Refrigerant choices for refrigeration systems are undergoing significant change, including choices for retrofits and new systems. This article is Part 1 of a 3 part series addressing such retrofits, and deals with the basics of refrigerant blends and temperature glide.
The phase-out of R-22 is well into its final few years. Severely restricted now, production and importation of new R-22 in the US will cease as of 2020. EPA Significant New Alternatives Policy (SNAP) Rule 20, issued in July 2015 also sets a timeline for de-listing R-404A, R-507, and others as acceptable alternatives for many commercial refrigeration applications
(1). Some of the first de-listings concerning system retrofits took effect July 20, 2016. EPA has also approved (or SNAP listed) several alternative HFC and HFO-blend refrigerants that may be used to retrofit existing systems, as well as for new systems
(2). Most of these blends are viable replacements for R-22, R-404A, and similar refrigerants, but have properties that require careful consideration to ensure a successful retrofit.
The refrigeration industry has been using refrigerant blends for many years. Some of these are azeotropic, in which the blend constituents form a mixture that changes phase at a constant temperature at a given pressure. This is the same behavior as a pure refrigerant, made up of only one type of molecule. In the evaporator, as the refrigerant boils off from a saturated liquid to a saturated vapor, no sensible heat is gained by the refrigerant until it is 100% vapor. This is a desirable characteristic for most refrigeration systems. Pure refrigerants and azeotropes have a single saturation temperature at a given pressure. Superheat and subcooling are simply calculated from this single-point relationship. Azeotropic blends are assigned ASHRAE R-numbers in the 500-series, such as R-502 or R-507.
Other blends are zeotropic, in which the blend constituents change temperature with change of phase at constant pressure. This behavior is commonly called temperature glide. As a zeotrope boils off in the evaporator, the constituent with the highest vapor pressure will evaporate first. While the first constituent is evaporating, the remainder of the constituents are gaining sensible heat. As the temperature of the refrigerant increases, the vapor pressure of the next constituent will then be reached, and it will evaporate, while the remaining constituents (including those already evaporated) will continue to gain sensible heat. This will continue for all constituents until all are boiled off. At a constant pressure, the saturated vapor of the blend will be at a higher temperature than the saturated liquid, thus the term “temperature glide.” This also means that at a given pressure, there are now two saturation temperatures to consider: saturated liquid temperature (also known as “bubble point”) and saturated vapor temperature (also known as “dew point”). Because of this behavior, superheat and subcooling for zeotropes must be calculated using the proper point. Zeotropic blends are assigned R-numbers in the 400-series, such as R-404A, R-407C, R-410A, etc.
Because having little or no glide is favorable for heat exchanger design, zeotropes with low glide have been more widely used in the past. In the case of blends such as R-404A or R-408A that have only around 1°F temperature glide, there is very little effect on heat exchanger performance. Even ignoring glide introduces only a small error into superheat and subcooling calculation. With the current movement toward more environmentally-friendly refrigerants, many blends with significantly higher temperature glide (on the order of 10°F) are commonly being used, and more are being introduced to the market. These include but are not limited to R-407A, R-407C, R-407F, R-448A, and R-449A. With high-glide zeotropic refrigerants, care must be taken to use bubble point to calculate subcooling, and dew point to calculate superheat.
Figure 1 shows saturation pressure-temperature relationships for R-507 (an azeotrope), R-404A (a low-glide zeotrope) and R-407A (a high-glide zeotrope) (3). This is a useful way to visualize temperature glide, and some of the differences between these refrigerant types. The focus is on typical medium temperature refrigeration evaporator range, from 10 to 30°F.
On this plot, temperature glide is the horizontal distance (meaning at fixed pressure) between the bubble point and dew point lines of the zeotropic blends. Azeotropic R-507 has a single saturation line, and has no glide. R-404A has just over 1°F glide, while R-407A has over 10°F temperature glide.
Temperature glide and subcooling and superheat calculations can also be visualized on a pressure-enthalpy (P-h) diagram. Figure 2 below is a P-h diagram for R-407A generated by NIST Refprop (3) and illustrates the refrigeration cycle and proper calculations for a high-glide refrigerant. It should be noted that an idealized cycle is illustrated, with no high or low side pressure drops. This simplifies the example calculations, but pressure drops in a real system should be accounted for by using temperatures and pressures measured in close proximity to calculate superheat and subcooling.
As with all P-h diagrams, the liquid saturation line forms the left side of the saturation dome, and the vapor saturation line forms the right side of the dome. The lines of constant temperature (or isotherms) slope downward left-to-right in the saturated region, showing that R-407A is a zeotrope, having temperature glide. For a pure refrigerant or azeotrope, these lines would have no slope: no change in temperature at a given pressure during phase change. Subcooling and superheat are always calculated from the nearest saturation state.
In Figure 2’s example, an R-407A refrigeration system is operating at medium temperature conditions. The measured high and low side system pressures are Phigh = 252 psia and Plow = 57 psia. Using the proper saturation states:
Liquid Subcooling Calculation:
Measured liquid line temperature = 80°F
From NIST Refprop, R-407A bubble point @ 252 psia = 100°F
Liquid Subcooling = 100°F – 80°F = 20°F
Suction Superheat Calculation:
Measured suction gas into the compressor = 65°F
From NIST Refprop, R-407A dew point @ 57 psia = 20°F
Suction Superheat = 65°F – 20°F = 45°F
Discharge Superheat Calculation:
Measured compressor discharge gas = 165°F
From NIST Refprop, R-407A dew point @ 252 psia = 108°F
Discharge Superheat = 165°F – 108°F = 57°F
The calculated numbers can be checked with the P-h diagram, by scaling them with the isotherms. In the case of liquid subcooling, that process extends from the bubble point, leftward into the subcooled liquid region. It extends from the 100°F isotherm to the 80°F isotherm, confirming our calculation of 20°F subcooling.
To show the effect of using the incorrect saturation state, consider Suction Superheat calculated using bubble point:
Measured suction gas into the compressor = 65°F
From NIST Refprop, R-407A bubble point @ 57 psia = 10°F
Suction Superheat = 65°F – 10°F = 55°F
An additional 10°F superheat would be calculated that does not actually exist. This amount of error in calculating superheat at the outlet of an evaporator coil would lead to an improperly adjusted TEV and liquid flooding in the suction line, possibly all the way back to the compressor. We also readily see this is incorrect by checking our calculation against the P-h diagram – the closest saturation state is dew point, not bubble point.
The thermostatic expansion valve (TEV) has only one function in the system: to meter the proper amount of refrigerant to the evaporator coil by controlling the superheat at the bulb location. In order to do this, the TEV is located at the coil inlet. However, conditions at the coil outlet indicate whether the TEV is working properly. For zeotropic refrigerants, superheat is calculated based on the dew point, the closest saturation condition at the evaporator outlet. It follows naturally that superheat is also controlled based on this condition. This means the vapor pressure or dew point of the refrigerant is used to control the TEV.
Based on the characteristics of high-glide zeotropic blends, we see the evaporator pressure referred to is in fact vapor pressure or dew point. Considering the three pressures acting on a TEV, P1 (bulb pressure) and P2 (vapor pressure) are acting on opposite sides of the same diaphragm. TEVs are designed for these two pressures to be very close to one another. These will then be the fundamental forces acting to position the pin in the port, producing the best superheat response. Since each of these two similar pressures is acting on the same area, only a small amount of additional force (via spring pressure, P3) is needed to fine tune the balance through adjustment. This is illustrated in the above figure by the small size of P3 in comparison to the other two forces. The stroke of the valve is relatively very small, and the force the spring needs to provide is very small, so the spring constant is low. The adjustment assembly is typically designed to provide 8-12 full rotations from minimum to maximum design spring force. This allows for virtually infinite tuning, but only within the design range of the spring. It follows that the thermostatic element charge should be designed to be close to the vapor pressure (dew point) of a zeotropic refrigerant to provide the best TEV performance.
What if the thermostatic element charge is significantly lower in pressure than the vapor pressure at the evaporator outlet?
The opening force on the TEV will be very low, and the closing forces will keep the pin and port closed until the bulb gets much warmer, resulting in an inadequate refrigerant feed to the evaporator and high superheat. Adjusting spring pressure lower will help compensate by balancing the forces at a more open position. If evaporator outlet pressure is too much higher, even removing all spring pressure may not be sufficient. The TEV will be operating outside its design parameters, desired superheat may not be obtained, and the evaporator will be fed too little refrigerant for its load.
What if the thermostatic element charge is significantly higher in pressure than the vapor pressure at the evaporator outlet?
The opening force on the TEV will be very high, moving the pin and port more open, and flooding will occur. Adjusting spring pressure higher will help compensate by balancing the forces at a less open position. However, if too much additional spring pressure is needed, the maximum design spring force may be reached, beyond which the valve will be operating outside its design parameters. Proper superheat adjustment may not be possible, and a superheat response will be compromised. Liquid flood back may occur, particularly at high condensing pressure conditions. TEVs, like most control systems, tend to operate best in the middle portion of their design range.
Part 2 of this series will address TEV thermostatic charges and capacity of some of the alternate refrigerants, and compare the sizing and selection of liquid line components for system retrofits.
Article contributed by John Withouse, senior engineer - Refrigeration, Sporlan Division of Parker Hannifin
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