Month: January 2018

The piston (fixed orifice) and TXV (Thermostatic Expansion Valve) are the two most common metering devices in use today, with some modern systems utilizing an electronically controlled metering device called an EEV (Electronic Expansion Valve).  It should at least be noted that there are other types of fixed orifice metering devices like capillary tubes, but their use is not common on most modern A/C systems though you will see them in refrigeration.

While the compressor creates the pressure differential to get the refrigerant moving, by decreasing the pressure on the suction and increasing the pressure on the discharge side, the purpose of the metering device is to create a pressure drop between the liquid line and the evaporator coil or expansion line (the line between the metering device and the evaporator when there is one). When the high-pressure liquid refrigerant is fed into the metering device on the inlet the refrigerant flows out the other side and the immediate pressure drop results in an expansion of a percentage of the liquid directly to vapor known as “flashing”. The amount of refrigerant that “flashes” depends on the difference in temperature between the liquid entering the metering device and the boiling temperature of the refrigerant in the evaporator. If the difference is greater, more refrigerant will be “flashed” immediately and if the difference is less than less refrigerant will be flashed.


A piston is a replaceable metering device with a fixed “bore”. It is essentially a piece of brass with a hole in the center, the smaller the bore the less refrigerant flows through the piston and vice versa. The advantage of a piston is that it is simple and it can still be removed, the bore size changed and cleaned if required.


Some piston systems also allow the reverse flow of refrigerant as shown in the diagram to the above. In a heat pump system when the reversing valve is energized (cool mode), the unit will run in cool mode and the refrigerant will follow the path indicated on the bottom.  This seats the piston so refrigerant must pass through the orifice.  With the reversing valve de-energized the flow reverses.  This unseats the piston and allows the free flow of refrigerant.  In this case, there is a metering device in the condensing unit (outside unit) that meters the flow of refrigerant in heat mode and one inside that meters in cooling mode.


The TXV can vary the amount of refrigerant flow through the evaporator by opening and closing in response to evaporator heat load.  compared to a fixed orifice a TXV operates more efficiently in varying environmental conditions (theoretically at least).

To operate, the TXV has a needle and seat that restricts the flow of refrigerant and acts as the orifice.  This needle, when opened, allows more refrigerant to flow and, when closed, restricts refrigerant flow.  There are three factors that affect the flow of refrigerant flow through a TXV.  A sensing bulb filled with refrigerant exerts force to open the TXV.  Since gas pressure increases with a rise in temperature, the bulb, which is attached to the suction line after the evaporator coil, “senses” the temperature of the suction line.  If the suction line becomes too warm, the additional pressure created by the heated refrigerant opens the TXV more to allow additional refrigerant flow.  A spring inside the bottom of the TXV exerts pressure to close the valve.  An external equalizer senses pressure in the suction line after the evaporator, and also works to close the valve. In essence, the TXV is a constant superheat device, it sets a (relatively) constant superheat at the evaporator outlet by balancing bulb, spring and equalizer pressures.

The primary method of charging a system changes based on the type of metering device. A piston system uses the superheat method of charging and the TXV uses the subcooling method of charging.

No matter what primary method of charging you use it is still important to monitor suction pressure (Evap temperature) head (condensing temperature), Superheat, subcool and delta t (or some other method of air flow verification).

While a TXV and a piston function differently the end result is a pressure drop and boiling refrigerant in the evaporator.

— Bryan

You have seen the C terminal on a dual run capacitor before. You have also seen the C terminal on a compressor.

It stands to reason that they would both connect together right?

Wrong, They don’t connect together and they aren’t even related, at least not in the way that you think.

In both cases, the C denotes a “common point” in the dual capacitor it is the common point between the fan capacitor (fan) and the compressor capacitor (herm). In the compressor it is the common point between the run and start windings (this is why R+C + S+C = R+S if you ohm a compressor)

The C terminal of a dual capacitor is actually fed from the OPPOSITE leg of power as the C terminal on the compressor. This is because you must power the start and run windings with the same leg and common with the other leg.

The way I always said it was “The same leg that feeds start feeds run” and the C terminal on a capacitor is actually the common feeds for the start winding of the compressor and fan (OPPOSITE side from the fan and herm plates on the capacitor)

So compressor terminals

C goes to one leg of power

R goes to the other

S goes to the HERM terminal on a capacitor with the other side of that capacitor (C) going to the same leg that feeds R.

C what I’m saying? Confusing

If you are new to the trade and you see the designation C or the word common don’t assume it is the same as other C and common terminals and start connecting stuff together… Unless you like creating smoke.

— Bryan

As we have mentioned in several previous articles, many blended refrigerants have glide, which simply means they boil and condense over a range of temperatures instead of just one temperature.

As an example consider refrigerant R407c, it is a zeotropic blend which means it has enough glide that it makes a big difference if you fail to take it into account.

For example, on an evaporator coil running R407c the refrigerant leaving the TXV will begin boiling at the bubble point, let’s say that the pressure in the evaporator is 80 PSIG that bubble temperature will be 40°.

Now as the refrigerant continues boiling the temperature will begin increasing towards the Dewpoint which is 50.8°. Any temperature gained ABOVE 50.8° on a R407c system at 80 PSIG is superheated, meaning the refrigerant is completely vapor.

So we calculate superheat as temperature above the dew point and subcool as temperature below the dew point and the condensing temperatures and evaporator temperature aren’t fixed but they GLIDE between the bubble and dew and back again when the refrigerant is changing state.

But what does this mean for evaporator and condensing temperatures when calculating target head pressure (condensing pressure) and suction pressure (evaporator pressure) also known as evaporator TD and condensing temperature over ambient?

The simplest way is to use the midpoint between the dew and bubble points to calculate CTOA and DTD.

In the case above you would simply calculate 50.8° + 40° = 90.8 | 90.8 ÷ 2 =  45.5° average evaporator temperature or midpoint

Emerson points out that evaporators would be better calculated using 40% of bubble and 60% of dew but the extra complexity generally doesn’t make enough difference to mention.

I made this video to demonstrate further

— Bryan

Keep in mind when reading ANY article about electrical theory or application is that it only scratches the surface of the topic. You can dedicate years of your life to understanding electrical theory and design the way many engineers do and still know just enough to be dangerous.  In HVAC we rarely need to have a DEEP understanding of electrical design but there are a few cases where a little understanding can go a long way to identifying issues before they cause trouble and that is the intent of this short article.

What is Three Phase Power?

Power is generated at the utility in three phases that are 120 degrees out of phase with one another at 60hz (Hertz). This simply means that 60 times per second each individual leg of power makes one peak and valley (a full circle), and all three of the phases together split the cycle into thirds (trisect).

This video is the best visual demonstration I have seen of three-phase power and how it works.

What Does an HVAC Tech Need to Know About Three Phase Power?

Three-phase motors don’t’ require a run capacitor because the 120-degree phase difference is ideal for efficiently spinning a motor so a “start” winding and phase shifting capacitor isn’t needed.

The biggest concern for the techs and installers with three phase is getting the phasing correct so that motors run in the correct direction. While this doesn’t matter for reciprocating compressors it is important for condenser fans and blowers and it is absolutely CRITICAL for scroll and screw compressors. Changing the direction of rotation is just as simple as swapping any two phases.

Keep in mind when installing replacement parts and equipment that if you keep the phases connected in the same way you will generally be in good shape. It is still a good practice to use a phase rotation indicator like the one above to confirm proper rotation. In most cases, clockwise phase rotation is what you are looking for but I’m sure there are exceptions to that. Alternatively, you can disconnect the compressor that could be damaged by improper phasing and start up the blower to see if it runs the correct direction before bringing the compressor(s) online. One caveat is that when motors use a VFD (variable frequency drive) the phase rotation will automatically correct making them an unreliable test in those cases.

Balancing Phases

Electricians are responsible for balancing the amperage of single-phase loads (both 120v single leg and 208v two leg loads typical on a wye three-phase system) both so the neutral doesn’t carry high amperage on the 120v loads and so the one leg of power doesn’t carry significantly more or less load than the other two. As the amperage load on a particular phase goes up, there is more opportunity for voltage drop depending on the size of the load, size of the transformer and service feeding the space and well as wire size and connection quality. This can become a challenge when there is a mix of single phase outlets, 208v appliances, and three-phase equipment.

Let’s say someone connects a bunch of space heaters on phase A, as well as a few smaller HVAC systems between phases A and B and almost nothing on phase C. If you have a large RTU that uses three-phases phase C will tend to have less load and therefore higher voltage while the load on phases A and B will fluctuate based on when the smaller systems and space heaters go on and off.

This can cause overheating of conductors and damage but it can also cause voltage imbalance which is a real cause for concern for an HVAC technician.

3 Phase Voltage Imbalance

Voltage imbalance is a motor killer. It causes poor motor performance and increased winding heat which leads to premature failure. In the case of HVAC blowers and compressors, this additional heat ends up in either the refrigerant or the air which must then be removed, further decreasing efficiency.

To test for three-phase imbalance always check from phase to phase not from phase to ground. You simply check the voltage from each of the three phases to one another and find the average (add all three and divide by three). Then compare the reading that furthest from the average and find the % of deviation. For most of you I know that sounds like a giant pain so we made this easy calculator for you.

The US Department of Energy recommends that the voltage imbalance be no more than 1% while other industry sources say up to 4% is acceptable. In general, you will want to make SURE the imbalance is below 4% and work to rectify anything over 1%.

What Can I Do About It?

You want to first look for the obvious. Melted wires, loose terminals and lugs, undersized wire, pitted contacts, poor disconnect fuse contact etc… Obviously, if you aren’t licensed or allowed to open a panel you won’t always be able to fully rectify the issue yourself but you can go a long way towards the diagnosis.

When checking voltage it is generally best to do it with the system running as close to the motor you are checking as possible. This is the actual voltage the motor is “seeing” and is what matters to the operation of the motor. You can then test back towards the distribution point, if you see a big increase in voltage as you test back towards the source you know you found a voltage drop and a cause or contributor to the issue.

From there the issues of amperage load imbalance in the panel, service size and utility issues must be considered once all the basics are covered. Most of all, if the imbalance is severe (over 4%) you don’t want to leave your motors running or you risk damage and expensive repairs.

— Bryan

Jim Bergmann and I recorded a podcast for HVAC School that covered when and how to check the refrigerant circuit without connecting gauges. Listener Joe Reinhard listened several times and wrote up this summary of what he gained from the episode. I edited it lightly but most of this is his work. Thank you so much Joe!

Keep in mind that when we make Fahrenheit to Celsius conversions we use K (Kelvin) to show temperature difference like splits and DTD and we use C (Celsius) to show measured temperatures.

Following mostly from two 45-50 minute podcasts from discussions between Bryan Orr HVACR, expert tech, teacher, & business owner, andJim Bergmann, renowned HVAC-R expert & teacher, from Redfish instruments and the MeasureQuick app, providing a detailed explanation of why techs should not connect gauges & hoses to system just to check refrigerant charge (in many cases).

Why Not Connect?

The benefits of NOT connecting gauges during every visit for HVAC-R business owners, technicians, and clients include:

  1.  Non-invasive measurements with only temperature data taken. Exact same way one checks if a typical refrigerator was operating properly which has no ports to attach hoses and gauges.  
  2. Just measuring DTDs (Design Temperature Differences) and line set piping temperatures are non-invasive, involve less liability both for the system and technician safety, and demonstrates technical knowledge and best practices.   
  3. Better for the refrigeration system and the environment (“green”) since it saves R22 and R410A released to atmosphere.
  4. Time savings at site so techs can concentrate on better and more preventative maintenance (PM) of the air flow system (including condensate drainage) and PM checking electrical characteristics of various control components (capacitors, contactors, sequencers, etc.,.)
  5. Eliminate more call backs and potential premature system cooling (and heating for heat pumps) performance problems and failures due to cross contamination, moisture contamination and lost refrigerant.
  6. Saves the customer money on refrigerant added due to connection losses.

Term Definitions 

  • Evaporator DTD (Design Temperature Difference) is the designed difference between the evaporator coil saturation/boiling temperature as measured on the suction gauge and the return air temperature. 35°f (1.66°C)of difference is considered normal for a typical system set at 400 CFM(679.6 m3/h) per ton airflow. Oversized evaporator coils and increased airflow above 400 CFM(679.6 m3/h) per ton will result in lower DTD and lower airflow with smaller coils will result in higher DTD.
  • Condenser CTOA (Condensing Temperature Over Ambient) is the temperature difference between the condensing coil saturation / condensing temperature as measured on the liquid line high side gauge and the outdoor temperature. This difference will vary depending on the efficiency of the system/efficiency of the condenser coil.

6 – 9 SEER Equipment (Very Old) = 30° CTOA

10 -12 SEER Equipment = 25° CTOA

13 – 14 SEER Equipment = 20° CTOA

15 SEER+ Equipment = 15° CTOA

  • Delta T (Evaporator Split) is the temperature difference between the return and supply air. Delta T will vary quite a bit depending on airflow and indoor relative humidity. This chart shown below is designed for a 400 CFM(679.6 m3/h) per ton system. Lower airflow will result in a higher delta t and higher airflow will result in a lower delta t. This is why Jim Bergmann does not prefer Delta T as a firm diagnostic or commissioning tool but rather as an approximation of airflow.
  • Target Superheat on a TXV system is dictated by the design of the TXV. Usually target superheat on a TXV system will be 5°f- 15°f (2.75°K – 8.25°K) at the outlet of the evaporator where the TXV bulb is located. On a piston system the target superheat is calculated using a superheat chart and measuring and plotting the outdoor dry bulb temperature and the indoor wet bulb temperature.
  • Target Subcooling on a TXV system will be listed by the manufacturer but is generally between 8° – 14°(4.4°K – 7.7°K)subcool. Subcooling will vary quite a bit on fixed orifice systems but 5°-20°(2.75°K – 11°K) is a common range.

DTDs (Design Temperature Difference) of the coils, after a system is newly commissioned orfirst-time assessed with gauges, should not change over the life of the sealed refrigeration system once a system has been charged correctly unless one or more of the following has developed:  

  1. Airflow restriction with dirt buildup as main cause – dirty outdoor coil, dirty indoor coil, dirty filter,  dirty blower blades/inside the housing, Return/Supply duct restrictions, blower motor speed or operation problems, and if the homeowner installs a so-called high efficiency, nothing-gets-thru-including-air filter.
  2. Critical component failure.
  3. Refrigerant flow restriction.  

So after the first-time visit performance assessment or a new system is commissioned, subsequent system checkups or maintenance visits should be performed without connecting gauges.    

The following risks, problems, and liabilities occur and eventually develop when technicians attach gauge hoses every time to check the system refrigerant characteristics versus just using measured system temperatures and knowledge of Return/Supply air TD, Evaporator/Condenser split, and refrigerant P/Ts.   Not attaching hoses & gauges to systems without good reason is actually correct practice and the following could be avoided or greatly minimized.

  1. Techs are inducing system contamination if, prior to connecting the hoses the techs didnotuse dry nitrogen to purge air, moisture, and/or old refrigerant out of their hoses & manifold from the prior system the gauges were attached.  Perhaps the prior system had a different refrigerant that may/may not have been contaminated with non-condensables and other refrigerant(s)
  2. Were the hoses on the gauges left open to the atmosphere in the back of the truck used for the prior R410A system?  If so, the coating of POE (polyester) refrigerant oil (highly hygroscopic) would have absorbed moisture which, if not correctly purged with dry nitrogen, would contaminate systems by inputting moisture which will cause TXV and liquid filter-drier freeze ups (blockages), cause contaminated refrigerant (making R22 recycle subject to high fees and fines), and cause acids which will attack and corrode compressor surfaces (copper plating), valves, and windings.  Hoses should always be tightly connected to the manifold parking ports to prevent moisture contamination.
  3. Are techs properly & carefully disconnecting gauge hoses while the system is running?   If not, perhaps a service call back will shortly occur since, every time hoses are connected and disconnected, some refrigerant is lost.  If the liquid hose is not charged back through the manifold and Suction hose, several ounces or more in the liquid hose are lost if techs inadvertently or on purpose blow or dump refrigerant by not properly disconnecting gauge hoses while the system is running.  This occurs if techs are inexperienced or decide not to take the time or are not equipped with low-loss-ball-valve hose end fittings to slowly, carefully, after purging hoses if needed, charge from the liquid hose (holds 7x the R410A as the vapor or suction line; 10X for R22) through the gauge manifold into the vapor or suction hose back into the running system.  If this procedure is not done correctly, air and moisture can enter the system.   After one, two, or three years of visits, techs can be chasing “leak(s)” created by multiple connects/disconnects.  
  4. Caps no longer inadvertently left off on Schrader valve ports leading to leaks.
  5. Reduced safety issues for techs since less chance of refrigerant in eyes and frozen-fingers and loss refrigerant to the atmosphere.

Data to recordduring first-time system performance assessments and new system commissioning using refrigerant gauges so that benchmarks exist to compare to future checkup visits but without attaching gauge hoses if no observed or reported system problem reasons.  

  1. TD or Temperature Difference between the Return air dry-bulb (DB) and Supply air DB.   TD level depends on the sensible & latent heat content of the inside air.  Higher TD for low RH% (Relative Humidity), lower TD for high RH%.  20°F (11°K)  TD is good if system operating properly at 75°F(23.88°C), 50% RH and set for 400 CFM/(679.6 m3/h)ton.  If reduce CFM/( m3/h) ton, TD increases, but if RH% increases, the TD decreases back-and-forth so the TD can range 16°f – 24°f(8.8°K – 13.2°K) (or more in extreme cases, see the Delta T chart)    
  2. Evaporator DTD (Design Temperature Difference), also called “Split”, is temp difference between the Return air dry-bulb (DB) temp and the refrigerant saturation temp of the coil – either 35°F(1.66°C) at 400 CFM/(679.6 m3/h) ton to 525 CFM/(891.98 m3/h)ton or 40°F(4.44°C) at 350 CFM/(594.65 m3/h) ton.  
  3. Evaporator outlet SLT (Suction Line Temp) and SH (SuperHeat) On a TXV system the superheat range 5°f(2.75°K) to 15°f(8.25°K) depending on factory setting +/– 5°F(2.75°K) of 10°f(5.5°K).   Fixed-bore or piston reading depends on inside heat load, Return air WB, and outside air DB temp
  4. TESP (Total External Static Pressure) inches WC of the air handlerbetween non-turbulent point in Return plenum before a clean filter and in the Supply plenum non-turbulent area.  With caution, drill 3/8”-1/2” holes to cover when done with vinyl or plastic professional looking plugs. On a furnace drill above the filter for the return reading and between the furnace and the coil for the supply reading. Note if the coil was wet or dry since TESP changes. 
  5. Pressure Drop  “wc across thefilter.
  6. Pressure Drop  “wc across theEvaporator coil, note if wet or dry coil, and plug holes.
  7. Indoor Blower motor (IBM) running load amps (RLAs) compared to nameplate Rated or Full Load Amps (FLA) with the panels on.
  8. SLT and SH at the Condenser (Compressor inlet).  SH within +/– 5°f(2.75°K)  is acceptable. For a TXV, superheat average 10°f(5.5°K) plus additional 1-3°F (.55°K – 1.65°K)of SH the Suction/Vapor line absorbs (as measured).  For a fixed-bore or piston Metering Device at the indoor coil, a total “target SH” is determined by outdoor DB and indoor WB temps.  
  9. Condenser DTD or Split is temp difference between the refrigerant saturation temp and the DB temp of air at entering middle of the coil. As SEER increases, condenser surface areas are larger but are limited by diminishing heat transfer capability as the temperature difference between the outdoor air and the coil temperature decrease.  
  10. LLT (Liquid Line Temp) and SC (SubCooling) at the Condenser outlet.   SC within +/– 3 °f(1.65°K)  is acceptable. Ex. for 85°f(29.4°C) ambient, 13 SEER with a 20°f(11°k) DTD split, and 10° (5.5°k) Subcool nameplate, the Liquid Line temp = 95°f(35°C)  = 85°f(29.44°C) outdoor + 20°f(11°k) CTOA  – 10°f(5.5°K) Subcooling ).
  11. Compressor and OFM running load amps (RLA) compared to nameplate Rated and Full Load Amps (FLA), respectively.  
  12. Measured suction temperature differential between the suction line leaving the evaporator and entering the compressor in °f. So if the suction line is 50°f (10°C) inside and 53°f (11.66°C) outside there would be a 3°f (1.65°K) temperature rise.
  13. Measured liquid temperature differential between the liquid line leaving the condenser and entering the metering device in °f. So if the liquid line is 95°f (35°C) outside and 92°f (33.33°C) inside there would be a 3°f (1.65°K) temperature drop.

Again, benchmarked DTDs, SHs, SC, and ESPs should not change during the life of the system unless one or more of the following has developed:  

  1. Air flow restriction
  2. Component failure
  3. Refrigerant flow restriction  

Data to recordduring follow-up seasonal checkup visits and compare to benchmark data.  See if problems have or are developing and show improvement after any services are performed which offers value to clients/customers paying for the service call or membership fee.  Service could be simple as a filter change, coil cleaning, and blower maintenance but, since have more time for PM, also identify potential electrical parts failures and inform clients to choose to fix now or later.

  1. TD between the Return air dry-bulb (DB) and Supply air DB.  Should be in 16-24°F(8.8°K – 13.2°K) range depending upon sensible & latent heat content of inside air (see chart).  
  2. Evaporator outlet SLT.  If a TXV, should be within +/– 5°f(2.75°K)  of benchmark reading.   Fixed-bore or piston reading depends on inside heat load, Return air WB, and outside air DB temp.  More practical SLT determine at outdoor coil Suction/Vapor line.
  3. TESP (Total External Static Pressure) “wc of the air handler and note if wet or dry coil.
  4. Static Pressure Drop “wc across thefilter and re-plug holes (or visually inspect / replace)
  5. Static Pressure Drop “wc across theEvaporator coil, note if wet or dry coil, and re-plug holes.
  6. SLT at the Condenser (Compressor inlet). For an indoor TXV, should be within +/– 5°f(2.75°K)  of benchmark reading.  For fixed-bore or piston indoor coil Metering Device, determine total “target SH” from outdoor DB and indoor WB temps.  
  7. LLT (Liquid Line Temp) and SC (SubCooling) at the Condenser outlet.   LLT using SEER-rating split, should be within +/– 3°f(1.65°K)  of benchmark reading. Outdoor air temperature + CTOA based on system efficiency – subcooling = target liquid line temperature

Other notes:

Always use pre-tested, calibrated (as possible) digital thermometers to measure air temps and line set pipe temps or insulated temp sensors.  Do not depend upon the space thermostat to accurately represent inside air temps since could be Return duct leakage, bypass ducts not dampered correctly, and air handler cabinet leaks e.g. holes/gaps at indoor coil line set inlet affecting the Return air temp.  

Air flow through/across Evaporator and Condenser coils will only decrease and not “magically” increase. The primary reason is dirt accumulation on air flow components e.g. coil fins, indoor filter, indoor blower blades, outdoor fan blades.  Other reasons include leaky air handler cabinets from gaps at Return & Supply duct connections, holes at line set inlet to Evaporator cabinet, and a bypass duct with no damper to close off air flow between Supply and Return in Cooling mode.

Systems should not be benchmarked with a wet Condenser coil or if the LLT is at or below the outdoor ambient air DB temperature.

Use a battery or cord operated leaf blower to dry out the coil in 5-10 minutes.  

The only action that increases airflow is increasing the fan or blower RPM or speed.   If Suction line supposed to be 54°F(12.22°C) (40°F(4.44°C) coil + 10°f(5.5k)  SH if TXV + say 2F SH addl to Vapor line length) but is 47-48°F(25.85°K – 26.4°K), look for indoor air flow restriction issues.   The evaporator is like a boiling pot of water but a sealed system so if the burner heat is turned up, pressures and temperatures increase. More than additional 24°f(1.1°K – 2.2°K)  Superheat at the Compressor inlet, probably better insulate the Vapor line.

Maximum inlet temperature Suction line at Compressor inlet should be below 65°f(18.33°C) .   If not, the Compressor will have the potential overheat and oil breakdown can occur do to excessive discharge superheat / temperature.

TXV designed to maintain 5-15°f(2.75°K – 8.25°K) superheat (10°f(5.5°K) given +/- 5°f(2.75°K) range) but only at the Evaporator outlet or where the sensing bulb is located on the suction line.  Some SH is added to the suction line before gets to the Compressor inlet.   However, if the line set is located in a 145°F(62.77°C) attic and Vapor line not well insulated, significant SH gain will be seen at the Compressor inlet. Vapor line needs good insulation (also for Heath Pumps in Heating mode) e.g. with thicker tubing insulation and/or using a foil-bubble wrap or “Reflectix” attached with foil tape since reflects IR heat.

Summary of the Jim Bergmann / Bryan Orr Podcast on checking the charge without using gauges by Joe Reinhard

P.S. – As mentioned in the podcast the Testo 605i and the 115i make a great pairing to check a system in the way described above

You can now do ALL of these calculations easily with the MeasureQuick app at

Before we convert temperature scales, let’s take a step back and think about what temperature is in the first place.

Temperature is proportional to the average kinetic energy of the random microscopic motions of the constituent microscopic particles, such as electrons, atoms, and molecules.

Translation: Temperature is the average “movement energy” of the molecules in a substance

Higher temperature means there is more average heat ENERGY when compared to the same mass of the same substance at a lower temperature. While there is no limit to how high the temperatures of matter can go (at least that science is aware of), there is a bottom limit and that is the point at which there is NO HEAT, and therefore no molecular or atomic motion. That point of no energy is called ABSOLUTE ZERO.

Absolute zero, is a theoretical point because it has never been (and likely will never be) achieved. For most of us, absolute zero has no real application and this is why our most common temperature scales are tied to freezing and boiling instead of absolute zero.

Anyone can make up their own temperature scale. All you need to do is pick a zero based on a known (say boiling or freezing water at atmospheric pressure) and then decide on a size of the degree.

In the “Fahrenheit” scale a guy named Daniel Fahrenheit decided to make a temperature scale, the coldest “constant” he had at his disposal was a water and brine solution, so he called that 0°F. He then used the body temperature of an “average healthy man” and somewhat arbitrarily called that 96°F. This dictated the “size” of the Fahrenheit degree as well as the zero point. From there he ascertained that on his scale water froze at 32°F and boiled at 212°F at sea level (14.7 psia).

The Celsius scale (often called Centigrade by old timers) logically used the freezing and boiling of water as the 0° and 100° points. This established a logical starting point at the freezing of water as well as LARGER degree size than the Fahrenheit scale.

This means that when you are converting a temperature… say 75°F to Celsius, you first subtract 32° to normalize for the different starting points of the scale and THEN you multiply by .555 (I prefer decimals because fractions don’t work on a calculator) so 75°F = 23.865°C

But be careful…

When converting a temperature comparison or differential you must skip the +/- 32° part of the process. In those cases you are only coverting the SIZE of the degree, not it’s location on the scale

This means that if we are discussing 10°F of subcooling, we would simply multiply it by .555 to see that it is 5.55°C of subcooling.


Scientists didn’t like the old Fahrenheit and Celsius  systems because they are scientifically nonsensical. There is NO SUCH THING as negative heat! exclaimed the angry Mr. Kelvin & Rankine (at least that’s how they do in my imagination). So they invented scales where zero is ABSOLUTE ZERO so there are no negative numbers. The Kelvin scale starts with 0 at absolute zero and uses the Celsius degree size and Rankine starts at absolute zero and uses the Fahrenheit degree size.

So whenever we make a conversion from Fahrenheit to Celsius on the SCALE we show the converted temperature as °C, but when making a conversion that is simply as comparison or a differential (DTD, Delta T, CTOA, Superheat, Subcool etc…) we show it as °K to help us differentiate.

Consfusing huh?

— Bryan

This article is written by Jeremy Smith CM, experience refrigeration tech and all around great dude. Thanks, Jeremy

A very common means of control seen on refrigeration equipment is the pump down control. Why do we use this rather than just cycling the compressor off and on like a residential HVAC unit?

Since most refrigeration equipment tends to be located outdoors, it comes down to ambient temperatures and the basic properties of refrigerant we all understand about temperature and pressure and how they can conspire to kill a compressor.

During periods of low ambient temperatures, if we were to just cycle the compressor off, it can easily get colder at the compressor than it is inside the space.   If the compressor cycles off for long enough as it would during a defrost cycle, refrigerant vapor will start to condense within the crankcase.  If we are lucky, the extent of this problem will be a unit that doesn’t start because the pressure of the refrigerant is lower than the cut in setting of the pressure control.  What typically happens, though, is that enough refrigerant will condense to start to settle under the lubricating oil causing a lack of lubrication on restart leading to bearing wear and premature failure.  If enough refrigerant condenses within the compressor housing, the resulting damage could cause valves, pistons and other internal parts to break if liquid gets into the cylinders.

How can we prevent this?

One thing that is applied across almost all sectors of our industry is crankcase heaters.   These small heaters, either immersion style heaters or wrap around style heaters add a small amount of heat to help keep the compressor oil warm and help to prevent vapor from condensing there. The effectiveness of these are limited by the wattage of the heater, the ambient temperature and the size of the compressor.   Too low an ambient or too large a compressor and they start to lose some effectiveness.

So, how else can we prevent condensation within the compressor?  Let’s look to the pressure/temperature relationship of refrigerant for the answer.   If we lower the pressure in the crankcase to a point where the saturation temperature of the refrigerant is below the ambient temperature the compressor is in, refrigerant cannot condense.   This is why we use a “pump down” type system.

In operation, a pump down control consists of little more than a liquid line solenoid valve, a thermostat control and a low pressure control.   When the thermostat or defrost control opens, the solenoid de-energizes, stopping the refrigerant flow and allowing the system to pump the suction pressure down before the low pressure control turns the compressor off.

How low should we set that cut out?   The Heatcraft installation manual has us setting the cut out as low as 1” Hg vacuum, depending on the minimum expected ambient.  I like to set the cut in just below the lowest expected ambient temperature so that you don’t wind up in a situation like I mentioned earlier.   If the ambient gets too low and the cut in is too high, your unit won’t cycle on until it warms up enough resulting in a preventable service call.

Combining a pump down control with a crankcase heater and ensuring that all controls work properly at all times can save your compressor from damage in cold weather.


Jeremy Smith, CM

Erich Vinson is a tech from Colorado and one of the most entertaining people I interact with online. He wrote this quick tech tip on stripping back the outer jacket properly on control wires and it happens to also be something I preach. Thanks Erich.

In the first picture (above), you can see what happens when you try to use a pair of wire strippers to remove the jacket. It damaged the wire underneath. Instead use the pull string when you strip the jacket off of a low voltage (control) cable.

Use your strippers to remove about three inches of the jacket (or cut into the end like I show above – Bryan), and use the pull string to peel away the jacket, as is shown in the below.

Then, cut off the wires just below where you used your strippers.

The result will be low volt wires with no damaged insulation, and no hard to find low voltage short circuits.

— Erich

The gas laws. We all learned about them in school and promptly forgot all about them. I really think that we need to dig our books out, dust that information off and work to understand and apply it.

Many will say that nitrogen pressure doesn’t change with pressure like other gasses. This is false but read on.

Let’s start by looking at pressure a little differently. Pressure is a measure of the force exerted by a gas within a container. It exerts pressure because the individual molecules of the gas are colliding with the walls of the container. Those collisions are happening because each molecule has a specific amount of energy. So, in this way, we can view pressure as a measure of the amount of energy contained within our container of gas. That might sound complicated, so let’s kind of unpack it and see if we can understand it better.

We have a container that has a fixed volume, for example, 1 cubic foot. So at 0 psig, there is a certain number of gas molecules contained within that container and a certain number of collisions with the container walls occurring.

Now, let’s take that container and we’re going to double the number of molecules inside that container without changing its size at all.. We know that the pressure increased, but what did it take to do this? Energy.

Adding those additional gas molecules required that we add energy to force that extra gas into the container. The addition of energy to force additional molecules into the container resulting in an increase in pressure. The thing to remember now is the law of conservation of energy. Energy isn’t created nor destroyed, it simply changes form.

Since heat energy is simply another form of energy so it stands to reason that adding or removing heat energy from our system will affect the energy level of the gas molecules and ultimately the pressure exerted by them. Let’s return to our sample container of 1 cubic foot internal volume. We’re going to expend enough energy to put enough molecules into this container to raise the pressure to 100 psig at a temperature of 70°F. If we add more energy not in the form of compressing more gas but in the form of heat energy, what will happen to the pressure in the container?

The heat energy is going to ‘excite’ the molecules in the gas, increasing the number and force of the collisions that are occurring that are the basis of pressure existing. Since we’re adding energy, the pressure will rise and it will rise in a predictable and consistent way. The reverse is also true if we remove energy, the pressure will drop in the same consistent and predictable

This is why we need to understand the gas laws as technicians. They allow us to predict and understand the pressure change caused by adding or removing heat energy from a sealed, pressurized system.

Practical application
Now that we understand how heat energy affects the pressure within a sealed system, we can apply this knowledge to pressure testing. A large number of factors are making proper leak testing at installation more important than ever and manufacturers are demanding more detailed leak testing procedures. Add to that the fact that our tools are more refined than old-school analog gauges and a leak of even 0.5 psi over a several hour period of time is easily something a technician can spot.

Let’s take a look at an imaginary but fairly realistic scenario to see how this works and what it means on the ground in the field.

New construction split system. Tonnage isn’t super important to this, but we just made the last brazed joint, it’s the end of a long day in the 90° heat and a nasty thunderstorm is brewing. Let’s get this thing pressurized and get home. Run the pressure up to 350 psig of nitrogen and get out of here. When we show up in the morning when it’s 65°F and find that the pressure has dropped almost 16 psig, that might make us a little nervous. We checked all of our joints with a mirror and with soap bubbles but we don’t see any leaks… where did the pressure go?

Before we get excited, let’s look at how the temperature change affected the pressure within this sealed system. We pressurized to 350 psig at 90°F and it’s now 65°F. With the gas law equations, we can know what the pressure in the system should be and eliminate time wasted looking for leaks that aren’t actually there. This is an expression of the gas laws known as Gay Lussac’s Law. In this, the system volume is a constant and can be disregarded. For our purposes, the copper piping we use to build systems is unchangeable, so we’ll use this equation.

The first step is for change the equation around to isolate the answer we wish to get.
P2= T2 (P1/T1)

Now, we have a simple equation we can plug our numbers into and get the answer, right? Not quite yet. We have one more step before we get the calculators out. We need to convert the pressure and temperature valves that we have to absolute pressure and temperature readings, so add 14.7 to the pressure and 492 to the temperature to get to absolute scales

Now, our numbers look like this:
T1 = 582°R (Rankine)
P1 = 364.7 psia
T2 = 557°R
NOW, let’s solve.
P2 = 557 (364.7/ 582)
P2 = 557 (0.6266)
P2 = 349.03.

But wait, our system dropped to 334 psig, so we have a leak…
We forgot one VITAL step. We need to convert our P2 reading back to gauge pressure.
349.03 – 14.7
334.33 psig

This says that the pressure loss within the system was due ONLY to the temperature change and was not due to a leak.
Time to get the vacuum pump out and finish this job up.

In summary, every gas responds to the gas laws in the same way. We use nitrogen because it is readily available (the air is mostly made of nitrogen), dry and it doesn’t readily combine with other molecules under normal circumstances.

It does change pressure with temperature and all you need to do to find out how much it will change is by changing both the before and after temperatures to absolute scales (Rankine for Fahrenheit or Kelvin for Celcius) and convert your before and after pressure readings from gauge pressure (PSIG) to absolute pressure (PSIA). Once you have your solution you can convert back to Celcius or Fahrenheit

— Jeremy Smith CMS

P.S. – I made a little before and after calculator HERE

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