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  1. #1
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    Ok Gary,
    lets take a typicalpackage rtu with 2 compressor's piped into
    evaps. Circuit 1 is the bottom 50 rows, and circuit 2 is the top 50 rows. Am I to believe that one circuit will provide more moisture removal? I mean I would think the greater surface area would
    extract more moisture than just half the coil.
    This set up is typical of most carrier rtu below 12 tons.



  2. #2
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    You added a compressor, Bernie...lol

  3. #3
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    Okay, let's see if I can explain this a little better:

    Increasing the airflow increases the coil temp.

    Decreasing the airflow decreases the coil temp.

    A lower temp coil removes more moisture than a higher temp coil.

    A bigger lower temp coil removes even more moisture.

    If we install a bigger coil, without changing the airflow, we now have a bigger higher temp coil, which will bring the space temp down faster, but will not do much for the humidity problem, and may in fact make it worse due to the shorter run time.

    If we then decrease the airflow, we will have a bigger lower temp coil, which will remove moisture by the bucketsful.
    Last edited by Gary; 15-03-2002 at 07:51 AM.

  4. #4
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    If instead the original compressor remians and so the coil is being starved somewhat but still has the same ADP then the SHR is maintained but, duty is maintained and dt is maintained.
    It all kinda depends on how much larger the coil is, whether the compressor can handle it, and how far we reduce the airflow.

    Because of the increased surface area, we can drop the ADP further without freezing the coil.

  5. #5
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    So larger coil, slower air but same SST results in reduced ADP and SHR affecting increased moisture removal and of course reduced sensible duty.
    And if we want NOT to remove moisture, we shrink the coil and pump up the volume.

  6. #6
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    I disagree.....

    If you want NOT to remove moisture, you increase coil surface area, in order to decrease TD (increase SST) yet maintain required capacity at resulting lower TD.

    I suggest, forget the airflow rate. That factor is inherant in the unit cooler capacity rating. You select air velocities according to the application.... but if you're going to argue the effect of "changing" airflow... don't go there... because we know what that effect is... we're talkin' factory spec, okay..... not in the feild engineering.

    Check your data on floral cooler specs.
    Last edited by herefishy; 16-03-2002 at 12:45 AM.

  7. #7
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    Marc,

    recheck my post.... I changed the exact quote you are referring.

  8. #8
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    I always post and then think... It's just more of a problem now, since WebRam put that damn instant reply window on the screen !

  9. #9
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    I was going through my Witt literature. I noticed that in the evaporator capacity data it indicated

    "capacity at 10degF TD.....NOT MTD"

    MTD?

    Is that what I think it is?

  10. #10
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    Gary... I think you're jousting me.

  11. #11
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    Marc,

    I expected a little more than that.

  12. #12
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    you were only hoping that I would explain what I think it is ...... SO YOU COULD NAIL ME TO THE WALL THAT YOU'RE WRITING ON !!

    MEAN TD?

    please..... don't punish me.......

  13. #13
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    Originally posted by Fridgetech
    If you had a pack maintain coil SST and the return air temperature was also maintained, then you slowed the fan speed down, you will see the leaving air temp drop and the TEV superheat drop. But the coil still has the same TD. The return air is constant and the SST is constant but we have slowed the air flow down and must therefore be doing less duty.
    I agree you are doing less duty. But you can't slow the fan down, and all perameters stay the same. The TD is not going to stay the same, because the coil temperature is going to drop.

    I think you're just trying to take up my time.

  14. #14
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    Dang, I come back from dinner and you guys have filled two pages. And now you're having controlled dreams. LOL

  15. #15
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    In answer to bernie's question regarding a two compressor split faced evaporator, I think you would get more dehumidification from operating only one of the coils simply because, as Gary suggests, run time. We find the same phenomenon in stores that have oversized A/C units, high humidity and cold dry bulb because temperature is too rapidly satisfied with both coils operating.

    Turning both stages on will withdraw more moisture for the short time the units run, but you will get more latent heat removal in the long haul by operating only one slab, or a/c unit, especially when one is controlled by a humidistat as well as a thermostat.

  16. #16
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    Dan would love this conversation.

    Manufacturers are re-thinking as we speak...lol

    I was thinking more along the lines of desert A/C, but okay, let's consider floral cases.

    We have two considerations, here. The humidity of the air and the humidity of the product. The product needs high humidity and minimal airflow.

    Keeping in mind that this is not my area of expertise, as I understand it, the latest strategy is big coil, low airflow, quick pulldown and re-humidify.

  17. #17
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    well that in fact is the argument Ive been having with my boss for awhile now. SO its true, size dosent matter.
    What if we add 10%fresh air thru that same coil?
    Its a 12 ton machine on a 7 1/4 ton load. The temp swings are rapid. on an 80F day I can drop the temp in the building 10F in 45 minutes. I was considering the RAWL device as I believe its called to solve my problem.
    I converted the drive package down from 6000 cfm to 3900 cfm, and dissable 1 compressor.
    Im sure when It gets hot again Ill need anither couple of tons,
    this is why im considering the above mentioned.

  18. #18
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    Speak of the devil...lol

  19. #19
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    What does the RAWL device do, Bernie?

    Why not disable the second compressor with a humidistat? Bring it back on when the humidity drops?

  20. #20
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    Dan would love this conversation.
    LOL. I miss one day and there are 40 new posts! I think what Gary was specifically alluding to was what competing theories even Hussmann apparently has to reduce humidity in fan coil service meat cases.

    The old theory was to operate with an EPR with a thermostat either as back up or as the primary control and the EPR to keep suction pressure as high as possible. The goal being to maintain desired temperature with highest humidity.

    Then a few years ago, the thinking seemed to change. That higher humidity can be maintained by letting the evaporator run colder and cycle a suction stop solenoid for best humidity/temperature control. This thinking follows the line that you refrigerate the case quickly and use the off-cycle to rehumidify the case. An interesting paradox, eh? It is perhaps design and application specific, but for product life, the latter seems better. It would likely further apply to a floral cooler.

    The EPR has a lower rate of dehumidification, perhaps maintaining 75 to 80% but an increased run time (often 80%) The t-stat and solenoid will have a 30 to 40% run time, even though RH drops to 50% during the run time. But RH rises to 90% during the off-cycle.

    I would hazard that the difference between a floral cooler or meat case operation compared to a/c operation is that there is a greater effect of rehumidification during the off cycle in the refrigeration applications than there is in an a/c application... perhaps attendant to the relative size of the evaporator to the space being conditioned. This might affect a ratio of runtime versus off-time as well as re-evaporation surface during the off-time.

    Very enjoyable thread, fellows.

  21. #21
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    It is my understanding that the RAWL device will allow be to run that other 6 ton compressor and it will by pass enough hot gas
    make up for the drive change I made to the system. While at the same time give me the extra btu's I need on the warmer days.
    From what Ive read it will unload down 1 ton equivelant capacity.
    This is a pet project of mine. Our company has never tried this before but we do have some latitude for experimenting.

  22. #22
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    I'm not sure I understand the A/C problem.

    I'm assuming the goal is to increase dehumidification, hence the airflow reduction and disabling the second compressor. Is this correct?

    Did making these changes solve the humidity problem?

    Did these changes cause other problems?

    For what purpose are we considering HGBP?
    Last edited by Gary; 16-03-2002 at 02:59 PM.

  23. #23
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    Marc I havent had time to read your post yet, Im off too work.
    ON the surface it looks very imformative.
    OH, its Bernie Glasgow from Chicago Illinios.

  24. #24
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    Were's the prof these days?
    Hes lurking somewhere
    He's been in DC giving a presentation at the ARI hvac intructor's workshop..... now he's trying to catch up on all these damn posts....

    Lol, lets all phone the Prof at the same time, ready... 1.... 2..... 3!
    No need, you've all got a handle on this problem...

    The prof might hate me, but zp charges should only be used when a manufacturer intentionally slights the motor size for the displacement.
    The Prof doesn't mind.... In fact, the 'Z' charge is the preferred charge if the compressor can handle the pulldown requirements. The 'ZP' charge has an MOP to limit pressure at the compressor, but the charge will migrate if the head becomes colder than the sensing bulb. The net effect of using a 'ZP' charge is it will slow the pulldown of the box or case temperature, and in doing so it will reduce the load on the compressor.

    This equipment is less than 6 months old,
    the element is a kt 43 zp
    I have also had problems with the 53 elements
    The #53 element is a lerger sized element. Perhaps you may be seeing more migration issues with it. Also, you can expect to see more charge migration problems with an internally equalized TEV with a 'ZP' charge, than with an externally equalized TEV. With an internally equalized TEV, you have the underside of the valve diaphragm exposed to evaporating temperature.

    80 more posts to go....
    Prof Sporlan

  25. #25
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    I'm not sure I understand the A/C problem.
    The problem is simple, I have a commercial office space that reqires 15 tons of cooloing. This would be split up bewwen 2 rtu
    each 7.5 tons
    Problem is I have each unit is 12 tons for a total of 24 tons.
    To further complicate the issue This building utilizes carrier vvt
    control and damper system with automatic changeover betwwen heating ang cooling demands.
    The run times for the air conditioning are so short in duration there isnt sufficient run time for dehumidification and with an open ceiling design the excessive air velocity just rains down on the employyes.

    I'm assuming the goal is to increase dehumidification, hence the airflow reduction and disabling the second compressor. Is this correct?
    Yes, and no. By reducing the total cfm output The system runs for a longer duration and has eliminated the rapid temperature swings and has made the winter months quite comfortable. This change was implimented in late december 2001. I have yet to determine the effects of high outdoor amients on this this change, thus the purpose for this discussion.

    For what purpose are we considering HGBP?
    If I end up needeing that other compressor due to humidty levels that are unacceptable once the weather changes, Ill need to be able to use that 2nd stage compressor with out undoing the drive change I made to the machine. This is why Im intrested in HGBP> This is all in an effort to save the customer the cost of replacing the roof top units as they were just installed in 1993.

  26. #26
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    It seems then that the ideal would be to bypass hot gas to the distributor on each coil, adjusting it to maintain SST just above freezing, thus minimizing ADP. Then if the total heat removal is insufficient, increase the airflow.

  27. #27
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    Please, Ill have to plead ignorrant. What is ADP? Im sure I know what it is if I knew what you meant. :>)

  28. #28
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    ADP = Apparatus Dew Point

    The lower the ADP, the more moisture is removed.

  29. #29
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    ok lets explore that a little furhter to ensure I ifact understand what you mean.
    In our case the apparratus is the evaporator coil.
    The dew point is when the temperature of the coil is optimum
    for extracting moisture from the air passing thru it.
    So I would assume that a slower fan speed would reduce the evaporator temp thus lowering the adp?
    Is this thought process correct?

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  31. #31
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    I really like your idea of splitting up the duty between both compressors by passing hot gas into both the evaporator rather than just one.
    Wit air conditioning return gas temps are higher and acceptable up to 65F in some cases.
    Im going to have to get the data for this machine and really scrutinize it to insure I size this properly.
    Of course I have to entertain the liquid injection aspect of it also,
    I have to find that web site again about that RAWL device.
    It so simple to install, there is alot less piping to do. Its an all in one from what I remember.

  32. #32
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    Hello marc,
    Could you provide me with what Lmtd means? Im sure I know what it is if I knew what you meant. :>)

  33. #33
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    I'm thinking to complete the picture, we need to know what brings in the second compressor. If it is brought in by suction pressure, we could have both compressors running regardless of load.

  34. #34
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    Well My guess is it would cut in to its product line but more
    importantly the rawl is a one size fits all concept.
    Ofcourse those in the field know better and I think thats why sporlan recommends the use of a DVB and a desuperheating
    TEV to achieve more accurate results that are application specific.

  35. #35
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    I'm thinking to complete the picture, we need to know what brings in the second compressor. If it is brought in by suction pressure, we could have both compressors running regardless of load.
    In this case its the system demand as it is communicated to the monitor thermostat. Either a large difference in space temp and set point, or the vvt system monitors duct temperature and stages the compressors arrodingly.
    This system is set to maintain 50 discharge air temp, if we are above that temp for say 5 min or so during a call for cooling the 2nd stage will be turned on.

  36. #36
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    Can someone post a link were I can download a goog coversion calculator? This k(kelvin) C (Celcius) make it difficult at a glance to keep up. Im terrible with algebra:>)

  37. #37
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    That control strategy should work well with the HGBV on both coils.

  38. #38
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    I really like your idea of splitting up the duty between both compressors by passing hot gas into both the evaporator rather than just one.
    If either of the coils do not have hot gas bypass, that coil could freeze because of the reduced airflow, so this must be done to both coils.

    The advantage to reducing airflow is increased dehumidification.

    The danger is coil freezing.

    The disadvantage is reduced system capacity. But since this unit is oversized, this does not apply here.
    Last edited by Gary; 17-03-2002 at 12:17 AM.

  39. #39
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    These units we are talking about (like to have a model #)
    are most likely of the horizontally split coil. I have some of them in service and have in some applications had to set them up with controls so that either circuit could run if one failed, so cooling would not be lost and various critical equipment that needs cooling starts shutting down and all hell breaks loose with the loss of production and more money down the tubes than it would cost to buy dozens of new units. What I found was that if the top coil ran without the lower coil was that any dehumidfying done by the top coil was re-evaporated by the lower, offline, dry coil, no condensate running out the drain eh! No latent cooling but enough sensible to keep the cooled machinery on-line, though it felt a bit wet in the space. If I had to sit in there for 8 hours doing office duty I would be miserable. Bernie probably has the carrier rooftops with 06D compressors and horizontal split coils? How many unloaders? How many heads or cylinder pairs? These units are slapped up on rooftops by unknowing engineers by the dozens and frequent short cycle and humidity control problems crop up rather readily for the mechanics to deal with. They size them for worst case loads and don't worry too much about about what happens when the load is not so great and we go past the unloading capacity of the machine and get into the short cycle or no latent removal zone. Get some model #'s up on here and let's look at exact original design and go from there, maybe a control scenario could be worked out to tweak cycle times and or off cycle time to maximize latent load control.
    Mike Hopkins

  40. #40
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    Carrier M# 48hjd014
    Nominal capacity 12.5 tons
    compressor 2 RECIP
    refrigerant type R-22
    Circuit 1 9# 8o.z.
    Circuit 2 9# 5o.z.
    condenser fan 2 propellar @ 22" diameter
    evaporator fan centrifugal 15"x15"
    Nominal cfm 5000 @ 942 RPM @ 0.2 extrnal static(in. wg)

    total face area of evaporator 11.1(sq ft)

  41. #41
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    An 014, you can throw unloading out, both compressors will be can recips. Either 1 or both compressors or nothing in between.
    Carrier did have an option on those units called a moisturemiser.
    Basically they called it a subcooling coil which was placed on the leaving side of the evap and took hot gas off the compressor instead of using the regular air cooled condenser to reheat the leaving air and reduce humidity without using electric strips to do the job. You may look to see if the units could be equipped with that option and field install it. Control would be by humidistat and bring unit on in reheat mode running 1 compressor and auxilliary condenser, if set up right you could put airflow back to original.
    You could still run wide open straight AC when the load hits in the summer. Probably similar to the RAWL device but designed for the unit by the manufacturer.

    Mike Hopkins

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