Technical Papers 37th Annual Meeting International Institute of Ammonia Refrigeration March 22 25, 2015 2015 Industrial Refrigeration Conference & Exhibition San Diego, California
ACKNOWLEDGEMENT The success of the 37th Annual Meeting of the International Institute of Ammonia Refrigeration is due to the quality of the technical papers in this volume and the labor of its authors. IIAR expresses its deep appreciation to the authors, reviewers and editors for their contributions to the ammonia refrigeration industry. ABOUT THIS VOLUME IIAR Technical Papers are subjected to rigorous technical peer review. The views expressed in the papers in this volume are those of the authors, not the International Institute of Ammonia Refrigeration. They are not official positions of the Institute and are not officially endorsed. International Institute of Ammonia Refrigeration 1001 North Fairfax Street Suite 503 Alexandria, VA 22314 + 1-703-312-4200 (voice) + 1-703-312-0065 (fax) www.iiar.org 2015 Industrial Refrigeration Conference & Exhibition San Diego, California
Technical Paper #4 Optimizing Evaporator Runtime and Defrost Frequency Bruce I. Nelson, P.E. President Colmac Coil Manufacturing, Inc. Abstract When cooling air to temperatures below freezing, the surfaces (fins and tubes) of evaporators unavoidably accumulate frost which must periodically be removed by defrosting. The defrosting process is inherently bad because it: a) reduces system refrigeration capacity by interrupting the air cooling process, and b) increases system power consumption by adding heat to the space which must then be removed by the refrigeration system. The room temperature and relative humidity in combination with the evaporating temperature defines the Sensible Heat Ratio (SHR) for the air cooling process. The SHR in turn indicates exactly how much moisture will be removed from the airstream and end up as frost on the evaporator (the frost load ). To date, manufacturers of evaporators have not really offered much guidance to refrigeration system designers regarding how to properly select evaporators for a given frost load. Most of the time, manufacturers application guidance consists of Well, when the frost load is heavy increase the number of defrosts, Oh and by the way, watch out. The paper quantifies the rate of moisture removal as a function of SHR and shows the effect of frost accumulation on evaporator airflow and cooling capacity. The predicted reduction in evaporator cooling capacity over time is used to determine an appropriate value for runtime hours (hours/day) used for equipment selection as well as selecting an appropriate defrost frequency (no. of defrosts/ day). IIAR 2015 1
Optimizing Evaporator Runtime and Defrost Frequency Sensible Heat Ratio (SHR), Room Relative Humidity (rh%), and Evaporator Ratings Accurate prediction of the refrigeration load, both sensible and latent components, is important to proper refrigeration system equipment selection and successful operation (Nelson 2012(a)). Various types of sensible cooling loads must be anticipated and included in the calculation, such as: lighting, electric motors, forklifts, product cooling/ freezing, transmission of heat through walls, ceilings, and floors, and cooling of infiltration air. Latent cooling loads are present whenever moisture is added to the air in the refrigerated space. Sources of introduced moisture typically include: infiltration air, respiring food products, surface moisture on products, packaging and other objects entering the space, residual water left on floors after wash down (process rooms), human respiration, and humidification equipment (above freezing). Room relative humidity (rh%), which is the indication of how nearly the air in the refrigerated space is saturated with water vapor, will be the equilibrium condition resulting when moisture introduced into the space equals the moisture removed from space by the evaporator coils (Cleland 2012). Whenever evaporator surfaces operate at temperatures below the dew point of the air being cooled, water vapor in the airstream is condensed to liquid (at temperatures above 32 F (0 C)) or deposited to form frost (below 32 F (0 C)). The cooling effect associated with this dehumidification of the airstream is termed latent cooling. The sum of the sensible cooling load and latent cooling load is termed the total load. The ratio of the sensible cooling load divided by the total cooling load is called the Sensible Heat Ratio (SHR) and defines the slope of the air process line on a psychrometric chart. (1) The SHR indicates how much (or how little) moisture is being removed from the airstream during the cooling process. A SHR of 1.0 indicates all of the cooling is Technical Paper #4 IIAR 2015 3
2015 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, CA sensible and no moisture is removed, in other words, no frost would be formed on an evaporator operating with a SHR of 1.0. A SHR of 0.8 means 80% of the cooling is sensible and 20% is latent. Somewhat counterintuitively, the smaller the SHR the faster frost will form on evaporator surfaces the greater the frost load. An evaporator operating with a SHR of 0.7 will accumulate frost faster (and require more frequent defrosts) than an evaporator operating with a SHR of 0.9. The SHR can be determined from a psychrometric chart using the straight line rule knowing the temperature and relative humidity of the air entering and the surface temperature of the evaporator (Stoecker 1988). Using the evaporating temperature instead of surface temperature is a close approximation and is used in this analysis. Warmer air can contain more water vapor than colder air. As seen in Figure 1 and in Table 1, for a given relative humidity the SHR goes down (frost load goes up) as temperature increases. Figure 1. Psychrometric Chart 4 IIAR 2015 Technical Paper #4
Optimizing Evaporator Runtime and Defrost Frequency Refrigerated spaces with product being transferring in and out through doorways will very typically have a relative humidity in the range of 80 to 90% due to infiltration and other sources of moisture. Lower room relative humidity may be found in some exceptional cases where traffic through doorways is very light, product is tightly packaged, dehumidification equipment is used at doorways, etc. The relationship between room relative humidity and SHR is shown in Table 1: Sensible Heat Ratio, SHR Room Temperature, F ( C) 65%rh 75%rh 85%rh 95%rh 45 (7.2) 1.0 0.84 0.67 0.56 32 (0) 0.98 0.84 0.73 0.64 10 (-12.2) 0.98 0.92 0.87 0.83 0 (-17.8) 0.98 0.95 0.92 0.89-10 (-23.3) 0.99 0.97 0.95 0.93-30 (-34.4) 0.99 0.99 0.98 0.97 Table 1. SHR For DT1 = 10 Deg F At Various Temperatures And Room rh% From Table 1 it is clear that the highest frost load (lowest SHR) will occur in high temperature (+32 F) rooms with high relative humidity. The lowest frost loads (highest SHR) occur at freezer temperatures, even when relative humidity is high. The room relative humidity and resulting SHR can have a large effect on evaporator cooling capacity, especially at higher room temperatures. Interestingly, the total amount of heat which can be transferred to the surface of an evaporator coil increases significantly as the SHR falls below 1.0. This increase in apparent heat transfer efficiency is due to the very large amount of energy released as water in the airstream changes state from vapor to liquid or solid (frost) on coil surfaces. This process of air cooling with simultaneous removal of water from an airstream is referred to as combined heat and mass transfer, or combined sensible and latent heat transfer. The effect is a much higher rated cooling capacity for an Technical Paper #4 IIAR 2015 5
2015 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, CA evaporator operating with a SHR of 0.8 compared to the same evaporator operating with a SHR of 1.0. It is the same phenomenon that allows an evaporative condenser to operate with much less surface area compared to an air-cooled condenser for a given condensing load, but in reverse. Using a computer model developed to accurately calculate fin efficiency and surface effectiveness for both sensible and combined sensible and latent heat transfer, a prediction of the increase in evaporator performance as a function of SHR has been made (Nelson 2012(a)). Results of the predicted capacity increase as a function of SHR for an ammonia refrigeration evaporator coil operating over a wide range of room temperatures (+45 F to -30 F) and having typical fin spacings and geometry with DT1 = 10F is shown in Figure 2. Capacity Factor, Qtotal / Qsens only 1.5 1.45 1.4 1.35 1.3 1.25 1.2 1.15 1.1 1.05 1 0.6 0.65 0.7 0.75 0.8 0.85 0.9 0.95 1 Sensible Heat Ra7o, SHR Figure 2. Total Cooling Capacity Factor vs SHR When it comes to publishing ratings, things become a bit confusing since some evaporator manufacturers include the effect of room rh% in their ratings while others do not. As shown in Figure 2, the lower the SHR the greater the total cooling 6 IIAR 2015 Technical Paper #4
Optimizing Evaporator Runtime and Defrost Frequency capacity of the evaporator. A manufacturer who shows their evaporator ratings as all sensible (SHR = 1) will be more conservative (have more surface area) than the manufacturer who shows their ratings at 85 or 95% rh. Selecting evaporators using 85 to 95%rh ratings will result in evaporators having less surface area and lower first cost compared to evaporators selected using all sensible ratings. The risk in this approach is undersizing the evaporators in the case where the actual operating room rh% is less than the rh% used during the selection process. Conclusion: The latent load should always be estimated and included in the total calculated refrigeration load. Size evaporators for the design total refrigeration load at the estimated room relative humidity. If room relative humidity is difficult to estimate or cannot be estimated, then a conservative approach is to select evaporators based on a low room relative humidity (i.e. 65 to 75% rh) or using sensible only ratings. Types of Frost and How It Forms Frost can accumulate on evaporator coil surfaces by one of two mechanisms: 1. By deposition, and/or 2. As air-borne ice crystals Deposition As mentioned previously, whenever the temperature of the evaporator coil surface is below the dewpoint temperature of the room air, moisture will condense and be deposited on the surface either as liquid water (above freezing) or as frost (below freezing). This mass transfer process, when related to the formation of frost, is called deposition, and is driven by the difference in water vapor pressure between Technical Paper #4 IIAR 2015 7
2015 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, CA the air and the surface of the coil. The amount of heat associated with this mass transfer process is termed latent heat and is quantified by the SHR (see previous definition). Whenever the SHR is less than 1.0, the deposition of frost will take place. An estimate of the rate at which frost will be deposited on the coil surfaces can be calculated as a function of the total cooling load, the SHR, and the surface area of the evaporator. In the interest of simplifying the model used in this analysis, an assumption was made that frost will be uniformly deposited over the surface of the evaporator. It is recognized that the rate of frost deposition from row-to-row will vary depending on a number of factors. For evaporators with counterflow circuiting, the rate of deposition immediately after a defrost will be greatest in the first rows of the coil where the temperature difference between the air and evaporating temperature is the greatest. However, as frost accumulates in the first rows the rate of heat transfer and frost deposition there will slow while the remaining rows will tend to continue accumulating frost at a relatively greater rate. The overall effect over time being a more or less uniform distribution of the total amount of accumulated frost. For evaporators with parallel circuiting ( thermodynamic counterflow ) the temperature difference between coil surface and air will be more uniform from row-to-row and so frost will generally accumulate more uniformly. Lastly, for air coolers operating with secondary refrigerants (i.e. glycol) the greatest rate of deposition could be on the last rows of the coil depending on fluid flow rate and temperatures. Given the above, it is believed the general assumption of uniform distribution of frost is a reasonable simplification. Therefore, the assumption of uniform frost deposition is made in the following equation: (2) 8 IIAR 2015 Technical Paper #4
Optimizing Evaporator Runtime and Defrost Frequency Where:, ft 2 ft 3 Example: An evaporator having 8 rows deep and fin spacing of 3 FPI is operating with a 10 deg F TD (DT1) in a +10 F/85%rh room. The evaporator has a cooling capacity of 240,000 Btu/h (20 TR) and outside surface area of 4100 ft 2. What will be the rate of frost deposition? Answer: From Table 1 the expected SHR at this room air temperature and rh% will be 0.87. Likewise, as frost is deposited on the evaporator coil surfaces, the local air velocity between fins will increase and result in increased air pressure drop across the coil. The increase in air pressure drop due to uniform accumulation of frost can be approximated by the following equation: 1 1 25.4 (3) Technical Paper #4 IIAR 2015 9
2015 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, CA Where: The rate of blockage of the coil with frost and associated pressure drop will result in a reduction in airflow and cooling capacity. This reduction in cooling capacity will ultimately determine defrost frequency and efficiency, and overall system energy efficiency and power consumption. Cooling capacity of the evaporator can be characterized as functions of face velocity, SHR, and frost thickness. Knowing how capacity changes with these parameters, combined with the relationships shown in equations 1, 2, and 3, allows the construction of a simple model which will predict the change in coil capacity over time. Simplifying assumptions in the model include: Frost is deposited uniformly over the surface of the coil. Frost density is uniform and of a fixed value. Suction temperature remains constant. Evaporators are available with many different tube patterns and fin spacings. Deciding which tube pattern and fin spacing is optimum for a given set of operating conditions is not always obvious to those trying to select air coolers. Tubes can be spaced very closely together to create a very compact design which occupies a relatively small volume. While this type of compact tube pattern may work well in packaged air conditioning equipment where size and volume are at a premium, it will tend to plug quickly with frost in freezing conditions and require many defrosts per day. Tubes can be spaced more widely to create a more open tube pattern. This type of pattern combined with wide fin spacing, while it occupies more volume than the compact one, will allow the accumulation of more frost between defrosts and reduce the defrost frequency. 10 IIAR 2015 Technical Paper #4
Optimizing Evaporator Runtime and Defrost Frequency Three very different tube patterns were selected for an analysis of frost accumulation and its effects on airflow, cooling capacity and defrost frequency using the model described above: 1. 5/8" Staggered Pattern (compact) a. Designation: RBL(Stagg) b. Tube Diameter: 5/8" (15.9 mm) c. Normal Tube Pitch: 1.5" (38.1 mm) d. Transverse Tube Pitch: 1.299" (33.0 mm) e. Pattern: Staggered f. Fin Spacing: 2, 3, and 4 fins per inch (12.7, 8.5, and 6.35 mm per fin) 2. 5/8" Inline Pattern a. Designation: RBL(I) b. Tube Diameter: 5/8" (15.9 mm) c. Normal Tube Pitch: 50 mm d. Transverse Tube Pitch: 50 mm e. Pattern: Inline f. Fin Spacing: 2, 3, and 4 fins per inch (12.7, 8.5, and 6.35 mm per fin) 3. 7/8" Staggered Pattern a. Designation: RBR(Stagg) b. Tube Diameter: 7/8" (22.2 mm) c. Normal Tube Pitch: 2.25" (57.2 mm) d. Transverse Tube Pitch: 1.949" (49.5 mm) e. Pattern: Staggered f. Fin Spacing: 2, 3, and 4 fins per inch (12.7, 8.5, and 6.35 mm per fin) Technical Paper #4 IIAR 2015 11
2015 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, CA In all cases, the face area (height x width) of the coil block was kept the same as well as the number, diameter, and pitch of the fan blades. Fin Height: 54" (Staggered) 1350 mm (Inline) Finned Length: 186" Rows Deep: 8 Number of Fans: 3 Fan Diameter x rpm: 36" (914 mm) x 1140 Nominal Fan Power: 2 Hp ea Refrigerant used was pumped ammonia. Temperature difference (DT1) in all cases was 10 F (5.6 C). Three different room temperature and humidity conditions (with the resulting SHR shown) were examined: a. -10 F/85%rh (0.95 SHR) b. +10 F/85%rh (0.87 SHR) c. +32 F/85%rh (0.73 SHR) The evaporator designs described above were modeled to answer the following questions: How quickly does frost accumulate on evaporator surfaces? How does frost accumulation affect airflow and cooling capacity? 12 IIAR 2015 Technical Paper #4
Optimizing Evaporator Runtime and Defrost Frequency What will be the correct defrost frequency for a given set of conditions? How much does defrosting cost? What is the optimum tube pattern and fin spacing for operating (and efficiently defrosting) evaporators? The Rate of Frost Deposition The model indicates that the rate of frost deposition is a strong function of SHR and of fin spacing. As already observed above, a lower SHR means more moisture is being removed from the airstream, i.e. a heavier frost load. Interestingly, the rate of deposition is not such a strong function of the tube pattern, all three show about the same mm/h slope for a given set of operating conditions. As will be seen later, however, even though the rate of deposition is about the same for the three tube patterns for a given fin spacing, the rate of decrease in airflow and cooling capacity over time is much different. See Figures 3 through 5. 2.5 8R3F, - 10F/85%rh Room (0.95 SHR) 2 Frost Thickness, mm 1.5 1 0.5 RBR (Stagg) 8R3F RBL(I) 8R3F RBL(Stagg) 8R 3F 0 0 4 8 12 16 20 24 28 32 36 40 Time, h Figure 3. Frost Thickness vs. Time for Various Tube Patterns 8R3F, -10 F/85%rh Room (0.95 SHR) Technical Paper #4 IIAR 2015 13
2015 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, CA 8R3F, +10F/85% rh Room (0.87 SHR) Frost Thickness, mm 1.8 1.6 1.4 1.2 1 0.8 0.6 0.4 0.2 RBR (Stagg) 8R3F RBL(I) 8R3F RBL (Stagg) 8R3F 0 0 1 2 3 4 5 6 7 8 9 10 11 12 Time, h Figure 4. Frost Thickness vs. Time for Various Tube Patterns 8R3F, +10 F/85% rh Room (0.87 SHR) 1.6 1.4 8R3F, +32F/85%rh Room (0.73 SHR) Frost Thickness, mm 1.2 1 0.8 0.6 0.4 RBR (Stagg) 8R3F RBL(I) 8R3F RBL (Stagg) 8R3F 0.2 0 0 1 2 3 4 5 6 Time, h Figure 5. Frost Thickenss vs. Time for Various Tube Patterns 8R3F, +32 F/85% rh Room (0.73 SHR) 14 IIAR 2015 Technical Paper #4
Optimizing Evaporator Runtime and Defrost Frequency Effect of Frost Accumulation on Airflow and Cooling Capacity As discussed above, over time the accumulation of frost reduces cooling capacities in two ways: a) increasing air pressure drop reduces airflow, and b) increasing frost thickness insulates the fin metal and reduces the surface effectiveness. Airflow The model was used to predict the reduction in airflow over time for the three tube patterns with three fin spacings operating at three different temperatures and SHR. For all operating conditions the 5/8" inline tube pattern produced the greatest airflow over time because of its low air pressure drop compared to the other two patterns. The 5/8" staggered (compact) pattern produced the lowest airflow because of its high air pressure drop. Figures 6 through 8 show the effect of frost accumulation on airflow over time. 8R3F, - 10F/85%rh Room (0.95 SHR) Airflow, cfm 50000 45000 40000 35000 30000 25000 20000 15000 10000 5000 0 0 4 8 12 16 20 24 28 32 36 40 Time, h RBR (Stagg) 8R3F RBL(I) 8R3F RBL(Stagg) 8R 3F Figure 6. Airflow vs. Time for Various Tube Patterns 8R3F, -10 F/85% rh Room (0.95 SHR) Technical Paper #4 IIAR 2015 15
2015 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, CA 8R3F, +10F/85% rh Room (0.87 SHR) Airflow, cfm 50000 45000 40000 35000 30000 25000 20000 15000 10000 5000 0 0 1 2 3 4 5 6 7 8 9 10 11 12 Time, h RBR (Stagg) 8R3F RBL(I) 8R3F RBL (Stagg) 8R3F Figure 7. Airflow vs. Time for Various Tube Patterns 8R3F, +10 F/85% rh Room (0.87 SHR) Airflow, cfm 8R3F, +32F/85%rh Room (0.73 SHR) 50000 45000 40000 35000 30000 25000 RBR (Stagg) 8R3F 20000 RBL(I) 8R3F 15000 RBL (Stagg) 8R3F 10000 5000 0 0 1 2 3 4 5 6 Time, h Figure 8. Airflow vs. Time for Various Tube Patterns 8R3F, +32 F/85% rh Room (0.73 SHR) Cooling Capacity As illustrated below, the model indicates that the 7/8" staggered tube pattern always produces the highest clean the condition immediately after defrosting cooling capacity of the three patterns examined. Over time, however, the cooling capacity 16 IIAR 2015 Technical Paper #4
Optimizing Evaporator Runtime and Defrost Frequency of the 5/8" inline pattern overtakes the capacity of the other two patterns due to its lower air pressure drop and greater airflow. This advantage of the 5/8" inline pattern under frosting conditions over time becomes more pronounced as fin spacing becomes closer (as fins per inch increases). This is shown in Figures 9 through 11. 90 80 8R2F, - 10F/85%rh Room (0.95 SHR) Cooling Capacity, kw 70 60 50 40 30 20 RBR (Stagg) 8R2F RBL(I) 8R2F RBL(Stagg) 8R 2F 10 0 0 4 8 12 16 20 24 28 32 36 40 Time, h Figure 9. Capacity vs. Time for Various Tube Patterns 8R2F, -10 F/85% rh Room (0.95 SHR) Cooling Capacity, kw 8R3F, - 10F/85%rh Room (0.95 SHR) 100 90 80 70 60 50 RBR (Stagg) 8R3F 40 RBL(I) 8R3F 30 RBL (Stagg) 8R 3F 20 10 0 0 4 8 12 16 20 24 28 32 36 40 Time, h Figure 10. Capacity vs. Time for Various Tube Patterns 8R3F, -10 F/85% rh Room (0.95 SHR) Technical Paper #4 IIAR 2015 17
2015 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, CA 8R4F, - 10F/85%rh Room (0.95 SHR) Cooling Capacity, kw 100 90 80 70 60 50 40 30 20 10 0 0 4 8 12 16 20 24 28 32 36 40 Time, h RBR (Stagg) 8R4F RBL(I) 8R4F RBL(Stagg) 8R4F Figure 11. Capacity vs. Time for Various Tube Patterns 8R4F, -10 F/85% rh Room (0.95 SHR) Knowing the rate of reduction in evaporator cooling capacity over time due to frost accumulation allows us to examine the relationship between defrost frequency, equipment first cost, and operating cost, and then to optimize defrost frequency. For defrost frequency calculations it is useful to represent the evaporator cooling capacity over time as a percentage of the clean coil capacity. An example of this is shown in Figure 12. 8R3F, - 10F/85%rh Room (0.95 SHR) Cooling Capacity, % 100% 90% 80% 70% 60% 50% 40% 30% 20% 10% 0% 0 4 8 12 16 20 24 28 32 36 40 Time, h RBR (Stagg) 8R3F RBL(I) 8R3F RBL (Stagg) 8R 3F Figure 12. Capacity vs. Time for Various Tube Patterns 8R3F, -10 F/85% rh Room (0.95 SHR) 18 IIAR 2015 Technical Paper #4
Optimizing Evaporator Runtime and Defrost Frequency Calculating the Cost of a Defrost Based on a cost of electricity of $0.15/kWh, the cost of a single 30 minute defrost was calculated for the evaporators used in the model and is shown in Table 2 below. Note that the cost of defrost shown does not include the actual energy required to melt the frost. Here the cost includes only the energy required to heat and cool the metal plus the convective heat loss to the room. Estimating the defrost convective heat loss from an evaporator during defrost is a bit of a challenge in that it is influenced by the specific configuration, size, orientation, and location of the evaporator. A simple approximation of the convective heat loss can be stated as follows (Aljuwayhel 2006): (4) Where:.5 Fin Spacing, Room Temperature (SHR) -10 F/85% +10 F/85% +32 F/85% Tube Pattern fpi rh (.95) rh (.87) rh (.73) 7/8" Staggered 2 0.09 0.05 0.02 7/8" Staggered 3 0.11 0.06 0.02 7/8" Staggered 4 0.13 0.07 0.02 5/8" Inline 2 0.13 0.07 0.02 5/8" Inline 3 0.16 0.09 0.03 5/8" Inline 4 0.17 0.10 0.03 5/8" Staggered 2 0.09 0.05 0.02 5/8" Staggered 3 0.11 0.06 0.02 5/8" Staggered 4 0.13 0.07 0.02 Table 2. Cost Of A Single 30 Min Defrost (Convective Losses and H/C Metal), $/Tr/Defrost Technical Paper #4 IIAR 2015 19
2015 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, CA Note that in Table 2, the units of $/TR/Defrost represents dollars per ton of refrigeration per defrost. Refrigeration Energy Balance During operation the amount of heat taken out of the refrigerated space by the evaporator must equal the refrigeration load plus the heat lost to the room due to convective air heating during defrost. Equations 5 through 7 define the amount of heat removed and added to the space. Note that equation 6 assumes the rate of reduction in evaporator cooling capacity over time is approximately linear. This assumption is shown to be valid by the calculated results of the model described above (see Figure 12). Where: (5) 1 2 (6) (7) (8) 20 IIAR 2015 Technical Paper #4
Optimizing Evaporator Runtime and Defrost Frequency Where: Figure 13 illustrates equation 5 where the area under each section represents Qevap, Qload, and Qdef. Heat Transferred, Btu/h Figure 13. Heat Transferred vs. Time Technical Paper #4 IIAR 2015 21
2015 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, CA Combining and rearranging equations 4 through 8 into equation 9 allows us to calculate the clean evaporator cooling capacity required to achieve the energy balance shown in equation 5 and Figure 13. 2 (9) Equation 9 can be restated in terms of the slope of the evaporator capacity vs time curve, such as is shown in Figure 12. Where: 2 (10) Using the model(s) described above, the slope of evaporator capacity over time for the various tube patterns and fin spacing selected for analysis is shown in Table 3. Fin Spacing, Room Temperature (SHR) -10F/85% +10F/85% +32F/85% Tube Pattern fpi rh (.95) rh (.87) rh (.73) 7/8" Staggered 2-0.017-0.045-0.111 7/8" Staggered 3-0.021-0.058-0.143 7/8" Staggered 4-0.031-0.069-0.175 5/8" Inline 2-0.014-0.037-0.086 5/8" Inline 3-0.015-0.041-0.094 5/8" Inline 4-0.019-0.052-0.119 5/8" Staggered 2-0.034-0.093-0.207 5/8" Staggered 3-0.046-0.126-0.304 5/8" Staggered 4-0.056-0.154-0.390 Table 3. Slope of Evaporator Capacity vs. Time (At DT1 = 10F), 1/h 22 IIAR 2015 Technical Paper #4
Optimizing Evaporator Runtime and Defrost Frequency Time to initiation of defrost can easily be calculated from the slope shown in Table 3 according to equation 11: (11) Example 1: An evaporator is operating in +32 F/85%rh (SHR = 0.73) conditions. Assuming the appropriate time to initiate defrost will be when evaporator capacity reaches 60% of initial clean capacity. What will be the time to initiate defrost for an evaporator with 7/8" staggered tube evaporator and 3 fpi? Answer: (0.6 1)/(-0.143) = 2.8 hours Example 2: Find the time to initiate defrost for a 5/8" staggered evaporator with 3 fpi operating in the same conditions as the evaporator in Example 1: (0.6 1)/(-0.304) = 1.3 hours If each defrost cycle lasts 30 minutes, then 1.5 hours between defrosts will be required in order to satisfy the 2 to 1 rule. That is, keeping 2 x evaporators running for each evaporator that is in defrost, i.e. 1.5 hours = 30 minute defrost x 3 evaporators. It is apparent that the evaporator with 5/8" Staggered tube pattern in Example 2 cannot keep up with the refrigeration load if it must be defrosted every 1.3 hours (less than the absolute minimum of 1.5 hours). Because evaporators with the 5/8" Staggered tube pattern used in this analysis lose capacity so quickly over time, this tube pattern is NOT recommended for applications Technical Paper #4 IIAR 2015 23
2015 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, CA where frost will form. This pattern is appropriate in high temperature rooms where the fins are wet and a compact design is desirable, but not in rooms with below freezing temperatures. Design Runtime Hours Per Day Normally, the refrigeration load used to select equipment is adjusted based on some assumed number of runtime hours per day. This number is typically between 16 and 20 hours/day. It is not clear when or why this rule of thumb came into practice. Perhaps it is a corollary to the 2-to-1 rule for hot gas defrosting. That is, two coils must be in operation while the third coil is in hot gas defrost in order to provide a sufficient quantity of hot gas for the defrost cycle. Perhaps it was intended as simply an additional catch-all safety factor. One industry historian mentioned that the 16-20 runtime hours/day rule came from split system applications where capacity drops out during defrost compared to a central system which would rebalance the TD on the remaining evaporators (Welch 2013). It makes the most sense to the author that the runtime hours adjustment to the design refrigeration load should be used to account for the degradation in coil performance over time due to accumulation of frost. In a large refrigeration system having a constant refrigeration load and compressor unloading capability, the compressors will unload to maintain a constant suction temperature as the coil capacity falls off due to frosting. As the compressors unload to maintain system suction pressure, they will run longer to maintain room temperature. Runtime hours per day can be calculated from the ratio of clean evaporator capacity to refrigeration load (equation 12) as shown in equation 13 below: 2 (12) 24 IIAR 2015 Technical Paper #4
Optimizing Evaporator Runtime and Defrost Frequency 24 (13) Optimizing Defrost Frequency and Design Runtime In a strict sense, the only useful work done during a defrost cycle is the work required to heat the accumulated frost from the evaporating temperature up to 32 deg F and then to melt it. The work required to heat and cool the metal in the coil, and the convective heating of air which escapes to the room is parasitic and can be considered a loss during the defrosting process. Losses due to defrosting are typically quite high on the order of 60 to 80% in most systems (Nelson 2011(1)) and represent a significant opportunity for improvement and cost savings. Running longer between defrosts minimizes these losses, but increases first cost since more evaporator surface is required to compensate for the reduction in cooling capacity. To determine the optimum defrost frequency for a given evaporator design, we will need to know both the first cost of the evaporators (in $/TR) and the cost of defrosting (in $/TR/y). The optimum will occur when the increase in first cost of the evaporators is offset by the savings in operating cost produced for a reduction in the cost of defrosting. The first cost of the evaporators (and other system components) will be inversely proportional to the runtime hours used in the design, i.e. as runtime hours go down, the size and cost of the refrigeration equipment goes up. The cost of defrosting will go down as defrost frequency goes down, i.e. fewer defrosts per day will reduce the cost of defrosting. Technical Paper #4 IIAR 2015 25
2015 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, CA The models were used to calculate changes in first cost of the evaporators along with the changes in cost of defrosting as evaporator capacity at the time of defrost was varied. Figures 14 through 19 show costs vs. evaporator capacity at the time of defrost. The curves at the top of each graph show evaporator first cost in $/Net TR while the curves at the bottom show cost to defrost in $/Net TR/y. Evaporator first cost is only for the coils themselves and does not include control valve costs nor the cost of structural steel. Cost/TR vs % Evaporator Capacity at Time of Defrost $2,000 $1,800 $1,600 Cost per TR $1,400 $1,200 $1,000 $800 $600 $400 8R2F Evap 8R2F Def 8R3F Evap 8R3F Def 8R4F Evap 8R4F Def $200 $0 40% 50% 60% 70% 80% 90% % Evaporator Capacity at Time of Defrost Figure 14. RBR -10 F/85% rh Room (0.95 SHR) Cost/TR vs. % Evaporator Capacity at Time of Defrost Cost/TR vs % Evaporator Capacity at Time of Defrost $2,000 $1,800 $1,600 Cost per TR $1,400 $1,200 $1,000 $800 $600 $400 8R2F Evap 8R2F Def 8R3F Evap 8R3F Def 8R4F Evap 8R4F Def $200 $0 40% 50% 60% 70% 80% 90% % Evaporator Capacity at Time of Defrost Figure 15. RBL(I) -10 F/85% rh Room (0.95 SHR) Cost/TR vs. % Evaporator Capacity at Time of Defrost 26 IIAR 2015 Technical Paper #4
Optimizing Evaporator Runtime and Defrost Frequency $2,000 $1,800 $1,600 Cost/TR vs % Evaporator Capacity at Time of Defrost Cost per TR $1,400 $1,200 $1,000 $800 $600 $400 8R2F Evap 8R2F Def 8R3F Evap 8R3F Def 8R4F Evap 8R4F Def $200 $0 40% 50% 60% 70% 80% 90% % Evaporator Capacity at Time of Defrost Figure 16. RBR +10 F/85% rh Room (0.87 SHR) Cost/TR vs. % Evaporator Capacity at Time of Defrost $2,000 $1,800 $1,600 Cost per TR $1,400 $1,200 $1,000 $800 $600 $400 8R2F Evap 8R2F Def 8R3F Evap 8R3F Def 8R4F Evap 8R4F Def $200 $0 40% 50% 60% 70% 80% 90% % Evaporator Capacity at Time of Defrost Figure 17. RBL(I) +10 F/85% rh Room (0.87 SHR) Cost/TR vs. % Evaporator Capacity at Time of Defrost Technical Paper #4 IIAR 2015 27
2015 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, CA Cost/TR vs % Evaporator Capacity at Time of Defrost $2,000 $1,800 $1,600 Cost per TR $1,400 $1,200 $1,000 $800 $600 $400 8R2F Evap 8R2F Def 8R3F Evap 8R3F Def 8R4F Evap 8R4F Def $200 $0 40% 50% 60% 70% 80% 90% % Evaporator Capacity at Time of Defrost Figure 18. RBR +32 F/85% rh Room (0.73 SHR) Cost/TR vs. % Evaporator Capacity at Time of Defrost $2,000 $1,800 $1,600 Figure 19 RBL(I) +32F/85% Room (0.73 SHR) Cost/TR vs % Evaporator Capacity at Time of Defrost Cost per TR $1,400 $1,200 $1,000 $800 $600 $400 8R2F Evap 8R2F Def 8R3F Evap 8R3F Def 8R4F Evap 8R4F Def $200 $0 40% 50% 60% 70% 80% 90% % Evaporator Capacity at Time of Defrost Figure 19. RBL(I) +32 F/85% Room (0.73 SHR) Cost/TR vs. % Evaporator Capacity at Time of Defrost Data shown in Tables 4 through 6 was derived using the information from the above discussion and Figures 14 through 19. 28 IIAR 2015 Technical Paper #4
Optimizing Evaporator Runtime and Defrost Frequency Fin Spacing, Room Temperature (SHR): -10 F/85% rh (.95) Evap Capacity at Defrost Freq, Required Tube Pattern fpi Time of Defrost No./day Runtime, h/day 7/8" Staggered 2 0.7 1.3 20 7/8" Staggered 3 0.7 1.6 19 7/8" Staggered 4 0.7 2.4 19 5/8" Inline 2 0.7 1.1 20 5/8" Inline 3 0.7 1.1 19 5/8" Inline 4 0.7 1.4 19 Table 4. Optimized Defrost Parameters for -10 F/85%rh (SHR 0.95) Fin Spacing, Room Temperature (SHR): +10 F/85% rh (.87) Evap Capacity at Defrost Freq, Required Tube Pattern fpi Time of Defrost No./day Runtime, h/day 7/8" Staggered 2 0.65 3.0 18 7/8" Staggered 3 0.65 3.8 18 7/8" Staggered 4 0.65 4.6 17 5/8" Inline 2 0.65 2.5 18 5/8" Inline 3 0.65 2.6 18 5/8" Inline 4 0.65 3.4 18 Table 5. Optimized Defrost Parameters for +10 F/85%rh (SHR 0.87) Fin Spacing, Room Temperature (SHR): +32 F/85% rh (.73) Evap Capacity at Defrost Freq, Required Tube Pattern fpi Time of Defrost No./day Runtime, h/day 7/8" Staggered 2 0.6 6.1 17 7/8" Staggered 3 0.6 7.6 16 7/8" Staggered 4 0.6 9.2 15 5/8" Inline 2 0.6 4.8 17 5/8" Inline 3 0.6 5.1 17 5/8" Inline 4 0.6 6.5 16 Table 6. Optimized Defrost Parameters for +32 F/85%rh (SHR 0.73) Technical Paper #4 IIAR 2015 29
2015 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, CA In no case does it make financial (or practical) sense to use the 5/8" Staggered (compact) pattern when frosting conditions exist. The above discussion and recommendations are based on frost accumulation by deposition and do not include the effects of air-borne ice crystals on fin spacing and defrost frequency. The effects of air-borne ice crystals are briefly discussed in the following section. The Other Type of Frost: Air-Borne Ice Crystals This type of frost is formed quite differently from the frost formed by deposition as explained above. It accumulates on evaporator surfaces by a different mechanism, and is more difficult to quantify and predict. Air-borne ice crystals as a type of frost that can be deposited on coil surfaces has been recognized and discussed for some time (Cleland 2002, Stoecker 1988). These ice particulates form when infiltration air mixes with refrigerated air to produce a supersaturated condition. On a psychrometric chart, a supersaturated condition is indicated when the mixed air condition falls to the left of the saturation (100% rh) line (think of fog that has frozen in mid-air). Rather than accumulate relatively uniformly over the entire coil surface as is the case with frost formed by deposition, air-borne ice crystals accumulate on the leading edges of the coil fins and have the primary effect of restricting airflow. This type of frost is more difficult to predict since its formation depends on not only the condition of the air outside the refrigerated space, but also on the condition of doorways and how they are operated. When evaporators are located directly above doorways where air-borne ice crystals are formed, this type of frost can accumulate very quickly and have serious 30 IIAR 2015 Technical Paper #4
Optimizing Evaporator Runtime and Defrost Frequency consequences in terms of degraded performance and inability to defrost effectively due to excessive accumulation of hoar frost and ice. In one particular case observed, two identical evaporators were installed in the same refrigerated space (a -10 F freezer) along the same wall, one directly over the doorway and the second offset between doorways. The evaporator directly over the doorway had chronic problems with rapid, heavy accumulation of frost, and with defrost issues related to accumulation of ice on the unit cabinet and in the drain pan. The evaporator that was located only 20 feet away between doorways, operated without accumulating ice on the cabinet and/or in the pan and defrosted normally and effectively. It is therefore recommended that evaporators not be located directly above doorways whenever possible. If it is known that the evaporator will be exposed to this type of frost, variable fin spacing is recommended. That is, a fin spacing arrangement which has fins on the first one to two rows on the air entering side of the coil spaced wider than in the remaining rows. Typical arrangements are 1/2 fpi, 1.5/3 fpi, and 2/4 fpi. Conclusions 1. A method has been described which quantifies rate of moisture removal (i.e. frost accumulation) by a refrigeration evaporator as a function of SHR. 2. It has been shown that airflow rate and cooling capacity are significantly reduced over time as frost accumulates. This reduction in airflow and capacity can be very rapid in rooms operating at warmer temperatures (approaching +32 F) and high relative humidity. 3. A model was developed to quantify the negative effects of frost accumulation on evaporator airflow and capacity and to determine appropriate defrost frequency and design runtime for a given set of operating conditions. Technical Paper #4 IIAR 2015 31
2015 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, CA 4. The model was used to demonstrate how tube pattern and fin spacing can be manipulated to manage defrost frequency and the cost of defrosting, and to optimize the design. 5. Of the three tube patterns and three fin spacings examined in this analysis, the 7/8" Staggered pattern with 3 fpi had the lowest first cost in $/TR. 6. Of the three tube patterns and three fin spacings examined in this analysis, the 2 fpi spacing had the lowest annual cost of defrosting (operating cost). 7. In the optimizing analysis, the 7/8" Staggered pattern with 3 fpi gave the best return on investment based on savings in the annual cost of defrosting compared to the lowest first cost design. 8. In all cases (operating conditions and fin spacings), the 5/8" Staggered (compact) pattern had significantly higher annual cost of defrost compared to the other two tube patterns and is not recommended for frosting conditions. References Aljuwayhel, N.F. 2006. Numerical and Experimental Study of the Influence of Frost Formation and Defrosting on the Performance of Industrial Evaporator Coils. PhD Dissertation. University of Wisconsin-Madison. Cleland D.J., O Hagan A.N. 2002. Performance of an Air Cooling Coil Under Frosting Conditions. American Society of Heating, Refrigerating and Air-Conditioning Engineers. ASHRAE Transactions 2002 V. 109, Pt. 1 Cleland, D.J. 2012. The Effect of Water Vapour on Food Refrigeration Systems. The Institute of Refrigeration. London, UK. Proc. Inst. R. 2011-12. 5. 32 IIAR 2015 Technical Paper #4
Optimizing Evaporator Runtime and Defrost Frequency Nelson, B.I., 2011(1), Optimizing Hot Gas Defrost. Technical Bulletin. Colmac Coil Manufacturing, Inc. Colville, WA. Nelson, B.I., 2012(a), Comparing Air Cooler Ratings Part 1: Not All Rating Methods are Created Equal. Technical Bulletin. Colmac Coil Manufacturing, Inc. Colville, WA. Nelson, B.I., 2012(b), Comparing Air Cooler Ratings Part 2: Why DTM Ratings Cost You Money. Technical Bulletin. Colmac Coil Manufacturing, Inc. Colville, WA. Stoecker, W.F. 1988, Industrial Refrigeration Handbook, Business News Publishing Company, Michigan. Welch, J. 2013. Personal correspondence. Technical Paper #4 IIAR 2015 33
2015 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, CA Notes: 34 IIAR 2015 Technical Paper #4