A persistent, methodic, and patient person will find a leak in a refrigeration system before anyone else. It may seem like it takes a long time, but it doesn’t hold a candle to the time involved with repeated evacuation and pressure testing, especially if the system is very large.
The following is an outline of how to approach leak testing to save time and effort.
1. Have a planned approach. Use a sketch of the system to mark the progress as you go. Start at one point and work along the system path, this prevents repetitive checking. Mark the sketch up as you go, for instance, note a section you can’t get at to test or got small hits, because once you eliminate the known area, the leak must be in the unknown area.
2. Know the tricks. Use tape to isolate flanged connections, wrap the connection and poke a small hole, this way you are testing one point rather than the entire flange. Another trick, test insulation at seams, the refrigerant will propagate under the insulation until it finds a way out. Look for oil, if there is oil stain there is a leak at some point. Put balloons over relief devices and plug the weep hole to find a slow leak valve. Put condenser or chilled water straight from the chiller in a closed container and let it sit, after a while use an electronic detector to sense if any residual refrigerant is in the container, could be a tube leak.
3. Use tags, chalk, or any removable marker to note leaks on the system as you go. Do not use paint or permanent markers because once the leak is repaired, the tag or mark needs removed.
4. Pick the right tool for the job and know how it works. There are many leak detection tools, and they all work, but there isn’t any one tool that will do it all.
I have listed the leak detection tools that I’m aware of below, they are not in any order, just as they occurred to me.
Ø Bubble solution: This can be homemade from dish-soap and water, or, commercially available, such as SNOOP®. It works by coating the area with a high surface tension fluid, when the gas seeps out of the pipe it is momentarily captured by the fluid and a bubble forms then pops. This is a tried and true method of testing for leaks. Couple draw-backs; it can freeze, and you do get liquid all over the place when you use it.
Ø Foaming shave cream: When applied to a leak, the gas will leave a little hole in the cream or blow it off completely. It also works well to find leaks on long seams or where a solution won’t hang long enough. Can find leaks on vacuum or pressure. Draw-backs; Is messy and short lived sometimes drying up or breaking before small leaks are found. Aerosol cans may not be permitted on premises and it isn’t good if its windy out.
Ø Ultra-sonic leak detectors: Several are available and all generally work to “listen” for the high frequency sound that occurs when gas leaks out. They are used by narrowing the search area based on the pitch, volume, or number of beeps increasing as you search. These can help isolate a leak in a wall. Draw-backs; they don’t work well in noisy or windy environments and lack the ability to “pin-point” a leak. If there are numerous leaks, you will hear them all.
Ø Sniffers: Or, electronic gas detectors, are designed to pick up trace amounts of gas in an air stream. They work by using a little air pump to pull samples across an LED light, a receiver, tuned to the specific frequency of the gas, compares the sample to its reference and alarms when it gets a match. These can be very accurate and are usually easy to use, hand held, and fairly rugged, but they have some draw-backs; get water or oil in the end and your done until you replace the filter. They will pick up false hits if you move too fast or get too close. They take a second or so to register a leak, so move slow.
Ø Halide and sniffers that use heated elements: they work, but do so by burning the sample. Draw-back; I avoid these since you don’t always know if there is a flammable atmosphere which could lead to an explosion.
Ø Helium leak detectors: These detectors work by identifying helium atoms in a sample stream, you can’t fool them, at least not ones that work on gas chromatograph basis. They will find the smallest leak possible. Draw-backs; the detectors are expensive, you need a supply of helium to add to the system, they are costly to maintain.
Ø Thermal imaging leak detectors: These are relatively new to the industry. Based, in part, on the fact that there will be a temperature drop associated with the leak. There is an image of the equipment you watch through a screen looking for the temperature zone that’s out of place, there’s the leak. Draw-backs; they are very expensive and costly to maintain/repair. If you have ever tried to take a picture of every aspect and all sides of a system, then you know how challenging these would be to maneuver in a congested mechanical room.
There are all kinds of ways to find leaks and new ones are invented every day. If you know of a clever way to spot a leak, please let me know!
Stall is when a centrifugal compressor reaches a point of maximum flow and minimum head. The wheel can deliver no more gas. (For a description of the centrifugal wheel and head, see “Centrifugal Refrigeration Compressor Surge”.)
Stall is common in aircraft engines, turbo-chargers, and other turbo machinery that operate in an open loop, but much less likely to occur in a closed loop refrigeration machine.
A theoretical wheel curve is displayed bellow, the far-right line labeled “stall”, is at the operating conditions the compressor must reach to induce “stall”.
For a refrigeration machine to reach the conditions at stall on the graph above, there would need to be almost no head across the compressor, and a flow, far exceeding the design.
Refrigeration machines are closed loop systems, take a vacuum cleaner and stick the suction nozzle into the discharge port, wa-la, closed loop system, if you kept the motor cool, the vacuum would run and run and never see a stall condition.
Now, I agree, not a fair comparison, vacuum cleaner and refrigeration machine, but the point is clear, without a mechanism to induce more volume flow into the loop, the flow can’t exceed what the compressor is capable of.
It is true, that a great deal of volume can be generated by the boiling of the refrigerant in the evaporator, and it does, but not more than the machine is designed to handle.
I guess you could have huge evaporator that, at times, gets overloaded, and is piped up to a compressor that runs forward on the curve, with a ginormous condenser that sits in the arctic; I just have never come across such a design or miss-applied machine to witness this perfect storm. However, if you have I would love to hear about it.
Centrifugal refrigeration compressors have an operational condition called “surge”. Surge can be identified by a loud squalling sound (shaft is thrusting against the physical stop or bearing) followed by a whooshing sound (the sudden reversal of flow of refrigerant) in a repeated sequence. Surge will eventually cause extensive damage to the internal components of the machine if not corrected. To help understand the surge condition, it will help to understand compressor “head”.
Centrifugal compressors impart the energy to the refrigerant by spinning bladed wheels at a relatively high speed. The energy imparted to the refrigerant develops a new potential in the gas called head.
The wheel diameter, speed, refrigerant density, and other physical design attributes affect the total head possible from any wheel design. These factors, for the most part, are fixed once put into place, somewhat fixing the surge point to a narrow range of operation.
The energy (or head), the wheel imparts on the refrigerant is responsible for the rise in pressure and temperature of the refrigerant during compression.
Head is measured in foot, the energy used to raise one pound of gas one foot, is one foot of head.
The head a compressor is developing can be found on the pressure-enthalpy diagram. The diagram below shows a compression curve for a R-22 compressor on a pressure-enthalpy diagram
At the compressor inlet, the red line is drawn down to the enthalpy of about 178, and the discharge enthalpy is about 188. The difference between these two numbers is a specific enthalpy of 10. We use 10 Btu/lb of gas multiplied by Joule’s constant, 778.16, to obtain a head of 7781.6 foot.
Varying the temperature of the inlet or discharge will directly affect the amount of head the compressor will produce.
Surge conditions will be in effect when the compressor is asked to exceed the design head. The point it begins is system independent and must be determined by testing actual components in operation, but most would agree, a 50% increase over design head will likely put a centrifugal into surge. This 50% mark is usually enough because the curve a wheel makes when plotting head versus flow, looks like the one below, the wheel is selected based on its pressure and flow requirements, then different wheels are looked at to obtain reasonable efficiency, the best efficiency can be seen close to the top of the curve where the surge break-over point occurs.
That means, if we ask this compressor to generate 11672.4 foot of head, instead of the design head of 7781.6, the compressor will likely surge. That seems like a lot, but that is only increasing the specific enthalpy of compression from 10 Btu/lb to 15 Btu/lb, and you’re in surge territory.
So, how does this happen? Those curved lines are temperature lines, in ten degree increments, the current discharge temperature is 160 F, enthalpy 188, an increase of 5 Btu/lb to 188+5 = 193 is an increase in temperature from 160 F to 175 F would push it to 193, this could easily occur if there was an increase in the heat going into the compressor.
The opposite is true also, a decrease in the compressor inlet temperature by 15 degrees will also push the enthalpy down 5 Btu/lb leading to surge. Conditions such as low load and high condensing pressures will get there very quickly.
The value here, is by charting your compressor operating conditions against design conditions, you can readily see what you need to address to keep the compressor out of surge.
Note on “design conditions”, if the design conditions are not known, you can use conditions where you know the machine runs normally (which is hopefully known) or refer to AHRI standards (on the links page) manufacturers stick pretty close to these standards.
All refrigerants have a pressure-enthalpy diagram. In short, the diagram shows property conditions of the refrigerant at various stages of state.
The properties inside the “thumbprint” shape is “saturation” properties (mixtures of liquid and vapor). The area to the left of the thumbprint are properties of liquid refrigerant, and to the right, properties of vapor refrigerant.
The left vertical axis is the index for pressure (psia, pounds per square foot atmospheric). The bottom axis is the index for enthalpy, or energy, the capacity of the refrigerant at that stage. Other lines, temperature (shown), volume (not shown), entropy (not shown), and quality (not shown), allow for the state of the refrigerant in a cycle to be plotted to find unknown values that can be used to calculate energy, cooling capacity, efficiency, horse power, and many more valuable bits.
Using the pressure and temperature from a system the cycle below was plotted. I want to know how much horse power I should expect the system to be using.
The two red lines in the diagram are drawn to be straight drops for the intersection of the cycle point, compressor inlet, and compressor discharge. The lines cross the enthalpy line at 177.5 and 188.5 (approximately).
By subtracting 177.5 from 188.5, I arrive at a specific enthalpy of compression of 11 Btu per pound of refrigerant compressed per minute of operation.
The formula to find our horse power is:
42.43 is a mechanical constant, the mass flow of refrigerant we will find next.
Let’s say we are working on a 2-ton air-conditioner, and since 200 Btu per minute equals 1 ton of cooling, our total cooling per minute needed is 200 X 2 = 400 Btu per minute.
Since each refrigerant, and set of conditions, have a specific cooling capacity called the “net refrigeration effect” we must find the specific cooling capacity of the system plotted above.
To do this, we need to drop two more lines as below:
The lines intersect the enthalpy scale at 107.5 and 175.5 (approximately). Once again, we take the difference of these two numbers, 175.5 – 107.5 = 68 Btu per pound.
68 Btu/lb is the “net refrigeration effect” for our system, or, the amount of cooling capacity per pound of refrigerant flowing.
To find the mass flow of refrigerant, in pounds per minute, we divide 400 (Btu per minute of cooling for the total system) by 68 Btu/min (net refrigeration effect).
400 / 68 = 5.88 pounds per minute is being circulated to achieve the 2 tons of cooling in this system.
Now that we have mass flow, 6.88 lb/min, enthalpy of compression, 11 Btu/lb, we can work the first equation:
(6.88 X 11) / 42.43 = 1.78 horse power.
Knowing that our system should use 1.78 horse power, we can now determine if the unit is overloaded, under loaded, drawing too high of current, etc.
Rapid dehydration can accomplish the removal of moisture and in a fraction of the time traditional techniques require by maintaining the pressure during dehydration at 6000 microns or about 37 degrees.
The process provides a means of “sweeping” the system with a dry gas that promotes the movement of moisture while maintaining an atmosphere that will vaporize water.
Required components: Vacuum pump, Rota-meter with needle control valve scaled to about 10 SCFM, source of nitrogen capable of 1~1.5% by volume of the pump, 6 cfm pump/~.06-.09 cfm of N2, 300 cfm pump/ 3-5 cfm, and an electronic vacuum meter of good quality. See below:
CAUTION: ALWAYS USE AN APPROVED NITROGEN REGULATOR, A NITROGEN BOTTLE IS AT 3000 PSI!
The regulator pressure must be set at the inlet design pressure determined by the Rota-meter, which can be found at the Dwyer website.
Installation of the components: There is little criticality to the orientation of the devices but placing the nitrogen injection as far from the pump nozzle produces better results.
Find a “neutral” point to attach the electronic vacuum gauge. Look for a port that is about half way in between the meter and pump connection. The preferred spot should also be low traffic and safe for the electronics in the meter. Avoid low points also, if water or oil gets into your vacuum meter it may be finished for good! I also suggest using a cold trap and filter to protect your vacuum pump from destruction.
1. Pull an initial vacuum on the system until the electronic vacuum gauge reading is essentially stable. This reading should be somewhere below 5000 microns. If the pump is suitably sized for the system, the approximate time is less than an hour. If the system will not achieve a vacuum below 5000 in this amount of time, the pump may be too small, or faulty, or a leak still exists.
2. After reaching a stable operating point, set the flow of the Rota-Meter to about 1% of the volume flow of the pump. Monitor the vacuum gauge until the reading is stable. Adjust the nitrogen flow to establish an environment of ~6000 -8000 microns.
3. For general dehydration the system should remain stable and dehydration will be complete when the gauge ceases to drop. Constant fluctuation and necessary adjustments to the nitrogen indicate extreme quantities of water and/or leaks. The required dehydration will be two to three times longer. Constant monitoring of the system will be the only way to determine termination of the dehydration. If the oil mist ceases to purge from the pump exhaust, it is a sign that the process is very near complete.
4. Repeat the pull-down in step 1 and compare it to the chart below. If the reading maintains a steady trend for one hour below the line in the chart, the system is dry.
If the vacuum readings over time increase sharply, there is a leak, if they trend up slowly or cross the line, it is water in the system.
Below is a vacuum conversion chart.
Once super-heated refrigerant gas leaves the compressor it travels through the suction line to the compressor. The compressor in this example is a reciprocating compressor and the cycle can be referred to as the vapor compression cycle.
To compress the vapor, it must be put in an enclosed container, then, the container must be made smaller, forcing the molecules of the gas closer together, or, compressing them.
A few things change when you compress a gas, the volume goes down, the pressure and the temperature go up.
If I took a frictionless compressor and put 1 cubic foot of R404A in it and compressed that gas to .33 cubic foot the following would be the outcome:
Compressing the volume in a ratio of 3 to 1 (1/.33=3), the pressure increased 110.5 psi (pound per square inch), and the temperature increased 72.5 degrees Fahrenheit!
This is a pressure ratio of 3.17 and a temperature ratio of 2.65
Considerable change in the properties for a relatively small compression ratio.
Notice the ratios are not the same, volume = 3, pressure = 3.16, and temperature = 2.65. These properties change based on the physical aspects of the individual refrigerant and will be different for every compression process.
A basic reciprocating compressor will have the piston and cylinder just like the example above except with a few more components to put work into the system.
The compressor needs a suction valve to open when gas is coming into the cylinder, and a discharge valve that opens when gas is leaving the cylinder.
The valves are reeds or spring loaded disks, reeds being more common due to cost. Below is a sketch of the example compressor with the suction and discharge reed valves.
Reed is a flat piece of spring steel which lays flat over the suction and discharge ports. As the picture indicates, the valves open when there is a difference in pressure, they spring closed when the pressure is equal.
If you lose the suction or discharge valve, the suction and discharge pressure will run about the same.
If you lose the rings in a compressor, but not the valves, oil sump pressure spike, discharge pressure will be low.
The refrigerant leaving the metering device is at a much lower pressure than the liquid line delivering it to the metering device. The liquid is pushing forward into the evaporator coil and rapidly expanding (hence the term “expansion-valve”) into droplets and vaporized refrigerant.
The droplets shed down the tube walls picking up the heat energy from the external load (heat) being applied to the exterior of the evaporator coil. As the droplets absorb enough energy (latent heat of vaporization) they transform into vapor (boil). The refrigerant will be in a state of “saturation” (equal parts of vapor and liquid) until the last drop of liquid is boiled away.
The pressure on the low-side is dictated from the vaporization process and will remain constant while the continuous boiling occurs. This is the saturation pressure and it has an associated saturation temperature that goes with it. If the low side, (suction) pressure is reading 92.7 (~93) psig on your manifold, and the refrigerant is 404A, then you know the evaporator is operating at 44 F per your TP (Temperature Pressure Chart).
But now you put your thermometer on the suction line leaving the coil (because you usually can’t check the coil directly) and the temperature is 54 F, what up with that?
Coils are designed with a function called “super-heat” built in. It is a section of coil exposed to the load that allows the refrigerant gas to be further heated up by the heat available in the load. This function, the super-heat function, protects the compressor from liquid flood back, a dangerous condition that will severely damage the compressor.
As a note, here, expansion valve manufacturers usually pre-calibrate new thermostatic expansion valves to operate at 10 degrees Fahrenheit of super-heat, I suggest you don’t adjust it unless you REALLY, REALLY, know what you are doing.
If you could measure the evaporator it still wouldn’t read 44 degree directly since you must account for all the interference from the blower/heat-load interacting with the coil surface, a likely reading would be between 46 F and 54 depending on where you take it and how well the reading is insulated from the process. Manufacturers use a device called a “thermos-well” to obtain an approximation of the coil temperature when needed. The thermos-well is a tube inserted into the refrigerant line which is partially protected from the load and can readily sense the coil temperature with a thermocouple or temperature gauge. (See below.)
After the refrigerant leaves the condenser, (covered in an earlier post), the refrigerant is delivered, via the liquid line, to the metering device.
The metering device has two jobs, provide the necessary pressure drop, and provide the correct amount of refrigerant to the evaporator to handle the design load.
‘Design’ is in bold italic because the metering device, no matter the type, is matched to the load conditions and will only perform properly in a small operating window.
I like to lump the metering devices into two categories, fixed metering devices, and dynamic metering devices.
Fixed metering devices operate in a very small range and are extremely susceptible to operating conditions. The basic design of a fixed metering device is based on a restriction of the flow causing a pressure drop in the line while maintaining enough of a through put that some refrigerant makes it into the evaporating coil.
Fixed metering devices are inexpensive solutions but also inefficient and have limited capability. You will likely see a decline of fixed metering devices as the demand for more energy efficient design rises.
Fixed metering devices do not respond to the load per say, only to the differential pressure between the condenser and the evaporator.
Some fixed devices rely on an orifice, or ‘seat’ and an adjustable ‘trim’ which is a tapered stem that, when held at the seat, creates a restriction that is adjustable.
Another type of fixed metering device is the capillary tube, a very common type of metering device in small appliances. The capillary tube is a long length of tubing with a small diameter bore inside. The small bore and length, the bore determines how much refrigerant will pass, while the length, determines the pressure drop. Word of caution, if you cut a capillary tube, its ruined, the length and continuous, unobstructed bore, are critical! Patching it back together will change the characteristics of the tube, usually resulting in a very different performance.
A third type of fixed device is an orifice tube, commonly found in cars. The orifice tube is a metal tube that sometimes contains a removable inlet screen and a plate with a very small hole in it (the orifice), the very small hole controls the amount of refrigerant and causes the needed pressure drop. Due to the itty-bitty hole, the orifice tubes plug up easily. Best way to spot this is to look for frost just after the tube or temperature below the expected evaporator temperature. (Assuming the unit has the proper charge in it.)
A fourth, yes there’s another, is the piston orifice, the piston contains the small hole just like the plate in the orifice tube, but the piston is free floating in a large close coupled nut assembly. The piston moves forward during operation, then relaxes when flow stops, this action allows for the system to equalize the high and low side pressures lowing the compressor starting torque. Warning, these assemblies are usually brass brazed onto copper tube, they will gal, strip, cross thread, and twist the tube if not properly serviced, never force it together, keep it straight, do not over-tighten, and ALWAYS use a backup wrench to hold the swivel nut!
Below are some sketches of the basic configurations of fixed metering devices:
Dynamic Metering Devices try to adapt to changes in load by one of several means, pressure, level, or temperature.
Pressure controlled metering devices are called automatic expansion valves. The valve is an adjustable spring loaded diaphragm against the evaporator pressure with a tension that corresponds to the middle of the expected range of operation. When the evaporator pressure fluctuates with load, the valve can open or close to compensate within a narrow range. These were found a lot in old window, PTAC, and purge units, and look like right-angle valves with a silver tank on top, the adjustment was usually a slotted screw sticking out the top. Note; if it leaks around the screw, its bad.
Metering devices that respond to level changes are used only in flooded evaporator arrangements and are usually configured with a float to respond to liquid level changes. (There were some electronic level sensors but they never became very popular.) The float is installed in a ‘side’ chamber of the evaporator, one for serviceability, and second, for level stability. As the level went up or down the float would move up or down and either feed refrigerant in, or cut it back. Floats could also be used to send a signal to a pneumatic or electronic controller that could actuate a remote valve to achieve the same purpose. The pivot pins in these designs wear very quickly and periodic replacement is a must. Most side chambers could be isolated and re-built, but some units must have the entire charge removed to work on the float.
Temperature controlled metering devices are called thermostatic expansion valves and are used in air-conditioners, spot coolers, window units, chillers, refrigerators, freezers, etc. The most common temperature controlled expansion valve (TXV) is controlled via a sensing bulb full of a captured refrigerant charge. The bulb senses the temperature of the suction line. An increase in suction temperature opens the valve and increases the refrigerant flow to the evaporator, a decrease in the suction temperature cuts back the valve due to the expansion and contraction of the captured refrigerant in the bulb and power assembly. There are both balanced (pressure compensated) and unbalanced (spring only compensated) valves.
EXV, or electronic expansion valves, are thermal expansion valves with motors attached instead of power assemblies. They have a controller that senses the super-heat and adjust the valve accordingly. EXV valves can modulate the refrigerant flow very accurately and are the most efficient solution for many applications.
First, let’s hit condensing. In a refrigeration machine, the condenser, has three jobs, de-super-heating the hot gas from the compressor, condensing the refrigerant back into a liquid, and sub-cooling the refrigerant to maintain a solid flow of liquid for the metering device (expansion device, TXV, capillary tube, EXV, etc.)
De-super-heating: The gas leaves the compressor at the highest temperature in the system. The excess heat from friction and compression must be removed to approach the saturated (condensing) temperature. This is usually done in the first section of the condenser. In the illustration below, from point 2 to point 3, the fan, forcing air across the condenser, lowers the refrigerant temperature to 100 Fahrenheit from a discharge temperature of 160 Fahrenheit.
The lowering of the temperature by removing the heat from the refrigerant is a ‘sensible’ cooling process where no change in state of the refrigerant occurs (it’s still 100% vapor).
Condensing: Once the refrigerant reaches its saturation temperature, the refrigerant condenses, during which, the energy being removed from the refrigerant is spent turning the refrigerant into liquid, the temperature and pressure do not change during the condensation phase. The condensation phase is a latent cooling process. The refrigerant exists in the condenser in varying states of liquid and vapor until only liquid remains.
The resulting condenser pressure is the saturation pressure and dictates the pressure of the ‘high side’ of the system.
Sub-Cooling: When only liquid remains, another sensible process can begin, it is called sub-cooling. Sub-cooling is further lowering the temperature of the liquid refrigerant below its saturation temperature, the pressure, does not change.
The TP table, temperature pressure table available free from most any refrigerant manufacturer website, is a list of the saturation pressures and temperatures for individual refrigerants. (There are also apps for that!).
In the above example, the refrigerant is 404a, and the condensing temperature was 100 F, the pressure would be 235 psig (pounds per square inch gauge or what your manifold would read). See below:
However, if the gauge was reading, 218 psig, then out condensing temperature would be 95 F.
Say the pressure is 218 psig, and you read the line temperature leaving the condenser and it is 100 F?
The higher than saturation temperature at the liquid line leaving the compressor indicates a problem, could be under-charge, air in the system, or some reason the unit isn’t condensing the refrigerant (double check your temperature reading device against boiling water to make sure it reads correctly should read 212 F in the water).
So, now what if the temperature is reading below 95 when the pressure is 218 psig? That would mean there is sub-cooling occurring, when the liquid is being cooled below its saturation temperature.
If we see about 90 F on the liquid line, the condenser is probably doing its job, marginally anyhow.
Condenser sub-cooling range varies based on design but it is required to have enough sub-cooling to maintain a liquid full liquid line regardless of the pressure drop between the condenser and metering device (expansion valve).
The typical range will be from 5 to 15 degrees depending on the type of metering device, with 9-12 being common for TXV and EXV systems (always check with manufacturer data) and higher for capillary tube systems (which vary with load considerably)
Low sub-cooling ranges are usually reserved for close couple units where there is no appreciable pressure drop.
The sub-cooling is always taken just after the condenser, if there is a receiver or suction heat exchanger the temperature after those devices indicates how they are working not the condenser.
Basic Vapor Compression Cycle, the Carnot CycleFour main elements are required to complete the basic refrigeration cycle, a compressor, a condenser, a metering device (expansion valve in this case), and an evaporator, all labeled below in the system diagram along with connecting components.
Below is what the perfect system, Carnot system, looks like compared to a practical system on a diagram called a “Pressure-Enthalpy Diagram”. (Enthalpy, for this brief introduction, can be considered energy).
The numbers in the above diagram coincide with the description numbered 1-8 above for each stage of the cycle and the refrigerant condition.
It stands out, that the practical cycle is more outside the saturated region and rises above and below the Carnot cycle. The differences are driven by the efficiency of the physical equipment and materials used to apply the process.
The lower line, dubbed 7, is the evaporator line, it is lower because the material used in evaporators isn’t 100% effective at transferring the heat. Evaporators must be designed below the cycle temperature that is wished to achieve.
Same as for the condenser, location 3,4, and 5, but in the opposite direction, condensers must run hotter than the theoretical cycle to work.
From point 1 to point 2 is compression, on the Carnot cycle, the compression takes place with no losses, on the practical cycle, it extends out past the Carnot due to inefficiency that generates heat and volume losses.
Point 5 must extend into the liquid region to ensure that the expansion valve function correctly.
At point 6, The Carnot cycle would deliver liquid at 100% quality to the evaporating cycle, where the practical system must accept liquid at a quality of 12-18% (vapor fraction) this reduces the over-all refrigeration attainable per cycle, which is less efficient.
That wraps up the basic refrigeration vapor compression cycle. More details will be explored in later posts.
I worked for over thirty years in the HVACR industry. I have designed, installed, serviced, and trouble shot units of various types throughout the years. The posts here are information based on that experience, I hope you find them useful. If you have a different experience, please comment.