Technical Papers. 32nd Annual Meeting. International Institute of Ammonia Refrigeration. March 14 17, 2010

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1 Technical Papers 32nd Annual Meeting International Institute of Ammonia Refrigeration March 14 17, Industrial Refrigeration Conference & Exhibition Manchester Grand Hyatt San Diego, California

2 ACKNOWLEDGEMENT The success of the 32nd 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. Board of Directors, International Institute of Ammonia Refrigeration 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 (voice) (fax) Industrial Refrigeration Conference & Exhibition Manchester Grand Hyatt San Diego, California

3 Technical Paper #5 Consider Two-Stage Systems to Save Power Richard A. Corbett, P.E. R.A. Corbett Engineering Houston, Texas Abstract In recent years, many low-temperature ammonia refrigeration systems using screw compressors have used single-stage compression to satisfy refrigeration loads in lieu of traditional two-stage systems. This paper compares performance data for both types of systems and estimated operating costs based on total brake horsepower (BHP) requirements for each type of system. The performance data shows that two-stage systems provide significant power savings without significantly increasing system complexity. The data also provides guidance for when to specify two-stage systems over single-stage systems. Performance charts are presented for low-temperature saturated evaporating temperatures ranging from 20 F to 45 F. The charts show total system BHP requirements and estimated costs for the required total BHP at each low-temperature. Total BHP includes compressor BHP, condenser BHP, and recirculation pump BHP. The typical system used for the comparison is based on 300 tons of low-temperature refrigeration load and 200 tons of high-temperature refrigeration load. The system used for comparison is comprised of one low-temperature recirculator operating at various low temperatures, one hightemperature recirculator operating at +25 F saturated evaporating temperature, rotary screw refrigeration compressors with thermosyphon oil cooling, and evaporative condensers operating at 95 F saturated condensing temperature. The system loads and operating temperatures are typical of many refrigeration systems. IIAR

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5 Consider Two-Stage Systems to Save Power Introduction Two-stage refrigeration systems can save significant energy and operating cost over single stage and economized systems, particularly for systems operating below 20 F ( 29 C). When there is ample high-stage load, as there is with most industrial ammonia systems, two-stage systems do not substantially increase the complexity and cost of the system. Cost savings are achieved at both low and high condensing pressures. At temperatures above 20 F ( 29 C), the costs savings diminish but still provide power savings over single-stage systems, particularly at condensing pressures above 150 psig (1034 kpa). To demonstrate the energy savings available with two-stage systems, a performance study was conducted to compare power consumption for three types of ammonia refrigeration systems typically used in industrial applications (particularly food processing and storage). The power consumption for a two-stage system was compared to the power consumption of single-stage and economized systems. Annual power cost savings were estimated from the results of the power-consumption comparison. The performance study was based on real-world compressor performance data rather than theoretical thermodynamic performance comparisons. Engineers should be able to reasonably duplicate the results for actual system designs and installations of these and other systems. Systems Studied Three ammonia refrigeration system types were considered for the study (Figures 1 3). All three systems are designed for liquid recirculation (liquid overfeed, mechanically pumped) for both low-temperature and intermediate-temperature Technical Paper #5 IIAR

6 2010 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, California users. The heat loads and operating conditions chosen are typical of actual operating refrigeration systems. The systems studied are typical of systems used in cold-storage and food processing facilities. The low-temperature and high-temperature heat loads chosen represent typical loads for those facilities. Obviously, actual system loads can vary widely, but the loads chosen provide results representative of cold-storage and food-processing operations. The low-temperature heat load and operating conditions used for the study were 300 tons ( KW) for the low-temperature users, with the saturated suction temperature (SST) varied from 0 F (17.8 C) down to 50 F ( 45.5 C). The intermediate heat load and operating conditions were 200 tons (713.5 KW) at 20 F ( 6.7 C), with the condensing pressure varied from 100 psig (689.5 kpa) to 200 psig (1,397 kpa). All three systems use rotary screw compressors, thermosyphon oil cooling, and evaporative condensers. The compressor and condenser symbols on the system flow diagrams (Figures 1 3) may represent one or more compressors and condensers. Although the systems studied herein are typical of many systems installed in the industry, there are other system arrangements that were not studied. Performance comparisons for those systems may provide similar results. Those systems, however, should be studied specifically to confirm the performance advantage of the two-stage system. Two-Stage System The two-stage system (Figure 1) is representative of systems typically used in the industry. The system booster compressors discharge into the high-temperature recirculator (HTR). The HTR serves as the intercooler for the system. The lowtemperature recirculator (LTR) receives intermediate-pressure liquid refrigerant (IPL) from the HTR. The liquid temperature is 20 F ( 6 C). The HTR receives liquid (HPL) 4 IIAR 2010 Technical Paper #5

7 Consider Two-Stage Systems to Save Power at condensing temperature from the high-pressure receiver (HPR). For this paper, the system will be designated as two-stage. Single-Stage System The single-stage system (Figure 2) is also representative of systems installed in the industry. However, when compared to the two-stage system, both low-temperature and intermediate-temperature compressors discharge directly to the condensers. The LTR receives sub cooled liquid refrigerant from the HTR with the liquid at 20 F ( 6.6 C). This is important to keep in mind because the system described here is not a classic single-stage system where the refrigerant liquid supplied to the lowtemperature evaporators would be at condensing temperature. The single-stage system used for the study provides significant thermodynamic advantage over the classic single-stage system. However, single-stage will be used hereafter to identify this system. Because the classic single-stage system high pressure liquid supplied directly from the condensers to the LTR would be significantly less efficient than the single-stage system evaluated here, the classic single-stage system was not evaluated for this study. Economized System The economized system (Figure 3) uses screw compressors equipped with economizer side ports. The compressor main suction is connected to the LTR. The suction from the HTR is connected to the economizer ports. The LTR receives liquid refrigerant at 20 F ( 6 C) from the HTR. Approximately 75 tons of the intermediatetemperature load from the HTR passes to the compressor via the compressor economizer ports. The 125 remaining tons of intermediate load must be handled by an additional high-temperature compressor because of limited capacity of the main Technical Paper #5 IIAR

8 2010 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, California compressor economizer port. economized will be used hereafter to identify this system. System Evaluation Comparison The system performance of each of the three systems described was evaluated to determine the total compressor power required for each of the systems at varying saturated evaporating temperatures and condensing pressures. The differences in power required among the three systems were determined from the total power requirements of each of the systems. The power requirement differences were then used to compare estimated power cost differences. Although the evaluation is for 300 tons (1070.3KW) of low-temperature load and 200 tons (713.5 KW) of intermediate temperature load, the evaluation results can be scaled up or down, as long as the ratio between low-side and intermediate-side loads remains essentially the same. Different load ratios will require re-evaluation at the actual loads to reach valid comparisons. Compressor Performance The power requirements were determined from compressor rating programs available from various compressor manufacturers, and from rating information and techniques developed by the author from historical experience. The results obtained from the various programs were combined and adjusted to represent average performance of the compressors at the operating conditions considered for the evaluation. The power requirements provided in this report do not represent the specific performance of any make or model of compressor, nor do they represent performance data specific to any make or model of compressor. 6 IIAR 2010 Technical Paper #5

9 Consider Two-Stage Systems to Save Power Performance Evaluation Basis The performance comparisons were evaluated at the operating conditions previously stated. Allowances for normal system losses were included in the evaluation and are equivalent for all three systems. Variations in recirculation pump power requirements and evaporator power requirements (air-unit fans), were not considered for the evaluation because the power consumption variations are the same for all three systems. Condenser heat rejection requirements are less for the two-stage system, resulting in slightly less power required for the condenser fans and water pumps. However, the power savings are small and, conservatively, not included in the comparisons. System performance is based on no liquid subcooling from the condensers and 10 F (5.5 C) of suction superheating for all compressor suction conditions. Suction and discharge losses were included in the ratings and were set the same for all compressors. Refrigerant pump performance was considered equal for all three systems, and pressure losses in liquid and suction piping to and from the recirculators were also considered equal for all three systems. For the two-stage system, the high-stage compressor capacity requirement is the booster heat rejection (low-temperature load plus power required) minus the booster oil cooler heat load. The oil cooler heat load of the booster compressor does not flow to the HTR, but directly to the condensers via the thermosyphon oil cooling system. For the single-stage system, both high-temperature and low-temperature compressors discharge to the condensers. The performance evaluation is based on 20 F ( 6.6 C) liquid supplied to the low-temperature recirculator. The high-temperature compressor capacity requirement, therefore, includes the heat load to cool the liquid flowing to the LTR from saturated condensing temperature to 20 F ( 6.6 C). Technical Paper #5 IIAR

10 2010 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, California For the economized system, the compressor performance is based on economized operation with the HTR serving as the flash economizer. The intermediate refrigeration load also passes to the HTR. Because the compressor economizer ports are limited to a high-stage:low-stage ratio of about 25%, an additional compressor(s) is required to handle the balance of the intermediate-temperature load. Power Cost Comparison The comparison of power costs among the three systems studied is based on system operation at varying condensing pressures throughout a year of operation. A year s operation totaled 8,760 hours of operation. To account for the performance effects due to condensing pressure variations throughout the year, 3 periods, each 4 months in length, were used. For the coolest period of the year, the system operated for 1,947 hours at 150 psig ( kpa) and 973 hours at 165 psig ( kpa). For the moderate temperature period of the year, the system operated 1,947 hours at 165 psig ( kpa) and 973 hours at 175 psig ( kpa). For the warmest months the system operated at 1,947 hours at 165 psig ( kpa) and 973 at 185 psig ( kpa). The variations are estimates only, but generally reflect reasonable operating hours of a typical system. Systems generally do not normally operate below 150 psig ( kpa) or normally above 185 psig ( kpa). The hours used reflect 16 hours per day at the lower condensing pressure for each period and 8 hours per day at the higher condensing pressure for each period. Power costs were estimated at $0.08 per KWH. The power cost per KWH is typical of power costs across the industry, however wide variations in costs per KWH have been observed due to the effects of long-term power-supply contracts and differences in demand-based contracts. Regardless, the $0.08 per KWH provides a reasonable base cost for comparison. 8 IIAR 2010 Technical Paper #5

11 Consider Two-Stage Systems to Save Power Study Results The results of the performance study are illustrated in Figures 4 through 9 and in Tables 1 through 6 for those desiring detailed values. Figures 4 through 8 show the power requirement comparisons of the three systems at varying evaporating temperatures and condensing pressures. Tables 1 through 3 provide the brake horsepower (BHP) required for the two-stage, single-stage, and economized systems respectively. The evaluation revealed an interesting result, in that the economized system required the highest BHP at all of the various conditions of evaporating temperature and condensing pressure. This can be clearly seen on Figures 4 though 8, and by comparing Table 3 with Tables 1 and 2. Because the economized system BHP requirements exceeded the single-stage system BHP requirements, the economized system was not included in the power-cost evaluation. It is apparent that the single-stage system, with sub cooling of the liquid flowing to the LTR is more thermodynamically efficient than an economized system. The lower efficiency can be attributed to the loss associated with the economizer port pressure falling significantly below 33.5 psig (233.9 kpa), the intermediate evaporating pressure, at low evaporating temperatures. As the compressor suction pressure falls, the economizer port pressure falls proportionately. The loss can also attributed to the limited economizer port capacity to handle the entire 200-ton side load of this evaluation. Table 4 shows the power saved by the two-stage system over the single-stage system for full-load operation of the systems. Table 5 shows the savings for an annual operation when the system operates for the hours stated previously at varying condensing pressures during three different periods of the year. Technical Paper #5 IIAR

12 2010 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, California Using the results in Table 5, the annual operating cost savings of the two-stage system over the single-stage system are shown in Figure 9. Table 6 shows the detailed savings. Power Requirement Savings The power savings of the two-stage system begin to become significant at 20 F ( 28.9 C) and increase dramatically at lower evaporating temperatures (Figure 9). The power savings for systems operating from 30 F ( 34.4 C) to 40 F ( 40 C) are significant with the power savings varying from 123 BHP (91.7 KW) for -30 F ( 34.4 C) and 150 psig ( kpa) to 464 BHP (346.1 KW) for 40 F ( 40 C) and 185 psig ( kpa) (Table 5). The results are typical of many systems that operate at these levels of evaporating temperature and condensing pressure. At evaporating temperatures above 20 F ( 28.9 C), the results appear to be less significant, particularly at the lower condensing pressures. Nevertheless, the twostage system offers power savings over the single-stage system for the entire range of evaporating temperatures and condensing pressures studied. Power Cost Savings Figure 9 shows the annual power cost savings vs. evaporating temperature of the two-stage system over the single-stage system based on annual operation at varying condensing pressures. Table 6 provides the details for each evaporating temperature. The annual power-cost savings range from $11,734 at 10 F ( 23.3 C) to $323, 487 at 50 F ( 45.5 C). At evaporating temperatures below 20 F ( 28.9 C) the annual cost savings provide significant annual savings that can more than make up for any additional capital cost of a two-stage system over a single-stage system. Overall, Figure 9 shows that the two-stage system should definitely be the system of choice for operation below at or below 30 F ( 34.4 C) evaporating temperatures. 10 IIAR 2010 Technical Paper #5

13 Consider Two-Stage Systems to Save Power Two stage systems should also be given significant consideration over the single-stage system for systems operating at or below 20 F ( 28.9 C). Typical System The power requirements and power cost savings were specifically examined for a system that is typical of many systems installed in the industry. The typical system operates at 40 F ( 40 C) evaporating temperature and 165 psig ( kpa) to 185 psig ( kpa) condensing pressures during the warmest times of the year, and at 150 psig ( kpa) to 165 psig ( kpa) condensing pressures during the cooler periods of the year. The 300-ton (1070 KW) low-side load and 200-ton (713 KW) intermediate load are typical of a real-world systems. Table 6 shows the power cost savings of the typical two-stage system over the typical single-stage system at 40 F ( 40 C) evaporating temperature and for the varying condensing pressures during a year s operation as previously discussed. The annual power-cost savings is $183,729. Over 10 years of operation, the total power cost savings would amount to more than $1.8MM. The cost savings definitely support choosing a two-stage system. Even if the typical system operates at 20 F ( 28.8 C) evaporating temperature, the annual savings is $41,938 per year and more than $400,000 over 10 years. Although the savings is not as dramatic, consideration should be given to the two-stage system. System Complexity and Cost Comparison of Figures 1 and 2 reveals that the two-stage system is not significantly more complex than the single-stage system. However, there will be some cost differences between the two-stage and single-stage systems. Because the economized system power requirements were higher than both the single and two-stage systems, Technical Paper #5 IIAR

14 2010 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, California the complexity and costs of the economized system were not evaluated. The cost differences include: The low-temperature compressors will be of the same displacement for the twostage and single-stage systems, but the drive motor for the single-stage system will be larger, resulting in higher cost for the motor and associated power supply equipment. The HTR for the two-stage system will have higher cost because it has to absorb the heat rejection from the low-temperature (booster) compressor. The HTR for the single-stage system need only absorb the 200-ton intermediate load. The high-temperature compressor will be higher cost for the two-stage system. Again, because it must handle the heat rejection from the low-temperature (booster) compressor plus the intermediate load. The piping associated with the larger compressor will also cost slightly more. Because the total heat rejection to the condensers will be slightly less for the twostage system, the cost of the condensers will be slightly less than for the singlestage system. The LTRs, low-temperature and high-temperature evaporators and associated piping, and the system controls will be essentially the same cost for both systems. Table 7 provides estimates of the cost differences between the single-stage and twostage systems. The single-stage system is the base system for comparison, with the cost differences shown for the two-stage system. The costs shown are estimates only, but are within reasonable range for the systems studied. It should be noted that a direct comparison of the actual costs of installed systems is not available. It is very difficult, if not impossible, to find examples of the two types of systems of identical design, with the same loads and capacities. Table 7 shows that the two-stage system estimated capital cost would be about $90,000 higher. The extra cost of a two-stage system operating at 20 F ( 28.8 C) would be paid back in slightly more than 2 years of operation. At 30 F ( 34.4 C) 12 IIAR 2010 Technical Paper #5

15 Consider Two-Stage Systems to Save Power evaporating temperature, the extra cost would be paid back in slightly less than a year. For operations at lower evaporating temperatures, the payback is significantly shorter. For perspective, the capital cost of the systems studied was estimated to range between $2.2 MM and $2.8 MM. Conclusions Two-stage refrigeration systems should be the chosen configuration for installations that will operate below 30 F ( 34.4 C). The performance comparisons of twostage and single-stage refrigeration systems, as described in this paper, support this conclusion. Two-stage systems should also be given serious consideration over single-stage systems for installations operating at 20 F ( 28.8 C) and below. However, the cost and complexity of these systems should be evaluated because the power-cost savings of the two-stage system over the single-stage system are less. At evaporating temperatures above 20 F, the evaluation shows insignificant advantage of the twostage system over the single-stage system. However, if longer payback periods can be tolerated, the two-stage system may have some advantage, considering that industrial ammonia refrigeration systems have operating lives exceeding 30 years. These conclusions are valid for the systems studied, i.e., low-temperature load at 300 tons ( KW) and intermediate-temperature load at 200 tons (713.5 KW). The loads can be scaled up or down, resulting in proportional results if the ratio of the low-temperature load to the intermediate-temperature load is reasonably close to the ratio used for the study. If the ratio is substantially different, an evaluation of the specific system should be made. Technical Paper #5 IIAR

16 2010 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, California Systems with substantially different temperature and load requirements must be specifically evaluated to determine whether the two-stage design will offer an advantage. For operations where no side load is encountered, two-stage and economized systems will likely show an advantage over classic single-stage systems (no subcooling of refrigerant to the evaporators) operating above 0 F ( 17.7 C). Again, a specific evaluation should be conducted. The overall conclusion offered by this paper is: two-stage systems should always be considered when operations require evaporating temperatures below 30 F ( 34.4 C). The two-stage system will likely always provide adequate power cost savings to justify the slight additional cost of the two-stage system. Two-stage systems should also be seriously considered at evaporating temperatures at or below 20 F ( 28.8 C) with the final decision made after a detailed performance evaluation of both two-stage and single-stage systems. It is important to note that the conclusions in this paper serve to demonstrate the potential for energy savings by considering two-stage systems. Detailed analyses for the system types considered in this paper should be conducted during the design phase of any system to confirm actual energy requirements and savings available. Energy modeling can be quite complex due to the number of variables involved, but efforts should be made to ensure appropriate evaluations are conducted. 14 IIAR 2010 Technical Paper #5

17 Consider Two-Stage Systems to Save Power Figure 1: Typical Two-Stage System, Liquid Recirculation Technical Paper #5 IIAR

18 2010 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, California Figure 2: Typical Single-Stage System, Liquid Recirculation 16 IIAR 2010 Technical Paper #5

19 Consider Two-Stage Systems to Save Power Figure 3: Typical Economized System, Liquid Recirculation Technical Paper #5 IIAR

20 2010 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, California Figure 4: Power Requirement Comparison at 100 psig (689.5 kpa) Condensing 2500 SST, C Total System BHP Required SST, F Single Stage Economized Two Stage 18 IIAR 2010 Technical Paper #5

21 Consider Two-Stage Systems to Save Power 2500 SST, C Total System BHP Required SST, F Single Stage Economized Two Stage Figure 5: Power Requirement Comparison, 125 psig (861.9 kpa) Condensing Technical Paper #5 IIAR

22 2010 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, California Figure 6: Power Requirement Comparison, 150 psig (1,034.2 kpa) Condensing 2500 SST, C Total System BHP Required SST, F Single Stage Economized Two Stage 20 IIAR 2010 Technical Paper #5

23 Consider Two-Stage Systems to Save Power 2500 SST, C Total System BHP Required SST, F Single Stage Economized Two Stage Figure 7: Power Requirement Comparison, 175 psig (1,206.6 kpa) Condensing Technical Paper #5 IIAR

24 2010 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, California Figure 8: Power Requirement Comparison, 200 psig (1,379.0 kpa) Condensing 3000 SST, C Total System BHP Required SST, F Single Stage Economized Two Stage 22 IIAR 2010 Technical Paper #5

25 Consider Two-Stage Systems to Save Power $350,000 SST, C $300,000 $250,000 $200,000 $150,000 $100,000 Power Savings, $ Per Year, $0.08 per KWH $50,000 $ SST, F Figure 9: Power Cost Savings, Two-Stage Over Single-Stage Relative to Annual Condenser Pressure Variations Technical Paper #5 IIAR

26 2010 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, California Table 1: Two-Stage Power Consumption Results Sat. Evap. T Condensing Pressure, psig (kpa) 100 (698) 125 (873) 150 (1048) 175 (1222) 200 (1397) F C BHP KW BHP KW BHP KW BHP KW BHP KW Table 2: Single Stage Power Consumption Results Sat. Evap. T, F Condensing Pressure, Psig (kpa) 100 (698) 125 (873) 150 (1048) 175 (1222) 200 (1397) F C BHP KW BHP KW BHP KW BHP KW BHP KW , ,392 1,038 1,662 1,238 1,968 1,468 2,345 1, , , ,493 1,114 1,735 1, , ,356 1, , IIAR 2010 Technical Paper #5

27 Consider Two-Stage Systems to Save Power Table 3: Economized Power Consumption Results Sat. Evap. T Condensing Pressure, psig (kpa) 100 (698) 125 (873) 150 (1048) 175 (1222) 200 (1397) F C BHP KW BHP KW BHP KW BHP KW BHP KW Table 4: Power Saved, Two-Stage Over Single Stage At full-load operation Sat. Evap. T, F Condensing Pressure, psig (kpa) 100 (698) 125 (873) 150 (1048) 175 (1222) 200 (1397) F C BHP KW BHP KW BHP KW BHP KW BHP KW Technical Paper #5 IIAR

28 2010 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, California Table 5: Power Savings BHP, Two-Stage vs. Single-Stage For annual condenser pressure variation Evap. T Condensing Pressure, psig (kpa) 150 (1048) 165 (1152) 175 (1222) 185 (1292) F C BHP KW BHP KW BHP KW BHP KW Table 6: Estimated Annual Power Savings Relative to annual condensing pressure variations Sat. Evap. T Savings. F C $ Per Year $323, $183, $91, $41, $11,734 Based on $0.08 per KWH power cost 26 IIAR 2010 Technical Paper #5

29 Consider Two-Stage Systems to Save Power Table 7: Estimated Cost Premium for Two-Stage System Over Single-Stage System Item Cost Difference for 2-Stage LT Compressors, Same Displacement for Both Single 0 and 2-Stage LT Comp Motor, Nominal 700 hp 2-Stage, 1200 Hp 15,000 1-Stage LT Recircirculator, Same for both Single and 2-Stage 0 HT Recirculator, Add for Intercooling 15,000 HT Compressors, Higher Capacity for 2-Stage 85,000 Condensers, Nominal 8 10% lower THR, 2-Stage 5,000 HT Piping, Increased HSS and HSD sizing for 2-Stage 10,000 LT Evaps and Piping, Same for Both Single and 0 2-Stage HT Evaps and Piping, Same for Both Single-Stage 0 and 2-Stage Controls, Same for Both Single-Stage and 2-Stage 0 Total Estimated Additional Cost for 2-Stage $90,000 Note: The above are estimated only, costs could vary by 10% based on actual system design Technical Paper #5 IIAR

30 2010 IIAR Industrial Refrigeration Conference & Exhibition, San Diego, California Notes: 28 IIAR 2010 Technical Paper #5