This is the second part of a three-part blog series on high-glide refrigerants for refrigeration system retrofits. Part 1 covered calculating superheat and subcooling with glide, and control of superheat with glide. Part 2 deals with the selection of proper thermostatic charges, capacities of several refrigerant blends relative to both R-404A and R-22, and examines how existing components on the liquid side of a system may be affected by a retrofit to a newer, high-glide blend.
Many alternates to R-404A, R-507, and similar refrigerants have been developed. Most of these are also suitable alternates to R-22 in refrigeration applications. At this time, the alternates that are generating the greatest interest and gaining market momentum include R-407A, R-407F, R-448A, and R-449A. R-407A and R-407F are HFC blends containing R-32, R-125, and R-134a. R-448A and R-449A are blends of HFC and HFO constituents, containing R-32, R-125, R-134a, and R-1234yf. R-448A also contains a small percentage of R-1234ze (1). These alternate refrigerants all have similar temperature glide, in the range of 10°F, due to the common constituents among them. Section 2 examines how these alternates compare to R-404A and R-22 from the standpoint of TEV thermostatic charges and volumetric capacity.
Let’s relate the design and operational characteristics of TEVs to some of the high-glide refrigerants. Over the past few decades, many different refrigerants have been used. Thermostatic charges have been developed to function properly with most of these. Since the charge in the thermostatic element does not function as a heat transfer fluid, the primary concern is that its pressure-temperature relationship is close to the vapor pressure-temperature relationship of the system refrigerant.
During the early stage of thermostatic charge development, designers will compare the vapor pressure-temperature relationship of the new refrigerant to that of a refrigerant with existing thermostatic charges. If they are a close match, that is a strong indication that an existing thermostatic charge will also function well with the new refrigerant. Figure 3 compares the vapor pressure-temperature relationships of R-407A and R-407F with R-404A, across typical ranges of low and medium temperature refrigeration.
The vapor pressures of R-407A and R-407F are like one another, but both are significantly lower than R-404A. Is there an existing refrigerant that may be a better match, for which thermostatic charges have already been designed? Figure 4 adds R-22 to the same chart.
The saturation curve of R-22 is a much closer match to the vapor pressure curves of R-407A and R-407F. Consider typical low and medium evaporator operating temperatures of -20°F and 20°F:
R-22 saturation pressure varies by less than +/- 1 psi at -20°F, and by +/- 2 psi at 20°F. This indicates thermostatic charges for R-22 will likely be applicable to R-407A and R-407F. However, the vapor pressure of R-404A exceeds those of R-407A and R-407F on average by 7 psi at -20°F and nearly 12 psi at 20°F.
Figure 5 is the same chart as above, but the vapor pressure curves of R-407A and R-407F are replaced with those of R-448A and R-449A. It should be noted that in Figure 5, the dew point curve of R-448A lies directly under that of R-449A. Though it may appear that it was omitted, these two are instead similar enough the curves cannot be distinguished at the scale presented.
The saturation curve of R-22 appears to be an even closer match to the vapor pressure curves of R-448A and R-449A. Looking more closely at the -20°F and 20°F points:
We find a variation of only +0.4 psi at -20°F and -1.5 psi at 20°F for R-22. R-404A vapor pressure exceeds those of R-448A and R-449A by 6 psi at -20°F and by 11 psi at 20°F. These straightforward comparisons strongly indicate that thermostatic charges designed for R-22 will function well for R-407A, R-407F, R-448A, and R-449A throughout the range of low and medium temperature refrigeration applications.
What is the effect of using thermostatic charges designed for R-404A with these refrigerants? We established in Part 1 that small variations in vapor pressure from one refrigerant to another, around 5% or less, pose no issue with TEV adjustment or operation. What if that variation is on the order of 20%? From the TEV balance equation (P1 = P2 + P3), if P1 becomes much larger than P2, P3 must increase to compensate. However, we also know that:
a) P3 should be kept small so that P1 and P2 are the fundamental forces acting to position the pin;
b) By adjusting P3 very high the design range of the spring may be exceeded;
c) The TEV will not be operating in its effective design range, where it will provide the best superheat response.
Thermostatic charges designed for refrigerants with much higher vapor pressure, including R-404A, should not be used with R-407A, R-407F, R-448A, and R-449A. Thermostatic charges designed for R-22 are the proper choice. Conversely, lower-pressure thermostatic charges for R-22, R-407A, R-407F, R-448A, or R-449A should not be applied with higher vapor pressure refrigerants such as R-404A.
The capacity (Q) of a direct expansion heat exchanger is calculated by the following equation:
Q = ṁ(hout – hin), where:
ṁ = mass flow rate of the refrigerant;
hout = the enthalpy of the refrigerant leaving the heat exchanger;
hin = the enthalpy of the refrigerant entering the heat exchanger
Refrigeration systems are almost always designed to a specific cooling capacity, so the Q of greatest interest is the capacity of the evaporator. The (hout – hin) part of the equation is also known as the net refrigerating effect or NRE. It is closely approximated by subtracting saturated liquid enthalpy into the TEV from saturated vapor enthalpy at the evaporator pressure. Refrigerant choice and high and low side operating conditions determine NRE. So for a given Q, ṁ will be determined by the NRE of the refrigerant. Let’s look at the NRE of some of the alternates, using 90°F saturated liquid, and saturated vapor at -20°F and 20°F, calculated by NIST Refprop 9.1. We will compare with R22 and R404A:
Using NRE, we will calculate the required mass flow of each refrigerant per ton (12,000 Btu/hr) capacity:
There is quite a significant variation in mass flow required per ton capacity at the same conditions for these refrigerants. The mass flow of R407F is very close to that of R22, and both are ~32% lower than R404A. The mass flow of R407A, R448A, and R449A are similar, and are ~23% lower than R404A, but still ~12% higher than R22.
It is also useful to compare the volumetric flow of these refrigerants, in order to better understand the sizing of liquid and suction control valves. All valves fundamentally control flow by volume, not mass directly. Density or specific volume is required to calculate volumetric flow. We will use -20°F low temperature saturated evaporating, and 20°F medium temperature saturated evaporating, and 100°F condensing with 10°F subcooling for both. So as not to overly complicate the analysis, high and low side pressure drops will be ignored. From Refprop 9.1, densities for -20°F saturated vapor, 20°F saturated vapor, and 90°F saturated liquid is:
Mass flow per ton is divided by liquid and vapor density to determine the liquid volumetric flow and vapor volumetric flow per ton for each refrigerant, at low and medium temperature system operating conditions:
This comparison highlights the differences on the liquid side and the similarities on the suction side of the system. The lower liquid density of R404A in comparison with R22, along with higher mass flow requirements means that an R404A liquid side valve must flow an even greater volume than a simple mass flow comparison of these two would indicate. R407A, R407F, R448A, and R449A are again in the middle, around 30% lower than R404A and 20% higher than R22. On the suction side of the system, the refrigerants are a much closer match. The suction volumetric flow of R404A and R22 are very similar, with R407A, R407F, R448A, and R449A being 4 – 10% higher than both at medium temperature and 9 – 17% higher at low-temperature conditions.
There is much interest and significant activity with these refrigerants for system retrofits, mostly due to regulatory actions. These include the ongoing phase-out of R22, which will no longer be able to be produced or imported into the US as of January 1, 2020, and the pending removal of R404A and similar refrigerants from EPA’s approved list of refrigerants for retrofit supermarket systems (July 20, 2016), new supermarket systems (January 1, 2017), and new condensing units (January 1, 2018.) Refrigerant will still be available for service into the foreseeable future, but prices will rise. Blends like R404A present more difficulties with reclaiming and recycling than pure refrigerants such as R22.
When considering retrofits of refrigeration systems, we will assume that one goal is to replace as few system components as possible, in the interest of reducing both cost and system downtime. Section 3 examines how R407A, R407F, R448A, and R449A may function with existing valves. The large differences in mass flow and liquid volumetric flow rate between R404A and R22 noted in Section 2.2 mean that retrofits of each to alternate refrigerants must be addressed separately.
The valve sizing and selections in all the subsequent sections of this document are carried out with Version 5.09.15 of the Sporlan Selection Program. This is proprietary software that employs the NIST Refprop DLL for all refrigerant property calculations. Flow and loading calculations are based on experimentally determined valve characteristics, and results have been experimentally validated. In the following sections, results that are outside the normally accepted range are highlighted in red.
These refrigerants are mostly used in systems having multi-circuit evaporator coils, which requires the use of a refrigerant distributor. The pressure drop across the distributor reduces the pressure drop across the TEV, so that must be determined before selecting TEVs. The Sporlan selection program will be used to calculate the distributor nozzle and tube pressure drop for use in selecting TEVs. We will use a 1 ton (12,000 Btu/hr) load, and assume a 3 circuit evaporator with 3/16” OD, 18” long distributor tubes. From this a nozzle is sized for R404A, then that nozzle size is applied to the alternate refrigerants. The selection program will calculate tube ΔP, nozzle ΔP, and distributor total ΔP. Distributor total ΔP is then subtracted from system high-to-low ΔP to obtain ΔP across the TEV, the driving force for flow across the TEV:
For a given distributor orifice size, the total pressure drop across it for the alternates is lower by 10 – 15 psi at low-temperature conditions, and 8 – 12 psi at medium temperature conditions. While these lower distributor ΔP’s are still adequate for the function of the distributor, there is more ΔP across the TEV for R407A, R407F, R448A, and R449A. Using the R404A distributor ΔP, a TEV is selected for R404A. The same TEV (same port size, pin angle, and stroke) is then applied to the alternate refrigerants, employing their distributor ΔP’s, in order to compare sizing:
From this comparison, we see that a TEV and distributor sized properly for R404A may be oversized for the alternates. Percent loading for R407F drops by around 25%, the others drop by around 20%. While there are no hard rules that apply to all TEVs on all systems, loadings below 50% begin to become a concern. Mass and volumetric flow requirements for R407A, R407F, R448A, and R449A are lower, yet there is greater ΔP across the TEV, to drive flow higher. Keep in mind the TEV is sized at a nominal condition. Under low-load conditions, the TEV will be even more oversized. When thermostatic charge are also considered (as discussed in Section 2.1), there is a high risk existing R404A TEVs will provide poor superheat control when retrofitted to R407A, R407F, R448A, or R449A. There is also a high risk of liquid flood back.
Replacement of TEVs will be the best choice in most retrofits from R404A to R407A, R407F, R448A, or R449A. Most existing distributors will perform acceptably well. Sizing of existing R404A distributors and TEV’s should be checked before replacement.
The same exercise can be performed with an R22 system. We will again use a one-ton (12,000 Btu/hr) load and assume a 3 circuit evaporator with 3/16” OD, 18” long distributor tubes. A nozzle is sized for R22, then applied to the alternate refrigerants:
With a distributor orifice sized for R22, the total pressure drop for the alternates is higher by 4 – 11 psi at both low and medium temperature conditions. Yet because of their significantly greater total high-to-low side ΔP, there is still as much as 47 - 49 psi greater ΔP across the TEV at both low and medium temperature conditions. Using the R22 distributor ΔP, a TEV is selected for R22. The same TEV (same port size, pin angle, and stroke) is then applied to the alternate refrigerants, in order to compare sizing:
With a TEV and distributor sized properly for R22, percent loading changes little for the alternates. Percent loading for R407F is somewhat lower at both medium and low-temperature conditions, but well within an acceptable range. Loadings for R407A, R448A, and R449A are just 2 – 4% percent higher at low-temperature conditions. At medium temperature conditions, loading with R448A and R449A increases by 6 – 8%. All these would be considered within an acceptable range. Coupled with the knowledge that thermostatic charges for R22 are the proper choice for these refrigerants as well, there is a high likelihood that existing R22 TEV’s and distributors will provide good superheat control when retrofitted to R407A, R407F, R448A, or R449A. Sizing should be checked to confirm this.
Another commonly used control valve in refrigeration systems is the solenoid valve. These are most commonly used to control liquid flow but are also sometimes used to control suction gas flow. We will first look at liquid line solenoid valves. Unlike an expansion valve, the purpose of which is to precisely meter refrigerant to the evaporator, the solenoid valve’s purpose is to simply provide electrically actuated on/off refrigerant flow control. Most solenoid valves are “piloted”. They are opened by a coil that applies electromagnetic pull on the stem but relies on a light spring and a small amount of pressure drop across the valve to close it. At sizing conditions, this pressure drop should be 1 psi or greater. If a solenoid valve is too large, there may not be enough ΔP across it to initially close it, particularly at low load conditions.
If a solenoid valve is too small, there will be more pressure drop across it than is desired. This causes loss of subcooling in a liquid line. If there is very little subcooling, the liquid can flash across the valve. A general guideline is that the pressure drop should be no greater than 1°F of saturation (bubble point) drop for the system refrigerant. For all the refrigerants in our discussion, 1°F of saturation is 3.0 – 3.5 psi at 100°F.
For TEVs and distributors, we used one ton (12,000 Btu/hr) loads in our comparisons. This is appropriate because there is typically one TEV per case or unit cooler in a system. Often one solenoid valve will control flow on a system branch supplying liquid to a few or several cases or unit coolers. For solenoid comparisons, we will size for a two-ton (24,000 Btu/hr) load at low-temperature conditions (-20°F evaporator, 100°F condensing) and a four-ton (48,000 Btu/hr) load at medium temperature conditions (20°F evaporator, 100°F condensing). A liquid line solenoid valve will be selected for R404A, then that valve will be applied to R407A, R407F, R448A, and R449A to see the effect on sizing. Selecting a solenoid valve for R404A that will have ΔP greater than 1 psi, but less than 3.5 psi:
An E6 liquid line solenoid valve is selected for the R404A low-temperature system, with 1.3 psi ΔP. This is on the low end of the preferred range, so what if the next smaller size E5 is selected instead? ΔP would be 4.2 psi instead, with 117% loading. In this case, either would likely function without issue. Usually, the lower ΔP would be chosen for an actual system, so the E6 will be used for comparison. As shown in the table above, that choice results in low ΔP for all the alternate refrigerants. The situation is different for the medium temperature system with a four-ton load. An E8 solenoid valve is chosen for R404A, with 2.6 psi ΔP. When applied with the alternate refrigerants, ΔP’s from 1.2 to 1.5 psi is calculated. These are obviously lower, but still acceptable.
When retrofitting an existing R404A system to R407A, R407F, R448A, or R449A, liquid line solenoid valve sizing should be checked. If ΔP will be lower than 1 psi with the alternate refrigerant, appropriately sized solenoid valves should be selected to replace the existing ones.
Let’s perform the same comparison with R22 as a baseline, again using a two-ton (24,000 Btu/hr) load at low-temperature conditions and a four-ton (48,000 Btu/hr) load at medium temperature conditions. A liquid line solenoid valve will be selected for R22, then that valve will be applied to R407A, R407F, R448A, and R449A to see the effect on sizing. 1°F saturation drop for R22 at 100°F is 2.9 psi, so we select a solenoid valve for R22 that will have ΔP greater than 1 psi but less than 3 psi:
An E5 liquid line solenoid valve is selected for the R22 low-temperature system, with 1.6 psi ΔP. This would be a common choice in an actual system. As shown in the table above, that selection results in ΔP’s from 1.8 to 2.3 psi for the alternate refrigerants. The E5 would function well, without excessive ΔP for the alternate refrigerants. Results for the medium temperature system mirror the low-temperature system. An E6 solenoid valve is selected for R22, with 1.8 psi ΔP. When applied with the alternate refrigerants, ΔP’s of 2.0 to 2.5 psi is calculated, again well within an acceptable range.
When retrofitting an existing R22 system to R407A, R407F, R448A, or R449A, existing liquid line solenoid valve sizing will in most situations be acceptable, though it should be checked. If ΔP will be excessively high with the alternate refrigerant, appropriately sized solenoid valves should be selected to replace the existing ones.
The third and final part of this series will deal with suction side components of systems in retrofits, as well as system protection, oil, and seal changes.
HVACR Tech Tip Article contributed by John Withouse, senior engineer - Refrigeration, Sporlan Division of Parker Hannifin
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