I’ll start this discussion with a plot of the system design perimeters. I do this a lot because if you don’t know how the system is supposed to operate, you can’t know when something is broke! Our unit is an air cooled 134a cycle with 10 degrees of superheat and 9 degrees of subcooling. The saturation temperature of the evaporator is 45 degrees Fahrenheit, and the saturation temperature of the condenser is 120 degrees F. The system capacity is 36,000 Btu/hour, or three tons. Below is the cycle plotted and some of the system characteristics calculated from the information on the diagram. There a varying degrees of low system charge, very low system charge will freeze the coil up and have ice hanging off the suction line. This condition doesn’t require much diagnostic effort, if the filter is okay, and the blower is okay, then low charge is probably it, and then you find the leak. The low charge condition that’s tough to find is one that leads to capacity problems with no other apparent symptoms. This condition leads to very high energy costs and should be corrected as soon as its determined. Refrigerant mass flow multiplied by the heat of compression and divided by 42.43 is the compressor horse power, so, (9.8 X 35)/42.43 = 8 horse power. To find the compressor amp draw, we multiply horse power by 746 (watts per horse power) and divide by the (volts X Efficiency X power factor). Efficiency and power factor can be considered 1 for most circumstances, so, (8 X 746) / 220 volts = 27.4 amps we would expect to see on a singlephase unit compressor. Next, we plot the low charge conditions against the design conditions. The plot will look like low load or low ambient, but one thing to remember is that low load and low ambient have ample cooling capacity, where low charge everything is low (pressures, capacity, subcooling) but two characteristics, the heat in the compressor and out. The superheat with an expansion valve system will be elevated depending on the extent of undercharge since the valve will try to compensate for the high suction line temperature. On fixed metering devices, such as pistons and capillary tubes, the superheat will be notably high. I have plotted a cycle of the same unit during the same operating conditions but the unit is under charged. The subcooling dropped to 1 degree and the suction superheat bumped up to 13 degrees. The compressor superheat (discharge temperature minus the saturation temperature) also increased from 108 to 111 degrees.
The evaporator enthalpy dropped to 60 Btu/pound lowering the net refrigeration effect and the inlet volume to the compressor changed to .99 cubic foot per pound. The volume of flow, 8.7 cfm, calculated on the first diagram, is the compressors physical displacement because that was the design flow at the design inlet volume. If we divide the fixed volume of 8.7 by our new suction density of .99, the mass flow is reduced to 8.7/.99 = 8.78 pounds per minute of refrigerant flow. 11% lower. The new mass flow multiplied by the lower evaporator enthalpy, 8.78 X 60 Btu/pound = 527 Btu/minute. There is 200 Btu/min in a ton of cooling, so, 527/200 = 2.6 Tons of cooling 14% less cooling than the 3ton design. That’s enough to cause problems. The compressor is running hotter because the gas coming back to it is hotter. The pressure ratio dropped down some so there is a little less work going on. The heat of compression, 33 Btu/pound, multiplies by the new flow, 8.78 pounds per minute, divided by 42.43 is 6.8 horse power. Finding the new motor current, (6.8 X 746)/220 = 23 amps, about 5 amps less than the fully charged system.
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Hot gas bypass (also called discharge bypass) is a feature in a refrigeration system uses to satisfy the mechanical needs of the system during low load conditions. Low load conditions can lead to frosting or freezing the evaporator, refrigerant flood back to the compressor, system shutdown, oil return problems, and several other undesirable conditions. The hot gas bypass utilizes a line from the discharge of the compressor to deliver high temperature, high pressure refrigerant either to the evaporator distributor (preferred) or the suction line to the compressor (which may overheat the compressor). The components usually are a solenoid valve and a pressure/flow control. Larger systems will likely only have an automatic valve, but no solenoid valve. Below, is a system sketch of a basic discharge bypass line on a simple aircooled system. Item 1above, is the solenoid valve, the valve is to prevent the addition of discharge bypass gas when it isn’t wanted. The solenoid valve is usually controlled by a pressure sensing switch which activates if the suction pressure drops below a certain pressure, for airover evaporators this should be the saturation pressure that coincides with 36 degrees Fahrenheit. The valve could just as easily be controlled by the evaporator temperature. Item 2 above, is the pressure regulator/flow control. The valve is a springloaded regulator that monitors the line pressure downstream of the valve and opens, or closes, to maintain a preset pressure against the internal spring. Some have sensing lines, some don’t. They are set to the normal operating pressure of the suction line. Item 3 above, is the distributor. The distributor is installed after the expansion valve and “distributes” the refrigerant to the various loops of the evaporator for even cooling across the coil. Some distributors have a tap specifically for the hot gas, others will simply be a tee in the line prior the distributor itself. As mentioned before, but not shown on the sketch, the line can be tapped into the suction line somewhere before the compressor (not recommended). The location should be as far from the compressor as possible. Air cooled systems that are going to operate during low load and/or low ambient conditions should have a head pressure controller or fan cycling control capable of keeping the condenser pressure at a saturation temperature of 80 degrees Fahrenheit. The system will not function correctly without the control. Now that we have a description of the system we can add the details. The refrigerant is R134a designed to operate at a 120degree saturated condenser and a 45degree saturated evaporator. There is 9 degrees of subcooling and 10 degrees of superheat. The design load is 1 ton, or 12,000 Btu/hour. We know when the ambient temperature is below 70 F, the system will be expected to maintain half the design load, or 6000 Btu/hour. First, I will plot the two operating conditions without the hot gas. So, the blue cycle is the design cycle, and the pink is the cycle at low load. The low load cycle evaporator pressure is below freezing (41 psia/ 30 F). This will cause the evaporator to freeze up and slug the compressor with liquid refrigerant. Hot gas works by modifying two conditions in the evaporator, one, it raises the saturation pressure above freezing, and two, it displaces some liquid refrigerant that does the cooling. So we next need to determine how much hot gas we need to displace 6000 Btu/hour of cooling. The design cycle has a refrigeration effect of 61 Btu/pound of refrigerant. The 12,000 Btu/hour design divided 60 minutes, is 200 Btu/minute. Divide 200 by 61, and the mass flow of refrigerant per minute is 3.28 pound per minute. This is the design flow rate. The flow at low load depends on the volume change of the refrigerant at the compressor inlet. For the design cycle the volume is .89 cubic foot per pound, or .89 X 3.28 pounds, 2.93 cubic foot per minute. For the low load inlet conditions the volume is 1.19 cubic foot per pound, so the compressor design cfm, 2.93, divided by the low load specific volume, 1.19, gives 2.46 pounds per minute flow of refrigerant at low load. (Volumes can be found on standard P&H charts). The refrigeration effect at low load is 72 Btu/pound. Multiply by the flow, 2.46, and the system capacity is 177 Btu/minute. As stated earlier, the system needs to operate at half load, or 6000 Btu/hour = 100 Btu/minute. We need to apply 77 Btu/min (177 – 100) of hot gas bypass to reduce the amount of liquid available for refrigerating with discharge gas. What happens when we inject hot gas into the evaporator inlet is we displace liquid droplets that are capable of cooling, the result is a change to the “quality” of the gas in the evaporator, and in turn, a lowering of the refrigeration effect. The specific enthalpy of the gas at the discharge of the low load compressor is 76 Btu/lb. We need 77 for the evaporator, or 77/76 = 1.01 pound per minute of hot gas bypass. Another advantage to this approach is that hot gas diverted from the discharge isn’t fed into the condenser and expansion valve so less liquid refrigerant is available for the evaporator. So, we now have 1.01 pound per minute of hot gas and 1.45 pound per minute of liquid feed to the evaporator at an unknown quality. We can approximate the quality using a mixing equation to find the new quality; {(Mass1 X Quality1) + (Mass2 X Quality2)}/ (Mass1 + Mass2), {(1.45 X .15) + (1.01 X 1)}/ (1.45 + 1.01) = .499 is the new quality of the gas in the evaporator. This looks like the process plotted below. Changing the quality of the gas reduced the refrigeration effect to 41 Btu/pound. We multiply 41 by the new evaporator flow of 2.46 lb/minute we get 100.86 Btu/minute or 6052 Btu/hour, pretty much half the load.
The introduction of the gas at the prescribed rate will also raise the line pressure to the regulator set pressure of 46 psia (36 degrees F saturation temperature mentioned earlier) getting the coil above freezing. The discharge pressure and temperature will also increase proportionately while the suction superheat will remain controlled by the expansion valve. When discharge gas is injected in the suction line the behavior is a little different. The effect is to remove gas that could pass through the condenser and evaporator then raise the back pressure to raise the overall saturation temperature above freezing. If the hot gas injected into the suction line is before the sensing bulb, the expansion valve will try to compensate by adding liquid to the evaporator to lower the superheat which will lead to a roller coaster ride between the expansion valve and the hot gas regulator. If the injection point is down stream of the expansion bulb, the heat from the discharge gas will raise the temperature of the gas in the suction line to about 75 degrees which could be above the manufacturer inlet temperature specification for the compressor. 
AuthorI 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. Archives
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