The total latent and sensible evaporator heat load needs to be optimized at your normal operating conditions. This will also optimize the condenser heat load.
Additionally, it is critical to have the supply air and return air near the ceiling where the warmest air is so the temperature differential increases both sensible and the latent heat of moisture condensation transfer to the outside condensing unit!
You must know & record the operating feet per minute (FPM) velocity & the CFM to each room & the Total CFM airflow!
Come on engineers, consumers need a wholly variable system to achieve optimal efficiency performance when functioning in variable humidity conditions. HVAC company engineers can do it, therefore the companies need to get them on the market at reasonable price levels, ASAP.
This computerized control system AC would have to be sized to the combined latent and sensible heat load targets (i.e., 76ºF/50RH) and the cubic foot volume of air changes that we would like per hour. This would need to be performed accurately to achieve the requisite run time and our combined comfort zone and unit efficiency goals.
Whenever you have a heavy latent-heat-load of moisture for the evaporator coil to condense -- it will reduce the evaporator coil's ability to reduce the sensible air temperature of the conditioned air.
Air Temperature Drop Through Evaporator Coil (1987 Period)
Indoor temperature and humidity load variations graph.
Refrigeration & Air-Conditioning (ARI) Second Edition,
Page 624, © 1987
Above: Just a 'rough' brief demonstration of how the latent heat capacity of the DX coil "increases with the increase in room relative humidity.
A modern 2-ton 13-seer system would produce around .70 of a ton or 8,400-Btu/hr, however at 70% Relative Humidity its capacity would increase to around 1.1 ton or 13,200-Btu/hr or over half of the 2-tons would be used for the latent heat-load. "That is around a 36% increase in latent capacity" and a 36% reduction in sensible capacity, --due to a higher humidity.
There are four main factors to humidity control. These are related to equipment selection and installation and the effects of the performance of the four equipment factors. The Four Factors are: evaporator coil selection, airflow, refrigerant control device, and superheat setting of the refrigerant cycle.
Proper duct sizing and location is important. Most older homes need reduced ambient (outside) air infiltration and more effective use of vapor barriers, coupled with adequate insulation. Windows and doors are special areas to work on. My upstairs windows around the pulley wheels for the weights, allowed air to blow through almost unrestricted from the attic area winter and summer.
The Supply Air & the Entering Return Air delta-T, - tends towards less & less as the EER goes higher,
7 EER or less
'Max' condenser air temp 'delta-T'
15 to 25
14 to 24
Max temp drop 'across' E-Coil
'Max' SA/Return Entering Air 'Delta-T'
therefore, dehumidification could become more difficult at the highest EER levels. The EER & SEER levels widen, as SEER sky rockets.
My response to an HVAC Forum question on BTU & Tonnage Ratings - Sizing & Humidity Removal:
Three ton is 36,000 BTUs.
The units are Rated in Nominal Tons per hour.
However, the nominal BTU/hr rating of some range from 36,000 down to around 34,000-BTU/hr.
Additionally, with high indoor temperatures & very high humidity a nominal 36,000-BTU/hr could go considerably higher.
Example, Goodman Expanded Data: a 3-ton condenser 13-SEER GSC130363A, with a 4-ton evaporator coil:
1434-cfm or 478-cfm per ton of cooling
85 OAT Outdoor Ambient Temp
80 IDB Indoor Dry Bulb
71 IWB Indoor Wet Bulb or 63% Relative Humidity
Nominal BTU/hr of 39,500
At 75 OAT outdoor Ambient Temp
other figures the same, nominal listed @ 40,500-BTU/hr. (At ARI Conditions)
Moderate outdoor temps coupled with high indoor temps results in a high latent humidity heatload through the evaporator coil which boils refrigerant at its fastest rate, which transfers more heat outdoors per unit of time.<>Size equipment for the Sensible Heatload at the mfg'ers Sensible Ratio design figures
That is why we should NOT be upsizing equipment for latent heat removal; because the A/C system increases its latent capacity to handle that load. When the unit is upsized the run-time operating-cycles can be way too short for effective latent heat (humidity) removal.
I would rather have my unit a little undersized than a little oversized.
Especially if your system is oversized or there are a lot of low AC load days use an Adjustable Differential Room TSTAT.
TH Differential: Differential is defined as the difference between the cut-in and cut-out points as measured at the thermostat under specified operating conditions. For example, if the thermostat turns the COOLING EQUIPMENT ON AT 78-F & OFF at 76-F that is a 2 degree differential setting. Some have half degree increment settings over several degrees of differential spread.
As you put more total heat-load through the evaporator coil, up to its capacity ceiling, both total capacity and efficiency increase for optimal Btu/hr, EER, and SEER.
When it comes to airflow, the laws of physics apply. Air follows the line of least resistance. So many of the duct systems are poorly designed that ductwork problems can seriously curtail proper system performance. These factors usually show up in uneven temperatures through the conditioned area. In addition, airflow across the cooling coil can affect humidity removal. Too much air will result in poor dehumidification. Too little air can cause the registers to sweat. The right amount of air is usually between 325 CFM and 400 CFM per ton. Lower airflow will produce increased humidity removal, but compromises sensible heat removal. Finding the right air flow and run time balance can eliminate most of the comfort zone humidity problems.
Condensation forming on supply air diffusers or registers is caused by insuficient airflow through the evaporator coil and the individual ducts that are sweating.Refrigerant Control Device
The device used to make a cooling coil evapoprate refrigerant and thus absorb senible heat and latent heat of humidity is called a Thermostatic Expansion Valve refrigerant control. The most effective refrigerant metering control in this application is a balanced port (TEV/TXV) expansion valve. The expansion valve provides consistent performance over a wide range of conditions that exist in any home. Without an expansion valve the entire system performance is compromised. Adding an expansion valve to an existing system can often improve performance, reduce operating costs, and extend the life of the overall system.
One factor that is often overlooked in trying to increase humidity removal is the superheat of the suction gas. Superheat can often be out of design conditions and the system seems to work fine. A five degrees warmer coil temperature can reduce humidity removal by 20% or more. Correct superheat should be about 13 degrees at the coil and about 16 to 17-degrees at the compressor. This measurement is best taken at 90+ outdoor temperature with indoor temperature around but no higher than 80 degrees. Since superheat is a measurement of heat these outdoor and indoor temperatures can affect the readings. The best time to set and adjust the superheat is on hot summer days.
If any of the above four (4) factors are not correct you can expect that humidity problems will occur. Other factors can affect the humidity levels. The way the house was built — the number of people that live there and the life style of the occupants. The correction of humidity problems in any residence can be accomplished simply by applying the above factors. This is why it is essential that you find a company that you can trust to solve your humidity problems. if neglected, humidity problems only get worse.
In my opinion all the major components of air conditioning systems should be engineered and specially selected for the specific climate conditions where they will be used.
In humid climate areas the use of a dehumidistat and variable speed programmable indoor blower motors that use half the electricity of conventional motors should be used. Also, ask your Utility company about their "high efficiency Rebate Program," these programs can add to your energy cost savings.
Arid climates can use a higher temperature operating oversized coil, coupled with 450 cfm or more, per btu/hr ton of cooling, -- cycling through the evaporator coil along with a lower rated btu/hr compressor to condenser ratio.
Let's say you had 0.20 IWC without evaporator coil and large oil furnace heat exchanger being near the coil 0.20-IWC was added, the coil adds 0.30-IWC for a total of 0.70-IWC . Check the static pressure with a wet coil then check the Units Blower Curve Chart to see if you are getting 350 to 400 or 450 CFM per ton of cooling, depending on the humidity removal requirements. Is the blower motor horsepower and blower wheel size and RPM up to the task?
There ought to be a code requiring every manufacturer of an airhandler or furnace to provide capped taps ahead of the evaporator coil and ahead of the blower for easy static pressure testing access.
All air conditioning condenser manufacturers' should publish the CFM and normal temperature rise range across the condenser coil, so that the service tech's can measure the heat transferred from the evaporator coil. Most high efficiency units will have temperature degree rises between 18 and 25ºF. Older lower SEER condensers can have temperature rises up to 28 or more degrees.
Such temperature rise data provides a guide to the actual heat transfer by the evaporator coil to the outdoor condenser coil, and therefore also, whether the proper design amount of (cubic feet per minute) CFM of indoor air/per ton of cooling BTU/HR, is passing through the heat absorbing -- cooling coil.
Most higher efficiency units are designed to operate at higher evaporator coil temperatures which results in less temperature drop, also the high outside and inside humidity levels will put a heavy latent heat-load on the evaporator coil and reduce and delay the conditioned area's temperature drop.
You may also have too much outside air infiltration into your home, check it out and reduce it, because warmer high humidity air will overload the evaporator coil with latent heat removal with the result being little if any reduction in humidity levels and no lowering of the actual sensible temperature readings in the conditioned areas.
In high humidity climates everything in the home is loaded with the latent heat of moisture. Humid air contains a lot of heat-loaded vapor. Some moisture is airborne but most of it resides or hides in the bricks, wooden furniture, carpeting, walls, and the concrete floors we walk on, etc. Dish washers, clothes washers and taking hot showers, etc., all add a lot of moisture to the air and to home materials.
This latent moisture in the conditioned area is vaporizing in the air, and as it is being conditioned it gives up its latent heat to the evaporator coil's liquid refrigerant causing it to boil into the heat absorbing refrigerant vapor. That heat-loaded vapor is then sucked back to the compressor where it is compressed into a high temperature gas in the condenser coils, where the outside air cooler outside air cools it below its condensing temperature point causing the vapor to condense into a liquid.
FINDING the LATENT HEAT OF CONDENSATION
The amount of heat energy in BTU's that must be removed to change the state of one pound of a vapor to one pound of liquid at the same temperature.
You don't have to know this, but to find the LATENT BTUH heat absorption transfer split, of the system, --measure the amount of condensate produced in a given time span and pro-rate it to an hour, or to BTU's per hour|(BTUH).
At a condensing temperature of 55ºF the latent heat transferred to cooling coil is close to 1,066 BTU per pound.
Most use 960-btu/lb, which is okay for our ball-park purposes, therefore we will use it.
For example, use an alarm timer set for 15 minutes, if you collect 18-ounces in 15 minutes, that's 72-oz/hour or 4.5-lbs an hour. (4.5-lbs X's 960-btu/lb 4,320-BTUH) of latent heat transfer per hour.
Or you can convert with ounces: 960-btu/lb / 16-oz = 60-btu/oz. | ounces X 60-btu/oz = BTUH of latent heat absorption transfer performed by evaporator coil. 72-ounces *-X's 60= 4,320-BTUH of latent heat transfer per hour. You can use a kitchen dietetic scale in ounces to weigh the condensate. Using an 18,000-btu/hr system:18,000 - let's say it is 4,500-btu/hr latent heat absorbed and 13,500-BTUH of Sensible heat transfer, at its Rated Tonnage Capacity.
Therefore, 1.5-ton 18,000-btu/hr example would be operating at 25% latent, 75% Sensible. What is the usual latent/sensible split on your A/C unit? If you see any errors let me know!
(Note: Immediately after passing through the evaporator coil the cooler air is near the 100% humidity level and will not absorb more vapor moisture until it mixes with the room air bringing it to a higher temperature. This is why if your returns and supply air are at the floor level, coupled with low air flow, it is important to use low amperage fans to mix the warmer ceiling air with the cooler floor level air.
This dehumidification process continues until a stabilized point is reached in the entire home or building. The air conditioner will continue to pull that moisture out of the surroundings for days before it reaches a point of equalization at a specific system design point. Eventually, it will reach your chosen temperature and humidity comfort zone set point level.
Since a home air unit may circulate all of the air in a space (the home) six to eight times an hour, then it would change the temperature and lower the humidity of that total volume only during each pass through the cooling coil, —of the entire conditioned space.
How much run-time combined with the amount of CFM airflow over a large area cooling coil represent the critical latent and sensible heat removal factors. Each pass through a coil that is below dew-point will remove additional moisture. Increasing the number of the conditioned space passages through a warmer coil can remove more moisture than lesser conditioned space passages through a colder coil. Therefore, CFM Vs room volume Vs run time and the area of the cooling coil are critical factors that directly affect the total heat removal equation.
Using a larger cooling coil with more area of coils and fins at below the dew-point of entering air will permit a bigger volume of air at a lower velocity through the coil. This will increase the systems rate of latent heat of moisture removal per hour that is of great benefit for improving the comfort zone at higher 78 degree temperatures permitting lower cooling costs.
As compared to the amount of air infiltration and increased rate of sensible heat load during an hour, —four (4) conditioned space air changes per hour compared to six or 8 changes per hour also makes a big difference in the amount of latent and sensible cooling a system will produce. With the coils and fins below the dew-point of air entering the evaporator coil, more airflow up to 450 CFM per ton of cooling Btu's will reduce humidity more rapidly than a 35 degree coil but with low airflow at the minimum of 350 CFM/Ton cooling.
Once you get the humidity down in your home, unless the air is very dry, don't open your home up and work to stop all the air infiltration you can. This is important because your AC system will have to run a long time again just to pull the new latent moisture heat load out of the air and materials in your home!
In humid climates, removing moisture should be a most important focus in the selection of your Air Conditioning System.
Simply matching coils doesn't mean you are getting the SEER rating or the rated BTUs of your new or older unit if the indoor airflow, the ductwork, and the entire refrigerant system is not balanced to design specifications! The evaporator coil has to be carrying its designed maximum heat absorption load or your system will not deliver its BTU/HR design rating or its SEER rating.
Page 618, Refrigeration & Air-Conditioning (ARI) Second Edition, C 1987
Those lower SEER units had higher condenser splits than 12-SEER and higher units.
Sorry, I defiled the graph, 90-db outdoor, 80-db indoors with 67 wet bulb or 50% RH represents the condenser splits shown above graph. Graph: 80-DB & 80-WB line-intersect is 100% Relative Humidity.
In order to achieve high SEER ratings, condensers and evaporators are generally btu/hr oversized and compressors are btu/hr undersized. Many 12 SEER 18,000 btu/hr condensers have 16,500 btu/hr compressors and 2 ton evaporator coils. This combination could make reducing humidity more difficult. The SEER increase is based on higher evaporator temperatures and suction line pressures for increased volumetric capacity, permitting the use of lower amp draw compressors. Manufacturer's should design units with a little more compressor btu/hr capacity for very humid climate regions, so TEV refrigerant temperature controlled evaporator coils could be set closer to the freezing point.
Subject: What design for lower duct static and lower blower motor HP?
Remember that many oil furnaces have a large round heat exchanger in the center up to near the top and if the evaporator coil is set too close this can cause extreme turbulence and back pressure which could be a huge factor in running the static pressure way up!
Try using 0.05 for the supply and return ducts & for filter size try using 2 CFM per sq. in. (1000 cfm = 500-sq/in. | A 20" X 25" is only 500-sq/in.
A Hart & Cooley seminar suggests using 0.08 for supply and 0.06 return. What do you use and why? Would it be better to use 0.06 for both? You can always adjust to lower RPM which will reduce the needed blower HP to move the same CFM of air through the cooling coil & diffusers.
Service techs' put your Magnehelic gages' and Digital Micromanometer to good use to measure the static pressure and then get and apply the blower curve charts on each system you are working on, then you know you're getting the proper evaporator airflow temperature and heat-load to meet the customer's desired humidity and temperature comfort zone. It is always very good practice to measure the external static pressure on all systems; you can do this with a simple magnehelic gauge or with a digital micromanometer. In any case, static pressures above 0.45 IWC should be investigated and reduced if at all possible. The Air handlers should be pre-tapped by manufacturer's for the static pressure probes.
The air moving fan law formulas:Required fan motor horsepower (hp) varies as to the cube of the rpm speed: hp2 = (rpm2 /rpm1)3 x hp1 = hp2
Belt driven blower: (new 800-rpm/old 700-rpm)3 | result X's old HP .25
(1.1428571)3 =(1.306122449)= 1.492711314 X .25 = .373177828 HP or 1/3 hp within SF of 1.3
.3333 X's 1.30 Service Factor = .43329 HP maximum load. (Many have a 1.35 SF.)
Brother's 1/4th HP furnace blower curve: at .45" SP @ 700-rpm only gets 500-cfm for 18,400 Btu/hr.
Even at only 350-cfm per ton (way too low for his situation) it requires 525-cfm.
Since his supply and return air is at the floor level and the diffusers are old style for a gravity flow furnace, his needs at least 475-cfm per ton or over 700-cfm.
According to the blower curve chart the 1/3HP would deliver 700-CFM at .55-SP.
This 1/3 HP belt drive blower motor at .65"SP would only deliver 500-cfm.
Finding the New Static Pressure:
SL11B.pdf PDF File: Blower curve lines show in (blasted) yellow, use "black and white darkest printing page settings" to get readable lines!
- SP2 = (rpm2 800/rpm1 700)2 = (1.142857143)2 = 1.306122449 X's SP1 .45-SP = .58-SP2 = or only 600-CFM, or an increase of 100-cfm.Figure ways to cut SP to .5 25" and it gets close to 700-CFM.
The total External Static Pressure (ESP) is the total of the positive and negative inches of water column air pressure between the negative suction return side before the blower, and pressure supply side taken before the A-Coil but after the furnace or air handler blower. If you check and know the Coil is clean, --it's usually much easier to take the reading after the A-Coil and add its pressure,the figure should be in the papers with your A-Coil, normally .3" to .4 inches to be added.
The negative or minus Static Pressure is added to the positive Static Pressure Knowing the operating static pressure is a first order essential to revealing the operating cfm. RPM and HP may need to be increased to achieve the most efficient CFM for the system. I.e., negative -.35"wc added to +.40"wc = .75"wc the total External Static Pressure (ESP).Below, PDF File: ThermoPride Blower-Curve-Chart - Click Print , Click on Properties, Click on Graphics, Slide Setting to the Darkest Setting, click OK, or blower curve lines won't show up on the printed copy!
Every manufacturer should furnish blower curve charts with their units and put them on the Internet for service tech's to download and print. Also, air conditioning codes should be updated in respect to proper sizing of the duct work which must include all the pressure inducing factors when sizing the supply and return ducts.Considering the restriction between the large oil burner heat exchanger and the bottom of the evaporator coil and supplies and returns at floor levels, —a study of the blower graph indicates these figures may be relevant. Due to return air and supply air being at the floor level, target cfm is 700 or 720-cfm for the 1.5-ton cooling system. Some manufacturer's use 720 cfm for a 1.5 ton system, or a high of 480 cfm per ton of cooling.
The unknown variable is the duct system static pressure. It is essential to check the static pressure with a magnahelic gage or manometer, and then you'll know what cfm you are getting with a specific rpm and hp. An RPM strobe light would also be helpful on belt drive blowers.
Compare those readings with the graphic blower-curve chart of that manufacturer's specific airhandler furnace blower. Every duct system varies; therefore it is helpful to figure the actual cfm the system is delivering. If you know the BTUH output that your heating system delivers, you can use the air temperature rise formula to figure the cfm being delivered.
Use the proper speed for cooling, the manual blower on switch will usually cycle the cooling blower speed on high speed. Subtract 50 cfm from the cfm derived from the formula for the wet cooling coil cfm. It may not be getting the requisite 375 to 450 cfm per ton of cooling, especially if there is a low demand for heating in your area and a high demand for cooling tonnage, or the cooling coil is too close to the oil furnace's heat exchanger causing extreme air flow turbulence resulting in high back pressure, raising static pressures to 0.75 IWC or higher (view and print blower curve chart).
In this 1.5 ton unit situation, if the compressor has the capacity to handle the demand, I would go for 480-cfm/ton cooling or 720-cfm, (supply & returns are at floor levels). If the refrigerant charge is then adjusted to match the new TEV metered refrigerant heatload levels on the evaporator coil. This will cause the 1.5-ton/18,000 btu/hr with a 2 ton cooling coil, to generate its optimal SEER and BTUH rating. With the right discharge air diffusers we will get better throw.
All furnaces and air handlers should have a static pressure/blower-curve line graph chart sticker on them for use my the AC technicians to use to accurately adjust required CFM for that residence's specific ductwork pressures.
Any of the HVAC companies I list on any of my web pages have nothing to do with the information I post on any of my Web pages nor do I assume any responsibility for how anyone uses that information.
All HVAC/R work should always be done by a licensed Contractor! This information is only placed on these pages for your understanding & communication with contractors & techs.
This information is for the edification of contractors and techs. I am NOT liable for your screw-ups, you are liable for what you do! - Darrell UdelhovenDarrell's Refrigeration Heating and Air Conditioning - Federal Refrigerant Licensed - Retired Licensed Contractor
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