Airflow, Airflow, Airflow…. when we setup and commission comfort cooling and heating systems we need to pay more attention to airflow before we worry about the fancy controls or the refrigerant circuit.

So as a thought exercise let’s consider a typical 2-ton, straight cool, TXV, residential system and think through what happens when we alter airflow and what impacts that has on the system.

Rather than talk in terms of advanced psychrometric math we will keep the math to a minimum and focus on “If this than that” relationships between airflow and system function

Mass vs. Volume

First let’s establish that it is the molecules or “stuff” that makes up air that contains and can move heat energy. While we often talk in terms of CFM (Cubic Feet Per Minute) that is a measurement of volume rather than mass. The air conditioner cares about the mass flow of air over the coil not the volume flow which is why more airflow in CFM is required in high altitudes where air density is lower.

In other words…

Mass flow is what matters and when air get’s less dense we need more air volume to move the same amount of heat

So when we speak in terms of CFM/ton (Cubic feet of air per ton of cooling) that is referring to typical air at sea level and needs to be adjusted as air density changes.

400 CFM/Ton 

The 400 CFM/ton design has been used for years and it is an adequate baseline airflow for many types of equipment and in many moderate climate zones. There are several issues with the 400 CFM/ton rule where it needs to be adjusted.

  • Higher altitudes where air is less dense and therefore more air is required to maintain the same mass flow rate over the coil
  • The nominal or listed tonnage on a piece of equipment is often NOT what the equipment produces at current load conditions. A 2-ton system that is designed for AHRI conditions (95° outdoor and 80° indoor return temperature) could easily produce under 20K btu/hr at 73° indoor and 97° outdoor temperatures, so 800 CFM would be well over 400 CFM/ton in that scenario.
  • Areas with higher latent (humidity) load will run lower than 400 CFM/ton on purpose to remove more moisture from the air and areas with arid (dry) climates will often run higher than 400 CFM/ton to remove less or no moisture from the air.

How The Evaporator “Absorbs” Heat

In my refrigeration circuit basics training I call the evaporator coil the “heat absorber” because its end goal is to take heat from where you don’t want it and move it somewhere else.

The heat gained in the evaporator in this scenario comes from the indoor air being moved over the evaporator coil. The air is warmer than the refrigerant so heat leaves the air as it impacts the tubing and fins of the coil because “hot goes to cold”.

The heat is transferred from the air though the walls of the copper tubing and into the refrigerant via conduction while the heat is transferred through the air and refrigerant itself via convection because they are both dynamic (moving) fluids.

The air temperature is decreased because heat is removed from it into the refrigerant. The refrigerant in the evaporator coil is at saturation (boiling) so the coil temperature doesn’t change directly as heat is added to the refrigerant but it does begin to increase indirectly because as the total heat energy in the evaporator increases so does the coil pressure and vice versa. This is similar to the pressure cooker effect where as the water boils in the pressure cooker the pressure increases and so does the boiling temperature of the water.

When the temperature of the coil is below the dew-point of the air moving over it there is also a transfer of latent energy from the air as some of the water vapor in the air condenses to liquid water (condensate) on the evaporator coil. This latent heat transfer does not result in colder air but rather lower moisture content in the air, this heat does impact the evaporator in the same way as sensible heat as it is added to total heat picked up in the evaporator.

Evaporator Coil TD

We use the term “coil TD” a bit differently in different parts of the industry but in air conditioning it is the difference between the air temperature of the return air entering the evaporator coil and the saturated suction temperature often called the “coil temperature”. In typical 400 CFM/ton applications this difference will be around 35° with a higher number meaning a colder coil and a lower number meaning a warmer coil. There are several things that can impact coil TD including refrigerant mass flow rate (how much refrigerant the compressor is moving), metering device performance, return air dew point (moisture content) and most commonly…. airflow.

What Happens When Airflow is Decreased?

In this theoretical system when the airflow is decreased and all else stays the same the following things will occur –

  • Mass airflow will decrease, meaning there are fewer molecules moving across the coil
  • Air velocity will decrease, meaning the air is moving over the fins and tubing more slowly
  • Bypass factor decreases, this means more of the air molecules will be touching the metal as a ratio
  • Air temperature decreases (to a point) due to the air moving more slowly across the coil with less bypass factor
  • Coil temperature decreases because less overall heat is being picked from the air
  • Coil drops further below dewpoint, causing more moisture to be removed from the air increasing dehumidification
  • Suction pressure decreases because less heat energy being picked up means less pressure and as the superheat falls the TXV also futher throttles the flow of refrigerant through the coil
  • Compression ratio increases as the suction pressure drops meaning the compressor moves less refrigerant as the refrigerant density entering the compressor falls
  • Coil TD increases as indicated by the colder coil in relationship to the return air

We all know that if you have far too little airflow a system can freeze up when the coil temperature drops below 32°F. The other consequence of dropping airflow is lower overall sensible capacity and therefore a drop in EER and SEER rating. On the positive side in humid climates, a system with lower airflow will remove more water from the air which can be desirable.

The lesson is, sometimes you need more airflow and sometimes you need less but no matter what, changing airflow changes a lot about how the system operates and should be done carefully and thoughtfully.

— Bryan

 

 

In my recent classes with my employees at Kalos, we are going over finding target pressures and temperatures for an air conditioning system with the goal being to get techs to have “target” readings in mind before they start connecting tools. This step is an important part of being able to “check a system without gauges” like we have talked about so often. Much of this list makes more sense if you are already familiar with our 5 pillars of diagnosis.

It is important that we start using these terms when speaking to one another, writing notes and diagnosing because these will translate better between systems and between A/C and Refrigeration. Some readings we take in HVAC like static pressure and delta t do not apply to refrigeration, while others like target condensing and evaporator temperature are key in both disciplines.

Keep in mind that when I give a “rule of thumb” you should always consider manufacturer specs, charts, panels, install and service manuals as superior to a rule of thumb. You will be amazed what you might learn reading the service and installation manuals of the systems you install and work on.

 

Target Evaporator Temperature or DTD (Design Temperature Difference) = The temperature the evaporator coil should be based on the return temp (A/C 35°F below ambient) or below box temp in the case of refrigeration 10° below box on a walk in 20°below box on a reach in). on a typical A/C system with a 75°F return temperature, the DTD would be 35°F which means the target evaporator temperature would be 40°F. The DTD will vary based on airflow and evaporator coil size.

 

Measured Evaporator Temperature or TD (Temperature Difference) = The suction saturation temperature (not pressure) as measured on the suction gauge for that particular refrigerant it can then be compared to box or return temperature to calculate your measured or actual TD

 

Target Condensing Temperature Over Ambient (CTOA) = This is the target temperature the liquid saturation (condensing temperature) measured on the gauge SHOULD be above the outdoor air temperature measured in the shade entering the condenser coil. This will be 30° over ambient on VERY old units, all the down to as low as 15° on new very high-efficiency units. This is LIQUID LINE ONLY the discharge line will be higher pressure.

 

Target Condensing Temperature – The outdoor temperature + the CTOA = Target condensing temperature Example: Outdoor Temp 95° + 15° for a 16 SEER system = 110° target condensing temp.

 

Target Superheat – This is the superheat you SHOULD have and it varies based on if the system has a TXV or Piston metering device. If a piston you MUST use a use a superheat chart as well as a thermo-hygrometer / psychrometer to measure the indoor wet bulb and dry bulb because those charts require those readings. If the system is a TXV then set your target at 5-15° if measured inside and 10°-20° if measured outside

 

Measured Superheat – The increase of the suction line temperature when compared to the suction saturation.

 

Target Subcool – The subcool you wish to achieve. Many units will have this marked on the data tag, if not then use 10° subcool on TXV systems and 5° – 15° on piston systems recognizing that on a piston system this rule will not always apply.

 

Measured Subcool – The measured difference between the liquid line temperature and the condensing temperature (liquid saturation temp) off of the high side gauge. This is liquid line only, not the discharge line.

 

Outdoor Ambient – the outdoor dry bulb temperature, in the shade entering near the center of the condenser coil

 

Return DB – Return dry bulb. The temperature of the return air without taking evaporation or humidity into account. Best taken in the return right before the unit and not in the space.

 

Return WB – Return wet bulb. The temperature of the return air + the evaporative effect. Lower WB, when compared to DB, means lower relative humidity. Wet bulb and dry bulb will be the same at 100% RH. Best taken in the return right before the unit and not in the space.

 

Return RH% –  The relative humidity of the air in the return. Relative humidity is the percentage of moisture in the air compared to how much moisture is in the air. Hotter air can hold more moisture than the same air at a lower temperature.

 

Target Supply Air Temperature – Target supply air temperature is calculated using a delta t chart and comparing the return DB and the return WB temperatures. The target supply air temperature is dry bulb and can be compared to the return DB to calculate the target delta t.

 

Target Delta T (Air Temp Split) – Don’t confuse TD or Evaporator split above with delta t or air temp split. Keep in mind that the 18°- 22° rule that many use only applies to homes with 45% to 55% relative humidity. As RH% goes up the target split will go down and as RH% goes down the split will go up. Delta T will also vary based on airflow with higher airflow resulting in a lower Delta T.

 

Measured Delta T – The measured difference between the supply and return air DB. Keep in mind this should be taken a few feet before and after the unit to allow for air mixing and reduce radiant gains/losses.

 

Delta H – Delta H is an advanced measurement that calculates the change in enthalpy (heat content) of the air between the return and the supply. You can do this with two digital thermos-hygrometers like the 605i and it takes into account the temperature and humidity of the air entering and leaving the evaporator.

 

Delivered Capacity – Delivered capacity is the calculation of BTUs of heat being removed from an air stream which combines the Delta H with the CFM of air to give you the total “work” being done across the evaporator coil.

 

Discharge Temperature – The measured temperature (with a line clamp) of the discharge line leaving the compressor, not the liquid line

 

Target Liquid Line Temperature  –  When “checking a system without gauges” the target liquid line temperature is the target condensing temperature minus the target subcool. This is usually measured at the condensing unit.

 

Target Suction Line Temperature –  When “checking a system without gauges” the target suction temp is the target evaporator temperature plus the target superheat. This is most accurate when measured inside but is also valuable when measured outside.

 

Approach – Approach is just another name for target liquid line temperature and it a reading that Lennox publishes a target for on many of their units in the installation manual and on the back of the panel of the condensing unit. Systems with larger or more efficient condenser coils tend to have a lower approach (cooler liquid line) while those with smaller, less efficient coils ten to have a higher approach (warmer liquid line)

 

Suction Line TD (Temperature Rise) – The difference between the suction line temperature inside after the evaporator coil and outside by the condensing unit. When a suction TD is more than 10°F compressor overheating and oil carbonization can occur under some load conditions.

 

Liquid Line TD (Temperature Drop) – The difference between the liquid line temperature outside by the service valve and inside before the metering device. ideally, the liquid would have VERY LITTLE temperature drop and any drop of more than a few degrees should be looked into. Long line length, vertical risers, running the liquid line through a low ambient space, contact between the liquid line and the suction line or restrictions can lead to higher than normal LL TD.

 

Static Pressure – The positive or negative pressure exerted on all surfaces equally within a duct system. Static pressure does not measure flow, it is like the pressure inside a balloon that inflates or deflates. Static pressure is generally measured in inches of water column in the USA (“wc)

 

Design TESP – This is the total external static pressure, both positive (supply) and negative (return) that a particular furnace or air handler are designed to work under external to the appliance. Most typical residential units are designed for 0.5”wc.

 

Measured TESP – This is the total external static measured using a manometer or magnahelic gauge and static pressure tips. On a furnace this would be measured inside the furnace before and after the blower but before the coil. In an air handler or fan coil it will generally be measured before and after the unit in the ducts. This is the total difference between the negative and positive reading so if return static was -0.2” and supply static was +0.3” the total would be 0.5”wc TESP.

 

Static Pressure Drop – This is the measured pressure change across a portion of the air system. For example across a coil, filter, duct etc… This is helpful in diagnosing airflow issues and changes over time.

 

While this may seem like a long list, most of it is pretty common sense. One thing to mention is the fact that if you do not have a thermos-hygrometer like the 605i and accurate temp clamps you cannot properly check a Delta T on any system or set the superheat on a fixed orifice system. In order to properly set a charge or diagnose a system, you need a way to accurately test line temperatures and measure return / indoor Wet Bulb, Dry Bulb and Relative Humidity.

 

Step one on diagnosing a refrigerant issue, checking or setting a charge should be to get an accurate return (or box) DB, WB and RH as well as the outdoor ambient temperature and then working from there taking appropriate readings. When calling a senior tech or your manager please be prepared will all relevant readings to make a quick and correct diagnosis.

— Bryan

 

We have discussed DTD (Design Temperature Difference) quite a bit for air conditioning applications, but what about refrigeration? Let’s start by defining our terms again

Suction Saturation Temperature

Saturation temperature is the temperature the refrigerant will be at a given pressure if it is in the process of changing state. This change of state would be from liquid to vapor (boiling) in the case of the low side (evaporator / suction line). When we look at saturation temperatures instead of pressures we can use similar rules and we will see similar saturation temperatures across all refrigerants when the application is the same. Experienced HVAC and refrigeration techs pay far closer attention to the saturation temperatures than they do pressures.

Evaporator TD and DTD

Evaporator TD (temperature difference) is the measured difference between the suction saturation temperature (evaporator boiling temperature) and the box temperature. DTD (design temperature difference) is the designed or expected TD.

Delta T

Many A/C techs will confuse TD with Delta T. Delta T is the difference between the evaporator AIR temperature entering the coil to the air temperature leaving the coil. The Delta T will vary based on the humidity in the box where TD will not.

Target Box Temperature 

The temperature the refrigeration box should maintain when the system is operating properly

Superheat

The increase in temperature between the suction saturation temperature and the suction line temperature leaving the evaporator. Superheat is the temperature (sensible heat) gained between the point that all of the liquid boiled off in the evaporator coil and the suction line at the outlet of the coil. in refrigeration, like HVAC 10°F(5.5°K) of superheat  is average with a range from 3°F to 12°F(1.65°K – 6.6°K) depending on the equipment type (10°F(5.5°K) for med temp, 5°F(2.75°K) for low temp, 3°F(1.65°K) for ice machines ).

Hot Pull Down

Refrigeration equipment is unlike HVAC equipment in that the evaporator will spend most of its life running in a very stable environment with minimal fluctuation in the box temperature.

On occasion a refrigeration system will see a huge change in load in cases where it was off and needs to “pull down” the temperature, or when doors are left open or when a large quantity of warm product is placed in the box. When a piece of refrigeration equipment is in hot pull down it cannot be expected to abide by the typical DTD or superheat rules and must be allowed to get near the design box temperature before fine adjustments are made to the charge, TXV superheat settings or to the EPR (Evaporator Pressure Regulator) if there is one.

Design Temperature Difference (DTD)

In air conditioning applications a 35°F DTD is a good guideline for systems that run 400 CFM(679.6 m3/h) of air per ton of cooling (12,000 btu/hr). In refrigeration the DTD is much lower than in air conditioning.

There are several reasons for this but one big reason is the desire to maintain relatively high relative humidity levels in refrigeration to keep from drying out and damaging product. Keep in mind that NOTHING is a substitute from manufacturer’s data but here are some good DTD guidelines for traditional / older refrigeration equipment. Keep in min dthat the trend is toward lower evaporator TD on newer equipment.

Walk-ins  10°F(5.5°K) DTD +/- 3°F(1.65°K)
Reach-ins  20°F(11°K) DTD +/- 5°F(2.75°K)
A/C 35°F(13.75°K)) DTD +/- 5°F(2.75°K)

You then subtract the DTD from your box temperature / return temperature to calculate your target suction saturation. You can then use this target saturation / DTD and compare it to your actual measured saturation and DT once the box is within 5°F – 10°F(2.75°K – 5.5°K) of it’s target temperature to help you set your charge, TXV and EPR as well as diagnose potential airflow issues when compared with suction superheat and subcooling / clear site glass.

For Example –

If you have a medium temp walk-in cooler with a 35°F(1.66°C) box temperature you would expect to see a suction saturation of  25°F +/- 3°F

When doing a quick inspection of a piece of refrigeration equipment without gauges you can use this data to do the following calculation –

35°F – 10°F DT + 10°F superheat = 35°F suction line temperature +/- 3°F 

In this particular case logic tells us that the suction line could be no WARMER than 35°F(1.66°C) because that is the temperature of the air the refrigerant is transferring its heat to. However by the time you factor in the the accuracy of your box thermometer and line thermometer and the assumed saturation temperature you would still expect a 35°F(1.66°C) suction line temperature +/- 3°F(1.65°K)

For a -10°(-23.33°C) box, low temp reach-in you would calculate it this way

-10°F- 20°F DT + 5°F superheat = -25°F suction line temperature +/- 5°F 

Clearly, this is NOT the way to commision a new piece of equipment or to benchmark a system you haven’t worked on before, but it can give you a quick glimpse at the operation of a piece of refrigeration equipment without attaching gauges, especially on critically charged or sealed systems.

The best practice is to know the equipment you are working on, read up on it and properly log benchmark data the first time you work on a piece of equipment or during commissioning.

It should also be noted as Jeremy Smith pointed out, in recent years TD’s have been decreasing as manufacturers seek higher efficiency through higher suction and lower compression ratios.

This means that TD’s as low as 5 can be designed into some units but keep in mind… the suction line can still be no warmer than the box so as DTD drops so does superheat and the critical nature of expansion valve operation.

— Bryan

This tech tip was written by a friend of HVAC School, Brian Mahoney HVAC instructor at Western Suffolk BOCES/Wilson Tech. Thanks Brian!


The podcast on delta T for A/C the other day got me to thinking about the formula I learned in school about calculating the GPM of a hydronic system using a handy formula. We will be using the following values:

Td – temp difference of your supply vs return

Net boiler output(btu) use the boiler plate rating or get fancy and do an efficiency test and multiply your rated input multiplied by your efficiency rating. On an oil system, the unit could be down-fired.

It may be rated for 1 gallon per hour (140,000 BTU per hour input, but it may be firing with a .85 gallon per hour nozzle. So you have to do the math:
1 gallon of #2 fuel oil contains about 140,000 BTUs. Multiply that by .85 (your nozzle size) and you get 119,000 btu/hr input. Input would be 119,000 x .80 efficiency = 95,200.

500 – a constant which stands for a pound of water times 60 minutes – 8.33 x 60 = 499.8 (we fudge a bit.)

This is the weight of water at 60 degrees. You could look up the weight at the temp you are working with and multiply by sixty but it wouldn’t be far off.

To find a system’s gallon per hour:
BTU/ (500 x TD)
100,000/(500 x 20)
100,000 / 10,000= 10 GPH

Nice, but is there anything else you can do with this? How about a room that’s not warm enough. Is your baseboard supplying enough heat? You could look up the specs for that product, maybe. But what if it has dirty fins or mud in the pipe that is affecting temperature transfer. How would you know?

By using your Testo temp clamps on either end of the baseboard you find your temperature difference and using the data from the last calculation you solve for net BTU output of the baseboard

Btu = GPH x 500 x td
10 x 500 x 2 = 10,000 btu/hr

Now you know what you are getting. So you can check the specs of that baseboard and see if it’s giving you its rated output. If it is you don’t have enough baseboard or you have a problem with the room; thermal bypass for instance.

If it’s not performing as rated and the fins are clean you have an internal problem such as mud in the pipe insulating it.

Just something for the wet-heads.

— Brian M.

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