PAPER NUMBER 64-GTP-16 S. T. ROBINSON J. W. GLESSNER. Solar, A Division of International Harvester Company, San Diego, Calif. Mems. ASME.

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1 PAPER NUMBER 4-GTP-1 Copyright 194 by ASME AN ASME PUBLICATION Total Turbine Energy in Refrigeration Cycles S. T. ROBINSON J. W. GLESSNER Solar, A Division of International Harvester Company, San Diego, Calif. Mems. ASME. The means of using total energy from a gas-turbine engine in various refrigeration systems are reviewed. Combinations of heating and cooling or electric power generation and cooling are discussed as well as combined centrifugal and absorption refrigeration systems. The economics of gas-burning turbine engines are investigated and shown to be attractive in these applications. $ I PER COPY S C TO ASME MEMBERS The Society shall not be responsible for statements or opinions advanced in papers or in discussion at meetings of the Society or of its Divisions or Sections, or printed in its publications. Discussion is printed only if the paper is published in an ASME journal. Released for general publication upon presentation THE AMERICAN SOCIETY OF MECHANICAL ENGINEERS 345 East 47th Street, New York, N. Y. 117 Contributed by the Gas Turbine Power Division for presentation at the Gas Turbine Conference and Products Show, Houston, Tex., March 1-5, 194, of The American Society of Mechanical Engineers. Manuscript received at ASME Headquarters, January 13, 194. Written discussion on this paper will be accepted up to April, 194. Copies will be available until January 1, 195. Printed in U.S.A. Downloaded From: on 11/2/217 Terms of Use:

2 Total Turbine Energy in Refrigeration Cycles S. T. ROBINSON J. W. GLESSNER The idea of using both the shaft power and exhaust heat of a gas turbine for total-energy utilization is not new. However, what is new -- and exciting -- is the use of small gas turbines which recover up to 7 or 75 percent of the input energy. The number of applications in the 2 to 1 hp range is many times that of the larger power range, which has been the traditional area for industrial gas turbines. In fact, up until the last 5 years, nobody had seriously thought of the small gas turbine as a contender in the industrial field. The major reasons why this now looks like a promising market are: 1 The increasing availability in nearly every section of the country of low-cost natural gas fuel. 2 The emergence of the simple-cycle gas turbine as a reliable, low-cost prime mover. 3 The utilization of waste heat from the turbine to offset its low efficiency when based on shaft power alone. The use of natural gas is intimately tied up in the success of the small industrial gas turbine. Not only is gas the least expensive fuel available, but its cleanliness makes it an ideal fuel for an internal-combustion gas turbine. It is well known that the natural-gas industry has excess capacity in the summer months and for this reason, gas-burning turbines used to supply refrigeration and air conditioning primarily in the summer look very promising. There are many combinations available with the turbine and all involve waste-heat recovery. There are straight refrigeration schemes and refrigeration plus heat or electric power. The purpose of this paper is to catalog a few of these various schemes and describe them. Since these applications are new, not much can be given in the way of economics, but where applicable, installations are described. It is the view of the authors' company that the best turbine engine for this kind of application should be the simplest one; that is, simple to build, simple to maintain and simple to operate. The kind of turbine which fits this description best is the single-shaft, single-stage centrifugal type with a single combustor. Using moderate turbine temperature of F, a specific fuel rate of 12, to 2, Btu/hphr is obtained. For this paper, 17, 5 Btu/hphr is used as a representative value. The performance of a typical small gas turbine is shown in Fig. 1. Since we are vitally interested in heat recovery, the steam which can be generated from it is also shown. This performance will be used throughout the paper in computing the heat rates for various refrigeration schemes. Fig, 2 shows the performance for a typical centrifugal refrigeration machine. By use of guide vanes in the compressor, the power required for reduced capacity has been minimized. The upper curve shows that at 3 percent capacity, more or 8 28 g o z 1 2 SHAFT HORSEPOWER Fig. 1 Typical single-shaft turbine performance ffs) RELATIVE POWER INPUT PERCENT OF FULL LOAD PERCENT RATED POWER INPUT PERCENT CAPACITY Fig. 2 Typical centrifugal refrigeration-system power requirements Downloaded From: on 11/2/217 Terms of Use:

3 H C.) 4 RELATIVE HEAT INPUT PERCENT OF FULL LOAD possible, these units are not used below 25 to 3 percent capacity. The same kind of plot is presented in Fig. 3 for an absorption refrigeration system. Using solution control, the specificheat requirement remains substantially constant over the whole capacity range of this absorption system. This performance is representative of the lithium-bromide systems capable of chilling brine to about 45 F. Using these three figures, the combination of turbine, centrifugal and absorption systems can be investigated under various circumstances. CENTRIFUGAL CHILLER WITH ABSORPTION REFRIGERATION SYSTEM 2 COOLING LOAD I PERCENT CAPACITY Fig. 3 Typical absorption refrigeration-system heat requirements WNMW CONDENSER EVAPORATOR V\A./\/\/VN VVVVVVVV\ CONDENSER VAPORAT111_ WVVVVW IE GENERATOR ABSORBER GAS TURBINE - Fig. 4 Centrifugal-absorption system for air conditioning less, the specific power requirements begin to rise rapidly, although the total power at that point is not high. If at all Driving a centrifugal refrigeration compressor with a gas turbine is the most direct way of obtaining refrigeration and affords the most flexibility in tonnage and temperature. For normal air-conditioning work, the power required is approximately 1 hp per ton and with a simple-cycle, gas-turbine drive, having a heat rate ranging from 12, to 2, Btu/ hphr, the gas rate would range from 12 to 2 cfh per ton. Although this lower value is quite acceptable, the part-load performance is quite poor, precluding its use in air-conditioning service where normally operation is at quite low load factors. The higher value, typical of low-power gas turbines, is noncompetitive with other systems throughout the load range. However, if the thermal energy in the exhaust can be recovered and used to drive an absorption machine to provide more cooling, design and part-load performance is improved, particularly if multiple units with boost burning are used to meet design tonnage. Absorption chillers can be driven by steam or hot water or directly by exhaust gas through a separate exhaust-heat generator. It makes little difference, so far as energy recovery is concerned, which is used; the final selection being dictated by economics or another use for a secondary heat-recovery fluid such as steam or hot water. As far as cost is concerned, a heat-recovery boiler is the most expensive. Hot-water heaters are about 8 percent of the cost of a 12-psi steam boiler of equivalent thermal output and use of a direct exhaust-fired absorption generator results in the most economical system. Fig. 4 is a schematic of a centrifugal-absorption system where the chilled water passes through the two systems in parallel. Fig.5 presents typical performance of such a system where boost burning is used to increase the absorption-system capacity. This is particularly useful in meeting design tonnage as specific gas consumption is fairly constant down to 5 percent load. Installation of two or more low-capacity chillers instead of one large unit is widely practiced in air conditioning. Such a system provides back up in event of machine failure and improves performance at low loads. Fig. shows gross and specific gas consumption versus tonnage for a turbine-driven centrifugal/absorption system of 1 tons capacity, consisting of dual 25-hp turbine drives and a 5-ton absorption chiller using boost burning to meet the design tonnage. As 2 Downloaded From: on 11/2/217 Terms of Use:

4 2 CENTRIFUGALS 3 '21 E9 2 OR CENTRIFUGAL z 1 a zo CENTR rfuga L -- BOOST BURNING ABOVE 8 TONS. 1 5, CENTRIFUGAL WITH ABSORPTION BOOST BURNING ABOVE 4 TONS FO x f9 412 O 11, 1 7 B 5 8 O N 4 CENTRIFUGAL ENTRI FUGAL WITH ABSORPTION z z TONS REFRIGERATION 45 F CHILLED WATER TONS REFRIGERATION 45F CHILLED WATER Fig.5 Performance of centrifugal chiller with absorption system Fig. Multiple centrifugal chillers with absorption system many units can be used as are required to meet the tonnage requirement, either with or without boost burning. Capital investment for a 1-ton system would be lowest using a single 5-ton chiller-condenser for two centrifugal units, a single boiler and a single 5-ton absorption chiller. Some mention should be made at this point about relative equipment costs. At the 5-ton level, the turbine-driven combined system including boiler, cooling tower, boost burner, and so on will cost about $85,. Of this cost, the centrifugal chiller, driver and associated equipment account for $45, and the absorption system for $4,. This total cost is about $15, higher than an electrically driven centrifugal system for the same capacity and the specific fuel rate is about 4 times the equivalent of the electric power required to drive a hermetic system. However, the rate spread between gas energy and electric power is such that the turbine total-energy system saves money every hour it operates. Choosing a range from 1 to 1 1/2 /kwhr for electric power and 3 to /mcf for gas, the savings is from $1 to $5 per full load hour for our 5-ton unit. At 1 hr full-load operation per year it would take from 3 to 15 years to pay back the $15, cost differential, neglecting interest charges. For 2 hr, the simple payback time is 1.5 to 8 years. Special circumstances can of course affect these payback times, but for an order of magnitude they are correct. Under very favorable circumstances, the payback period could be as little as a year, and under very unfavorable conditions, 15 years or more. Normally, a payback, including interest, should be made in no more than 5 to 1 years in order to be practical. An example of a centrifugal plus absorption system is illustrated by Washington Natural Gas Company's installation at its new Seattle office building. This 11, -sq-ft building is scheduled for completion in May 194. In this installation, a 22-ton centrifugal chiller is shaft driven by a 225- hp recuperative gas turbine. Since a recuperator is used, there is less waste heat from the engine and at full load the turbine exhaust will provide enough heat for only 5 tons of absorption refrigeration. The absorption unit itself is sized for 1 tons, the remaining 5 tons coming from additional heat sources. The waste heat in this case is recovered by a hot-water heater having an outlet temperature of 24 F. Fig. 7 shows the gas turbine and centrifugal compressor during installation. This unit is to go into service about the first part of 194. The average cooling load during the summer is expected to be about 175 tons with swings between 1 and 28 tons. The range up to 5 tons will be met by the absorption system alone with heat from another source. Here the turbine and centrifugal chiller are started, and by modulating their capac- 3 Downloaded From: on 11/2/217 Terms of Use:

5 Fig. 7 Washington Natural Gas Company plant ity and the capacity of the absorption unit, the range up to 1 tons is covered. At this load the absorption unit is at full capacity and the centrifugal chiller at 3 percent capacity. The remaining range, from 1 to 32 tons is covered by modulating the centrifugal chiller unit. Typical of totalenergy units, this one is expected to pay for itself through savings on energy cost in a period of about 8 years. CENTRIFUGAL CHILLER PLUS STEAM TURBINE Another way of using the waste heat is to generate steam in a waste-heat boiler and use this for a steam turbine which helps the gas turbine drive a centrifugal chiller compressor. In order to use the energy in the steam more effectively, the steam turbine should be condensing. Using a steam rate of 5 lb per hphr, the 2-hp gas turbine of the previous example supplies enough steam for a 55-hp turbine. The centrifugal compressor is sized at 315 tons to absorb all the power from both turbines. The full-load heat rate is 14, 4 Btu/ton-hr, using the same conditions as before. Note the total tonnage is smaller and the fuel rate higher than for the centrifugalplus-absorption system. Comparing this system with the centrifugal-plus-absorption system, we have substituted a steam turbine, condenser and a large centrifugal chiller for the absorption system. So in terms of equipment and performance, the steam turbine - gas turbine - centrifugal system is less attractive than the centrifugal-absorption system. 7 g 5 E 2 Z 4 F. (1 3 2 N. m. a.m. 2,,,,,.., 1/ /"',. irdlipem'' _...,,,migiss 15,srmsertro, 125 ffery,o. aim_a. 15 roll.2.1 Pr" KW Fig. 8 Combined electric power and absorption refrigeration system,. CENTRIFUGAL CHILLER PLUS STEAM TURBINE AND ABSORP- TION SYSTEM By combining these two systems, higher tonnage and a better fuel rate can be obtained. Using a 15-psig back pressure, the steam turbine not only furnishes the absorption system with steam but eliminates the condenser. Using a 55-1b/ hphr steam rate as typical for a small turbine like this results in 5 hp for 27 lb per hr of steam and an additional 5 tons of centrifugal refrigeration. The absorption system adds another 13 tons for a total of 44 tons or 97 Btu/ton-hr. For the addition of the steam turbine and a larger centrifugal 4 Downloaded From: on 11/2/217 Terms of Use:

6 Fig. 9 McAllen High School machinery room ELECTRIC POWER AND REFRIGERATION Thus far we have considered a straight refrigeration system producing chilled water for air conditioning. The heating, ventilating and air-conditioning system of a modern building requires from 1/2 to 1 kw of electrical energy per ton of air conditioning to provide for the operation of the refrigeration system and the air-supply and return fans. This total-energy requirement can be met by a gas turbine driving an electrical generator and using the exhaust heat to raise steam or hot water for building heating and/or air conditioning, the latter via an absorption chiller. The electrical load will size the equipment needed and by supplementing the exhaust heat by boost burning any heating or air-conditioning load usually encountered can be met. Fig.$ presents the performance that can be obtained from a 17-kw turbine-generator set using the exhaust heat with boost burning to raise steam for heating and/or air conditioning. While the performance shown is in terms of tons of air conditioning, all or any fraction is available as steam for heating in the quantity of 2 lb per ton. Fig. 1 Centrifugal-absorption cascade system for low temperature refrigeration compressor, an additional 1 percent output is obtained with the same fuel. The incremental cost is about $/ton installed or about $3 for this example. Figuring $. 45/mcf fuel, the reduction in operating costs amounts to about $. 3/hr at 4 tons so the pay back would take 1, hr. This scheme might be used for the sole purpose of improving efficiency, since the payback is 5 years at 2 hr per year. Even if the pay back time were considered too long, the addition of the turbine does increase the total refrigeration capacity and at a reasonable incremental cost. Many facilities such as hospitals, theaters and auditoriums, telephone centers and radio stations require a source of emergency power. Use of a combined system provides this source by transfer of part of the electrical load required for the heating, ventilating and air-conditioning system to more essential services, providing uninterrupted or "no-break" power with but minor wave distortions. The equipment for a combined electric power and refrigeration system is like that in Fig. 9, which shows the machinery room at the McAllen, Texas, High School. In this application, which is different than the one described, all of the electric power plus heating and cooling is supplied by a turbine system. Fig. 9 shows the two turbines driving two alternators each, one for -cycle power and the other for 84-cycle power which is used for fluorescent lighting only. 5 Downloaded From: on 11/2/217 Terms of Use:

7 COEFFICIENT OF PERFORMANCE OF CENTRIFUGAL CHILLER SYSTEMS 5' F CONDENSER TEMPERATURE 75% COMPRESSOR EFFICIENCY Fr pi. II, Pl. 'R-13 4 ;MM MI Q. il 3 A A / O; / -1 -BO VAPORATOR TEMP -- F Cle/Q8 Fig. 11 Turbine exhaust to shaft energy ratio F. Z 5 EP: 4 cl C.) F. At the time of this writing this system has been in operation only a little over 3 hr. With this combined system, peak electrical loads may not coincide with the maximum steam demand, so boost burning may be used if the steam demand cannot be met with the turbine-exhaust heat. On the other hand, if the electrical load exceeds the cooling load, then the excess cooling capacity can be used to cool the combustion air. This increases the turbine's shaft-power capability E.. co o TONS REFRIGERATION Fig. 12 Specific heat consumption - MBtu/ton-hour C.) 34. O Use of the high frequencies for lighting reduces the heat load due to lighting, lessening the amount of cooling required. The turbine exhaust is directed to an exhaust-heat boiler, one of which is used either for heating or for cooling. The generators are rated at 53 kva for the -cycle machine and 32 kva for the 84-cycle machine. At full load the turbine exhaust will generate 7 lb of steam per hr which can be used for heating or for driving a 47-ton absorption system. Co CASCADED CENTRIFUGAL AND ABSORPTION SYSTEMS Low-temperature refrigeration can be provided by a gasturbine system wherein the shaft output is used to drive a centrifugal chiller in cascade, i. e., in series, with an absorption chiller which in turn uses the exhaust heat from the turbine as an energy source. A schematic of such a system is shown in Fig. 1. In this system the centrifugal chiller provides the primary refrigeration, rejecting its heat to the absorption chiller which in turn rejects its heat to ambient. A desuperheater will undoubtedly be required to remove the superheat from the compressed primary refrigerant and reduce the load on the absorption system evaporator surface. Looking at the total system it will be seen that for efficient operation it is necessary to match the power and thermal requirements of the three individual systems involved; the absorption system must be sized to receive the heat rejected from the centrifugal chiller; and the gas turbine shaft and thermal output must match the power and thermal requirements of the centrifugal and absorption systems. Primarily, the characteristics of the two refrigeration systems define the ratio of usable turbine-exhaust energy to shaft energy. Designating this ratio by Q e/q s and using the coefficient of performance (COP) of the refrigerating systems Downloaded From: on 11/2/217 Terms of Use:

8 Fig. 13 Kitchens of Sara Lee as their performance index, we have Q e Ra/COPa = QsRc/COPe where R is the refrigeration load and subscripts a and c stand for absorption and centrifugal. The tonnage of the absorption system will be larger than that of the centrifugal system as the former must also reject the shaft power required by the later, so R a 1 + COPc ReCOPc substituting in the above Qe 1 + COP c Qs COPa Primary refrigeration provided by the centrifugal system would be in the range from to -1 F evaporator temperature. Fig. 11 shows the coefficient of performance that could be obtained over this evaporator temperature range with a 5 F condensing temperature in a system having a 75 percent compressor efficiency using refrigerants suitable for the temperature range involved. This same figure also shows the required value of Q e/qs for this range of evaporator temperatures where the centrifugal chiller load is rejected to single effect and double effect absorption systems having values of COP of. and.8, respectively. The total energy that can be recovered in a system of this type will range from 5 to 75 percent of the heat input, depending upon turbine-shaft efficiency and stack-gas temperature. Using 7 percent recovery as an average value, Q e/q s will vary from. to 2.5 for turbines covering the range of 1 to 2 percent shaft efficiency. Thus an evaporator temperature range from -23 to -9 F can be covered for a heat rate varying between 3 and cu ft of 1-Btu gas per ton refrigeration. Fig. 12 shows the part-load performance to be expected from a -5 F centrifugal refrigeration system powered with the 2-hp turbine and rejecting the primary refrigeration load to a double-effect absorption machine receiving its energy from the turbine exhaust. For this particular power plant a -5 F evaporator temperature results in shaft power and absorber heat requirements matching the gas-turbine characteristics. A small departure from this optimum will not seriously effect the heat rate per ton of refrigeration; but a much lower evaporator temperature results in a higher exhaust heat requirement, i.e., a higher ratio of Q e/q s, which can be met by boost burning. A higher evaporator temperature calls for a lower value of Qe/Q, which for optimum performance would require incorporation of a recuperator in the gas turbine system. Its worth is a question of economics, i. e., the fixed charges on the additional investment versus the saving in fuel. HEATING AND COOLING Not all turbine refrigeration systems are required to pro - duce refrigeration only. Prominent among other uses is heating and cooling where the turbine drives a centrifugal chiller and the exhaust heat is used for space heating or process use. 7 Downloaded From: on 11/2/217 Terms of Use:

9 TABLE 1 COMPARISON OF 2-HORSEPOWER GAS TURBINE -- REFRIGERATION SYSTEMS Turbine Fuel Rate Btu/hp-hour Refrigeration Systems Centrifugal Absorption HP/Ton Btu/ton-hour Refrigeration Tons Centrifugal Absorption Total Specific Fuel Consumption Btu/ton-hour , , , , , Depending upon the application, the heat might be used directly, or to generate steam or hot water. The Kitchens of Sara Lee furnish an example of this use. Fig. 13 is an artist's rendition of their new plant. In this famous bakery, the turbine shaft power is used to drive a lowtemperature centrifugal chiller and its exhaust heat to raise high-pressure steam. Each compressor produces 535 tons of -42 F refrigeration with 1 hp input. This refrigeration system is cascaded with another + 25 F centrifugal system, the 25 F being used for storage refrigeration and the -42 F used for blast freezing. In this application at total of 2 hp is required to effect the 535 tons of low-temperature refrigeration. The turbines have a heat rate of 12, Btu/hp hr so the specific fuel consumption is 45, Btu/ton-hr. This can be compared with the cascaded centrifugal plus absorption system in Fig. 12. FUTURE IMPROVEMENTS All the examples shown are current state-of-the-art and are typical of the performances that can now be obtained. Improvement can be expected as experience is gained with these systems. Improvement of the turbine efficiency will have a major effect on the overall fuel rate. Considering the centrifugal plus absorption system, an improvement in the turbine's fuel rate to 12, Btu/hphr will reduce the fuel needed to drive the centrifugal but also reduce the exhaust-heat content. A typical turbine of 2 hp having this fuel rate would produce about half the steam and half the absorption tonnage of the one used in these examples. This results in an overall fuel rate of 95 Btu/ton-hr at a total rating of 33 tons. Improving the efficiency of the centrifugal system will also improve the fuel rate as pointed out earlier. A centrifugal refrigeration machine requiring only. 9 hp/ton will improve the specific fuel consumption to 1, Btu/ton-hr at a total rating of 43 tons. The absorption system can also be improved and perhaps accompanied with a reduction in first cost, too. An obvious step is to use a direct-fired absorption system so that the generator of the system is directly heated by the turbine-exhaust gas. This eliminates the boiler, feedwater pump, condenser, water treatment, and so on, and the losses associated with the boiler. It is, however, a rather major redesign of the absorption system, but nevertheless, some development in this direction is occurring. Use of a double-effect, direct - fired absorption system promises to bring the heat requirements down to about 15, Btu/ton-hr. This will increase the absorption tonnage so that, on the same basis as before the specific fuel consumption will be 1, 2 Btu/ton-hr at 447 tons. Recapitulating, improvements in the gas turbine and the centrifugal and absorption refrigeration systems which are all technically within reach can reduce the fuel rate, which is already very good, to something even better. Table 1 shows these improvements singly, as discussed above, and finally, all incorporated in the same package. Acknowledgment The authors would like to express their appreciation to all who helped in preparation of the data and performance in this paper. 8 Downloaded From: on 11/2/217 Terms of Use: