It is essential to understand why the local weather environment dictates what SEER level air conditioning equipment you should choose.In choosing equipment and its SEER level, it is important to understand the design engineering behind in its functional capabilities.
First, when the engineers designed for higher Seer levels, they increased the volume and the BTU per hour capacity of the condenser coils and the evaporator coils; however, they reduced the BTU per hour capacity of the compressor. The volumetric capacity of that smaller compressor depends on the absolute suction and discharge pressures under which the compressor is operating.
The lower volumetric capacity ratio of the compressor to the higher coil capacities only works well in an 82°F laboratory weather environment with a 50% relative humidity level, which is never a stable operating condition in the real world environment.When you have the high Seer units in a climate where you have high outdoor sensible temperatures along with a high humidity, the temperature pressure ratio of the evaporator coil skyrockets upward as does the condenser coil pressures and temperatures, therefore the smaller BTUH capacity compressor in its relationship to the coils becomes overloaded.
Here comes the
engineering caveat, if you
are in a high
temperature high humidity climate zone the evaporator pressure
ratio will be so high that there will be very little condensation of
moisture in the air. Additionally, the volumetric capacity of the
compressor will not be able to handle the increased volume of vapor.
this the condensing pressure will be much higher with also reduces the
volumetric capacity of the compressor which is rated at less than the
hour of both coils.
||7-EER or less
condenser air temp delta-T
temp drop 'across' E-Coil
|Max SA/Return Entering Air Delta-T
The Supply Air &
the Entering Return Air delta-T, - tends towards less & less as the
EER goes higher,
therefore, dehumidification can become more difficult at the highest EER levels. The EER & SEER levels widen as SEER sky rockets.
Those other options permit a reduction in the A/C tonnage with better airflow, etc.
I normally would measure the airflow with a flow hood, also called a capture hood. You should normally have around 400 CFM (Cubic Feet per Minute) per ton of cooling. Half of the systems I measure have [a mere] 200 CFM per ton, OR LESS. This will be aggravated by a dirty air filter. Also by Restrictive high efficiency air filter's &/or grilles closed in rooms that you are not using. Normally, do not turn the thermostat down below 70º [74º 76º -better] degrees. says A/C Tech guru, 'Stretch'
Summer Comfort Zone
Maximum Comfortable Temperature
Minimum Comfortable Temperature
The above comfort zone was found to be acceptable to 90% of test subjects drawn from a range of age groups and genders, with work and life-styles involving varying levels of activity and clothing. An air conditioning system that establishes and maintains indoor conditions within this zone will provide thermal comfort. It will produce a neutral sensation, occupants will feel neither too hot nor too cold. Above chart and findings From: Home Energy Magazine Online September/October 1996) Sizing Air Conditioners: If Bigger Is Not Better, What Is? by John Proctor and Peggy Albright Toward Optimal Occupant Comfort
If you over pay for over capacity equipment, --you will be paying more every month and will not be as comfortable as you would sizing it right to also achieve the appropriate humidity levels!"Select an air-conditioning system and Service Tech to achieve optimal comfort, efficiency and savings." A lot of A/C systems with older furnace air handlers and duct systems are not delivering your A/C systems' BTUH and SEER Ratings.
If you do not absolutely know whether the metering device is a TXV, or a fixed orifice device or cap tube.
Hook up your manifold gauges, block off considerable condenser air intake for a short time.
If the suction pressure starts rising, you have a piston, or a cap tube.
If only the high side goes up, you have a TXV.
Have things with you in your van or truck to block-off the condenser air for a short time.
Check every time you are not certain what metering device it has.
There will be a lot of guessing in the future.
Do this procedure on known metering devices to observe the difference.
Report back to me how well it works for you.
In some situations, that could save you from cutting a hole in the plenum.
Squirrel cage wheels with forward curved blades on residential systems
unload when discharge air is blocked off too much & will overload
when there is no static pressure.
There is a preferable ESP range for each Air Handler blower design, that ought to be listed on the blower; they vary at the point of serious unloading.
If you amp-probe check enough of those blower motors, if the amp draw is too low according to its rating, you can begin to tell that the External Static Pressures (ESP) is too high.
Additionally, mfg'ers could list the amp draw at various design ESP numbers, then we could amp-probe & know if it was too far above the amp rating, a duct maybe off,
if amp reading is too low, it is time to check all static pressures & delivered CFM to each room.
I lot of us used to set a nearly empty R-22 cylinder on top of a condenser to warm it a little. Back then fan motors had more HP
& higher amp draws, therefore it didn't seem to cause any harm, just more noise.
Back in the 1960's & 1970's there were a far number of TXV metering devices & some table top condensers' that had the fan underneath blowing up through the coils.
Well, where there were cottonwood trees, nearby clothes dryer lint vents, or a lot of leaves or other debris under the unit, the fan motors would be blocked overload & burnout.
I don't understand the engineering genius of that moronic design.
However, on hot days & a heat-loaded E-Coil,
You could move your wrist over the condenser from outlet up to inlet, & tell if the liquid was taking up too much area of the coils; an overcharged system. - udarrell
Always get the CFM airflow correct, first, if it is a piston or cap tube, use the superheat method to charge it.
If it is a TXV, subcooling is the way to charge it, but check the Superheat to verify the TXV is holding within its known specs
At normal room temperature settings, most evaporator coils are "heat-absorption under-loaded", due usually to Supply-Air and Return-Air being at the floor level and/or inadequate CFM airflow through the indoor evaporator coil! Lint plugged indoor coils are also a major problem! Also, unbalanced heatload on evaporator coil circuits.
Optimizing the Evaporator BTU/hr Heat-Input Important!"A lot of older furnace air handlers and duct systems, are not delivering anywhere near the AC Unit's BTUH and SEER Ratings. This is primarily due to inadequate cubic feet per minute (cfm) of a balanced air flow through the evaporator coil circuits, and/or dirty fins/coils and lint filled blower wheel blades. Also, improper location of supply diffusers and return air grills can result in inefficient floor level recirculation of the cold conditioned air, resulting in a lack of a proper heat load through the evaporator coil.
An unbalanced airflow through the evaporator coil (DX coil) circuits can cause a large reduction in heat absorption capacity. The non heat loaded vapor or ultra cold liquid will cause a TEV to shut down the flow of refrigerant to the coil. Superheat charging will be inaccurate when the coil is fed by flow rater pistons. Total BTUH capacity could drop 15 to 30% or more. An 18000-BTUH unit losing 30% of design capacity would be delivering only 12600-BTUH.
Thermostatic Expansion Valves (TEV / TXV) systems should be set for a minimum 13-Degrees Superheat. A/C Trouble Shooting Chart New!
You could rent a power sprayer with a long 90-degree sprayer wand as “it must be in perfect alignment with the fins or you will bend then.” Start with low pressure and work up. Some power sprayers have water heaters and evaporator coil cleaner can be added to the spray.
If possible loosen some screws and raise the coil — placing a large cardboard, thin ply-board, or large flat pan under it. Have someone use a wet-vac while you are cleaning it. Use a flash light above the coil and a mirror under it to make certain you get all lint from between the fins cleared.
I would pull it and take it to a car wash, but you can not legally do that. Never use anything but evaporator coil cleaner as some detergents contain oil based agents that can insulate the coil and fins.
The blades on the squirrel cage blower will also need to be thoroughly cleaned. Do not knock any of the balancing weights off, or leave any dirt on a blade that will unbalance the wheel.
The evaporator coil should not have frost on it! It is DIRTY and/or has a LOW airflow heat load on it! The lint and scum will be on the air intake side of the coil and up between the fins. Use a good evaporator coil cleaning fluid. Check to see if the blower wheel blades are dirty! Take the blower out and clean the blower too.
Additionally, always measure the temperature rise "split" off the outside condenser coil as that will tell you how much heat it is transferring outside.
The performance of a conventional split system residential air conditioner is highly dependent on adequate air flow through a clean evaporator coil to achieve the btuh heat absorption rating of the AC unit. The Air Conditioning Contractor's Association of America recommends selecting cooling equipment (Manual S) based on its stated sensible and latent performance (from Manual J), designing ducts to accommodate the necessary air flow (Manual D) and adjusting air handler (motor HP if needed) fan motor speed to match and achieve their rated btuh loads.
Reference: (Rutkowski and Healy, 1990; ACCA, 1995a,b,c). (0.05-IWC sizing returns is better.)
As I view it, the problem with this approach is that manufacturer's seldom provide blower curve charts to contractors, therefore service and installation techs seldom check duct static pressures in the field to determine if design flow rates, --with a specific duct system is being achieved. All furnaces and air handlers should have an external static pressure blower-curve line graph chart sticker on them for use my the AC technicians for use to accurately assess and achieve the required CFM for a specific AC unit and its unique ductwork system.
Each residence will normally have an entirely different duct system, most of which were never optimized for efficient cooling performance.
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 for accurately adjusting the required CFM airflow for that residence's specific ductwork IWC pressures.
Oil furnaces with large heat exchangers near the top of the furnace which then have the evaporator coil installed at the top of the furnace may have excessive turbulence and back pressure elevating static pressures reducing air flow. This also eliminates the kinetic force of direct air velocity through the coil. Let's say you had 0.30-IWC of external static pressure without evaporator coil pressure drop and due to the large heat exchanger near the coil add 0.20-IWC, the coil adds 0.30 IWC for a total of 0.80 IWC (Inches Water Column) . Check the static pressure with a wet evaporator coil then check the Unit's Blower Curve Chart to see if you are getting 400 to 450-CFM per ton of cooling with a wet coil, depending on the humidity removal needs. At the tested static pressures, --does your blower's motor horsepower and RPM deliver the required CFM of airflow?
Return Air Filter Rack Grille Sizing
For Example on a 4-Ton A/C or Heat Pump:
Let's look at two, Return Air rack/grilles 625-sq.ins., each for 1250-sq.ins *X .75% = free-air-area of 937.5-sq.ins., / 144 is 6.5-sq.ft., free-air-area; then 1700-cfm / 6.5-sq.ft. is 261.5-fpm velocity without a filter in the rack. Another method for 2-cfm per one sq.in. of figured free-air-area: 1700-CFM/850-sq.ins.= 2*144-sq.ins.= 288-fpm velocity.
The filter will reduce the sq.ft. free-air-area, thus increasing the fpm velocity, as it loads.
*All filter mfg'ers should print the free air area of the clean filter on the edge of the filter (we need that data) along with the pressure drop data.
Divide the rated CFM the duct is carrying by the free area sq.ft. of the filter for airflow velocity in FPM. (Or use above formula with the duct's sq.ft. area for duct airflow velocity)
Velocity in FPM = Known designed CFM to room divided / by Sq. feet of duct area.
I.E., 8" duct 8x8 = 64 x .7854 = 50.26-sq. In. area / 144 = 0.3490666-sq.feet | designed CFM to room is 173-CFM X's .3490666 = Velocity of 495.6-FPM, you can use a ductulator to get the actual Friction Rates (FR).
Formula for finding CFM Airflow and/or Velocity in FPM & BTUH
If you can measure the air velocity coming from a duct, here is a rough ballpark formula to get the CFM:
CFM = (velocity in (FPM) Feet per Minute times the square footage of the duct area)
I.E., 16" Rd duct 201-sq.ins. / 144 = 1.3958333-sq.ft. X's Velocity of 800-fpm = 1116-CFM
Times 1000-FPM = 1395-CFM. Branch ducts: 7" Rd duct 38.48-sq. ins. = 0.2672222-sq.ft. X's 500-fpm=133.6-cfm X's 30-BTUH per cfm ratio = 4,008-BTUH X 6 branch runs = 24,048-BTUH, or 2-Ton.
Additionally, all air conditioning condenser manufacturers' should publish the CFM and normal temperature rise range across an upflow wraparound coil air discharge condenser coil, so that the service tech's can measure the total latent and sensible heat absorbed by and transferred from the evaporator coil.
CONDENSER TEMP-SPLITS - My Brother's Heil 12-SEER Condensing Unit
1.5-Ton - Rated at 18,400-BTUH, Condenser fan CFM 1400 (Total Cond. Watts 2221 X's Power Factors 0.85 X's= 1887 X's * 3.413 = 6,443-BTUH Motor Heat additive +18400= Motor Power "Rated Gross Heat Ejection" is 24,843-BTUH / 1400= 17.7-F = 17.5-F Temp Rise Cond/Split. His condenser only gets a 10 to 12 temp rise split, the evaporator appears to be under heat-loaded or, an unbalanced heatload on the DX coil's circuitry.
Brother Don’s 18,400-Btu/hr Heil central A/C unit.
1400-cfm (outdoor) condenser *Xs 1.08 *Xs mere 12-F split = 18,144 minus 8,591-Btu/hr motor heat = 9,553-Btu/hr net Xs .80 sensible = 7,642 sensible 1,911 latent.The probable cause is "an unbalanced airflow heatload through the evaporator coil. "It's a (Thermo Pride OL 11 oil furnace). Those oil furnaces have a very large round heat exchanger that goes to near the top of the furnace, --due to a low basement ceiling the DX coil sets perhaps illegally close to the heat exchanger causing a few of the coil's circuits to be under heatloaded. Since the liquid refrigerant is not completely evaporated it will cause the outlet line that the TEV sensor bulb is on to be too cold and the TEV will shut-down the flow, which greatly reduces the BTUH capacity of the DX coil and the system. On piston refrigerant control systems, they may flood back liquid which could damage the compressor, unless the system is way under-charged.
Also, appears to be low on refrigerant with a mere 12-F indoor SA/RA split. Also, could be an unbalanced load on the evaporator circuits causing the TXV to shut down the refrigerant flow; among other things.
Thermo Pride could install airflow turning vanes just above the heat exchanger to funnel the air directly into the DX coil, instead of most of the airflow hitting the bottom of the DX's drain pan causing extreme turbulence back-pressure and an imbalanced DX coil circuitry heatload! This will cause a dramatic drop in the design capacity of the system. Checking the condenser-split, as outlined below, will reveal whether you are within 5 or 10% of design capacity!
These are ballpark figures. Unit owners can just check the outdoor condenser split and call a tech if it is not within a ballpark of say 15% of design. The Tech will have to be experienced and sharp to locate the cause or multiple causes of the capacity deficiency.
Do your own figuring based on this formula. Motor BTU/hr additive = Watts X's PF x's 3.413 for Btu/Watts additive added to rated BTUH, divided by condenser fan CFM X's 1.08 = condenser Temp-Split. Get the Motor Power Factors (PF) of the compressor and fan motor from the manufacturers. (A 0.80 factor could be close.)Some of the temp-split figures need correcting, will do ASAP. Most Splits rounded off.
CONDENSER TEMP-SPLITS 12-SEER units - Comfortmaker® | Heil® | Temp Star® - used 0.88 Motor Power Factors5-Ton 59,000 25-F Temp-S Cond. CFM 3400 WATTS 6969
ARI Conditions are: 95ºF-OAT; 80ºF-IDB; 67ºF-IWB or 50%RH | TVA conditions; 95-OAT; 75ºFIDB; 63-IWB or around 50%RH | Try 85ºF-OAT | Outdoor Ambient Temperature (OAT)
1.5-Ton 18,000 21-F Split Cond. CFM 1400 WATTS 1536 1.5-Ton is from actual published DATA - Only ARI Rating Conditions
1.5-Ton 18,000 @ 95ºF OAT; Indoors 75-IDB; 63ºF-IWB or near 50%RH; @ 600-CFM; Outdoor Ambient Temperature (OAT); 18ºF condenser split | @ 85ºF OAT; 67-IWB or 66.5%RH; +20ºF cond. split.
To figure this; units pressure chart, the Temps, instead of IWB the %RH, & CFM, For users, No gauges required, to check if your A/C is near specs! However, the temperatures & indoor humidity make a big differenence in the condenser split. (Airflow & proper load on evaporator!)
Take the both the indoor Supply Air & Return Air DB, WB or %RH , too! If you have an accurate airflow CFM, I can Ballpark the BTUH your A/C or Heat Pump is delivering in the cooling mode.
1.5-Ton 18,000 18 to 20-F Split -Cond CFM 1400
2-Ton 24,800 24-F Temp-S Cond. CFM 1400 WATTS 2659
2.5-Ton 30,200 21-F Temp-S Cond. CFM 2000 WATTS 3404
3-Ton 35,600 18-F Temp-S Cond. CFM 2800 WATTS 4117
3.5 Ton 42,500 21-F Temp-S Cond. CFM 2800 WATTS 4554
4-Ton 48,500 19.5-F Split Cond. CFM 3400 WATTS 4761
The chart split listed above is at Condenser Design conditions: Indoor Return Air 80-F dry bulb 67-F Wet Bulb or 50% Relative Humidity as you go up to 99% RH the condenser split could increase by up to 6-F; down as much as 4-F at a low humidity of 55-F Wet Bulb. We are only trying to get a figure to go by for a comparison.
IE Browser's Click for Graph
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/50%RH represents the condenser splits shown above.
Typical matched units from major manufacturers have Sensible Heat Ratios (SHR) in the 68% to 80% range (or 32% to 20% Latent) when it is 95-F outside and 75-F with 50% relative humidity inside. Proper mixing of the air and proper distribution to individual rooms is critical for comfort.
When new condensers and Evaporator-coils "are installed on older air handlers" the new, or old, evaporator coils are usually under heat-loaded. (Always, check voltage and amp draw!)
The Base Spec sheets 12-SEER part no. 421 41 33301 03, Feb 2001. These are the Comfortmaker® units, which are nearly identical to Heil® units. I used the first rating on each tonnage class. While the "Performance Cooling Data" is listed at a 95-F outside ambient temperature, you can adjust the indoor airflow to get the Nominal BTUH Rating at the customer's normal indoor stat' temp' setting and the most outside temperature/degree operating hours.
Take the "listed watts" of the compressor and Condenser fan and multiply that wattage by 0.85 X's 3.413 to get the BTUH heat additive of the motor then add the listed BTUH of the condenser to it, and then divide by the condenser fan's CFM.
By using the various units' "base specification sheet data" from the dealer, you can determine if it is operating near its BTUH capacity rating. Some packaged units run a very high condenser discharge CFM airflow!
Some "Condenser Makes" will have different temp-splits. The 2-ton 10-SEER, Janitrol; GMC; Goodman; with the U-29 E-Coil delivers less btuh, or 23000-btuh, I subtracted a reasonable amount from the total of the wattage and come up with 19 to 20-F temp-split. That is "if" its CFM is 1400, --get the figures on the "different Makes." The figures are used to provide an idea of what the condenser temp-split should be for use by the unit's owner and the service tech.
When the CFM of the condenser is known the temperature rise data equation can be applied to provide a guide to the actual heat transfer by the evaporator coil to the outdoor condenser coil. (I prefer 425 to 450-cfm/ton dry coil.)
Fan delivery in cfm varies directly as to the rpm fan speed:
- cfm2 -new = (rpm2 / rpm1) *X cfm1 = cfm2 -newFinding the New Static Pressure:
- SP2 new = (rpm2 / rpm1)2 *X SP1 = SP2 -newRequired fan blower motor horsepower (hp) varies as to the cube of the rpm or cfm:
hp2-new = (rpm2/rpm1)3 *x hp1 = hp2 -newThere 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.
Solving the Mysteries of ESP - External Static Pressure
ESP is the static pressure "external to the air handler." This is the reading that manufacturers' refer to in their fan performance data & is usually taken prior to the entry to the evaporator coil.
Take your measurements on both the Return and Supply Plenums of the furnace, as it was shipped from the manufacturer (including the filter).If you leave out the area up to & including the A-Coil where does that leave you? The area to the coil & including the coil can represent major Static Pressure problems!
This means that if it was a gas or oil fired furnace, the measurement would NOT include the AC coil. If a heat pump is being tested, the coil would be included.
Drill two holes large enough to insert the static pressure tip, one on the supply side and one on the return. Pressure measurements are then taken at each location. The measurement on the return side will be negative with a positive reading on the supply but you disregard the positive/negative and just add the two numbers together.
Once the ESP has been determined, look at the fan curve for that particular blower and determine the CFM from that chart.
If the air flow is not per manufacturers' recommendations, it is near impossible to get the refrigerant charge correct.
Read the static pressure in fractions of an inch on the gauge on the supply side and return side. Using separate pos+ neg- tests, Use a (+) sign before the positive or supply side reading to show where it was taken, and a (-) sign before the negative or return side reading. Disregard the positive and negative signs before the pressures and add them together to determine the total resistance the fan has to overcome. For example a +.40" and a -.30" equals a total static pressure reading of .70" I.W.C. Make needed changes to keep it within .50-IWC, if possible!
Record the pressure readings on a diagnostic report or on your service ticket. Write the pressures on the cooling coil plenum sticker for future reference and use. Any future changes in static pressure reveals a change in the system that should be addressed.
Keep the ESP, taken before air enters E-Coil (below .50"), and Total SP (taken after blower & before blower), as low as possible to keep the blower horsepower low and yet deliver the required CFM and E-Coil BTUH heat load. (Darrell)
(Air Conditioning and Refrigeration News, 1989). "Any assessment of refrigerant charge when the air flow rate is outside the stated 350 to 450-cfm/ton wet coil range-—will be invalid."
The air flow produced by an air handler is governed by the indoor unit's fan performance characteristics against the duct and other system components frictional air flow resistance. The blower fan curve is typically available as tabular data or graph. If test and balance data on duct air flow and external static pressure is available before the evaporator coil (positive) and before the blower (negative) in the return air stream it will give you the Total duct system (TSP) External Static Pressure.
Keeping the Total SP low is critical to the HP performance required and amp draw of the blower motor.
Add the negative to the positive unless your gauge does this for you by taking the two test readings at once.
The two curves can be plotted against each other to determine the system operating point and the corresponding cfm of achieved air flow. The operating point, where the fan curve intersects with the duct system pressure resistance curve. Since often there are three or more speeds available to the blowers, there are a corresponding number of operating points.
Field-Tests on 70 AC systems in Arizona discovered that "as found" EER was 40% lower than rated (Kuenzi, 1988). A field evaluation in North Carolina concluded that three of ten examined air conditioning systems had low evaporator air flow (Neal, 1987) and seven of the random sample had improper refrigerant charge.
A detailed study of 37 existing installations for Southern California Edison found an average evaporator flow rate of 300 cfm/ton; 80% of the systems were below the 350 cfm/ton level (Proctor et al., 1995). In this study, repairs were effected to increase flow. This involved opening or enlarging grills, replacing dirty filters, cleaning evaporator coil and increasing blower speed. The average post-repair flow rate increased to 322 cfm/ton. The study also noted that HVAC contractor's techs who had been previously called out to the homes had not identified or solved the actual problems with the systems!
Similar research was conducted by the same firm for utilities to examine new air conditioner installations. Testing of 37 new residential AC systems for Nevada Power Company in the Las Vegas area found an average flow of 345 cfm/ton (Blasnik et al., 1995a). Half of the units were below 350 cfm/ton and 30% were below 300 cfm/ton. Similarly, the average measured flow in 28 new installations tested for Arizona Public Service Company was 344 cfm/ton with more than half the unit below 350 cfm per ton (Blasnik et al., 1995b). Testing of ten new installations for Southern California Edison averaged only 319 cfm/ton, with all but one unit below the 350 cfm/ton action level (Blasnik et al., 1995c).
(Dry climates use up to 450-cfm per ton 'wet coil' of rated cooling! Darrell)
In nearly all cases, the new system had air handlers that were capable of delivering the necessary cfm. However, manometer measurements revealed high external static pressure of the duct system (averaging 0.584 in. wc. or 145 Pa). Undersized returns and filters were identified as the main culprits responsible for the low air flows.Average air flows for all three groups of 27 tested homes ranged from 130-CFM to 510-CFM Per Ton with an average of only 270 CFM/Ton. Air flow should be between 350 and 450-cfm per ton of cooling or per 12,000-btuh of cooling. 270-CFM/TON would cost you a fortune, along with shortened compressor life!
Several recurring factors were found to account for the inadequate flows:
Typical static pressure difference before the fan to after the coil in existing installations averaged 0.54 inches of water column (134-Pa). The Air Conditioning Contractor's Association of America (ACCA) duct design manual (Manual D) suggests that typical static pressure difference before the fan to after the coil should be approximately 0.40 inches of water column (100-Pa).
We also observed the impact of newer higher efficiency pleated filters. These filters often have a higher pressure drop than standard "disposable" models. In one case, flow was observed to drop by 25-cfm (4%) when substituting for a new conventional filter. In a second test at another home the change was a 50-cfm (5%) flow reduction from adding a higher efficiency filter.
Heavily soiled filters were observed in a project to add up to 0.23 IWC to system static pressure.
It was also shown that proper duct design aimed at reducing system static pressure has the potential to reduce fan energy by half and to improve overall system EER by 12%. Over sizing of AC system capacity should be avoided to provide adequate dehumidification through long duty cycle run-times.
Another reason to provide emphasis to low flow resistance duct and component design is the tendency in modern residential AC systems to add increased air filtration.
In existing installations, constrictions or restrictions which increase return or supply duct pressure drop should be addressed -- and fan speeds adjusted "according to measured air flow." In instances where existing air flow is deficient, installation of a larger indoor unit or modification to the duct system may be necessary.
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.
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" filter is 500-sq.in., this helps reduce the air velocity and pressure drop through the filtering media.
A Hart & Cooley seminar suggests using 0.08 for supply and 0.06 return. What do you use and why? It may be better to use 0.05 for both! You can always adjust to lower RPM which will reduce the needed HP to move the same CFM of air through the cooling coil.
Measure the actual Return Air ducting to determine if it equals or exceeds the Supply Air measurements. Oversizing the Return Air Filter Areas and a mild over sizing the (RA) duct area will help insure adequate airflow delivery through the cooling coil, reducing required blower motor rpm and horsepower.
blower curve charts with their units and also put them on the Internet
tech's to download and print. Also, air conditioning codes should be
in respect to proper sizing of the duct work in accordance with manual D, which
must include all the
pressure inducing factors when sizing the supply and return ducts.
Also, illustrate best furnace to evaporator coil transitions,
especially on oil furnaces!
The bottom of the evaporator coil must be
inches above this model oil furnace to achieve a proper airflow!
first use an air flow anemometer to check actual air delivery to the
then 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 and temperature and heat-load through the E-Coil to meet the
and temperature comfort zone. It is always very good practice
to measure the Total Static Pressure (TSP) on all systems
each side separate for comparison, then add-to + for the total ESP;
do this with a simple magnehelic gauge.
In any case, static pressures above 0.5"-IWC should be investigated and
reduced to specifications.
Get efficiency rating of the motors from the equipment manufacturer if possible. Condenser Power Consumption Heat Additive Formula: Using Companies Data | 1415-watts *X's 3.413 (normal power factor watt conversion to BTUH) = 4829-BTUH heat production. (Ask the equipment manufacturer for the power factor) 18,400 + 4829 BTUH plus 18400-btuh = 23229 Total BTUH divided / by condenser 1400-cfm = 16.6ºF or 17ºF condenser temperature rise/split. 90 degrees in, should give close to 107ºF condenser discharge air temp or, around a mere 216-psig head pressure. Always use an accurate amprobe to make sure you are not overloading the compressor or raising compressor discharge line temps above 225-F.
condenser air temperature rise/split is simply not acceptable. I want a
split at 78ºF
room temperature with the described unit above. Service techs, check
voltage at the unit and use your amp probe to see if the compressor is
within its service factor wattage rating.
Proper duct sizing and location is important. To achieve a proper evaporator heat absorption load levels with a floor level SA/RA system, depending on humidity load levels an increase to 475-CFM per ton of cooling capacity may be necessary.
Before you make all the recheck tests, it is very important that your condenser coil and evaporator coil and indoor blower wheel be squeaky clean.
the return air and the
supplier ducts are at the floor level, causing a recirculation of
discharge air back into the return air, resulting in a very low load on
evaporator coil. This is the reason that we do not get the BTU per hour
of the installed equipment. Additionally, when you oversize the
equipment, in relationship to the ductwork, you have a system that is
only giving you
a third to half of the BTU per hour rating of the equipment. - Darrell
I assume NO responsibility for the USE of any information I post on any of my Web pages, in E-Mails or News Groups.
All HVAC/R work should always be done by a licensed Contractor & properly licensed Techs! This information is only placed on these pages primarily for your understanding & communication with contractors & techs. This information is also for the edification of Contractors and Techs. Never attempt anything that you are NOT competent to do in a SAFE manner! I am NOT liable for your screw-ups, you are liable for what you do! - Darrell Udelhoven
Darrell's Refrigeration Heating and Air Conditioning - Federal Refrigerant Licensed - Retired Licensed Contractor
|Please write me if you have anything you'd like to contribute! - Darrell|
- Air Changes per Hour - AVERAGING INFILTRATION RATES - New