RESEARCH ON MICRO GAS TURBINE EXHAUST HEAT RECOVERY AND USAGE

Transcription

1 RESEARCH ON MICRO GAS TURBINE EXHAUST HEAT RECOVERY AND USAGE Atsuya Tajima, Urban Energy R&D Department, TOHO GAS CO., LTD Yoshiharu Ito, Urban Energy R&D Department, TOHO GAS CO., LTD. Tatsunobu Amemiya, Urban Energy R&D Department, TOHO GAS CO., LTD Micro gas turbines (referred to below as s) have received attention recently as smallscale distributed power sources. With characteristics such as their small size, lightweight, and minimal vibrations during operation, they are expected to become widespread in a wide range of applications, including residential and small-scale industrial use. Also, the exhaust heat emitted by the is in the form of an exhaust gas that is about 300 C, allowing heat recovery and usage systems to be compact. Because city gas is principally methane, it burns cleanly and therefore direct use of this exhaust gas is possible. Development is being carried out to maximize these characteristics and to develop a variety of heat recovery and usage systems relevant for a wide range of applications. To respond to a wide variety of heat recovery and usage needs, we have been conducting field testing and developing several types of heat recovery and usage systems. This paper introduces the findings of this development work, and describes a newly developed heat recovery and usage system that have advantages such as increased ease of use. 1. Recovered Heat in a Water Heating Application Since March of 2001, we have conducted field testing of a cogeneration system using a hot water package with an electric output of 26KW for the public bath Hot Spring Shiratori no Yu. (1) Installed package The specifications for the installed package are shown in Table 1.1; the outer view of the installation is shown in Figure 1.1. Manufacturer Package Capstone (US) Takuma TCP-30LH Net Electrical Output KW 26 Electrical Specifications Net Electrical Efficiency Fuel Consumption (13A) Manufacturer Heat Recovery Rate (Hot Water) Boiler Ratio of Actual Used Heat Input 200V, Grid Connect % 23 Nm 3 /h 9.7 KW Takuma 56(from 60 to 70 C) % 50 Total Efficiency % 73 Table 1.1 The specifications for the installed package Figure 1.1 Installation, Outer View (2) System Outline The schematic of the system is shown in Figure 1.2. The electricity generated by the meets 15% of the electrical load on the premises through the grid connect. Hot water enters the package at 30 C, exits at 40 C, and is used to heat the well water storage tank. This well water is sent to the hot water tank where it is heated by the hot water heater, and is used as tap/shower water. To prevent the formation of scale, a heat exchanger is used to separate the well water from the. Because there is sufficient heat and electrical load, the is operated with a schedule timer, allowing the system to be operated with a 100% load factor.

2 City Gas Hot Water Tank Hot Water Heater Shower Tub Hot Spring (Source) Exhaust Gene Gas Boiler rator Well Water Storage Tank Heat Exchanger Package Figure 1.2 Flow Diagram for Water Heating System Well Water (3) Operating Performance The operating data for the cogeneration between April 2001 and March 2002 are shown in Table 1.2. The operating hours of the public bath were 14 hours daily, between 10 AM and 12PM, and the was operated at this time. Total operating time was 3,697 hours; this figure was slightly lower than planned because regular operation started in May, and also because of downtime for investigation and repair of etc. Total electrical energy output was 83.7 MWh. Partly because intake air temperature averaged 40 C in the summer months and 30 C in the winter months, average electrical output was 22.7KW. Total recovered heat was MWh, and average recovered heat rate was 68KW. This value is higher than the specified recovered heat rate of 56KW because the circulating hot water was relatively low 40 C. Also, because of factors such as the inlet air temperature, the electrical efficiency was 20.5% but because of the high-recovered heat, net total efficiency was high value of 82.2%. (*1) LHV: Based on Lower Heating Value of [KWh/Nm 3 ] Item Total Operating Time Net Generated Electrical Energy Fuel Consumption (*1) Total Recovered Heat Net Electrical Efficiency Total Efficiency Units h MWh MWh[LHV] MWh % % Performa nce Value 3, Table 1.2 The Operating Data For The Cogeneration (April 2001 and March 2002) (4) Energy Savings Based on the data in Table 1.2, the energy saving rate of this cogeneration system was calculated by comparing primary energy of cogeneration system with that of previous system (grid electricity + boiler). Assuming the primary energy conversion factor of converted primary energy to grid electricity is 2.85 to 1(KWh), and 90% boiler efficiency, the energy saving rate was calculated to be 12.9%.

3 Item Units Cogeneration System Previous System (Grid Electricity + Boiler) Fuel Consumption MWh[HHV] 451 Generated Electrical Energy Equivalent Generated Electrical Power (Primary Energy Conversion) MWh 83.7 MWh 239 Remarks HHV: Based on Higher Heating Value of [KWh/Nm 3 ] Primary Energy Conversion Factor :2.85[KWh/KWh] Heat MWh Boiler Efficiency 90% Total MWh Percentage of Saved Energy to Total Energy of Energy saving rate % 12.9 Previous System Table 1.3 Calculation of the Energy Saving Rate (5) Summary Because the circulating hot water was at a low 40 C, and there was sufficient electrical and thermal load that ensured a high load factor of, this system was able to realize a high-recovered heat rate. In 2001, there were some problems resulting in system downtime, but these have been resolved and the operating time of this system is expected to increase in Recovered Heat in a Drying Oven Operation (1) System Outline This system is composed of a 75KW class and an existing paint drying oven and duct burner that uses recovered heat from the. They are located in the factory of a machine manufacturer A. The existing dry oven system maintained the drying oven interior at 145 C by using a gas blower to blow the drying oven exhaust gas mixed with outside air from an air supply port through a duct burner (referred to below as DB ) and back into the drying oven. When the recovered heat is being used, the exhaust gas is directed through a series of newly constructed ducts, shown as a dotted line in Figure 2.1, into the exhaust gas blower s inlet side. The decision to locate the exhaust gas inlet on the exhaust gas blower s inlet side was because the exhaust gas replaced the outside air from the air supply port, and because the maximum tolerated turbine back pressure is 0.25kPa. Output KW 75 DB Drying Oven Voltage V 200 Output Type Grid Connect Exhaust Gas Outlet Temp C 250 Maximum Combustion Rate KW 350 Combustion Type Pre-Mix Type Combustion Range Nm 3 /h 8.5~20 Oven Type Indirect Heating Oven Temp. Setpoint C 145 Time of Operation/Day 8:30~18:00 Table 2.1 The Specifications For the Drying Oven Heat Recovery and Usage System

4 Exhaust Gas Blower Exhaust Air Supply Port Paint Drying Oven Duct Burner Combustion Air 75kW City Gas Figure 2.1 System Flow for Drying Oven System (2) Operating Performance From June to December 2001, six months of energy-saving proving tests were carried out. Table 2.2 shows the DB fuel consumption before and after recovered heat usage. Originally, though the DB turndown ratio was low and resulted in the DB fuel flow rate falling below the lower control limit of 8.5Nm 3 /h, this problem was settled by adding a on/off control to the DB. As a result, we could demonstrate that the introduction of exhaust gas resulted in a significant decrease in DB fuel consumption, while slight temperature fluctuations were seen in the drying oven interior, but these fluctuations were small enough to cause no effect with quality of the products. Also, there was almost no change in the composition of the exhaust gas (NO x, CO, CH 4 ) and it was confirmed that there were no negative repercussions to the DB burn quality. Figure 2.2. DB Fuel Consumption Before and After Recovered Heat Usage (3) Energy Savings Table 2.2 shows results of summer time and wintertime tests measuring energy usage before and after heat recovery and usage. During winter, without heat recovery and usage (the existing system) the DB fuel consumption was 183KW (HHV). After the heat recovery and usage of at the rate of 64KW, the DB fuel consumption fell to 47KW, a reduction of 136KW. The sum total reduction in DB fuel consumption can be broken down as follows: exhaust gas input into the circulating duct contributing to the elimination of the input of outside air from the air supply port and the corresponding reduction in heat losses, and the effectively used heat amount represented by the difference in enthalpy between exhaust gas inlet and outlet of circulating duct. As a result, taking the primary energy conversion factor as 2.85 [KWh /KWh], an energy savings of 7.3% was calculated. During the summer, the higher duct inlet air temperature leads to lower heat losses in existing system. The higher intake air temperature also results in a lower electrical efficiency, resulting in an

5 increased fuel consumption. For these reasons, the achieved energy saving rate is at a low 0.5%. Assuming the same electrical output, the lower the intake air temperature is, the lower the rotation speed of the turbine is needed. As a result, the exhaust gas flow rate decreases. In the summer, the intake air temperature is higher and the exhaust gas flow rate is higher. It is thought that the exhaust gas blower was unable to suck in the increased volume of exhaust gas, resulting in the system being unable to fully use the exhaust heat. Testing Period Winter Summer Item Units Existing System (No Heat Recovery) With Heat Recovery Existing System (No Heat Recovery) With Heat Recovery Net Electrical Output KWh Fuel Consumption DB Fuel Consumption KWh [HHV] KWh [HHV] Equivalent Generated Electrical Power (Primary KWh Energy Conversion) Total KWh Energy Saving Rate % Table 2.2 Energy Saving Rate Calculated before and after Heat Recovery and Usage (4) Summary This system is an addition to an existing system and as such is not by any means optimized. However, it has confirmed that DB fuel consumption can be cut and energy savings realized in such a system. It is thought that with further system optimization, a high level of energy savings can be achieved. The heat emitted by the is in the form of exhaust gas, a gas that is clean and very close to air. Thus this application of the for an industrial oven effectively exploited the characteristics of the, and the large-scale adoption of this technology will be pursued aggressively. 3. Recovered Heat in an Air Conditioning Application Figure 3.1 cogeneration system Installation, Outer View In March of 2002, a cogeneration system combining a 50KW and an exhaust gas driven absorption chiller was put into place at Howa Sportsland, a general use sports facility located in Minato-ku, Nagoya City. Because this system recovers exhaust heat into an exhaust gas driven absorption chiller, it can efficiently provide air conditioning, ideal for the needs of a customer that has minimal water heating demands.

6 (1) System Outline The specifications of the cogeneration system are shown in Table 3.1, and the outer view is shown in Figure 3.1. The has been running continuously 24 hours a day, and the generated electrical power is used by the whole of Howa Sportsland through a grid connect. The exhaust heat created as a byproduct of generation is recovered by an exhaust gas driven absorption chiller (referred to below as an absorption chiller ); this heat is used to provide cooling and heating water for air conditioning in the training/lodging facilities in the sports complex. The system flow diagram is shown in Figure 3.2. Exhaust Gas Driven Absorption Chiller Manufacturer Toyota Turbine and Systems Output KW 50 Electrical Specifications V 210V, 60Hz, Grid Connect Net Electrical Efficiency % 25.5 Fuel Consumption Nm 3 /h 17.0 Method of Heat Recovery and Use Designed Output (Auxiliary Burning + Heat Recovery) Heat Recovery Only Cooling/HeatingWater Temp. KW KW C Exhaust Gas Driven Direct Investment Type Cooling 140.7(40 USRT) Heating Cooling 52.7(15 USRT) Heating 63.4 Cooling Inlet 12.5/Outlet 7.0 Heating Inlet 55.5/Outlet 60 Coolant Water Temp. C Inlet 32/Outlet 37.5 Table 3.1 The specifications of the cogeneration system Purchased Electricity Electrical Load Generated Electricity Cooling Tower [50kW] Exhaust Gas Driven Absorption Chiller [40RT] Cooling Water Cooling and Heating of Training Facility City Gas Existing Absorption Chillers (Auxiliary Burning, 60RT) x 2 Units Howa Sports Land Figure 3.2 Flow Chart of this System The system responds to cooling/heating water load and operates the with thermal output control. Inside the absorption chiller, there is an auxiliary burner, and even when the is not running it is able to provide heating and cooling. In addition, when the is insufficient to meet the heating/cooling load, the burner has a supplementary burning function. The system responds to different loads in the following way burner operation only for light loads, operation only for medium loads,and operation of both burner and for heavy loads at lower than 40RT.

7 (2) Operating Performance The operating data for the cogeneration system are shown in table 3.2. The data comes from May to September Item Total Operating Time Total Generated Net Electrical Output Total Cooling Energy Average Electrical Output Average Cooling Output Units hr KWh KWh KW RT Total 2,742 91, , Table 3.2 Cogeneration Operating Data (May~September 2002) In Table 3.2 the total generated net electrical output is divided by the total operating time to obtain an average electrical output of about 33KW, a lower number than electrical specification value of 50KW. The intake air temperature is part of the reason, but based on the average cooling output of 26.4 RT ( 1RT is equal to 3.52KW ), operating the system with thermal output control during night-time and at mid-season, periods with low air conditioning loads, is thought to be the primary reason. The average electrical output and the average cooling output on a typical day in a given month are shown in Figure 3.3 as well as the operation mode on a typical day in a given month are seen in Table 3.3. The great month-to-month differences in cooling load are apparent. The average cooling load on a typical day in May is around 14RT, but at night, there is a further reduction in cooling load and the comes to a stop, and the system runs on burner alone. In June, September, and October, the recovered heat is sufficient to meet cooling load. The cooling output at this time is close to 20RT, confirming that this value is far above the specification value of 15RT. Though in July and August and the burner operate jointly, because existing absorption chiller is concurrently operated to meet increased cooling load at higher than 40RT, the cogeneration side absorption chiller load factor was slightly lower. Because the control system currently does not include the existing chiller and is not optimized, this is a future improvement to make. Month May June July August September Operation Mode Night: Burners Only Day: Heat Recovery Only Day and Night: Heat Recovery Only Day and Night: Both Heat Recovery and Burners Day and Night: Both Heat Recovery and Burners Day and Night: Heat Recovery Only October Day and Night: Heat Recovery Only Table 3.3 Absorption Chiller Operating Modes in a Typical Day, by Month Average Electrical Output(kW) Absorption Chiller Cooling Load(RT) Average Cooling Output (RT) Average Generated Electrical Output (kw) May June July August September October Figure 3.3 Average Daily Cooling Output, by Month

8 (3) Energy Savings With this system, the calculated energy saving rate is shown in Table 3.4 by energy saving rate by operation Mode. The energy saving rate was calculated taking the primary energy conversion factor as 2.85 [KWh /KWh], and assuming the COP of the existing absorption chiller as 1.0. When the and burner is operated jointly, the achievable energy saving rate is around 2~3%, though this number changes with intake air temperature. When the is operated alone with its rated cooling load, an energy savings rate of 2-3% was achieved; when the cooling load was 16.5 RT, the energy savings rate was -1.9%. This is thought to be because the lower cooling load caused the rotation speed of turbine and electrical output to decrease, resulting in a decrease in not only electrical efficiency, but exhaust gas temperature and flow rate, causing a decrease in absorption chiller COP and a subsequent decrease in total efficiency. Item Electric Output Cooling Output Energy Saving Rate Intake Air Temp. Electrical Efficiency Ratio of Actual Used Heat Total Efficiency Unit KW RT % [HHV] C % [LHV] % [LHV] % [LHV] Heat Only Heat and Burners Table 3.4. Energy Saving Rate by Operation Mode (4) Summary Though not to the extent of the previously mentioned water heating application, the air conditioning system using recovered heat was able to realize energy savings. In this type of thermal control system, the decrease in cooling load causes a decrease in electrical output, causing a negative energy savings for one fixed cooling load domain. In this domain, it would have been more desirable to stop and do an auxiliary burn with the absorption chiller burner alone. Also, for further energy savings, the amount of heat recovery must be increased, and the performance of the absorption chiller side improved. Next we will conduct testing with heating operation and evaluate the energy savings possible with a year-round operation of this system. 4. The Development of a Metal Hydride Freezer - A Cooling Application of Recovered Heat. (1) Introduction Up until now, recovered heat has been developed and commercialized for hot water at C, steam, heating and cooling applications, drying ovens where the exhaust gas is used directly, and for general heating applications. This time, the aim was to evaluate a recovered heat driven freezer using metal hydride alloy (referred to below as a MH freezer ) developed by The Japan Steel Works, Ltd. With this system, the is expected to be used more widely by the food retail industry, starting with supermarkets, which have a high cooling and minimal heating needs. (2) System Outline The freezer system is shown in Table 4.1. Figure 4.1 shows the principle that the MH freezer system is based on. This device has two alloy pairs, each composed of an alloy on the high temperature side and an alloy on the low temperature side. Each side of alloy has different ingredient ratio of metal hydride that affect absorption and desorption characteristics of hydrogen. In one pair, the recovered heat from the releases hydrogen from the alloy on the high temperature side, which is absorbed by the low temperature side alloy cooled by coolant. (This is the regenerative process.) In the other alloy pair, the high temperature side alloy absorbs hydrogen, as the low temperature side alloy releases hydrogen. As the hydrogen is released from the low temperature side alloy, it absorbs heat from the surroundings, resulting in a cold alloy. Brine is put through this cold alloy

9 for heat exchange to produce cooled brine. (This is the cooling process). These two processes are switched every 16 min. Regeneration High Temp. Side Process 270 C Low Temp. Side Cooling Tower Exhaust Gas Cooling Process Hydrogen Release (160 C Heating) Hydrogen Absorption (45 C Cooling) 32 C Hydrogen Cooling Tower Freezer 32 C Hydrogen Absorption (45 C Cooling) Hydrogen Release (-10 C Heating) -5 C Figure 4.1 Refrigeration Principle of MH Freezer Unit. Figure 4.2 Outer View of MH Freezer. Manufacturer Package Net Electrical Output Electrical Specifications Fuel Consumption MH Freezer Freezing Output Manufacturer Capstone Active Power Corp. KW 28 V 220V, 60Hz, Grid Connect Nm3/h 9.7 The Japan Steel Works, Ltd. Dimensions W940 x D1930 x H 1932 Brine Outlet Temp. -5 C(30% Methanol Solution) Coolant Inlet Temp. 32 C KW 7.7 Table 4.1 The Specifications for /MH Freezing System (3) Operating Testing Since March 2002, in collaboration with the Japan Steel Works Ltd., we have conducted endurance and performance testing on the /MH freezing system. The diagram for the test unit is shown in Table 4.3. The cooling load was created during the test by varying the output of an electric heater attached to the brine tank between 1~10KW.

10 Exhaust Exhaust Gas 28kW Freezer Inlet Coolant Temp. T Cooling Coolant Tower 190L/min MH Freezer Brine Tank Freezer Inlet T Temp. T Brine Temp. Electrical Brine Tank Heater - Heating Brine 200L 1~10kW 45L/min Freezer Outlet Coolant Temp. T T MH Freezer Outlet Brine Temp. Figure 4.3 The Diagram for the Test Unit of /MH Freezing System Measured results are shown in Figure 4.4 and 4.5, from which the following conclusions were reached. (A) Based on a coolant temperature of 33.4 C and an average MH freezer unit outlet brine temperature of -5.4 C, a freezing performance of 7.7KW, a value almost identical to the specification freezing value, was confirmed. (B) The freezing output of the MH freezer is based on the amount of hydrogen absorption, which is in turn affected by the coolant temperature in both processes. As Figure 4.4 shows, a decrease in coolant temperature is accompanied by an increase in MH freezer output. Thus, during the summertime, with its higher coolant temperature, is accompanied by a decrease in freezer performance. Also, a change in brine temperature affects performance because brine inlet temperature affects the rate of hydrogen transport between alloys during the cooling process. (C) During the summer, the intake air temperature increases and the output decreases, resulting in a decrease in recovered heat by MH freezer, in turn resulting in a decrease in freezing output. (D) Figure 4.5 shows that there is a great change in brine temperature every time there is a process switchover. This is because hydrogen absorption and release is the greatest at the beginning of the cycle and gradually decreases with time. Also, because the MH freezer is operated on an on-off control, the fluctuation of the brine temperature is still large. Thus depending on the application to some refrigeration use, a brine tank may be necessary to stabilize the brine temperature. (E) The outlet temperature of exhaust gas from the MH freezer changes from 60 C to 150 C every time there is a process switchover. For this reason, if more heat recovery and usage devices are added downstream from MH freezer, design issues with controls and the like must be considered. (F) As of the end of September, the MH freezer has completed 2,358 cycles, 629 hours, with no malfunctions or decrease in performance. We plan to continue to evaluate the performance and durability of the unit.

11 10.0 Freezing Output [kw] Coolant Average Inlet Temp.20 C Coolant Average Inlet Temp.26.3 C Coolant Average Inlet Temp.33.5 C Device Exit Brine Temp. [ C] Figure 4.4 Freezing Output as a Function of Brine Temperature, and Freezing Output Figure 4.5 Coolant Temperature, Brine Temperature Change (4) The Development of an MH Freezer - - Desiccant Air Conditioning System To further increase energy efficiency and usefulness, we are developing a MH freezer - - Desiccant air conditioning system that incorporates an exhaust gas-driven desiccant air conditioning device downstream from the /MH Freezing System. Figure 4.6 shows the flow diagram of the system. The -10 C brine produced by the MH freezer is supplied to a refrigerated showcase via a brine tank. The remaining exhaust gas from the MH freezer is sent to the desiccant air conditioner, and the dehumidified dry air produced by it is sent to the lower portions of the showcase, where it is used to adjust the humidity in the store and to keep the aisles around the showcases warm for comfortable shopping. When the electrical load is minimal and the cannot be operated, a separate auxiliary burner is run, and the hot air sent to the MH freezer. In 2003, monitoring testing at a supermarket to evaluate the energy savings and the economy of this system is planned to start.

12 City Gas(2KPa) Capstone MH Freezer Exhaust Desiccant Air Conditioning Machine Regeneration Side Operation Side Exhaust Outside Air Outside Air Auxiliary Burner Cooling Tower Backup Freezer Brine Tank Showcase Showcase Dehumidified Dry Air Figure 4.6: Schematic of an MH Freezer - - Desiccant Air Conditioning System (Planned) (5) Summary The MH freezer is the first device in the world to refrigerate in the minus temperature regime using exhaust heat. Also, because it uses hydrogen as a refrigerant, it produces almost no noise and vibrations, and therefore is an excellent machine from an environmental standpoint. The Japan Steel Works, Ltd. is currently developing improvements in freezing output by increasing alloy performance, and improvements in economy and energy savings are expected. 5. Development of a Supplementary Burning System for Exhaust Gas (1) Introduction The thermal output of heat recovery and usage devices that do not have an internal heat source are greatly affected by stops in operation, and by when the entire system is operated with electrical output control. We have developed a supplementary burn system for exhaust gas that features a duct burner (referred to here as a DB ) that has the ability to supply heat alone even when the is not in operation. (This is referred to below as auxiliary burning ) (2) Supplemental Burning System - System Outline Absorpti on Chiller Manufacturer Capstone Packager Meidensha Corp. Output KW 28 Net Electrical Efficiency % 25 Output Type Manufacturer Grid Connect Sanyo Electric Air Conditioning Co., Ltd. Cooling/Heating Output KW 42.2/42.6 Ratio of Actual Used Heat Input (Cooling/Heating) Total Efficiency (Cooling/Heating) % 38 % 63 Heating Performance KW Max 116 DB Turndown Ratio Greater than 1:10 Combustion Type Pre-Mix Type Table 5.1 The specifications for Supplementary Burning System

13 The specifications of the supplementary burning system are shown in Table 5.1 and the schematic of the system is shown in Figure 5.1. The exhaust gas from a 28KW received supplemental burning from a DB with a 116KW maximum combustion. This gas is then put into an exhaust gas driven absorption chiller (referred to below as and absorption chiller ). The exhaust gas from the absorption chiller is partly recirculated to the upstream side of the DB; the rest is exhausted. We conducted evaluation tests from October 2001 to September 2002 at our test site. Circulation Duct Duct Burner Combustion T1 Combustion Blower Air Exhaust Gas Temp. (Less than 360 C) Controller1 City Gas M M Controller2 Cooling Water Control Control Temp. Controller Valve 1 Valve 2 Cooling:7 C Heating: 60 C T2 Exhaust Gas Blower Absor ption Chiller Cooling Water Indoor Unit Cold Air/Warm Air Exhaust Figure 5.1. Flow Diagram of Supplementary Burning System (3) Supplementary Burning System - Characteristics and Performance Test Results (A) Increase in Heating/Cooling Output The supplementary burning of exhaust gas resulted in a dramatic increase in AC cooling/heating output. As Figure 5.2 shows, when this system is operated with only, as the output decreases, resulting in a decrease in heat content and volume flow rate in the exhaust gas, the heating output of the absorption chiller is reduced. With a supplementary burning from the DB, the heating output of the AC was stabilized to a near-constant value of 70KW that was unaffected by changes in electrical output. Cooling tests gave similar results. (B) Reduction in and Burner Back Pressure To reduce the turbine back pressure caused by high exhaust volumes, an exhaust gas blower was installed on the outlet side of the absorption chiller, keeping the turbine back pressure and the burner back pressure below their maximum tolerated values of 2kPa and 0.5kPa, respectively. (C) Adoption of Recirculation Flow Because the capstone is operated with rotation speed control, its exhaust gas volume decreases with decreasing load, which could possibly lead to a reduction in DB burn quality. To prevent this possibility, shortages in exhaust gas volume were augmented by recirculated exhaust gas; this stabilized the inlet DB exhaust gas volume to a near-constant value. The blower flow rate was selected to keep the exhaust gas pressure at the circulation duct junction located upstream of the DB at or below the chosen value, thereby also preventing bypass of exhaust gas via the circulation duct. (D) Development of a Control Method for the DB Burn The DB burn must be controlled such that it is proportional to the absorption chiller outlet cooling water temperature, and at the same time, the temperature of exhaust gas being supplied to the absorption chiller must not be too high (less than 360 C). As Figure 5.1 shows, an inexpensive

14 solution was found to optimally control the fuel flow based on these two temperatures by adding two control valves to the fuel inlet lines. (E) Making the Supplementary Burning System Cost-Effective and Multipurpose The DB developed by us used a low-cost package burner (list price 700,000yen), achieving a system-wide cost decrease. Also, the supplementary burning system could be applied to exhaust gas driven freezers, desiccant air conditioning systems, steam boilers, drying ovens, and other devices as well as the absorption chiller. Heating Output[kW] Auxiliary Burning Heating Output - Supplementary Burning Heating Output - Only DB Fuel Consumption - Supplementary Burning DB Fuel Consumption[Nm 3 /h] Output[kW] Figure 5.2. Heating Output Characteristics As a Function of Output. DB Fuel Consumption[Nm 3 /h] DB Fuel Consumption - Auxiliary Burn DB Gas Consumption - Supplementary Burn 24(kW) Output Absorption Chiller Output [Nm 3 /h] Figure 5.3. DB Gas Consumption as a Function of Heating Load (4) Summary The development of the supplementary burning system resulting in a system with the following characteristics: able to increase thermal output by further heating the exhaust gas, able to be flexible in adopting to both electrical and thermal output control respectively, applicable to all types of heat recovery and usage devices.

15 Also, we are currently developing supplementary burning systems with further energy savings. One system has a drastically reduced combustion air volume. Another uses an exhaust reheat type burner so that exhaust gas can be used as combustion air. We plan to establish practical applications of these systems as soon as possible. 6. Conclusion exhaust gas is exhaust gas that is around 300 C in temperature, and is an extremely clean gas that is similar to air. Thus it is possible to design and construct various heat recovery and usage systems to use this exhaust. On the other hand, because of the high volume of the exhaust gas, design work is necessary to decrease back pressure imposed on and exhaust heat loss. It is critical that heat recovery and usage systems take these considerations into account while maximizing the advantageous characteristics of these systems. Though this paper introduced test data from heat recovery and usage applications including water heating, drying, and air conditioning, we were also able to confirm with these tests the energy-saving effect of each system and to develop a MH freezer and supplementary burning systems that widen the range of applications for recovered heat. We would like to achieve an increased level of adoption of these systems, and by developing new systems that exploit the characteristics of exhaust gas, to contribute to the spread of s in the world.