Correct Sizing of Residential HVAC Air Conditioning Systems & Ductwork Systems 

HVAC Procedures for Proper Ductwork Sizing for Residential Air Conditioning Systems - TEL ASP FR CFM
- with Darrell Udelhoven - HVAC RETIRED - udarrell

* Customers A Simple A/C Check you can do! Cynergy Home HVAC Energy Raters Listen SIMPLE BTUH AIRFLOW MATH | Formula for finding CFM Airflow
  System Evacuation Procedures | Ductwork Retrofitting Graphed Blower-Curve-Chart Measuring Low Airflow Air Infiltration! *Duct & Diffuser Sizing
Air Infiltration Sources Up-to 50% of Load|ROOM-DUCT CFM - Formula for finding CFM  | Use "Manual D"  |*TOTAL DUCT CHART |* DUCT Main SIZING CHART
Square Inches to Round Duct
DUCTWORK BASICS  Basic AC Overview - Specifications VS. Reality  | TEL - Solving for ASP Available Static Pressure | TXV or Piston Test | IWC_to_Pascals
| SA/RA locations | Gurgling@TXV | Sealing Rim Joists 
| Return Air - External Filter Rack Sizing  |*Sizing Units to Existing Duct Airflow |* Condensing Temp CT

The first procedure is always to doa manual J room by room heat-gain calc with the option shown so you can do everything possible to reduce all sources of heat-gain & heat-loss, greatly reducing both heat & cooling BTUH equipment sizing.

A reduction in equipment sizing will usually greatly improve the duct system performance.

Before you do anything else, educate yourself enough to "ensure that you request the proper things be done in the proper order of sequence." Checking ductwork & Airflow Checking Static Pressures  is Critically Important. As is knowing the operating feet per minute (FPM) velocity, the CFM & BTUH to each room along with the Total CFM airflow & BTUH.

Find the best way to save energy in your home, use your zip code. Profitabiity of Energy efficiency Upgrades | Hidden cost of Home energy Use

Then have a manual "J" heat-gain / heat-loss load calc done, followed by a manual "S" for equipment sizing & then a manual "D" for proper ductwork sizing.  All registers, grilles, filter racks, & diffusers must be located & sized for optimal efficient performance of the system. Order Turning Vanes for long radius 90-ells; use long radius 45's.
The correct sizing of residential air conditioning systems & ductwork is crucial to ensuring proper indoor space conditioning, equipment performance, and economical operation. Unfortunately, many A/C contractors measure “correct equipment sizing” by the system's ability to meet any indoor thermostat setting at any outdoor temperature. That method of sizing will lead to inefficient operation for the vast majority of conditioning time. Air conditioning systems “must be sized to meet typical or average indoor and outdoor conditions to ensure proper air mixing, filtration and dehumidification of indoor air across seasonal variations.”
A Major Oil Furnace Airflow Problem Fix
Regal & Hallmark & nearly all Oil Furnaces - Installation manuals
Download the installation & service manuals from ABOVE LINK BELOW LINK MAY NOT WORK!
To find the information below; Use within the pdf search:  at least 6” above
Or use down arrow to P-8 & scroll down a-ways...

"If the oil furnace is used in connection with summer air conditioning the evaporator coil must be installed at least 6” above the oil furnace for proper airflow. Distances less than 6” will result in decreased airflow."
Make sure outlet supply takeoffs are NOT blocked by the coil. In all cases, refer to the manufacturers’ data for static pressure losses to ensure the total system static pressure does not exceed 0
.5” WC.

Around 2005, my brother had a 12-SEER 1.5-Ton Heil condenser installed with a 2-Ton evaporator with a TXV metering device, it was erroneously installed directly on top of the thermo Pride OL11 Oil furnace.

I checked the airflow with a low cost anemometer, it was below 300-cfm, I also checked the subcooling & it was only 1-F, indicating low on refrigerant.
My electronic leak detector says there is a leak in the evaporator area. With my new digital pocket thermometer I checked the supply & Return wet bulb temps & found the enthalpy change on my chart to get the extremely low BTUH performance of his A/C unit.

The required  FIX: Pump much of the R-22 down into the condenser, it's a Scroll compressor, then close service valves & recover the R-22 from the lines & E-Coil, remove coil & repair leak, reinstall at least 6" above the furnace preferably on a transition rather than rails. Purge with nitrogen then flow dry nitrogen at under 3-PSIG, then evacuate to 500-microns & see if it holds under 1000-microns.

Replace quarter HP belt drive blower motor with a third HP motor. Check airflow & adjust RPMs, in his situation, until 700-CFM is achieved.

Open service valves & charge until subcooling reaches around 12-F, look up target subcooling for that unit. Also, check superheat to see if TXV is holding within its requisite perimeters. Then recheck its BTUH performance output compared to company charts at those conditions.
Download these Energy Saving PDF Graphics It will provide ways to cut your monthly Energy Bills, hopefully in HALF. Thanks for the link - TEDKIDD

Get a low cost Testo Tester & ballpark figure actual BTUH & EER - the information on it:

Everyone, very low cost anemometer to get airflow FPM Velocities, get it:

Also, get a low-cost digital flat-headed pocket Thermometer to use flat on the piping; these test instruments will PAY big returns!
This should be helpful: (Edited 4/5/10)
Duct system CFM X* 4.5 @sea-level, or use X* 4.25 if 1000' above sea-level, X* change in enthalpy = BTUH (Ballpark) Operating Performance.
"U Must Right Click Link & open in New Tab," look-up wet bulb enthalpy figures on chart," & figure enthalpy change.
Wet Bulb Enthalpy Chart

Rules of Thumb for Duct Systems  - Hart&Cooley

Look at the ducting, if it is not to code; make hard copies of this code & give it to whoever does the ducting work
Make sure they redo it right!
Never have flex duct interiors commercially cleaned, I just viewed Home Inspection photos showing the interior damaged & insulation plugging the duct.
Home Inspectors warn people because the duct cleaner's tell them it won't damage the ducts. Some HI's look into the boot areas for clues of problems...

Rules of Thumb for Duct Systems  - Hart&Cooley

Identifying your registers/diffusers & their (Ak) sq.ft. area, so you can multiply the FPM Velocity times the Ak to get the (CFM) Cubic Feet per Minute airflow from that register.
Have or do a manual J heat-gain calc for each room. If a room calls for 3,000-BTUH; first divide 12,000-BTUH by the CFM PER TON you want to use.

I.E., Wet coil, 12,000/400=30-BTUH per each CFM; Wet coil 12,000/425=28.235294; 3000/28.235-= 106.25-CFM;  Dry coil, 12,000/450= 26.6666-BTUH; 3500 / 26.6666= 131.25-CFM
If register/diffuser has the same (Ak) free-airflow-area, as the duct run!

Room calls for 3,500-BTUH, using 450-cfm per/ton dry coil or 26.6666-BTUH per CFM= 131.25-CFM.
I.E., 6" rd duct .6*6=36*.7854=28.2744sq.ins/144=0.19635-sq.ft.; 131.25-cfm / 0.19635-sq.ft= or 668.4-fpm velocity.

  •  SP2 = (rpm2/rpm 1)2 X SP1 = SP2

Required fan motor horsepower (hp) varies as to the cube of the rpm speed:

  • hp2 = (rpm2/rpm1)3 x hp1 = hp2

CFM Fan delivery varies directly as to the fan RPM speed:

  •    cfm2 = (rpm2/rpm1) X cfm1 = cfm2

Duct retro-work can solve the problem, increasing blower HP alone won't usually work! A few calculations
plus retro-work and presto, a matched airflow with your systems' heat absorbing coil capacities, delivering near its BTUH, EER, and SEER ratings at normal room temperature settings! (Most don't)
Click on the categories to see the diffusers & Return-Air Grilles then find them on your downloaded pdf's engineering data.
Hart & Cooley:
Do a lot of Hart & Cooley engineering data searches, look at the registers & the Ak sq.ft. data to figure register's delivered CFM.
LOW AIRFLOW - this will help to open your eyes!
WARNING: on units with a Thermostatic Expansion Valve (TXV), you cannot use the suction pressure to check the charge; many appear to be doing this; it tells you nothing. Only after you have verified that all the coils are clean & the airflow is right-on, can you begin to check the system's charge using Subcooling method with a Superheat check. Always check the actual airflow CFM before checking the charge, get it Right!

* There is a TXV system that has very low airflow, actually less than 200-cfm per-ton of cooling, they're only checking the suction pressure & saying the charge & everything is okay! * That system has a TXV & shows; 98-F condenser saturation temp & 97-F liquid line temp near E-Coil, a mere 1-F Subcooling, it's undercharged even with a mere 200-cfm per-ton cooling load! Unbelievable, but it's happening out there...
Use my Superheat Subcooling Charging page!

Air Conditioning Performance Diagnosis using listed (CT) Condensing Temperatures
Using Goodman 16-SEER "Expanded Performance Data"

  Find the correct (CT) Condensing Temperature with the following known mfg’ers data.

Outdoor Ambient Temp (OAT) 85-F; IDB 75-F; IWB 63-F or 50%-RH.
Listed pressure is 316-psig, or 99-F CT; that is 99-F -85-F is a 14-F SPLIT.

The delta T or temp-split should be within a 10-psig range or, +/- 2-F degrees; 97 or 101-F.

The mfg’ers Supply Outlet should be able to provide Contractors & Techs with those performance data charts. Goodman has their “Expanded Performance Data” on the Internet. BTUH = CFM X enthalpy difference from Chart X 4.5

Size for cooling; size the equipment and system based on 100% of the Total Cooling Load (both sensible and latent loads) at actual outdoor design conditions. Size the duct system properly & make certain that the proper airflow & optimal heatload is passing through the evaporator coil during most of the operating runtime. To Optimize Payback and lower operating costs always do as many things as you can to Reduce the heatgain-heatloss "BEFORE doing the manual J load calc & manual S for sizing the equipment."

To select the “proper size” heating, air conditioning, and duct system for a home, seven factors must be considered and all changes made prior to sizing equipment:

1. Improving Insulation Values and Reducing Infiltration, including e-windows, doors, etc.

2. Reducing Air leakage - air leakage accounts for between 25 percent and 50 percent of the energy used for heating and cooling in a typical older residence.

3. Solar orientation - and ways to reduce the radiant heatload should be considered.

4. The Internal BTU heat generation of appliances and people must be added.

5. Design conditions (typical outdoor and indoor weather conditions, humidity levels, etc.) A scientific calculation (manual J) called a Heat Loss/Heat Gain Calc, tabulates these factors into a load scenario for heating and cooling based on summer & winter outdoor design conditions for the climate where the home is located.

By comparing the heat loss/gain to manufacturer's equipment performance data, a properly sized heating and cooling system is selected. Use indoor design of 75-F dry bulb and 63-F wet bulb, around 50% Relative Humidity.

6. Proper & thorough ductwork testing and design Is Extremely Important for efficiency & BTUH performance - The evaporator coil needs to have an optimal heatload passing through it most of the time in order to approach achieving its Rated BTUH Capacities & its EER & SEER Ratings.

7. Study the diffuser data in respect to room CFM, the required throw & a diffuser Face Velocity of around 600-fpm, & a Terminal Velocity at the human occupant level area of 50 to 75-fpm. This is critical toward achieving an optimal human comfort zone.

What should you expect from the average heating and cooling contractor?

When a typical heating and cooling contractor quotes the efficiency of the equipment (SEER or AFUE) and leads you to believe the new equipment will automatically deliver that efficiency, think again. Typical installed equipment only operates at 55% to 70% of rated capacity.

It is important to understand that equipment ratings are only "the potential efficiency of that component of the system under perfect conditions." "Over half of the system's efficiency depends on the duct system and the field-installation." Check for Return Air drawing Hot Air from attic areas, etc.!

An A/C's system efficiency can often be increased by a skilled tech from 10% to 50%. The biggest benefit is the increase in comfort and lower utility bills!

The heat loss/heat gain calculates the amount of heat transfer by component, based on the surface area, and then tabulates the total transfer for all of the components.

Air leakage > air infiltration: Up to 50% of an average home's heat loss/heat gain is attributable to air leakage/air infiltration. Therefore, determining the proper leakage/air infiltration rate for a specific home is paramount. Design leakage/air infiltration rates are based on dwelling size and projected efficiency or actual measured performance.

Solar orientation, or the amount of window surface area and the direction a home faces can have tremendous impact on the cooling needs (heat gain). Similar houses with different solar orientation will have different cooling loads. Glass facing east/west has more heat gain than glass facing south or North.

The service techs should use a manual D to properly size the ductwork supply air mains and runs to outlet diffusers, as well as the “sizing of the return air ducting in relationship to the BTUH/CFM requirements of the various rooms.”

Qualified personnel must perform all installations and services. The duct system sizing and load sizing calculation should follow the design standards of Air Conditioning Contractors of America (ACCA) - Manuals D & J -or the American Society of Heating, Refrigeration & Air Conditioning Engineers, Inc. (ASHRAE) Fundamentals Volume (latest edition).
When sizing ducts, the use of one value throughout will assure incorrect duct size for many branch runout duct segments. If 0.05 is used, nearly half of the runouts of the system will be oversized, resulting in those zones being too cold in summer, resulting in the remaining runouts furthermost from the air handler being undersized, creating zones that are too warm in summer.

The result is a human comfort design failure. The modified equal friction method of Manual D, "requires that the available static pressure from the fan be 'consumed' by the duct through its run from fan to outlet/inlet, with no shortage or excess at the end."

Pressure drop per 100 ft is not an input, it is a calculated intermediate value. A contractor who knows how to do the calculations to determine the available static pressure, and correctly allocate it to the supply runouts and return, is a rare tech.

 This page does not explain everything you need to know about proper duct sizing a system for optimal comfort, but provides some general guidelines and concepts.
*You could ballpark the CFM using the static test & the air handler's graph. You could measure the CFM delivered to each room with a hood Alnor Balometer, it's usually the best instrument to use, but not cheap. Measuring the air velocity from diffuser's is a bit tricky because you should use the diffuser mfg'ers data which you should always have with you.

You can usually get the diffuser mfg'ers data, say its a 1.5-Ton system that already has 6"rd branch duct runs, to achieve enough CFM airflow, you need 700-FPM velocity in the ducts.
I would want to use a diffuser with a little more free open sq.ft area than the 6" duct  which area is 0.19635-sq.ft., say middle of the room in the ceiling;
Hart & Cooley 2-way curved blade 12x6 has Ak .235-sq.ft. Free-Air-Area delivering 140-CFM at 600-FPM - diffuser face velocity.

This would help lower the velocity of the duct
through the diffuser & reduce air noise.
Throw is 7.5-ft toward each wall. Terminal velocity at the occupant level is 75-FPM.

One duct yields 140CFM times 5 outlet runs yields 687-CFM *X's 30-BTUH per CFM = 20,616-BTUH, about right.
Whatever CFM you need for that room or area, divide the Sq.Ft.
free-air-area into the requires CFM to get the FPM velocity.

Say we're using 450-cfm/ton of airflow; 12,000-BTUH / 450= 26.6666-BTUH per CFM. You need 3600-BTUH for that room, 3600 / 26.666 that's 135-CFM / by .235-sq.ft. diffuser area is  only 574-fpm face velocity. 
Using 400-cfm per/ton / 12000-BTUH is 30-BTUH per CFM.

Taking the manifold gage head pressure & gage condensing temperature is very important data. Coupled with a condenser air discharge temp-reading, if the condenser gage pres/temp is too high compared to the TH reading, there may be non-condensibles in the system.

Also, there is a legitimate formula I use to determine the operating BTUH it is delivering at all the data taken. All the mfg'ers ought to list the condenser temp-split (it varies with EER & SEER) just like they list the indoor split, it is valuable trouble shooting info.

You can also use the condenser temp-split (it contains both Latent & sensible heat) combined with the indoor data to plot the indoor CFM. I was never good at math, but those equations have to balance, & they do work!
Quick Check for Sizing Units to enough Airflow
Actually, even on service calls where there are cooling problems the ductwork should have a quick Manual D performed.

Then take the ESP static pressure & compare to blower graph or chart, also take the FPM duct/diffuser velocity.

Have or do a manual J heat-gain calc for each room. If a room calls for 3,000-BTUH; first divide 12,000-BTUH by the CFM PER TON you want to use.
I.E., Wet coil, 12,000/400=30-BTUH per each CFM; Wet coil 12,000/425=28.235294; 3000/28.235-= 106.25-CFM;  Dry coil, 12,000/450= 26.6666-BTUH; 3500 / 26.6666= 131.25-CFM
I.E., 6" rd duct .6*6=36*.7854=
28.2744sq.ins/144=0.19635-sq.ft.;using 450-cfm per/ton dry coil: 131.25-cfm / 0.19635-sq.ft= or 668.4-fpm velocity.

Then do a quick estimate of airflow per equipment tonnage.

To find area of a round duct; Duct diam is 7"; 7"X7"= 49-sq.ins., X's .7854 = 38.04845-sq.ins divided/ by 144= 0.2672541-sq.ft. area X's FPM Velocity 600-FPM = 160.35246-CFM X30 = 4,810.5738 each 7" run X's 6 branch runs = 28,863-BTUH, or airflow for 2.4-ton.
(12,000-BTUH /400-cfm per-ton = 30-BTU per cfm ratio | / 450 = 26.666-BTUH per-cfm)

That would also be good for 2-ton; at 550-FPM velocity X's 0.2672541= 147-CFM X 30 = 4,410-BTUH each run X 6-runs = airflow for 26,460-BTUH.

*Never sell units requiring more airflow than the duct system will support! - Darrell udarrell

An affordable test instrument you need!
All I had was the Sling Psychrometer & spinning it was a bit time consuming, but I used it religiously, it is information you need. 

The Testo 605-H2 Humidity Stick (wet bulb), displays relative humidity, air temperature and wet bulb temperature.

It is very affordable & because of its potential to help deliver tons of other data everyone should have one!

For more information on it:

The other test data you need is the system's CFM airflow through the evaporator coil, then with software I have you can peg the BTUH the operating unit is delivering under those conditions.
Add to that a low cost Magnehelic gauge to read static pressures to compare with mfg'ers blower performance charts; plus a velocity meter & you have a ballparked CFM to plug into for the BTUH.

We could easily provide a detailed psychrometric print out of exactly what the operating system is delivering including condensate lbs/hr, & actual sensible & latent cooling BTUH & Ratio, every data detail imaginable.

Think about what that would mean to you & those you serve. - Darrell

Determining which metering device the system has without physically looking

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.

Supply Air (SA) and Return Air (RA) -- near the ceiling or near the floor
For cooling both SA and RA work better near an 8 foot high ceiling. The SA diffusers should distribute the air near the ceiling to the walls on all sides of each of the rooms.
Whether a single return system, or a multiple return system is used, there must be a low resistance path between every room and the nearest return air opening. This can be done by using wall transfer grilles, door grilles, or jump ducts.

To quickly aid you in evaluating existing duct systems, review the chart below. To insure the necessary air handling capacity of a duct system, each of the system's components (trunk lines, takeoffs, runs, diffusers, registers, and grill free areas) must be properly sized and matched together. A 12x8 duct with a 350-CFM capacity, for example, WILL NOT flow 350-CFM if the boot & diffuser(s) combo can only flow a total of 200-CFM within specs.

This Hart & Cooley pdf might help you select the right diffuser for the particular application:

The boot size & register size is determined by the throw, spread, drop, Terminal Velocity of 50 to 75-fpm at the occupant level, & noise class according to the CFM that must be delivered to the room. Those specifications are located in the technical engineering pages that the diffuser mfg'er has available to the dealer distributors. 

When sizing the return air duct system, the air handling capacity MUST BE EQUAL TO the SUPPLY SYSTEM at a minimum, I would oversize the return ductwork. It is recommended that contractors follow design parameters established by ACCA or ASHRAE on the return air duct systems.

If at all possible, use insulated metal ductwork.  For Heat pumps, Figure 450-
CFM per BTUH ton CFM | Round Duct | Square Inches | Rectangular Ducts | SA use 0.06" Friction Rate (FR) per 100 foot, 0.05" FR for Return  Air.


Best Method:
CFM formula per room:
Use the load-calc for that room divided by the total system BTUH for cooling. Say that room's heat-gain load-calc calls for 3500-BTUH / 24000-BTUH = 0.1458333 * 800-CFM = 116.6-CFM to that room, a 6" metal duct should be around 600-fpm velocity. (24,000-btuh / Ave of 3500-btuh = 7 Runs.)

I.E.,  a 4-Ton System: 48,000-btuh @400-cfm per-ton= 1600-cfm | Load calc says room requires 3600-btuh / 48000=  0.075 * 1600-cfm = 120-cfm, a 6" RD metal duct, close to 611-fpm velocity.

You must know & record the operating feet per minute (FPM) velocity & the CFM to each room & the Total CFM airflow! Every tech should be using an anemometer to check operating airflow velocities in FPM & then figure the CFM airflow's.

Another quick method: The heat gain and Btu/hr of cooling is done for each room.
At 400-CFM per/ton of cooling, 12,000-BTUH / 400-cfm = 30-BTUH for each (CFM) Cubic Foot per Minute of Airflow.
At 450-CFM per/ton cooling use 12,000 / 450-cfm = 26.66-BTUH per each CFM, etc.

Then Select Supply duct size by CFM, velocity, & optimal Supply Air *(FR) Friction Rate.

A Room requires 3000-Btu/hr divided / by 30 equals 100-CFM, or around a 6" dia. RD metal duct.
You need-> Five duct runs for 1.5-ton unit, 18,000-BTUH: (Equal room loads,ha!) 18000 / 5-runs= 3600-BTUH / 30 = 120-CFM each 6" duct velocity 611-fpm Velocity
18' branch runs 611-fpm velocity at a Friction Rate 0.03" per 18'.

For Room Return Air balancing, i.e., -.01" IWC = approximately -2.48 Pascals, which is a more precise easier incremental scale to read.  One inch water column (IWC) equals, rounded to > 250 pascals; 0.5" IWC is about 125 pascals; 0.25" WC = 62.5 pascals; 0.125 = 31.25 pa.; 1 / 250 pascals =0.0040322 *X's  -2.50 pascals = -0.01003657696655" IWC or make it  - 0.01" IWC for low Return Air room pressure differentials; - use pascals.

Formula for finding CFM Airflow
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. X's 0.00694 = 1.39494-sq.ft. X's Velocity of 800-fpm = 1116-CFM
Times 1000-FPM = 1395-CFM. Branch ducts: 7" Rd duct 38.48-sq. ins. X's 0.00694 = 0.2670512-sq.ft. X's 500-fpm=133.5-CFM
Example, times a velocity of 600-FPM X's 0.00694 = 160-CFM, the velocity is a big CFM & BTUH number changer for rooms.

It is better to use a ductulator to enable the use of the appropriate Velocity Friction Rate balances to achieve the correct CFM on each Branch Run, etc.

This chart is in accordance with the tonnage of the unit.
Residential metal Duct Design varies from as low as .02" on Returns, to -
Supply Trunk runs down to 0.05" Friction Rate pressure drop per 100 feet of duct.

Refrain from using Flex duct, if unavoidable, use .03 to .05 pressure drop or lower per 100 feet of duct run.
A 3.5-Ton main run at 450-CFM Per Cooling Ton would require an 20" rd main trunk run, which will also work for heat pumps.

1575-CFM | 20" rd duct | Gross sizing of Return Air (RA) filter grilles: 200 sq. ins. Per Ton of cooling.

A 5-ton system should have two 500 sq. in. Return Air Filter Grilles. (Or,  (two) 25" x 20" Return Air filter grilles for 5-ton.)
This is to try to reduce the air velocity through a clean filter to 300-FPM and to reduce the resultant pressure drops as the filter loads. The free air area of a filter should be stated on the edge of the filter by the mfg'er.

It is very important to size each duct to the CFM and velocity needed for the room served in accordance with the manual J load and manual D duct design. Dampers’ on too large a duct, if dampered too much, would result in too much velocity loss.

Once you know the manual J loads, number of duct-runs and the Velocities, Friction Rates, & CFMs for each room, you can use the register/diffuser data to get required throws, etc. (i.e., Hart Cooley from your supplier).

Example: One-Way - Adjustable 10x6 diffuser; select a duct size that provides 80-CFM @ 600-fpm, which will provide an 8.5-foot throw. Hart & Cooley lists a Pressure Loss through the diffuser at .022” of an inch.

Consider the register/diffusers you are going to use and the various CFMs’ & Throws’ you need according to your design layout. – Darrell 

DUCT SIZING CHARTS Residential Main Trunk Runs for Approximately 450-CFM Per Ton - USE Manual D!
Residential Supply Air - used 800-FPM as MAX for Main Runs - SA: Supply Air; RA: Return Air

1.5 ton 700-CFM
13"metal Main133-sq."SA 760-fpm FR .07 | RA 16" 201" 500-CFM  FR 0.01" (FR. 25' run)
1.5 ton 750-CFM
14" Branch 154-sq."  702-fpm FR .05 | RA  16" 537-CFM per ton SA/RA floor level
2-ton 900-CFM
14-SA Rd 154-sq." 854-fpm | 18"-RA 254 @ 510-fpm
2.5-ton 1125-cfm
16" metal  Rd 201-sq"  806-fpm | RA 20" RA 516-fpm 
3-ton 1350-cfm
18" metal 254-sq" 764-fpm | RA 22"  512-FPM
3.5-ton 1575-cfm
20"metal 314-sq"  | RA 22" Rd 314-sq" 531-FPM 
4-ton 1800-cfm
20"-SA 314-sq" 825-fpm | RA 24" Rd 380-sq" 573-fpm
5-ton 2000-cfm
22" SA 380-sq" 758-fpm  | RA 24" Rd 452-sq" 637-fpm

Tonnage ChartsFormula for finding CFM Airflow from Velocity in FPM above for Main Run Sizing, check chart below
If you can measure the air velocity coming from a known size duct or open area of a SA register, here is a rough ballpark formula to get the CFM:

450-cfm per ton | Maximum 900-fpm Trunk Supply Air (SA)
Runouts 500 to 600-fpm, 500-fpm or less on Returns |
Return Air 350 to 600-fpm Lower is better - use large ducts
The filter rack area sized for free-air-area to achieve an initial 300-FPM velocity through the filter
Friction Rates will be very low!
Main Trunks 0.03 to 0.1" Supply Air (SA) Runs.
Figuring duct size - cfm from required BTUH of each room. Using 450-cfm per 12,000-BTUH (one ton) 12,000 /  450 = 26.66-BTUH per CFM

A room  requiring 4,000-BTUH / 26.66 = 150-cfm | Chart 150-cfm = 7" round metal duct or 38.48-Sq.In. Square duct. Using Metal Duct!
I would use a 7" duct, or a duct. Using 425-cfm per ton 12,000-BTUH 4,000-BTUH is a third of 12,000-BTUH.
Therefore, 4,000-BTUH room requirement, using 450-cfm per ton (airflow 1/3 of a  ton) X's .3333 =
150-cfm or  7" duct | 25' with (2) 90's, Vel. 562-FPM  Friction Rate (FR)  0.04"

CFM = (velocity in (FPM) Feet per Minute times the Square Footage of the duct area). To convert sq.ins. multiply by 0.00694 for sq.ft., or divide sq.ins. by 144.

Converting square duct
inches to round duct size, Figuring the Square Inches of Round Ducts, an 8" x 8" duct = 64-sq.ins. x .7854 = 50.26 sq. ins. You round off to 50 sq. ins. for an 8" duct.  Or, simply getting the square inches of round ducts: a 7" duct; 7" x 7" = 49 x .7854 = 38.48-sq.ins. or divide / by 144 = .2672222-sq.ft. X's a velocity of 500-fpm = 133.6-cubic feet per minute delivered to the room; 133.6-cfm x 30 = 4,008-BTUH.

Sized for in the chart below - BTU/hr per CFM figures "are figured for heatpumps at 450-CFM per ton of cooling."
Use 800 to 900-FPM MAINS' VEL. Use an optimum of 500-FPM VEL for Supply Branch Runs | Air speed Face of Return. 

Air Filter Rack Sizing
ACCA Manual D specifies on Return-Air Filter grilles a maximum of 300-FPM velocity:
Recommended Main Supply-Air velocities for rigid duct should be 700-fpm & a Max of 900-fpm, for flex duct 600-fpm.
Recommended Branch Supply-Air velocities rigid & flex 600-fpm.

Recommended Return-Air main duct velocities should be 600-fpm or less, on Branch Return Air ducts try for
550-fpm or less velocities.
Keep air velocities through the RA Filter(s) as low as possible.

Andythedrew 10/08/09 | Quote:
Is a 24x30 filter grille large enough for a 3.5 ton system?
Yes & No...
A filter grille is to be sized not to exceed Manual D's 300-fpm maximum velocity.
Use the tonnage's CFM /divided by filter grille's free-air-area = FPM velocity.

You really can't have too much filter grille area; more is always better because they are always loading.

It all depends on the type of filter! Typical disposable 1-inch capacity return air filter is 2 cfm per square inch of gross filter area.
Recommended actual initial filter velocity is 300-fpm, lower is better. 

As the Filter loads Velocities higher than 500 fpm will decrease filter performance; Increase flow resistance, and possibly blow off collections of debris.
Measure Velocity 1” from RA grille face.

Average Free Air area of most Return Air grilles about 75%.
75% of 720-sq.ins., is 540-sq.ins. of free air area.
So, what is the filter's free-air-area?

The filter mfg'ers ought to be required to list the free-air-area of their filters, & then state on the packaging or on the filter, the limit of CFM to stay within 300-fpm velocity through their clean filter.

To achieve 300-fpm through a clean filter, you would need 700-sq.ins of free-air-area / 144 = 4.886111-sq.ft., | 3.5-Ton @1400-cfm /  4.86111= 288-fpm velocity through a clean filter.

Then how much free-air-area does that filter design have?
That depends on the type of filter(s) used.

Most 3.5 to 5-Ton systems require two filter racks exterior of the airhandler to achieve a low enough air velocity through those filters. Most 4 & 5-Ton systems nearly always have too much air velocity through initial clean filters, let alone when they begin loading.

Average Free Air area of most Return Air grilles about 75%.

Most 4 & 5-Ton systems require two filter racks
exterior of the airhandler to achieve a low enough air velocity through those filters. Most 4 & 5-Ton systems nearly always have too much air velocity through initial clean filters, let alone when they begin loading.

For a heat pump system, I would go for 4-ton sizing on the Return-Air Filter Grilles for your 3.5-Ton condenser & 4-Ton indoor cool.
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., X's .75% for the filter = 703-sq.ins., / 144 is 4.88-sq.ft., 
free-air-area; then 1600-cfm / 4.88-sq.ft. is 328-fpm velocity with a clean filter in the rack. 

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.
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 /  .3490666 = Velocity of 501-FPM, you can use a ductulator to get the actual Friction Rates (FR). See using the diffuser's Ak, below.

Formula for finding CFM Airflow and/or Velocity in FPM & BTUH

If you can accurately measure the air velocity coming from a duct, here is a rough ballpark formula to get the CFM:
 (velocity in (FPM) Feet per Minute times the free-air-area (Ak) square footage of the supply-air diffuser.)
I.E., diffuser Ak is .225-sq.ft., times say 600-FPM velocity = 135-CFM
Also, (12,000-BTUH /400-cfm per-ton ratio = 30-BTU per cfm ratio)


When sizing ducts, the use of one Fiction Rate value throughout will usually guarantee incorrect duct size, velocities & CFMs for some duct run segments. If Supply Side is 0.05" per 100 ft of duct run, or less is used, some runouts of the system may be oversized, creating zones that are too cold in summer. Usually the furthest from the air handler will be undersized creating zones that are too warm in summer. Install dampers in all Supply Air ducts for some balancing.

That is an incorrect design. The modified equal friction method of Manual D requires that the Available Static Pressure (ASP) from the fan be "consumed" by the duct through its run from fan to outlet/inlet, with no shortage or excess at the end. Also, pressure drop per 100 ft is not an input--it is a calculated intermediate value.

A contractor or Tech who knows how to do the proper calculations to determine the available static pressure - ASP and correctly allocate it to supply and return, is a very rare Tech indeed. First, select the CFM & velocities you want, that may result in small Friction Rate variations on the various Branch Runs & run lengths.

Always use "Manual D" for proper FRs & duct sizing for required CFMs to each room.
I used figures & formulas below for the table, & wanted 450-Btuh wet coil per ton airflow

Converting square duct inches to round duct size, i.e., an 8" x 8" duct = 64  -sq.ins. x .7854 = 50.26 sq. ins. You round off to 50 sq. ins. for an 8" duct.
Round duct diameter to Sq. Ins., duct diameter 6"x6" = 36 X's .7854 =
28.27-sq. ins.
CFM = (velocity in (FPM) Feet per Minute times the square footage of the duct area)
Quick method: The heat gain and required Btu/hr of cooling is done for each room.
Cubic Foot per Minute (CFM) of Airflow times 30-Btu/hr will ballpark the Btu/hr delivered to each room.
Duct Sizing Chart Residential some Branch Runs some Main Runs - for Approximately 425 to 450-CFM Wet Coil Per Ton on Main Runs
Residential Supply Air - used 800-FPM as MAX for Main Runs - 600-FPM MAX for Branch Runs - SA: Supply Air; RA: Return Air
Rd. DIA. | Sq.Ins.
Vel / FPM

4" Sm. Rm

5" Bd. Rm





7" Br

8" Br

10x6 8x7

9" Br


10" Br
14x6 12x8

12" SA-Main
RA 14"
113 RA-154 RA 430 RA 12X14

14"Metal SA RA-16"
201 RA
RA 502
RA 14x12


14" Metal SA RA-16" 201 RA
RA 645-fpm
RA 14x14

32000 1125
16"SA/RA 20"314
201SA RA 20"/516 RA 10x18
20x16 24X14

18"SA | RA 22" 380"
254 SA
RA 22" /512 24X16

1400 20" SA/RA 22"
RA 597
SA 642-fpm
18x22 16x24

22" SA/RA 24"

2000 24" SA 452-sq.ins
RA 24"
452SA/ RA
RA 14x32

Solving for Available Static Pressure (ASP) - When Designing or Redesigning Duct Systems, Finding TEL, FR:
Total Equivalent Length (TEL)  Find the Total Return length, then find the longest Supply Equivalent Length (EL) by finding the longest measuring duct length, number of EL in the turning elbows, trunk take-offs, boots, etc.

Once you have all the correct Device Pressure Losses (DPLs) on the longest Supply Air run, evaporator coil, diffuser, etc.
Use the manufacturer's nameplate pressure (IWC) or .5" ESP and subtract all airstream device pressure losses in the longest TEL duct run (supply diffuser, damper, wet coil, etc. all available in manual D) from that given value. That will leave the "Available Static Pressure" - ASP for duct & blower design purposes.

You can figure the Total Equivalent Length (TEL) by using the Manual D length additives for the various fittings, then use your Duct Designer ductulator to properly size the duct system to meet the required CFM, velocities, FR as required in respect to the blower's Nameplate ESP, normally, 0.5" and its performance graph in relationship the remaining (ASP) Available Static Pressure.

There are charts available to determine what the total pressure drop will be then when you figure the "Total Equivalent Length" run of the longest Supply duct runs;
ALL lengths of duct and ALL fittings trunk duct take-offs, etc., to get the Effective Length (EL) additive to that duct run for the 'Total Effective or (TEL) Total Equivalent Length.'

Subtract the total Device Pressure Losses (DPL) from the AH equipment's "Available Static Pressure," usually .5” ESP because most furnaces are designed for .5" ESP to achieve the desired high speed CFM, if needed, you can always use a lower blower speed.

Then get a total equivalent length of your ductwork most ductulators have this on the back of them.  Friction Rate = Available Static Pressure times X’s 100; divided by the TEL, that is the friction rate per 100 ft of SA &/or RA ducting.

TEL ASP FR Chart Graph  Loads slow using dailup - Save both the pdf to a quick access PC folder for review

Designing or Redesigning Duct Systems Chart  Print

Variable Speed Motors and Static Pressure

It will deliver the required CFM with some blowers, but at a higher operating cost.
If there are any ductwork air leaks, it will throw everything off.

Nearly all duct systems have a percentage of air leaks.
Check for & minimize air leaks.
Blower wheels, evaporator coils, etc., everything has to be clean.

Get copies of the ACCA Worksheet; below is my math involved with the linked Graph:
On older furnaces, the .45-Device Pressure Loss (DPL) should be subtracted from a furnace Nameplate Max ESP .50-ESP, leaving only .05-ASP (Available Static Pressure).
According to the graph, (using a .45-DPL), ASP only .05 X 100 = 5 / by 330 = .015 Friction Rate, way too low & off the graph resulting in Inadequate Airflow!

It would need a much shorter TEL or less DPL's to meet design functionality!
Air Turning vanes in 90-ELLs greatly reduce the TEL.
Other changes could also help reduce the TEL.

First find 330-TEL graph, then look at the ASP at the bottom the the FR shown.
Also, a .06-FR is the lowest Friction Rate shown on that graph; though many use .05-FR on the Return Side.
Once the ESP has been determined, look at the fan curve for that particular blower and determine the CFM from that chart.

Looking at the product data chart, what temp-rise does the mfg'er recommend?
Measure the temp rise & see what you get.

Unless the proper CFM heatload goes through the evaporator coil it is nearly impossible to achieve an accurate & proper refrigerant charge, and BTU/HR along with efficiency will be way below Ratings! 
Take the static pressure measurements on both the Return and Supply Plenums of the furnace (with filter(s) in place.  

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.

When using two tubes (neg. & Pos.) on a modern gauge you will read what the gauge indicates!

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.

7 EER or less
10 EER
12 EER
13 EER
'Max' condenser air temp 'delta-T'
ave. less
Max temp drop 'across' E-Coil
ave. more
ave. more
'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 could become more difficult at the highest EER levels. The EER & SEER levels widen, as SEER sky rockets.

Optimizing the "Evaporator HeatLoad" will Optimize the Condenser BTUH HeatLoad Output from your Home

Most evaporator coils are under-loaded when operating at the normal room temp setting!

The airflow should be adjusted to fully load the evaporator coil at the normal room temperature setting! This airflow adjustment will optimize your air conditioner's BTUH and SEER performance. Most air conditioner's have an underloaded evaporator coil at the room temp thermostat setting, where the vast majority of its run time will take place! In 8 foot ceilings, Return Air (RA) should always come from the warmer ceiling air areas.

On TXV metered systems the Subcooling should be within +/- 2-F of the mfg’ers installation instructions.

Air Conditioning Performance Diagnosis using listed (CT) Condensing Temperatures

Using Goodman 16-SEER "Expanded Performance Data"

What is the correct (CT) Condensing Temperature with the following known mfg’ers data?

Outdoor Ambient Temp (OAT) 85-F; IDB 75-F; IWB 63-F or 50%-RH.
Listed pressure is 316-psig, or 99-F CT; that is 99-F -85-F is a 14-F SPLIT.

The delta T or temp-split should be within a 10-psig range or, +/- 2-F degrees; 97 or 101-F.

The mfg’ers Supply Outlet should be able to provide Contractors & Techs with those performance data charts. Goodman has their “Expanded Performance Data” on the Internet.


My Scan of My ThermoPride OL 11 Graphed Blower-Curve-Chart
Thermopride OL 11 Graph ipg image - Thank you Dave Staso, CA. for the better expandable image!
"After it loads Right click "Show Original Images" - Move cursor arrow over graph - Click + when 'over graph' for expanded image," then print on the highest quality setting.

MOLO Plumbing & Heating sets the A-Coil at least 6" above a Thermo Pride OL 11 Oil Furnace. They know the importance of unrestricted airflow through the evaporator coil!

Every manufacturer should furnish blower curve charts with their units and also 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. Also, illustrate best furnace to evaporator coil transitions, especially on oil furnaces!

The evaporator must be mounted 4 to 6 inches above this model oil furnace to achieve adequate airflow!

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 and temperature and heat-load to meet the customer's desired humidity 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 SP; you can do this with a simple magnehelic gauge. In any case, static pressures above 0.5"-IWC should be investigated and reduced to specifications.

DUCTWORK BASICS - Solving the Mysteries of ESP - External Static Pressure:
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 Velocity -FPM losses & huge Static Pressure problems in Oil furnace applications. If there are any existing heating or cooling airflow problems, you can & should also measure & record each branch run static pressure, & velocity, then use the formula, 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
IF Times 1000-FPM = 1395-CFM. Branch ducts: 7" Rd duct 38.48-sq. ins. = 0.2672222-sq.ft. X's 500-fpm=133-cfm
Deliver the exact CFMs to & through each diffuser that the application calls for!
Always look at the amount of CFM the diffuser's will actually pass at a specific velocity!
 DTI Corp
Air Infiltration Sources Up-to 50% of Load

DTI Corp Air Infiltration

Measuring Low Airflow

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, Some 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'

First, before doing anything else check the sizing, and thoroughly seal and properly insulate all the ductwork!

In the linked pdf above: from pages 8 through 11 do NOT use those rules of thumb for sizing equipment & airflow to rooms! Use Manual J for heat-gain/heat-loss for each room & thus the total. Use the Manual S for heating & A/C equipment sizing! Then figure the airflow & ductwork sizing required for each room according to the airflow requirement for the cooling load, or Heat Pump load using Manual D.

Ductwork Retrofitting - An Excellent Economic Opportunity - Don't Miss IT

"More than 80% of the duct systems in residential and light commercial applications 'do not' work as designed." Do your service agreements include the duct system? If not, this is an important & significant business opportunity that you are missing!

Your best access into the duct renovation market is to include the duct system in your service agreements. What this includes is having the service tech measure static pressure on each service visit. Remember, this takes five minutes or less. If pressures are very high, or very low, send out a tech competent salesperson with a flow-hood and a manometer to identify the problem, and propose a bid to do the necessary ductwork fix.

One reason we have service agreements is to gain additional income from repairs, "so start repairing the real problem with the system," and not just the equipment. In most areas of the country there is very little competition for quality duct renovation and air balancing.

Prescribing HVAC repairs without competition from other contractors in a way that will greatly improve performance and efficiency, will delight your customers - that's our definition of opportunity.

Below is an outstanding PDF "Basic AC Overview - Specifications VS. Reality"
by John Proctor, P.E., Proctor Engineering Group, LTD:


"AC Specs Vs Reality" PDF - It's Worth Your Time

FREE HVAC Resources for Professionals
Quotes from linked PDF:
"Proper sizing, installation and maintenance of HVAC equipment are major
factors in operating efficiency. In fact, the potential energy savings from a quality installation are greater than those gained from the installation of high efficiency equipment.

Proper sizing and installation can result in energy savings of up to 35 percent for air conditioners and 16 percent or more for furnaces. Moreover, energy-efficient installation and proper maintenance practices also provide substantial non-energy benefits, such as greater comfort, lower maintenance cost and longer equipment life." Specification_of_Efficient_Installation1-36.pdf

Always do Blower Door Test for air infiltration rates with the central air blower off & on, as infiltration could be a lot higher with blower on!

Home Energy Magazine Online September/October 1993

Raising Standards and Savings
New Group Hunts Bad Ducts

Does 40 billion kWh sound like a lot of energy? How about 4 billion therms? Researchers believe that's how much electrical and gas energy this country "could save by fixing inefficient ducts using current techniques." "Refining those techniques could reap savings of 90 billion kWh" plus 9 billion therms! Peak loads would be reduced too. To pursue these tremendous savings, national, state, and utility research laboratories, the U.S. Department of Energy, utilities, and energy service companies are collaborating. Their consortium is called "Residential Energy Efficient Distribution Systems," or REEDS.

These techniques, along with reducing air infiltration & heatgain/heatloss calcs, ought to be taught in all our schools as part of the Science & math curricula. Half the heatgain/heatloss can be due to a high air infiltration rate!
ASHRAE standard 62-1989 is 0.35 ACH (Air changes per Hour) or 3-hours for a total interior air INFILTRATION Rate change.

Darrell Udelhoven | 

Gurgling sounds at TEV: Low evaporator heat-loads lead to reduced liquid line mass and increased evaporator mass could be due to airflow problems. Eliminate low evaporator heat-loads before looking into adjusting the refrigerant charge.
Gurgling and pulsation noises at the expansion device can be caused by low charge, and/or non-condensibles and moisture in the system. Unbalanced airflow through the various distributor circuits of the evaporator coil will cause the TEV to close down refrigerant flow starving the coil. Piston-flow-rators will make it impossible to properly charge the system and cooling will be greatly compromised unless you eliminate the cause!

On every Rheem condenser cover it lists "non-condensibles and or moisture" as causes for a gurgling or pulsating noise at the expansion device. The entire evaporator circuits, may not become active for various reasons, - "the entire coil must become fully active for efficient performance."

The purpose of these recommendations is to provide liquid refrigerant at the expansion device and provide efficient operation. Hopefully, this will aid your research.  If I can be of additional assistance, contact me.

Evacuation Procedures - Pulling a Deep Vacuum pdf Thanks for the link Shophound.
Too many do not properly purge & evacuate contaminated central air conditioning systems.
First, the piping should be checked for proper oil return to the compressor, if not adequate do that before proceeding.

The Triple Evacuation Method is normally done on refrigeration systems, R-410a systems require it on central air conditioning systems:

Another Viewpoint pdf: Triple Evacuation Method
First, remove any valve cores with a special  valve core remover this will speed up the evacuation time. Read & follow pdf procedures above.

1) Re-claim unit charge (Recover all the refrigerant)

2) Charge system to 150 PSIG with dry nitrogen and leak test

3) On contaminated systems replace the filter dryers. Then Repair all leak(s)

4) Evacuate system to 500 microns valve off & see if it holds 500 microns for ten minutes, if it holds, break the vacuum with dry nitrogen

5) Evacuate system to a deeper 400 microns, valve off vac pump, & again break the vacuum with dry nitrogen

6) Evacuate system to 400 microns and & then Check to see if it holds. (Recharge with fresh clean refrigerant)

7) Check to see if the Supply and Return air ducts were correctly sized & sealed by the original installer.

If a vacuum pump will not evacuate a system below 1500 microns there is a problem with the pump itself, a leak in the system, or moisture in the system. Moisture is most likely because water vaporizes at 1500 microns.

Many HVAC contractors will consider this excessive time & effort for contaminated residential air conditioning systems, however it is a must for low temp applications.

The “micron” is a metric unit of measure for distance. The micron is a unit of linear measure; one micron equals 1/25,400ths of an inch. Modern high capacity vacuum pumps help speed up the evacuation process.
When a system has been evacuated below 500 microns, the pump is valved-off with the micron gauge connected, if the vacuum rises to 1500 microns and stops, there is moisture remaining in the system. If it rises above 1500 microns & continues to rise there is a leak. You should allow at least 15 minutes after the pump has been shut off an accurate micron gauge reading. When a system will not evacuate below 1500 microns there is either a lot of moisture in the system or there is a refrigerant system leak.
Sealing Sealing & Insulating the basement rim joists  - Excellent DIY Project - Make certain there is adequate combustion air for those type appliances!
Insulation & Weatherization Costs Forum
HVAC Techs
This should be helpful.
CFM X change in enthalpy X 4.5 = BTUH (Ballpark) Operating Performance
"U Must Right Click Link & open in New Tab"

Wet Bulb Enthalpy Chart 

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

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Edited: 04/05/10

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