ENERGY EVALUATION OF A

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1 ENERGY EVALUATION OF A FROZEN FOODS PLANT Final Report - March 2000 Prepared by University of Wisconsin HVAC&R Center Funded by SOUTHWESTERN ELECTRIC POWER COMPANY

2 DISCLAIMER OF WARRANTIES AND LIMITATION OF LIABILITIES THIS REPORT WAS PREPARED BY THE ORGANIZATION(S) NAMED BELOW AS AN ACCOUNT OF WORK SPONSORED OR COSPONSORED BY THE ELECTRIC POWER RESEARCH INSTITUTE, INC. (EPRI). NEITHER EPRI, ANY MEMBER OF EPRI, ANY COSPONSOR, THE ORGANIZATION(S) BELOW, NOR ANY PERSON ACTING ON BEHALF OF ANY OF THEM: (A) MAKES ANY WARRANTY OR REPRESENTATION WHATSOEVER, EXPRESS, OR IMPLIED, (I) WITH RESPECT TO THE USE OF ANY INFORMATION, APPARATUS, METHOD, PROCESS OR SIMILAR ITEM DISCLOSED IN THIS REPORT, INCLUDING MERCHANTABILITY AND FITNESS FOR A PARTICULAR PURPOSE, OR (II) THAT SUCH USE DOES NOT INFRINGE ON OR INTERFERE WITH PRIVATELY OWNED RIGHTS, INCLUDING ANY PARTY'S INTELLECTUAL PROPERTY, OR (III) THAT THIS REPORT IS SUITABLE TO ANY PARTICULAR USER'S CIRCUMSTANCE; OR (B) ASSUMES RESPONSIBILITY FOR ANY DAMAGES OR OTHER LIABILITY WHATSOEVER (INCLUDING ANY CONSEQUENTIAL DAMAGES, EVEN IF EPRI OR ANY EPRI REPRESENTATIVE HAS BEEN ADVISED OF THE POSSBILITY OF SUCH DAMAGES) RESULTING FROM YOUR SELECTION OR USE OF THIS REPORT OR ANY INFORMATION, APPARATUS, METHOD, PROCESS, OR SIMILAR ITEM DISCLOSED IN THIS REPORT ORGANIZATION(S) THAT PREPARED THIS REPORT - HVAC&R CENTER This version of the report has been revised to remove identifying information and specific references to the client.

3 Table of Contents Introduction... 1 Summary of Recommendations... 1 Facility Description... 2 Refrigeration System... 2 Implement Floating Compressor Head Pressures... 3 Add a Compressor to Handle Flash Gas... 5 Evaluate Suction Risers at the Spiral Freezer Evaporators... 7 Safety Considerations... 7 Space Conditioning Systems... 8 Makeup Air... 8 HVAC Units... 9 Desiccant Dehumidification Glycol Loop Space Conditioning Steam Steam Jet Refrigeration Ideas for Identifying Energy Conservation Projects Acknowledgments... 12

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5 Frozen Foods Plant Introduction This is a report on an energy evaluation of a frozen foods plant in northwestern Arkansas. The HVAC&R Center at the University of Wisconsin performed this evaluation at the request of plant management and Southwestern Electric Power Company (SWEPCO), to identify opportunities to reduce energy use and costs while maintaining or increasing product quality and production capabilities. Mr. Jim Elleson and Mr. Jim Denkmann from the HVAC&R Center visited the plant on November 8, 1999, met with plant staff and observed the layout and operation of the plant. This report describes the existing conditions in the plant and recommends measures for improving the efficiency and effectiveness of the systems. A detailed assessment of all the systems in this large plant was not possible with the limited time available. Our evaluation emphasizes identifying opportunities in the refrigeration and spaceconditioning systems that have good potential for success. The goals of this report are to recommend specific improvements that we identified in our site visit, and to provide guidance to the utilities staff in developing a comprehensive, prioritized list of opportunities for further analysis. Summary of Recommendations Our major recommendations for the ammonia refrigeration system and other systems are: 1. Implement floating head pressure control. Dedicate one evaporative condenser to serve the thermosiphon oil coolers, convert high-pressure receivers to controlled-pressure receivers, and install liquid drainers on the evaporative condensers. 2. Install a variable-speed reciprocating compressor, with appropriate controls, to compress flash gas and allow three stages of liquid expansion and vapor compression. 3. Investigate the impact of static pressure penalty at the suction risers at the spiral freezer evaporators. Make appropriate ammonia piping revisions if a significant temperature penalty is evident. 4. Modify the evaporators in the storage freezer vestibule to feed them with high-pressure liquid. Control the evaporators to operate with a surface temperature of F. 5. Evaluate the exhaust air-makeup air balance in each area of the plant. Relocate or add makeup air as needed to provide an appropriate balance in each area. 6. When an exhaust air-makeup air balance has been established, consider conditioning additional production areas for increased comfort and productivity. 7. Identify and repair leaking steam traps. 8. Evaluate using heat recovery from exhaust air to preheat makeup air. Additional recommendations include: 1

6 1. Install a solenoid valve in each spiral freezer evaporator's hot gas relief line, to open only at the end of the defrost cycle. 2. Consider moving one accumulator from a sub-basement to an outdoor location to improve safety and accessibility. 3. When HVAC units are being replaced or upgraded, consider serving the HVAC loads from the ammonia refrigeration system. 4. Investigate using high-temperature exhaust air to regenerate desiccant dehumidifiers. 5. Where possible, reduce the energy cost of glycol cooling systems by decreasing the brine concentration, raising the brine supply temperature, or raising the ammonia evaporating temperature. 6. Install suction/liquid line heat exchangers on any HVAC units that use refrigerants R-404A or R-507A. 7. Re-evaluate the use of steam jet refrigeration in light of its high energy cost. Finally, we recommend that energy conservation projects be evaluated and prioritized using a systematic approach. The energy efficiency program should consider the relative energy costs of each process and product, as well as the effects of refrigeration and HVAC system performance. The program should also include measurement of the savings achieved by projects. Details of the systems and additional information on these recommendations are provided in the following sections. Facility Description The frozen foods plant produces frozen dinners, breakfasts, and pot pies. The plant, which encompasses 375,000 ft 2 under roof, has a peak demand of roughly 8.5 MW with a 70% annual load factor. The power cost averages 4 per kwh. The utilities staff has aggressively pursued energy efficiency improvements in the plant. Refrigeration System Overall, the ammonia refrigeration system is very well designed, operated, and maintained. It is unusual to see a design that takes total system pressure losses into account. The use of butterfly valves as control valves on suction lines and at compressors in the place of suction checks and stop valves is an excellent feature, especially when working with evaporating temperatures below -40 F. The engine room consists of 17 twin screw compressors, most of which are 350 hp, in a two-stage configuration. All compressors are Vilter twin screw packages with thermosiphon oil coolers. Compressor head pressures are normally held between 150 and 160 psig but sometimes rise as high as 190 psig during summer months. The high summer head pressures are evidence of insufficient heat rejection capacity. Factors that could be contributing to this shortfall include: Evaporative condenser capacity not matched to the load Waterside fouling or scaling of condenser tubes Uneven water or air distribution in evaporative condensers Noncondensables not being purged from the system 2

7 Liquid ammonia hanging up in the condensers The utilities staff has changed to tighter coalescer filters on the oil eliminators to try to reduce the head pressure. In a recent test the pressure was successfully dropped to 135 psi. However, oil cooling problems arose, so the head pressures are currently being maintained at a minimum of 150 psig. All suction accumulators and their respective refrigerant pumps are located remotely from the engine room. All but one of the accumulators (total of 5) are located outside the plant in the ambient. This provides an additional element of safety in that all valve servicing is done outdoors. Multiple remote accumulators also simplify the refrigerant piping and decrease system pressure losses due to shorter wet suction mains. One accumulator is located in a sub-basement and was experiencing a small ammonia leak at the time of our visit. We did not notice any air supply and exhaust system for this room to move air in and out should a leak occur. The possibility of moving this vessel to an outdoor location should be evaluated. Product freezers include five spiral freezers and some Amerio plate freezers. The nominal evaporating temperatures for the low-side loads are -40 to -50 F at the Amerio freezers and -51 to -56 F at the blast coils. At the time of our visit, the low-side compressors were running at 14 in Hg at the compressor suction, or about -50 F. The capacity of the Amerio freezers sometimes runs short in the summer, and the product throughput must be reduced. The storage freezer vestibule is conditioned by dedicated evaporators. Ideally, these units should be controlled to operate with a surface temperature just above 32 F. This is the most efficient way to remove moisture from the air, since no defrost is required. However, the vestibule units are currently served by the same liquid pump that serves the freezer units, which does not allow them to operate at the desired temperature. A project is currently under consideration to re-pipe the vestibule evaporators to feed them with high-pressure liquid. This modification will allow these units to operate at the correct temperature, and will result in a large energy saving by eliminating defrost for the vestibule and by reducing the latent load and the defrost requirement of the freezer units. The hot gas piping for the spiral freezers has a relief line to the suction main, to relieve the hot gas pressure at the end of the defrost cycle and prevent slamming when the main valve opens. There is a manual valve in this line that is always open 1 turn, which allows hot gas to bypass to suction whenever the hot gas solenoid is open. Installing a solenoid in this relief line, to open only at the end of the defrost cycle, would save refrigeration energy and capacity. Major recommendations for the refrigeration system are discussed in the following sections. Implement Floating Compressor Head Pressures A floating compressor head (or discharge) pressure is one that is free to move up and down with seasonal swings in the ambient wet-bulb temperature. Taking advantage of low wet-bulb temperatures by reducing the head pressure results in tremendous energy savings. As a rule of thumb, compressor energy will be reduced about 1% for every 2 psig reduction in saturated condensing pressure. 3

8 The following system enhancements are needed to implement this desirable feature on this particular system: Place all thermosiphon oil cooler loads on a separate evaporative condenser. Currently, the minimum head pressure is determined by oil cooling considerations. The oil temperature in the thermosiphon oil coolers must be held at about 130 F on a year-round basis. This requires a minimum refrigerant temperature in the thermosiphon of about 85 F, which prevents the condensing pressure from being reduced below 150 psig. In most cases it is possible to reconfigure the existing hot gas piping and thermosiphon receiver so that the oil coolers are placed on one of the existing evaporative condensers, re-piped for this purpose. This dedicated condenser can control the oil temperature directly by varying the condenser fan speed. The remaining condensers serving the refrigeration loads can then be controlled to provide much lower head pressures without affecting oil cooling. An added benefit is that the problem of draining oil out of thermosiphon oil coolers is eliminated altogether because the oil cooler piping circuit is separate and distinct. No oil resides in this circuit, only pure liquid ammonia. This relatively new technique has been implemented in other refrigeration systems with great success. Install liquid drainers on each evaporative condenser outlet line. Currently, the capacities of the individual evaporative condensers are equalized by liquid backing up from the outlet manifold. All the condensers on one manifold must discharge to the manifold pressure. If one condenser has more capacity than the others, liquid will back up into the coil until its capacity is reduced to match the others'. This effect limits the capacity of any one condenser to the "lowest common denominator" of the worst-case condenser on the manifold. This reduction in condenser capacity contributes to the high discharge pressures during the summer. It can also create considerable liquid hang-up during cold weather operation. It is common for high-pressure receivers to completely run dry then, with little warning, completely overfill with liquid a few minutes later. Installing a liquid drainer on each evaporative condenser outlet line allows each condenser heat exchanger to function at its maximum available capacity, independently from the other condensers. This revision necessitates insulating all piping downstream of the drainers, since it is now at a lower pressure and temperature. Convert the existing high-pressure receivers to controlled-pressure receivers. Installation of liquid drainers at the condensers means that the existing high-pressure receivers will become controlled-pressure receivers. The receiver pressure would become a variable, probably running in a range between 90 psig during hot summer weather to 60 psig during the winter. The present gas vent equalizer lines from the high pressure receivers should be disconnected from the hot gas inlet piping at the condensers and be reconnected to either: A high-stage suction main, or A new optional reciprocating compressor (discussed later). These receivers will now require insulation because the ammonia temperature inside the receivers will be in the range of 40 to 60 F. 4

9 Add a Compressor to Handle Flash Gas Adding a compressor to handle the flash gas produced in throttling to the new controlled-pressure receivers will increase the net refrigerating effect. This change will provide additional capacity and increase the efficiency of the system. At the present time, ammonia vapor is compressed in two steps or stages, first at the booster compressor and again at the high-stage compressor. Liquid ammonia is expanded twice after it leaves the condenser, once at the intercooler and again at the respective low temperature accumulator. Installing liquid drainers and converting the existing high-pressure receivers to controlledpressure receivers, as discussed above, adds an additional step of expansion. The full benefit of this modification can be achieved by adding a new compressor to compress the associated flash gas from the new intermediate pressure. In the modified configuration, liquid ammonia leaving the condensers passes through high-side floats (liquid drainers), reducing its pressure to the new intermediate pressure. The liquid, and the flash gas vapor formed in the throttling process, then flow into the controlled pressure receivers (CPRs), where the vapor separates from the liquid. The vapor flows from the CPRs to a new reciprocating compressor, where it is recompressed back up to the condensing pressure. This portion of the total system mass flow now no longer flows to the lower temperature screw compressors. The system is now functioning in a three-stage liquid throttling configuration, all done without having to revise the existing screw compressors and associated intercoolers or accumulators. Figure 1 illustrates the recommended modifications. Compressing the flash gas at a high suction pressure provides additional refrigeration capacity at a very high efficiency. The flash gas compressor operates with a compression ratio less than 2:1 during all running hours of the year The new compressor should be variable-speed, with a constant speed oil pump. The variablespeed capability does not save additional energy, but it allows much better control of the CPR pressure, which in turn helps stabilize system pressures overall. Any evaporators now being fed with high-pressure liquid would be reconfigured to receive either subcooled medium-pressure liquid from a controlled pressure receiver, or pumped mediumpressure liquid from a new liquid feed pump. These two options should be evaluated for each evaporator to determine which is preferred for each particular load. The pressure in the CPR would be set at a maximum of about 90 psig, and would be reset with the condensing pressure. The CPR pressure should be maintained at least psig below the condensing pressure, to ensure that liquid continues to flow to the CPR. The exact setpoint and differential should be determined by a more detailed analysis of the system. 5

10 Recip. Compressor 195 HPR To oil coolers CPR 105 Liquid drainers To oil coolers Intercooler 45 High-temperature loads Intercooler 45 High-temperature loads LPR 12 Medium-temperature loads LPR 12 Medium-temperature loads LPR 6 Low-temperature loads LPR 6 Low-temperature loads Existing System Revised System Figure 1. Recommended modifications to accomplish an additional stage of expansion and compression. Approximate pressures are shown in psia at design conditions. 6

11 Evaluate Suction Risers at the Spiral Freezer Evaporators The current piping arrangement for the spiral freezer evaporator coils may be causing a temperature penalty in the lower coils, which could have a significant effect on the freezer capacity. The suction piping for the spiral freezers is configured so that the line from the lower evaporator forms a riser wherein both liquid and vapor must be carried up to the suction main. The line for the upper coil has no such vertical run. Figure 2 illustrates the layout. When this system was designed, it was common practice to run the vertical suction riser in the same pipe size as the coil connection, with no pipe reducers. This sizing approach may actually increase pressure losses instead of decreasing them. If the velocity in the riser from the lower evaporator is not sufficient to entrain liquid up to the suction main, a column of liquid will develop in the riser. This places a back pressure on the lower coils, raising their operating temperature. At -50 F, a column of liquid ammonia imposes a static pressure penalty of 0.30 psi per foot of liquid. A ten-foot liquid column will cause a 3.0 psig pressure difference, which results in a saturated liquid temperature of -39 F at the bottom of the riser. This is a significant temperature penalty. 300 tons what is the impact of this riser length? ~50 ft 300 tons how much warmer is the bottom coil than the top coil? Figure 2. Section through spiral freezer. We recommend that plant personnel measure the actual operating temperatures in the spiral freezers, to determine whether there is a significant temperature difference between the upper and lower coils. If such a difference is found, resizing of the suction risers or installation of double suction risers should be investigated. Further information on this topic is provided in the 1998 ASHRAE Refrigeration Handbook, pp In addition, research currently underway at the HVAC&R Center is leading to improved methods for sizing suction risers. Safety Considerations The risk of injury to maintenance personnel goes up with increased ammonia pressures. Our recommendations reduce the quantity of running feet of high-pressure piping and the total 7

12 ammonia system charge contained within the main refrigeration system. Instead of high pressure receivers, the system now has medium (controlled) pressure receivers. High-side vessel pressures are reduced by half or more. Draining oil from a thermosiphon oil cooler is a dangerous task because of the risk of an ammonia release from a high-pressure source, namely from the oil cooler itself. This is also a particularly dangerous place to drain oil because of the risk of an atomized oil spray. An oil mist released within a confined space carries with it a very high risk of explosion because the ignition temperature of the oil mist is so low (around 300 F, compared with 1600 F for ammonia). Serving the oil coolers from a dedicated condenser completely eliminates the task of draining oil from the oil coolers. While high-pressure ammonia will continue to circulate to the oil coolers, it is oil-free, and in its own separate piping circuit. Regular maintenance access to the oil coolers is unnecessary because no moving parts exist anywhere in this piping. In addition, the total ammonia charge in the main refrigeration system is reduced by the amount of ammonia in the thermosiphon circuit. The EPA RMP worst-case release scenario may be favorably impacted by this modification. Space Conditioning Systems Makeup Air Any air that is exhausted from a space must be balanced by an equal volume of makeup air. If the makeup air is not mechanically supplied to the space, it will flow from other areas or from outside the building. Uncontrolled makeup air has a detrimental effect on temperature and humidity control. In the winter, cold makeup air entering from loading docks and doorways causes uncomfortable drafts. Excessive heating system operation is often required to offset the effects of these drafts. In the summer, hot humid air from docks and doorways increases the loads in conditioned areas, and adds to defrost requirements. Makeup air flowing from one production area to another can carry odors and other contaminants, as well as interfering with temperature control in conditioned spaces. We noticed substantial airflows between areas, and from loading dock areas into production areas. This is an indication that exhaust air is not properly balanced by makeup air. For example, there was significant airflow into the pie-casing area from nearby loading docks. This uncontrolled makeup air increases the space-conditioning load and contributes to moisturecontrol problems in the pie-casing area. We also noticed air flowing between conditioned and unconditioned areas. While this airflow may help to moderate temperatures in unconditioned areas, it increases the air-conditioning loads and diminishes the ability to control space temperatures and energy use. We recommend that the exhaust-makeup air balance in the entire plant and in each individual space be evaluated to ensure that airflows within the plant are properly controlled. The first step in this evaluation is to tabulate all exhaust and makeup air quantities for each area. The exhaust tabulation should include flue exhaust from combustion processes. Makeup air quantities for conditioned areas include only the outside air intake volume, not the recirculated air. Equipment ratings can provide a rough estimate of airflow volumes, although these estimates should be verified by measurements of actual airflow before implementing system modifications. 8

13 Any area that is short of mechanically-supplied makeup air is drawing its makeup from adjacent areas or from outdoors. The flow of air in the plant can be controlled by the placement of makeup air inlets relative to the exhaust outlets. Areas with high generation rates of heat, odors or contaminants should be maintained at negative pressure, with more exhaust than makeup air. The additional makeup air quantities should be supplied to adjacent areas. Conditioned areas, especially those with special temperature or humidity requirements, should be maintained at neutral or slightly positive pressure, to avoid increasing the space loads and to minimize the loss of conditioned air. If new makeup air units are added, exhaust air heat recovery for preheating makeup air should be considered. HVAC Units The facility is served by a large number of packaged rooftop heating, ventilating and airconditioning (HVAC) units, which use a variety of commonly applied halocarbon refrigerants. There are four McQuay rooftops with capacities of 85 to 125 tons, and many smaller 5- to 35-ton units. The total capacity of the HVAC equipment is approximately 1,000 to 1,500 tons. With a few exceptions, the halocarbon refrigeration equipment is air-cooled. Converting the halocarbon equipment to serve the HVAC loads from the ammonia refrigeration system could provide large demand and energy savings. This equipment typically has 20-30% higher peak demand, and uses 30-60% more annual energy, than an ammonia system serving the same loads. Much of this increased energy use can be attributed to the difference between aircooled and evaporative-cooled heat rejection. The air-cooled halocarbon units use thermal expansion valves and reject heat at about 130 F saturated condensing temperature, while the evaporative-cooled ammonia system rejects heat at about 95 F saturated condensing temperature. In addition, the efficiency of the relatively small packaged units declines considerably at low-load conditions, while the ammonia system's efficiency would not be affected by variations in HVAC loads. Converting 1000 tons of HVAC loads from halocarbon equipment to the ammonia system could result in 400 to 600 kw of peak demand reduction, and up to 2 million kwh of energy savings. The conversion would require installation of chilled water or glycol loops, as direct ammonia evaporators in the airstream would not be allowable for most areas. While it is probably not practical to convert all of the loads at one time, conversion should be considered when HVAC units are being replaced or upgraded. For units that use refrigerants R-404A or R-507A, we have found that installing suction/liquid line heat exchangers can result in appreciable capacity increases while simultaneously saving energy. This option should be studied for any present rooftop equipment using either of these two refrigerants where these loads cannot be reassigned to the ammonia refrigeration system or where adding a glycol heat transfer loop is not possible. Establishing a balance of exhaust and makeup air will allow better control of space conditioning in individual areas. When this balance has been established, the possibility of conditioning additional areas should be considered, since the conditioned air can now be targeted to the area where it is needed. Makeup air can be cooled and dehumidified to a dewpoint of F quite efficiently, providing increased comfort and productivity with minimal energy expenditure. 9

14 In some areas where exhaust air quantities are determined by heat removal requirements, adding mechanical cooling can reduce the exhaust volume and the associated makeup air volume. The reduction in fan energy further reduces the energy impact of mechanical cooling. Desiccant Dehumidification A desiccant dehumidifier uses a desiccant medium to adsorb water vapor from a moist air stream. The medium is regenerated at intervals by heating it to drive off the adsorbed moisture. A desiccant dryer can effectively reduce the latent load in a space, reducing or eliminating defrosting and preventing other problems with frost and ice. However, the desiccant dehumidification process also imposes an increased sensible load, because of the heat that is added to regenerate the desiccant medium. Our studies to date indicate that neither desiccant nor vapor-compression dehumidification is clearly preferred from an energy-efficiency standpoint. The best approach for a given application is determined by site-specific details such as infiltration loads, equipment limitations, and special conditions in the particular space. The pie-casing area and the service room immediately outside the pie spiral freezer formerly suffered from excessive humidity, to the extent that ice would build up along the service corridor common wall and the spiral access doors. A gas-fired desiccant dryer was installed to maintain this space at -4 F dewpoint. This has solved the humidity problem and eliminated the ice formation. The desiccant unit has a chilled glycol coil to cool the discharge air if the space temperature rises too high, but it has rarely if ever been used. The pie-casing area is open to the adjacent dinner-casing area. It is likely that the additional sensible load from the desiccant unit is being taken up by other units in the vicinity. It is also likely that the pie-casing desiccant unit is meeting latent loads from adjacent areas. This crossloading could represent unnecessary energy use if the dehumidification requirements in the other areas are not as strict as in pie-casing. The availability of relatively high-temperature exhaust air offers an opportunity for desiccant dehumidification. Packaged desiccant units typically heat outside air to F to regenerate the desiccant medium. Exhaust air in this temperature range could be used for regeneration, drastically reducing the desiccant system's operating cost. If the exhaust stream contains grease, particulates, or other impurities that could contaminate the desiccant medium, an air-to-air heat exchanger could be used to transfer heat from the exhaust to the regeneration air. Glycol Loop Space Conditioning Some production areas are conditioned with propylene glycol secondary heat transfer fluid, chilled in plate-frame heat exchangers. We noticed frost forming on the heat exchangers (see Figure 2), which indicates that the ammonia evaporating temperature is in the vicinity of F, or about 40 psig. Assuming a typical approach of 4 F, we could expect to see glycol supply temperatures to the area coolers in the range of F. This is an unusually low brine supply temperature for a space held at 75 F. While the coolers have relatively high fin spacing (4 fins/inch), the coils are deep enough that they should not need a 32 F supply medium. Sometimes space coolers, using brine as the heat transfer medium, are erroneously circuited for parallel flow. This is common practice when circuiting for ammonia but inapplicable when brines are used. If the coolers were circuited in parallel, this would probably 10

15 explain why the brine supply temperatures are held so low. Any parallel-flow brine coils should be reconfigured for counterflow heat transfer. In any event, the total connected brine pump horsepower is not insignificant. It may be possible to reduce this horsepower. We recommend studying the impacts of: Decreasing the brine concentration, Raising the brine supply temperature, and Raising the ammonia evaporating temperature The viscosity of propylene glycol varies directly with an increase in its concentration and inversely with its temperature. Anything that can be done to decrease the glycol concentration and increase its temperature will result in significant pump energy savings. Figure 3. Frost on plate-frame heat exchanger. Steam Steam for process needs, space heating, and makeup air heating is produced in two water-tube boilers. Boiler pressure has been reduced from 125 psi to 95 psi. Some low-pressure condensate is returned to the boilers. In some areas, condensate return has been abandoned because of leaking steam traps and excessive hammering. The utilities staff has considered installing thermocouples on the condensate returns from each zone to monitor their temperatures and detect leaking traps. Projects to return condensate to the boiler and correct leaking steam traps have large energy savings and are generally very cost-effective. Table 1 shows the cost of a leaking steam trap or other leak, at 100 psig, with a steam cost of $4.50 per 1000 lbs. Table 1. Cost of Steam Leaks Equivalent Orifice lbs steam lost Cost per year Diameter per year 1/16" 115,630 $520 1/8" 462,545 $2,080 1/4" 1,848,389 $8,300 1/2" 7,393,432 $33,200 11

16 Steam Jet Refrigeration Steam jet refrigeration is used for cooling some ingredients. The steam jet process allows some loads to be met without affecting the ammonia system, and without providing a chilled-water loop and plate-frame heat exchanger. However, this refrigeration process is very inefficient. Steam consumption is approximately 26 lbs of steam per ton of refrigeration, which is a coefficient of performance (COP) of In contrast, the COP of mechanical refrigeration to meet this load is 1.8 to 2.3. The energy cost of steam jet refrigeration at this plant is about $0.12 per ton-hour, while the cost of mechanical refrigeration would be $ per ton-hour. The utilities staff has made a concerted effort to reuse water from once-through cooling processes, by converting to closed-loop cooling, or by using the waste water for makeup to evaporative condensers. Steam jet systems currently using once-through water for condensing could be served by cooling towers, reducing the waste and expense of city water. Ideas for Identifying Energy Conservation Projects The following basic guidelines are useful for identifying and prioritizing energy conservation opportunities: Develop an end-use breakdown of the energy used in the plant. Measure, calculate, or estimate the amount of energy used in each major system: refrigeration, space conditioning, steam. Break these down further into usage for specific subsystems and processes. Account for all the energy purchased by the plant. Determine the energy cost for each process, in terms of cost per hour of operation or per pound of product. Calculate the energy cost of each product. Identify those processes that add the most energy cost to the products. Identify those products whose energy cost is a significant percentage of their total production cost. Consider the effects of space conditioning on comfort, productivity and labor costs. If productivity goes down during very hot weather, estimate the value of improved comfort conditions in terms of increased productivity. Focus on the products and the processes that have the most significant energy costs for possible modifications and improvements. Measure the impacts of conservation projects on energy use and productivity. Acknowledgments We would like to thank Mark Mobley of SWEPCO, and members of the plant utilities staff, for their assistance in gathering information for this report. 12