my head began to explode so bear with me.

    “Say your leaving recovery outside air is 80/67 with an enthalpy of 31.52 and you have to cool this to 55 degrees saturated so as not to impose a space latent load. “ 80/67: Is that 80 degrees F and 67 percent humidity for the example? Is the “leaving recovery outside air” the conditioned air leaving through the HRV? If so, does transfer of this energy to the incoming air (which is say 90F/71percent humidity) lower the incoming air 80/67 (in the example)?

    In any case, in this example, it would lower the outside air load from 5 tons to a little over 3 tons. I know that this is just an example, but this is something that I will pursue with my engineer along these lines.

    Lastly, when Brad gets to 5.1 tons for the total load, he says the sensible load, again in the example, is 47,220 BTUH, or almost 4 tons. What is an 80 percent SHR unit? (sensible heat ratio?) that is, the 4 tons sensible load divided by 80 percent to get at the 5 ton figure?

    Overall, adding recovery seems like a way to significantly increase the efficiency of the system. I will never work in the building but, given the greater awareness of global warming these days, I see designing this system as efficiently as possible as a responsibility for me in the situation I find myself in.

    Anyway, I will understand if very few Wall readers, including Brad, are still reading this post at this point. Again, thanks Brad. Have a good night and a good week, everyone.

    Tom


  • Tom_99
    Tom_99 Member Posts: 9
    IMPACT OF VENTILATION FRESH AIR ON AC LOAD CALC.

    Brad,

    One quick clarification about your comment about the ventilation load.

    In my first post, I set out my engineer’s calculation of the sensible and latent loads of 880 CFMs of outside air, without recovery:

    I wrote: “The biggest part of this is the sensible load and latent load of the 880 CFM fresh air ventilated from the outside: sensible load of 19,400 BTUH ((1.1)x880CFMx20dt) and latent load of 28,960 BTUH (880CFMx4840x.0068).”

    In this calculation, the total of sensible and latent load is 48, 360 BTUH. About 4 tons, right?


    In your post, you wrote:

    Let me take the 880 CFM at face value"

    “That air has to be taken down to ambient conditions plus wrung out to dry it. Normally say it goes from 90 degrees and 74 Wet Bulb (your local conditions may vary), to 55 degrees saturated at least. That gives me about 5 tons right there, without recovery.”

    How did you get to 5 tons? It looks like you are using a smaller temperature delta (90-74) than the one from the original calculation (delta 20). So, all things being equal, I would have expected a total load for the OA that was less than 48,360 BTUH. What is different? Are you using different equations than the ones set out above?

    Thanks again for taking an interest in this.

    Tom
  • Brad White
    Brad White Member Posts: 2,393
    Tom

    Good Morning-

    Thanks for the kind words!

    I am also going to write to you off-line so as not to burden too many with background.

    The way I calculate "tonnage" from outside air (or any air mixture for that matter) is by using "enthalpy" or the total heat capacity of the air in question. Enthalpy is expressed in "BTU's per Pound of Dry Air" or just "BTU's per Pound", by the way.

    Warm air that is dry may have less energy than cooler air that has more moisture.

    For example, air at 95 degrees and 30% RH has an enthalpy ("h") of 34.48. This air also has 74.11 grains of moisture per pound and a wet bulb temperature of 67.2F (Expressed "95/67.2" (dry-bulb/wet-bulb), which goes to another of your questions by the way.)

    Air at 90 degrees but at 50% RH (and a 74.9 degree wet bulb temperature) has an enthalpy of 38.35. Cooler air for sure but about 11% more energy. It also is holding 106.52 grains per pound of moisture compared to 74.11 in our warmer but dryer sample.

    In short, the entirety of the process is taken into account.

    When I calculate tons using this method I start with the enthalpy of the entering air and the enthalpy of the desired leaving air. This "Delta-h" is multiplied by air flow rate in CFM and a constant 4.45 (specific heat of air in this range expressed over an hours time) to arrive at BTUH total heat. Divide this by 12,000 and you have tons. Of that, some portion is sensible, the rest latent.

    So, for your conditions corrected, 880 CFM at 95/74 (guessing at the wet-bulb design condition here for now) has an enthalpy of 37.45. I want to drop that air to 55 degrees saturated so that it is "moisture neutral" to a 75F/50% RH space. The enthalpy I seek is 22.32. The difference is 15.13 h.

    Here we go:

    (880 CFM x 4.45 x 15.13h)=59,249 BTUH.

    59,249/12,000 = 4.937 tons, really now. Call it 5 tons.

    Different equations but it gets the same result; total heat all in one.

    To get this sensible portion of this, I take the 880 CFM dropped in temperature from 95 to 55 as follows:

    (880 x (95-55) x 1.085 constant) = 38,192 sensible heat. This is a 64.4 sensible heat ratio by the way.

    To get the "post recovery" tonnage, I used the same enthalpy formula but from a different (and admittedly assumed) starting condition.

    Oh, those "constants"? They can vary, a contradiction in terms I know. They vary with density of the air at those points but are close enough for our work.

    Hope that helps!

    Brad
    "If you do not know the answer, say, "I do not know the answer", and you will be correct!"



    -Ernie White, my Dad
  • Brad White
    Brad White Member Posts: 2,393
    More Thoughts

    Tom-
    This is some of what I wrote to you off-line but given here should it benefit anyone.

    These are other supplemental thoughts as to what the calculation basis is and what the terms and numbers mean. This may help you understand in a different way because, indeed, the terms and numbers seem also interchangeable! I may as well post these on the Wall too for the benefit of others.

    "Room Sensible Load"- This is the number of BTU's per Hour (BTUH) which affects temperature within a given space or system. It does not involve moisture removal, just the raising and lowering of temperature.

    THIS is the number which determines how much airflow (CFM) you need to cool a space and how the ductwork serving the space will be sized. This in turn is based on how cold the air is below the desired space temperature. Normally it is about 20 degrees temperature difference (55F going in to a 75 degree space for example). The quick rule of thumb is to divide your room sensible BTUH load by 21.6 or 21.7 to get CFM. Long-hand this is the room sensible BTUH divided by (1.085 x Temperature Difference)

    "Room Latent Load"- This is the moisture produced in the space, usually by human respiration and perspiration, open fish tanks, kitchen processes, coffee makers, etc. In a typical office these numbers are no big deal usually. Select a unit to handle the sensible load and there is more than enough reserve capacity to handle the latent. Hydrotherapy suites and swimming pools are at the other end of the spectrum. For you, no worries.

    That takes care of the space. The rest of the "tonnage", meaning heat and moisture from outside air, are taken care of at the air handling unit, all in order to get the room air supply at the appropriate condition. This is where talk about sensible and latent becomes more serious and seems to overlap.

    The "ventilation load" consists of the dry heat (temperature) of the outside air coupled with the moisture within it. To remove moisture, you have to cool it below it's dew point. To make it acceptable to be supplied to your space, you have to cool it at least to, if not below, the space dew point. If you want the air to do some positive dehumidification, you have to cool it further still, to dry it out.

    If outside air had no moisture, you could cool it to room temperature and be done with it. Because air does have moisture in it and more than your space wants, you have a challenge. If you cool muggy outside air only to space temperature, it may be close to saturation and contain a lot of moisture. This will be dumped into the room and the RH will absolutely soar.

    Instead, we cool that air down to at least the space dew point, say 62 degrees which corresponds to 50% RH in a 75F room. (Hint: The need for 55 degree supply air? That magic number is based on the noted dew point just mentioned. I did not make this up!)

    So now your hypothetical air handling unit has a job to do. Take in outside air, mix it with return air and cool it down to at least 53 degrees F. (The supply fan motor heat and heat of compression raise this, proximate to brake horsepower, to about 55F.)

    In your specific case, your unit if supplying say 1,100 CFM (to absorb the room sensible load) also has an 880 CFM outside air load. This is a LOT of outside air as a percentage. Most package units only go to 25-30% and that can be pushing it.

    Companies like AAON make rooftop units which are semi-custom and can match "low cfm per ton" systems. You may have to make a match-up of a split system (indoor coil and separate condenser) to meet your total cooling demand.

    Or you may have a separate OA conditioning system which would incorporate energy recovery and be in parallel to the normal packaged AC system. You and your engineer may have some details to work out.

    Brad
    "If you do not know the answer, say, "I do not know the answer", and you will be correct!"



    -Ernie White, my Dad
  • Tom_99
    Tom_99 Member Posts: 9
    IMPACT OF VENTILATION FRESH AIR ON AC LOAD CALC.

    Brad,

    Again, thanks for all your help.

    Tom
  • Brad White
    Brad White Member Posts: 2,393
    EDIT

    In the fifth paragraph from the bottom I mentioned cooling the air to a 62 degree dew point. I meant "wet bulb". Room air at 75 degrees F and 48% RH has a dew point of about 62 degrees. The air coming off the coil would be 55 degrees with a very close (54 to 54.9 wet bulb temperature). Sorry for the mess-up!
    "If you do not know the answer, say, "I do not know the answer", and you will be correct!"



    -Ernie White, my Dad
  • Mike T., Swampeast MO
    Mike T., Swampeast MO Member Posts: 6,928


    While I've yet to use, I often recommend the Rawal APR

    While I don't claim to fully understand how it works, I'm pretty sure that it turns the system into one of "constant enthalpy" that is especially useful for systems with a high percentage of make-up air.

    As I understand systems such as the one you describe, the "ventilation load" is extremely changeable and it's even quite possible for nearly the entire load to be the latent component of the "ventilation load". When this happens the indoor RH will be rising much faster than the indoor air temp.

    Now say you have 1 1/2 tons of total load with 1 ton of that coming from the added moisture of the outside air and you're running a 5-ton condensor off of a standard thermostat. If such conditions are sustained for very long you'll wind up with quite high indoor RH (at a normally comfortable room air temp) because the greatly oversized condensor will only operate for short periods (even though the outside air ventilation is constant) and consequently you'll never be able to bring down the indoor RH to your design level at your design indoor temp. As a result, the occupants will tend to crank down the thermostat just to keep the space from feeling "stuffy".

    The Rawal APR valve in such a situation will variably unload the condenser (and apply a varying degree of superheat I believe) such that you're continually removing the latent portion of the "ventilation load" even if such does not result in a reduction of room air temperature.

    Such conditions routinely happen here in Swampeast MO--even in residences whose "ventilation load" comes mainly from natural infiltration.

    Oversizing of A/C equipment around here is probably even worse than that of heating equipment. I believe much of this is the result of customer requests, "I want to be able to cool the place FAST!", or "If it's 100F with high humidity I want to be able to cool the place to 68F (or lower) with a house load of people." When sized this way it results in people--including me--continually turning down the thermostat at night even when sleeping in the nude with the lightest of bed coverings.

    Again, I've never used the Rawal APR but I have heard from three people who have--including the chief engineer at Unico. All said that the valve "does what it says" and will often result in significant energy savings and higher comfort. When I do install, I'll let everyone know the results. They're a bit pricey especially since I'll really need 4 for my house (attic, master suite, guest rooms, ground floor) and due to piping problems can't do the multiple evaporator thing. For a single unit installed at the time of system installation however, the cost won't be bad and should repay itself rather rapidly in your situation.
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