Refrigeration Case ControlsQUESTION:
From Steve Sloan
I am looking for factual information related to the energy savings resulting from the use of refrigeration case control boards with electronic expansion valves.
By Bruce Schneider
Danfoss has conducted a controlled test in a Buffalo, NY, store using Danfoss Smart Case Controller’s and our AKV10 electronic expansion valves (EEVs).
The test consisted of one two-stage and one medium-temperature rack operating 48 evaporators. A contractor field-installed EEVs in parallel with factory-installed thermostatic expan-sion valves. Solenoid valves were also installed to allow switching between the electronic and thermostatic valves.
The valves were alternated every second day to neutralize day-to-day load variations due to shopping, stocking, and cleaning activities. Test data were recorded for four months.
The average daily usage was found to be 703 KWh on thermostatic days and 639.4 on days when the EEVs and Smart Case Controllers were used, for a savings of 9.05%. Savings increased steadily as the weather warmed, from 7.5% in March to 9.9% in June.
Savings from 4.3 to 50.8% were realized from reduced compressor run time, reduced defrost duration, and improved recovery. Average defrost was quicker by 28.7% on one rack and 17.4% on the other. Also, the test revealed higher average suction pres-sures for all three suction groups tested.
It should be noted that the savings generated from EEVs are the result not of the valves, but from the software algorithms within the Smart Case Controllers. Algorithms also determine the stability of the fixture temperatures achieved. Both of these facts were proven by the test results. The entire test’s data collection and test report were audited by the public utility serving the store. Based on the results, a rebate program was established.
Finally, the test was documented in a three-part report, available from Danfoss.
By Mike Baird
Via the Internet
I would like to find information on the specific gravity and cubic feet of gas per pound of R-134a. Could you please let me know where I may get this information?
By Daniel Kramer, P.E.
Patent Attorney and Former Chief Engineer, Kramer-Trenton
The specific gravity of R-134a at atmospheric pressure and 60Â°F is 0.2753 lb/cu ft, or about 3.6 cu ft/lb.
I got this from a Klea program at www.dircon.co.uk/klea/downld.html.
It is free and provides all the technical and thermodynamic and cycle calculations you could want for both the old and new refrigerants. You can pick the language and the units and even print tables of the results.
From John West
My question involves refrigeration compressor racks using water-cooled condensers with an inside cooling tower. This is a very old system.
What would be the recommended Delta water temperature on the tower? What would be the maximum temp on the inlet of each circuit? What would be an easy method to confirm you are getting the correct tonnage out of these units?
By Daniel Kramer, P.E.
Patent Attorney and Former Chief Engineer, Kramer-Trenton
Let me try the easier part of your question first. You want to “confirm you are getting correct tonnage out of these units.” I would suggest that you want to find out what capacity the compressors are delivering.
To do this, you should:
First, copy down the model number of each compressor. Then get the rating curves for each from the manufacturer or dealer.
Second, make a chart with the compressor model on the left and columns for head pressure, suction pressure, compressor amperes, and rated capacity. Then, for the record, measure the drybulb and wetbulb temperature of the air entering your indoor tower, the water temperature off the tower, and the water temperature leaving each condenser.
Third, measure the suction, head, and amp of each compressor, entering the data carefully into the chart.
Fourth, for each compressor, locate the rating curve of your observed suction and head pressure (or corresponding temperatures), and read off the predicted capacity and amperes. If the amperes from the chart agree with your observed amperes, then the chart capacity is probably close to correct.
However, if your observed amps are much higher or much lower than the chart amps, you might have a compressor problem. For instance, on the old reciprocating compressors I’m used to, higher than normal amps suggested a cracked discharge reed and lower than normal amps, a cracked or broken suction reed, or more major compressor damage.
The compressor or rack manufacturer should tell you then maximum allowable entering water temperature into the condensers.
The performance of cooling towers is pretty complicated, since the driving force is not drybulb temperature, as in air-cooled condensers, but enthalpy difference between the entering air and the tower water temperature.
Since wetbulb temperature of air is almost perfectly proportional to air enthalpy, cooling tower ratings are generally specified in terms of 78Â°F entering wetbulb air temperature and water flow rate of 3 gpm per heat rejection ton (15,000 Btuh) cooled from 95Â° to 85Â°.
Cooling tower manufacturers sometimes publish charts allowing an engineer to estimate the water temperature drop at different entering air wetbulb temperatures and water flow rates and heat rejection loads.
Missing Data PlateCOMMENT:
From Wayne Shoemaker
Sr. Research Maintenance Technician
University of Missouri
I was unsatisfied with Paul Reed’s answer to David Cox in the Jan. 8, 2001 issue of The News relative to charging a system when the data plate is missing.
Just because a sight glass is present in a system fails to give adequate reason to fill the sight glass when charging. On many pieces of equipment I have found a warning label by the manufacturer to NOT fill the sight glass.
It is best to measure superheat at the suction outlet of the evaporator to ensure efficient operation and to measure superheat at the compressor to ensure a longer compressor life.
On small systems with short suction lines, one can measure suction pressure at the compressor and monitor suction line temperature at both the evaporator outlet and at the compressor.
Leaving 3Â° to 5Â°F of superheat at the evaporator may give up some efficiency but will ensure compressor life. Avoiding liquid floodback to the compressor is a major concern. Having 10Â° to 15Â° of superheat at the compressor is a good general goal.
On large systems with long suction lines, install a service line at the evaporator to allow measurement of suction pressure at the evaporator outlet. Again 3Â° to 5Â° of superheat may give up some efficiency; but what good is a 100%-efficient evaporator when the compressor has died due to liquid floodback? The goal should be 15Â° of superheat at the compressor where long suction lines are used.
On low-temp systems (0Â° to -40Â°), the superheat at the compressor may be higher than most techs are used to seeing, sometimes as high as 50Â°. In these cases, the best measurement is taken at the evaporator.
One might call this charging to performance. Performance is determined by measure of superheat at the evaporator to ensure all liquid is returned to vapor before leaving the evaporator.
Also, watch for systems that hot gas both as a method of defrost and as a method of temperature control. Run the unit in hot gas mode and measure superheat at the compressor to ensure no liquid is returning to the compressor while hot gassing. One must also check compressor amp load during hot gas to ensure the compressor motor is operating within design amp conditions, a.k.a full load amps (FLA ).
An overcharged system will overload during hot gas while running quite well during liquid cooling phase. If an overload condition exists during hot gas, then remove enough refrigerant to stay within the FLA rating. If superheat is too low during hot gas, then remove enough refrigerant to maintain safe superheat conditions during hot gas phase.
In each case you must trade away higher efficiency for longer compressor life.
Superheat can be measured whenever vapor is present in a refrigeration line. Simply measure line pressure and line temperature at the same point. Then find the saturated temperature for the pressure on a temp-pressure chart for the refrigerant in use. Finally subtract chart temp from the measured temp. If the result is less than or equal to 0, then no superheat exists and liquid does exist at this point. If the result is greater than 0, then superheat does exist and no liquid exists at this point.
Remember, you are breathing a superheated vapor. I know the oxygen and nitrogen we are breathing are both well above their boiling points. Superheat during all phases of operation can never be overstressed.
From Joseph Bebout
We service restaurants in the Panhandle of Florida and have been running into a problem with humidity. We’ve had heat indices of 110Â° to 115Â°F (97Â° and 90% rh). As we close the fresh air from the 25% to 10% on the a/c to increase cooling capacity, problems occur. There is sweat on the grilles and the ductwork above ceilings.
The systems are new, clean, and still have some fresh air makeup.
My boss said that I should not have partially closed the fresh air on the a/c unit because the hood system is balanced. He said that is why we have the humidity problem. He said I should open the a/c unit’s fresh air so we won’t have a negative pressure on the building.
Can you help me understand why this can cause a humidity problem?
By Daniel Kramer, P.E.
Patent Attorney and Former Chief Engineer, Kramer-Trenton
As you know, air conditioners cool the incoming air. When that cooled air is conveyed through ducts that are in a high humidity environment, they will sweat. Technically that means that the temperature of the cooled air leaving the air conditioner and traversing the ducts and the outlet grilles is lower than the dewpoint of the air surrounding the ducts.
There are several ways to cope with this problem.
The first and simplest is to let the air conditioner run until the cooled, dehumidified air discharged by the air conditioner has had a chance to lower the humidity and the dewpoint in the cooled space sufficiently. This can be determined when the dewpoint of the air in the cooled space is lower than the temperature of the air discharged by the a/c.
However, if there is a high humidity load in the restaurant from high patron density and/or infiltration of humid air from the kitchen or the outdoors, that solution will not always work.
The second solution is to insulate the sweating ducts and employ “non-sweat” grilles.
The third solution is to raise the temperature of the air flowing through the ducts and the outlet grilles. You could do this in several ways. One way is to open the fresh air dampers (as your boss suggested), allowing more of the hot, humid outside air to mingle with the cooler, drier air recirculated from the restaurant. This should raise the temperature of the air discharged by the a/c unit, probably above the dew- point of the surrounding room air.
Swimming pools and other humid applications which need little cooling but lots of dehumidification frequently employ electrical reheat in the cold, dry air discharged by the a/c. But I’m sure your restaurant has both a high humidity and a high sensible load. Therefore, I suggest that reheat would not be appropriate in your case.
While I have never seen it done, I suggest that you could employ a separate blower to draw warm air from the restaurant and blow it into the discharge air duct from your a/c. In this way, you would lose no cooling at all. You could reduce the outside air recirculation rate if you wished. The mixed temperature in the duct would probably be higher than the dewpoint in the room, thereby eliminating duct and grille sweating.
You would have to do a little psychometric engineering first. You would have to perform some air temperature and humidity measure-ments. You would have to know the quantity of air discharged by the a/c unit to determine the correct amount of recirculated restaurant air you had to mix with the colder air leaving the a/c unit to raise the mixed temperature in the duct above the dewpoint in the room. If the restaurant humidity was near 100%, you might get condensation inside the ducts and you would have to provide a drain at the low point in the duct.
From Anthony Lalli
I would like to know if there is a way to tell if a refrigerant is 100% leak-free. Normally, I always pull the system down to 400 to 500 microns and 29.5 hg vacuum. Usually there is a slight increase on gauges 40 to 60 minutes after the pump and gauge manifold valves have been shut. I have even performed this test with new hoses, gauges, and manifold, and experienced the same increase.
By Adam Seymour
TIF, Promax, Amprobe
There are several ways to leak-check a system, and your description of a vacuum leakage test is an excellent starting point. However, it is also recommended to check the system for pressure leaks once it has been charged with refrigerant, as it is possible that leaks may exist under pressures that are not present under vacuum.
With respect to a vacuum “rate of rise” check as you describe, there are, unfortunately, different opinions as to what is an acceptable rate of rise. For many years, an old “rule of thumb” was to pull to 500 microns and not rise above 1,500 microns. ASHRAE Standard Guideline 3, Sec. 220.127.116.11 states that a vacuum should be pulled to 300 microns, and an acceptable rise would be up to, but not over, 800 microns in a 30-minute period. The Trane Manual, Chapter 20, recommends pulling to 500 microns and seeing no rise for 12 hours. ‘No rise’ is probably more stringent than you can expect to see in most real-world applications. Some rise, as you describe, is almost inevitable.
The real key is how quickly the rise occurs and whether or not it stabilizes. A very quick and continuous rise is almost always an indication of a leak. If you see rise that then stabilizes, this is more often than not an indication that there is moisture remaining in the system which is boiling off and increasing the pressure (reducing vacuum).
The size of a system is also significant. The same size leak in a small cooler and a multi-ton a/c unit will cause much different responses. The smaller the system, the quicker the vacuum level will rise. In a very large system, a small leak could take hours to cause a noticeable rate of rise.
In conclusion, do not be deceived by small rises. These do not always mean leaks. What is important is that you do not see a rise above about 1,000 microns within an hour, and that the rise does not continue beyond that level.
From Sam Waldner
Mountain Lake, MN
Why is steel tubing used on so many condensers instead of copper or aluminum? It seems to me that copper and aluminum are better conductors.
By George Riggs
You are correct. Copper and alumin-um are better conductors than steel.
In the refrigeration industry, the standard for all manufacturers is copper tubing and aluminum fins in both condenser and evaporator coils.
Ammonia and copper don’t get along, so aluminum tubing often is used in applications with ammonia as the refrigerant. Sometimes, usually for environmental reasons, variations from standard may occur. In salt spray applications, for example, copper fins may be used.
The standard in refrigeration, though, is copper tubing and aluminum fins for good heat transfer characteristics and reliable long life.
I have a question about near-azeotropic mixtures. Are there any special considerations when calculating superheat or subcooling with these mixtures, since there is a temperature glide during change of state?
For example, I assume that using R-401A, you would look at your evaporator pressure and calculate superheat from the dewpoint corresponding for that pressure.
As for subcooling, I assume you would look at your condensing pressure and calculate subcooling from the bubble point corresponding for that pressure.
After talking to a few contractors, I found there is some confusion on this issue.
By Craig Barnett
Setpoint Heating & Air Conditioning
Both of your assumptions are correct. As the three components of a near-azeotropic mixture (ternary blend) move through the tubing of an evaporator coil, the component with the highest vapor pressure boils off first, followed by the component with the middle vapor pressure, followed by the component with the lowest vapor pressure.
In the last passes of the coil, the highest and middle vapor pressure components exist as vapor at the saturated temperature of the lowest vapor pressure component.
Somewhere near the outlet of the evaporator, the last droplets of the lowest vapor pressure component vaporize (at the dewpoint) and all three components exist in the vapor phase.
From this point on toward the compressor inlet, the components undergo the sensible heating process known as superheating. Thus, the near-azeotropic refrigerant with 10Â°F of superheat would exist at a location where the temperature of the refrigerant is 10Â° greater than the dew-point at a particular suction pressure.
Conversely, 10Â° of subcooling would exist at a location in the condenser or liquid line where the liquid refrigerant is 10Â° less than the temperature where the last few bubbles of vapor phase refrigerant are condensing into liquid (i.e., the bubble point at the particular high-side pressure).
From Albert Gallucci
New York, NY
With regards to commercial air-cooled systems in excess of 10 tons located in the Northeast United States, when is it necessary to incarcerate a pumpdown-type shut down in a split system?
Where is the proper location for the solenoid valve or valves for pumping down a split system?
When is a receiver required in a split system?
I have asked these questions to many service technicians and receive just as many answers.
By Daniel Kramer, P.E.
Patent Attorney and Former Chief Engineer, Kramer-Trenton.
I expect you will find a variety of answers to your questions, depending on whom you ask.
There are situations where large, air-cooled systems might not require a pumpdown. However, as a cautious engineer, I would always provide an automatic recycling pumpdown cycle to keep the pressure in the compressor low enough so that liquid refrigerant could not accumulate either in the evaporator or the compressor during extended off cycles.
Yes, others will say that crankcase heaters will keep liquid out of the crankcase during off cycles. That’s true, but heaters fail or get shut off. Also, even with crankcase heaters, refrigerant liquid can accumulate in the evaporator during off cycles, flooding back to the compressor on start-up.
Yes, automatic recycling pumpdown can generate compressor short-cycling. However, this is a clear symptom of leaking compressor discharge reeds, liquid solenoid, or oil separator float. Fix it.
Yes, auto-recycling pumpdown demands that the service person understands the correct law pressure switch settings and their logic.