Energy Performance Comparison between Power and Absorption

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Energy Performance Comparison between Power and Absorption Refrigeration Cycles for Low Grade Waste Heat Recovery MENGYING WANG, Yufei Wang, Xiao Feng, Chun Deng, and Xingying Lan ACS Sustainable Chem. Eng., Just Accepted Manuscript • DOI: 10.1021/ acssuschemeng.7b03589 • Publication Date (Web): 09 Feb 2018 Downloaded from http://pubs.acs.org on February 10, 2018

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ACS Sustainable Chemistry & Engineering

Energy Performance Comparison between Power and Absorption Refrigeration Cycles for Low Grade Waste Heat Recovery Mengying Wang†, Yufei Wang* ,†, Xiao Feng‡, Chun Deng†, Xingying Lan† †

State Key Laboratory of Heavy Oil Processing, China University of Petroleum, Beijing 102249,

China ‡

School of Chemical Engineering & Technology, Xi’an Jiaotong University, Xi’an 710049,

China *Corresponding author: Yufei Wang. E-mail: [email protected]

ACS Paragon Plus Environment

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ABSTRACT: Power cycles converting waste heat to electricity and absorption refrigeration cycles converting waste heat to cooling are widely used for low grade waste heat recovery in the process industry. It is significant to choose an appropriate cycle to recover waste heat at different temperatures. In this paper, a comparison is performed between a Kalina cycle and a LiBr-H2O absorption refrigeration cycle considering low temperature waste heat sources at different temperatures, and the optimal conditions of these two waste heat utilization technologies are determined. For power and cooling are different kinds of energy with different energy levels, a novel indicator, the amount of mechanical work, is used to evaluate the cycles to reasonably compare the energy performance of the cycles. The cooling generated by an absorption refrigeration cycle is converted into mechanical work required by a compression refrigeration cycle under the same cooling output. Results show that when the waste heat source temperature is from 100 to 175 ℃, it is proper to choose a LiBr-H2O absorption refrigeration cycle; when temperature range is 175 to 200 ℃, Kalina cycle is preferable.

KEYWORDS: Kalina cycle, absorption refrigeration, energy performance, low grade waste heat recovery


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INTRODUCTION There is an enormous amount of waste heat released in industrial processes. At present, industrial waste heat recovery rate is only about 30% in China1, and the energy utilization efficiency is low. According to the waste heat temperature, waste heat can be divided into hightemperature, medium-temperature and low-temperature waste heat, and low-temperature waste heat is normally below 200 ℃ with lower recovery efficiency than high and mediumtemperature waste heat. Therefore, improving the energy utilization efficiency of a waste heat recovery system is an effective way to increase the whole energy efficiency of the process and reduce CO2 emissions. Common industrial waste heat recovery technologies can be classified according to the characteristics of energy transfer or conversion, so they are mainly divided into direct heat recovery via heat exchange, power cycles, refrigeration and heating. Generally, direct heat utilization is the primary consideration of waste heat recovery. When there are no suitable heat users, waste heat can be recovered to generate power or cooling. However, when a power cycle and an absorption refrigeration cycle both have corresponding users, the aim of this paper is to study which cycle has better energy performance at different waste heat temperatures. It is of practical significance for the selection of waste heat recovery technology in the process industry. Low grade waste heat sources come from many process streams, and the characteristics of these streams are quite different in an industrial process. So, the system is too complicated if the working fluid of the recovery cycle exchanges heat directly with multiple waste heat streams. In practical processes, indirect heat transfer is usually adopted, namely, an intermediate medium (mostly hot water) exchanges heat with process streams first, then with the working fluid. In this case, the heat source from the intermediate medium is sensible heat. Our former work2 has found


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that the energy performance of a Kalina cycle is better than that of an Organic Rankine cycle (ORC) when a waste heat source is a sensible one. Therefore, the power cycle in this paper only considers Kalina cycles. As for refrigeration technologies, the single effect LiBr absorption refrigeration cycle has better economic performance when generating cooling from low temperature waste heat. Therefore, the single effect LiBr absorption refrigeration cycle is taken into consideration in this paper. The Kalina cycle proposed by Dr. Kalina3 uses the zeotropic mixture NH3-H2O as working fluids, which is principally a “modified” Rankine cycle. The mixture brings about a non-constant evaporation temperature and makes a good thermal match between the sensible heat source and the working fluid temperature profiles, so that less irreversibility occurs during the waste heat recovery process4. A LiBr-H2O absorption refrigeration cycle is a technology using aqueous LiBr solution as the absorbent and pure water as the refrigerant, as LiBr-H2O pair can have a refrigeration effect when mixing and separating continuously. In the evaporator, refrigerant is evaporated by receiving heat from a low temperature heat source, producing refrigeration effect. The LiBr-H2O absorption refrigeration system has a good refrigeration effect because of a large boiling temperature difference of working pair and large evaporation latent heat of pure water. Compared to the traditional compression refrigeration technology, the absorption refrigeration technology can recover waste heat, and has a small amount of power consumed by solution pump and heat rejection system for the condenser and absorber. Recently researches for waste heat recovery technologies have mainly on thermodynamic performance, influence of operating parameters on system efficiencies, and development and application of system configurations. Wang et al.5 proposed a new combined cooling and power system using ammonia–water mixture as working fluids to recover low grade waste heat,


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achieving both power and cooling supply for users, and analyzed the thermodynamic performance of the whole system. Saffari et al.6 conducted a thermodynamic analysis for a geothermal Kalina cycle and investigated operating parameters to optimize thermal and exergy efficiencies. Recent research indicated better thermodynamic performance can be achieved when using such mixtures as working fluids with low heat source temperatures than with high heat source temperatures7. In regard to the development of new systems, Hua at al.8 proposed a modified Kalina cycle, acquiring higher power recovery efficiency than a traditional steam Rankine cycle. Similarly, in the study of absorption refrigeration cycles, many studies have conducted thermodynamic analysis mainly based on the coefficient of performance (COP) or/and exergy efficiency. Kaynakli and Kilic9 studied the effect of various operation parameters on the COP indicator. Gong et al.10 analyzed the parameters of an absorption refrigeration using advanced exergy analysis. In this paper, the analysis of energy performance of absorption refrigeration cycles is also based on the first and the second law of thermodynamics. In order to investigate the effect of cycle parameters on the system performance, Karamangil et al.11 developed a visualized software to simulate the cycle performance and found that the operation temperatures in the generator, absorber, evaporator and condenser affect the cycle performance. The COP of the cycle increases with increasing operation temperature in the generator and evaporator but decreases with increasing absorber and condenser temperature. Kilic and Kaynakli12 adopted the COP and exergy efficiency indicators to evaluate the effect of each component’s temperature on the system performance. They also found that COP of the cycle increases with increasing generator and evaporator temperature. However, the energy performance of power and absorption cycles cannot be compared based on these indicators, since the energy produced by these two technologies is of different levels. Besides, the working fluids


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of absorption refrigeration cycles have a central effect on the performance. Sun et al.13 and Luo et al.14 carried out comprehensive reviews on the working fluids of absorption refrigeration cycles and studied the thermodynamic properties and application of the working fluids. The LiBr-H2O pair is preferred for residential air-conditioning applications due to its high coefficient of performance and easy management, but corrosion and crystallization issues exist with high working temperature15. Many researches also have been conducted to improve the absorption process. Yang et al.16 proposed a cascade refrigeration technology (CRT) that combines a LiBr absorption refrigeration with a NH3 absorption refrigeration. Yang et al.17 proposed a novel combined power and ejector-refrigeration cycle using the zeotropic mixture as the working fluid. Shi et al.18 proposed a new absorption refrigeration cycle with a simple construction to make full utilization of waste heat with large temperature glide and conducted a thermodynamic analysis of the system. Typical waste heat refrigeration cycles can also recover solar energy19, which have a wide application. Basically, the development of novel system prototypes also requires a uniform indicator for thermodynamic analysis when comparing with power cycles. There are some comparative studies on the same type of waste heat utilization technologies at present. Wang et al.20 conducted a thermodynamic performance comparison between ORC and Kalina cycles considering multiply waste heat streams. They divided multi-stream waste heat into three kinds, straight, convex and concave waste heat, and found that the thermodynamic performance of these two technologies is related to the shape of the waste heat composite curve. Walraven et al.21 investigated the performance of different types of organic Rankine cycles (ORCs) and Kalina cycles for low-temperature (100~150 ℃) geothermal heat sources. They found that transcritical and multi-pressure subcritical ORCs are in most cases the best performing cycles. Farshi et al.22 studied and compared three classes of double-effect lithium


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bromide–water absorption refrigeration systems (series, parallel and reverse parallel) with identical refrigeration capacities thermodynamically. Their results reveal the advantages and disadvantages of different configurations of double-effect lithium bromide–water absorption refrigeration systems. Little and Garimella23 analyzed five thermodynamic cycles to generate cooling, higher-grade heating, or mechanical work from waste heat, and compared cycles with the same output. The comparative criteria for the cycle performance are based on conservation of energy quantity and footprint required by each equipment. The performance comparison for power cycles is based on the net power output per unit heat input, while that for absorption chillers is based on the net chilling produced per unit heat input. They found that the organic Rankine cycle performs better than the Maloney-Robertson cycle, and the absorption cycle is better than the organic Rankinevapor compression cycle. Furthermore, they analyzed specific application of each cycle for waste heat recovery. However, they did not compare cycles with different energy outputs (e.g. power and refrigeration cycles). And they considered only two heat source temperatures (60 ℃ and 120 ℃). The comparative study on low temperature waste heat utilization cycles with different purposes is less, and the comparative indexes for different waste heat utilization technologies have limitations. Van de Bor et al.24 investigated different kinds of heat pumps and power cycles based on economic benefits. They found that heat pumps show better economic performance than power cycles in the temperature range of 45~60 ℃, while organic Rankine and Kalina systems become competitive at waste heat temperature of 130 ℃ or higher. But the study did not compare energy utilization performance.


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Gbemi et al.25 developed a screening criterion, the exergy degradation, to measure the deviation from the ideal performance of different technologies. They compared waste heat power recovery, refrigeration and heat pump systems as well as direct heat recovery, in which the waste heat power recovery system only considered ORCs, the absorption refrigeration cycle was considered in refrigeration systems, and heat pump systems include mechanical heat pumps, absorption heat pumps and absorption heat transformers. Then they determined feasible heat source temperature range for technology options using the exergy degradation. Finally, the obtained chart tool can be applied to screen and select technologies for waste heat recovery at different temperatures. Although different technologies at different waste heat temperatures were considered in the work, they only considered ORC for power cycles and they did not optimize operating parameters of different technologies. Obviously, the above researches on Kalina and LiBr-H2O absorption refrigeration cycles are mainly about performance and efficiency of a single system, but there is no comparison between them to recover low grade waste heat. Because of different energy levels of power and cooling, previous researches on the selection from power cycles and refrigeration at different waste heat temperatures were lack of a uniform standard, and the standard should be easily understood and accepted by the industry. The aim of this paper is to analyze the energy performance of a Kalina cycle and a LiBr-H2O absorption refrigeration cycle to recover low temperature waste heat at different temperatures. The optimal conditions of these two cycles are determined. In order to reasonably compare the energy performance of these two cycles with different energy outputs, a novel indicator, the amount of mechanical work, is used to evaluate the cycles. For the absorption refrigeration cycle, the cooling generated by absorption refrigeration is converted into mechanical work required by


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compression refrigeration under the same cooling output and compared with the mechanical work generated by Kalina cycle. In this way, the better waste heat utilization technology under different waste heat temperatures is obtained, which can provide a reference to waste heat recovery in process industries.

SIMULATION MODELS To evaluate the energy performance of Kalina and absorption refrigeration cycles, the parameters of each cycle should be obtained. So, the two cycles are first modeled in the process simulation software. Kalina cycle. The Aspen Hysys software is used to simulate the Kalina cycle, and the thermodynamic properties are calculated by REFPROP Equation of State26. Because the Kalina cycle has different configurations, each configuration corresponds to its most appropriate application. In this paper, two kinds of Kalina cycle configurations are established according to different waste heat temperatures. The model of Kalina cycle (a) (Figure 1a) is established based on that of Bombarda et al.27, which can recover the waste heat at 100~150 ℃. The model of Kalina cycle (b) recovering the waste heat more than 150 ℃ is established based on that in Ref.28.


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4 6

Turbine

Wtur H1

Evaporator 3

2

Separator H2

H3

Perheater

5 1

Recuperator 10

11

R

VLV

Mixer

11-1

RCY Qcon

9

7

8

Pump

Wpump

Condenser

(a) H2

Evaporator

Preheater 1 1

2

Recuperator 2 4

3 H1

Turbine

Wtur 17

H3

5

14

15

Preheater 2

Recuperator 3 Separator

16 H4

Recuperator 1 19

12 18

20 9

13

Mixer 2

21

Splitter

Wpump2

Pump 2

10

Condenser 2

8

Wpump1

21-1

RCY Mixer 1

Qcon1

Qcon2 11

R

VLV

7

Pump 1

6

Condenser 1

(b) Figure 1. Schematic diagrams of Kalina cycles


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In Figure 1b, the NH3-H2O mixture (1) is heated by the waste heat stream in preheater 1 and the evaporator. The heated vapor (3) is expanded in the turbine to generate mechanical power. After expanding, the hot gaseous working fluid (4) is used to preheat the cold liquid mixture (14, 15) in the recuperator 2, mixed with the weak concentration solution (21-1) from the throttle, and then condensed in condenser 1. The liquid mixture (6) is pressurized by pump 1, and then splits into two parts, one of which (13) is heated in recuperator 1 and recuprator 2, and the other (9) mixed with the strong concentration mixture (18) in mixer 2. The heated working fluid (15) is then heated to the separation temperature by waste heat stream (H3) in preheater 2, then separated in the separator, and the separated strong concentration mixture (17) goes into recuprator 3. The mixture (10) from mixer 2 is condensed in condenser 2, pressurized through pump 2, heated in recuprator 3, and again goes into the waste heat recovery processes. When the waste heat temperature is low, the waste heat recovery is carried out in Kalina cycle (a). Different from Kalina cycle (b), the NH3-H2O mixture is separated directly in the separator after being heated in the preheater and evaporator, and then the separated strong concentration mixture goes into the turbine. In order to simplify the process simulation, the Kalina cycle is assumed to be as follows: 1, the cycle is operated under steady condition; 2, pressure loss due to the friction and heat loss in heat exchangers and connecting pipes are taken as negligible; 3, working fluid in turbine inlet is saturated vapor; 4, working fluid in pump inlet is saturated liquid. 5, the isentropic efficiencies of the turbine and working fluid pump are assumed to be 80 %.


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The models for each component and the calculated parameters in each block are given in Table A1 and A2 in Appendix. Kalina cycles are cut off in Aspen Hysys for further simulation, and the break is inserted at state point 1 (Figure 1). Stream 1 is given an initial value. If these two broken fluid streams have the same results, the model converges. If they are different, the initial value of stream 1 needs to be adjusted. The cycle model of Kalina (a) requires only the following inputs: (1) the supply temperature of the heat source; (2) the separation temperature; (3) the inlet and outlet pressure of Turbine; and (4) quantity of heat available. In addition, the cycle model of Kalina (b) also needs the inputs of the evaporator exit temperature (related to waste heat temperature) and the splittter ratio. Absorption refrigeration cycle. A traditional single effect LiBr-H2O absorption refrigeration cycle consists of a generator, condenser, throttle, evaporator, absorber, solution heat exchanger and pump. The absorbent is LiBr solution and refrigeration fluid is pure water. The LiBr solution absorbs heat from waste heat source in the generator, and the refrigeration fluid (water) is evaporated and condensed in the evaporator and condenser. At the same time the LiBr solution in the generator becomes strong solution and flows into the absorber through the throttle. The saturated refrigeration fluid flows into the evaporator at evaporation temperature by the throttle to get a refrigeration effect. The LiBr solution is turned into dilute solution after mixed with the steam from the evaporator and absorbed by the steam. In the process, the generator pressure is higher than that in the absorber, so the dilute solution is pressurized in the pump and brought to the generator again and a cycle is done. The whole cycle is divided into a solution cycle and a refrigeration fluid cycle. Figure 2 shows the schematic diagram of the single absorption refrigeration cycle.


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8 CW3

Generator 4

H1

QCON

Separator 9

H2

Condenser REFV

5 3

COOLIN

10

QSHE

QEVAP

Solution Heat Exchanger 6

Evaporator COOLOUT

SOLV 2 7 1

11

Pump

WPUMP

CW2 1A

Absorber

12

QABS

MIXER

CW1

Figure 2. Schematic diagram of the single absorption refrigeration cycle Because LiBr-H2O belongs to electrolytes, ELECNRTL Equation of State is selected for the process simulation via Aspen Plus29. In order to calculate thermodynamic parameters easily, assumptions for each state point of the single effect LiBr-H2O absorption refrigeration cycle is given as follows. 1, pressure loss due to the friction and heat loss in heat exchangers and connecting pipes are taken as negligible; 2, refrigeration solution in the solution pump inlet is saturated liquid; 3, refrigeration fluid in the condenser outlet is saturated liquid;


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4, refrigeration fluid in the evaporator outlet is saturated vapor. 5, the isentropic efficiency of the pump is assumed to be 80 %. The model for each component and the calculated parameters in each block are given in Table A3 in Appendix. Because Aspen Plus uses a sequential solver, and the absorption refrigeration cycle is a closed cycle, it is necessary to break the cycle. In this paper, the break is inserted at state point 1 (Figure 2). In other words, the exit of the absorber (stream 1A) and the inlet of the pump (stream 1) are not connected. Stream 1 is given an initial value. If these two fluid streams have the same results, the model converges. If they are different, the initial value of stream 1 needs to be adjusted. The cycle model requires only the following inputs: (1) the supply temperature of the heat source; (2) the effectiveness of solution heat exchanger; (3) the evaporator exit temperature (related to desired cooling temperature); (4) the supply temperature of cooling water; and (5) either quantity of heat available or desired cooling load. For pure water in the Kalina and the absorption refrigeration cycles, the STEAMNBS property method is used, which is based on the International Association of Properties of Water and Steam (IAPWS)30.

OPTIMIZATION AND ANALYSIS OF CYCLE PARAMETERS When comparing different waste heat recovery technologies, the comparison should be made under optimal conditions. Therefore, the optimization of these two cycles is carried out in this section. Optimization of Kalina cycle. The mass fraction of ammonia water, turbine inlet pressure, separation temperature of separator and separation ratio of splitter are four main operating parameters that can influence performance of Kalina cycle. Optimal operating parameters are


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different at different waste heat temperatures. This paper considers the Kalina cycle recovering waste heat at different temperatures with the same heat duty of 1,000 kW. Heat source can be cooled to 55 ℃. The minimum temperature difference is set as 10 ℃. The optimal parameters at 200 ℃ of a waste heat source are determined as an example. To evaluate the performance of the cycle, the net power output and exergy conversion efficiency as indicators will be used. The exergy conversion efficiency for the Karina cycle is defined in Eq. (1).  EC _ KC =

Singh and Kaushik

31

Net power output Total exergy of waste heat

(1)

has shown that when the turbine outlet pressure is certain, the system

thermal efficiency increases with the increase of the mass fraction of ammonia water and turbine inlet pressure, but the mass fraction of ammonia water will not increase indefinitely in practical production. So, the mass fraction of ammonia water is set at 90 %. When the composition of the working fluid at the turbine inlet is constant, the net power output and exergy conversion efficiency increase first with the increase of the separation temperature. To a certain extent, the trend is slowing down and in decline, then decreases with the increase of separation temperature. The initial separation temperature is set at 96 ℃. The inlet stream of the pump is saturated liquid, so the turbine outlet pressure is set at 640 kPa, and an initial value of flow ratio of splitter (i.e. splitter ratio=mass flow 9/mass flow 8) is set at 0.1. The variation of the net power output and exergy conversion efficiency at different turbine inlet pressure is shown in Figure 3. In Figure 3, when the turbine inlet pressure is 9000 kPa, the net power output and exergy conversion efficiency are the maximum. In this case for the temperature difference of preheater 1 does not reach the limit of 10 ℃, the net power output can be increased further by changing the separation temperature and splitter ratio to recover more waste heat. Therefore, the influence of the separation temperature and splitter ratio on the energy performance of the system can


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continue to be investigated under the pressure of 7000 ~ 9000 kPa. Because the change of operating parameters has the same effect on the net power output and exergy conversion efficiency, only the net power output is investigated when considering different pressures. Figure 4 shows the variation of net power output with the separation temperature and splitter ratio under different turbine inlet pressure.

Figure 3. Energy performance of Kalina cycle at different turbine inlet pressure Figure 4 shows that the net power output increases first and then decreases with the splitter ratio under the same turbine inlet pressure and separation temperature. It is because the temperature difference of preheater 1 reduces gradually with increasing of the splitter ratio. When it reaches the limit of 10 ℃, the mass flow of working fluid cannot continue to increase, resulting in the decrease of the net power output. Under different turbine inlet pressures, the optimal separation temperature increases with increasing of the turbine inlet pressure, but the optimal splitter ratio is the opposite. Therefore, the optimal parameters can be determined in


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Figure 4. The turbine inlet pressure is 8000 kPa, separation temperature is 60 ℃, and splitter ratio is 0.16.

Figure 4. Energy performance of Kalina cycle at different operating parameters

Figure 5. The variation of net power output and exergy conversion efficiency at different waste heat temperatures


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The above optimization method is adopted to analyze the systems recovering waste heat sources at different temperatures with the same heat duty of 1,000 kW, and the simulation conditions for Kalina cycles at different temperatures are shown in Table A4 and A5 in Appendix. The variation of the net power output and exergy conversion efficiency at different waste heat temperatures can be obtained as shown in Figure 5. The simulated results of the net power output at different temperatures can be verified in Ref.32, though there is little deviation. It is shown in Figure 5 that the net power output and exergy conversion efficiency increase as heat source temperature increases both for Kalina cycles (a) and (b). When the temperature is 150 ℃, the net power output of Kalina cycle (a) is slightly less than that of Kalina cycle (b). The curve of these two systems can confirm that when heat temperature is greater than 150 ℃, the energy performance of Kalina cycle (b) is better, while Kalina cycle (a) is better at 100~150 ℃. Optimization of absorption refrigeration cycle. Wang et al.33 investigated the influence of effectiveness of solution heat exchanger, condenser temperature and mass fraction of LiBr-H2O solution on the performance of the absorption refrigeration cycle. They found that reducing condenser temperature, increasing effectiveness of solution heat exchanger and mass fraction of LiBr-H2O solution can increase the coefficient of performance (COP) and exergy efficiency at the same time. But the condenser temperature cannot be reduced unlimitedly, often controlled at more than 35 ℃. The mass fraction of LiBr-H2O solution should be also controlled within a reasonable range to avoid crystallization problem in practical operations. The same heat source can generate cooling of different grades using absorption refrigeration, by only changing evaporator temperature. In this section, we assess how the system performance changes with the heat source temperature at a fixed evaporator temperature of 5 ℃ according to the common demand of cooling in the industry.


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The heat source temperature is an important factor affecting the efficiency of the refrigeration cycle, so the influence of the heat source temperature on the system performance needs to be analyzed. The heat source temperature cannot be higher than 180 ℃, as the use of the absorption refrigeration cycle is not possible due to crystallization of the working fluid25. The waste heat temperature range is 100~180 ℃ under consideration for the absorption refrigeration cycle, and heat duty is taken as 1000 kW, which is the same with the Kalina cycle. The outlet temperature of the heat source is the outlet temperature of the solution from the generator plus 10 ℃34. Inlet and outlet temperatures of the refrigerant (water) are 12 ℃ and 7 ℃, respectively. The inlet temperature of cooling water is set as 25 ℃, which is a typical value. The concentration of the solution is taken as 0.5834. Cooling water in absorber and condenser is in a series mode. The effectiveness of solution heat exchanger is specified as 0.6435. For the exergy calculation, the environmental state is taken as 25 ℃ and 1 atm. To evaluate the performance of the cycle, two indicators are used as follows. Coefficient of performance (COP): Exergy conversion efficiency:

COP =

 EC _ AR =

Cooling output Waste heat recovered

Cooling exergy increased Total exergy of waste heat

(2) (3)

The temperature difference between the saturated refrigerant water from condenser and cooling water is 3 ℃ (generally 2~5 ℃)36. The saturated refrigerant water temperature is set at 40 ℃, so the condenser pressure is 7381.39 Pa, which is the high side pressure. The evaporation temperature is set at 5 ℃，and the evaporation pressure is 872 Pa, which is the low side pressure. Then the absorber pressure can be determined to be 800 Pa (generally lower than evaporator pressure by 27-80 Pa)36. The LiBr-H2O absorption refrigeration cycle is simulated at different waste heat temperatures based on the operation conditions above shown in Table A6 in Appendix. The energy


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performance of the cycle at different temperatures is shown in Figure 6, which is in consistent with Ref.33. With increasing heat source temperature, the COP increases gradually. When the heat source temperature increases over a certain extent, the increase in the system exergy loss will be greater than that of exergy output. In this case, the exergy conversion efficiency will decrease. Therefore, there is a maximum exergy efficiency, as shown in Figure 6. From the change of COP and exergy conversion efficiency, the effect of increasing the heat source temperature on the system performance is not obvious, and the exergy of waste heat energy is not used effectively.

Figure 6. Energy performance of the absorption refrigeration cycle at different waste heat temperatures ENERGY PERFORMANCE COMPARISON BASED ON CONVENTIONAL INDICATORS After the determination of optimal operation conditions of these two cycles at different waste heat temperatures, a comparison of the energy performance of the cycles is investigated first


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based on conventional indicators (i.e. waste heat conversion efficiency and exergy conversion efficiency). Waste heat conversion efficiency. Waste heat conversion efficiency is based on the first law of thermodynamics, the definition of which for the power cycle and refrigeration cycle is in Eqs. (4) and (5), respectively.  HC _ KC =

Net power output Total waste heat input

(4)

 HC _ AR =

Cooling output Total waste heat input

(5)

Figure 7. Waste heat conversion efficiency of Kalina cycle and absorption refrigeration cycle at different temperatures The variation of waste heat conversion efficiency of the Kalina and absorption refrigeration cycles with heat source temperatures is shown in Figure 7. The waste heat conversion efficiency of the Kalina cycle is always higher than that of the absorption refrigeration cycle. When waste heat duty and temperature are the same, cooling generated by absorption refrigeration will be far greater than power generation.


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Exergy conversion efficiency. Due to the different grades of cooling and power, the result is not reasonable based on the waste heat conversion efficiency, which cannot be used to guide the selection of waste heat recovery technologies. When comparing energy of different grades, the conventional method is based on exergy analysis. The conversion of exergy can be used to characterize the capability and technical usefulness of the whole system in converting waste heat sources into mechanical work. Exergy conversion efficiency is based on the second law of thermodynamics, the definition of which for the power cycle and refrigeration cycle is given in Eq. (1) and Eq. (3), respectively. The exergy conversion efficiencies of these two technologies at different heat source temperatures show in Figure 8. The exergy conversion efficiency of the Kalina cycle increases slightly and then rises gradually with the heat source temperature, while that of the absorption refrigeration cycle increases first and then decreases with the waste heat temperature, and it changes smoothly. Contrary to the result of the waste heat conversion efficiency, the exergy conversion efficiency of the absorption refrigeration cycle is always lower than that of the Kalina cycle. It can be found that these two waste heat recovery technologies show different performances when comparing by the waste heat conversion efficiency and exergy conversion efficiency. The main reason is that the energy level of cooling and power is different. Power is useful work, but a part of exergy will be lost when cooling converts into useful work. So, the exergy conversion performance of the Kalina cycle is better, even though the cooling output of the absorption refrigeration is higher than the power generated by the Kalina cycle. However, the results from the exergy conversion efficiency is still inconsistent with common sense, which also cannot be used to guide the selection of waste heat recovery technologies.


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Figure 8. Exergy conversion efficiencies of Kalina cycle and absorption refrigeration cycle at different temperatures ENERGY PERFORMANCE COMPARISON BASED ON MECHANICAL WORK Introduction of a new indicator. The mechanical work generated by the Kalina cycle and the cooling generated by the absorption refrigeration cycle are energy with different grades, the users of these two cycles are different. From the previous section whether using the waste heat conversion efficiency or the exergy conversion efficiency as indicator, it is impossible to guide the selection of waste heat recovery technologies. Therefore, it is necessary to adopt a new indicator. In this paper, the amount of mechanical work is proposed as a uniform indicator to compare the energy performance of the cycles. For the power cycle, the amount of mechanical work is the net power output recovered from the waste heat. For the absorption refrigeration cycle, it is converted from the mechanical work required by the equivalent compression refrigeration cycle under the same cooling output. When the power cycle can generate more mechanical work, the


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power cycle is better, otherwise the absorption refrigeration cycle is preferable. Obviously, such indicator has advantages of reasonable, intuitive in understanding and easy to operate in industries. Model of equivalent compression refrigeration cycle. A vapor compression refrigeration system is a technology for refrigeration by consuming electric energy, which is used as an equivalent cycle to convert the cooling output of absorption refrigeration into mechanical work. A vapor compression refrigeration system is composed of a compressor, condenser, expansion valve and evaporator. The role of the compressor in this system is the same with that of the solution cycle in the absorption refrigeration cycle. The coefficient of performance of a vapor compression refrigeration cycle is the cooling output per unit power input. And the coefficient of performance for a vapor compression refrigeration cycle is higher than that of the absorption refrigeration cycle. The vapor compression refrigeration cycle is modeled in Aspen Plus software, as shown in Figure 9. Relevant parameters, such as the condensation temperature, evaporation temperature and cooling output of the compression refrigeration system are the same with that of the absorption refrigeration system at different heat source temperatures shown in Table A7 in Appendix. The isentropic efficiencies of the compressor are also assumed to be 80 %. Water is selected as the refrigerant to recover the given waste heat, as water can maintain higher coefficient of performance over other refrigerants37. To show the validation of this model we developed, we compared our model with the model in the literature38. The results of the COPs and operating conditions mentioned in the literature are shown in Table 1. Then the vapor compression refrigeration system is simulated at different cooling outputs. Parameters and results of two refrigeration systems are shown in Table 2.


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COOLIN

Evaporator

COOLOUT

QEVAP 2

WCOMP

Compressor 1

3

4

VLV

Expansion valve

Condenser

QCON

CW1

CW2

Figure 9. Vapor compression refrigeration model in Aspen Plus Comparison results based on mechanical work. According to the simulation results, mechanical work input of the equivalent compression refrigeration system and the net power output of the Kalina cycle are in comparison shown in Figure 10. In Figure 10, when the heat source temperature is in the range of 100~175 ℃, the mechanical work input of the vapor compression refrigeration system is more than the net power output of the Kalina system. But the result reverses when temperature is higher than 175 ℃. As a result, when waste heat source temperature ranges from 100 to 175 ℃, the LiBr-H2O absorption refrigeration cycle should be considered first to recover the waste heat; when the temperature range is 175 to 200 ℃, Kalina cycle is preferable.

CONCLUSION In this work, a comparison of energy performance is performed between a Kalina and a LiBrH2O absorption refrigeration cycles for low temperature waste heat recovery at different temperatures, and the comparison is under the determined optimal conditions. The conventional


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indicators, the waste heat conversion efficiency from the first law of thermodynamics and the exergy conversion efficiency from the second law of thermodynamics, cannot give a guide to select a suitable technology from the two cycles. In order to reasonably compare the energy performance of these two cycles, a novel indicator, the amount of mechanical work, is used to evaluate the cycles. For the power cycle, the amount of mechanical work is the net power output recovered from the waste heat. For the absorption refrigeration cycle, the cooling generated by the absorption refrigeration cycle is converted into mechanical work required by an equivalent compression refrigeration cycle under the same cooling capacity. Table 1. Comparison of the recent compression refrigeration cycle model with the model developed in this work COP operating conditions In the literature Refrigerant: water

In our work

3.7

3.61

3.25

3.05

2.75

2.60

2.386

2.23

Te=0 ℃, Tc=40 ℃ ηis=0.7 Refrigerant: water Te=-5 ℃, Tc=40 ℃ ηis=0.7 Refrigerant: water Te=-10 ℃, Tc=40 ℃ ηis=0.7 Refrigerant: water Te=-15 ℃, Tc=40 ℃ ηis=0.7


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Table 2. Parameters of absorption refrigeration and compression refrigeration

Waste heat temperature/ ℃

Absorption refrigeration Compression system refrigeration system Cooling output/kW

Mechanical required/kW

100

350.00

70.41

110

407.97

82.38

120

449.11

90.69

130

480.13

96.95

140

504.48

101.87

150

524.12

105.65

160

540.65

108.74

170

553.82

111.62

180

565.85

114.04

work

Figure 10. Mechanical work input of compression refrigeration and net power output of Kalina cycle


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Results show that when the heat source temperature ranges from 100 to 175 ℃, it is proper to choose the LiBr-H2O absorption refrigeration cycle; when temperature range is 175 to 200 ℃, Kalina cycle is preferable. As our work is mainly about thermodynamic performance comparison of power and absorption refrigeration cycles, we do not consider the investment recovery period. In future work, the economic analysis will be considered for further improvement.

ACKNOWLEDGMENT The financial support from the National Natural Science Foundation of China (21576286) and Science Foundation of China University of Petroleum ，Beijing (No. 2462017BJB03 and 2462018BJC004) are gratefully acknowledged.


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APPENDIX. Table A1. Simulation model of each component of the Kalina cycle (a) Component

Block

Calculated parameters

Pump

Pump

Power requirement, temperature, enthalpy and entropy of State 9 are calculated.

VLV

Valve

Temperature and entropy of State 11 are calculated.

Heat Exchanger

Heat duty of Preheater, temperature, enthalpy and entropy of State 2 and H3 are calculated.

Evaporator

Heat Exchanger

Heat duty of Evaporator, temperature, enthalpy and entropy of State 3 and H2 are calculated.

Recuperator

Heat Exchanger

Heat duty of Recuperator, temperature, enthalpy and entropy of State 10 and 1 are calculated.

Condenser

Cooler

Heat duty of Condenser, temperature, enthalpy and entropy of State 8 are calculated.

Separator

Separator

Mass flow, enthalpy and entropy of State 4 and 5 are calculated.

Turbine

Expander

Power output of Turbine, temperature, pressure, enthalpy and entropy of State 6 are calculated.

Mixer

Mixer

Mass flow, enthalpy and entropy of State 7 are calculated.

RCY

Recycle

Temperature, enthalpy and entropy of State 11-1 are calculated.

Preheater


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Table A2. Simulation model of each component of the Kalina cycle (b) Component

Block


Pump 1 and 2

Pump

Power requirement, temperature, enthalpy and entropy of State 8 and 12 are calculated.

VLV

Valve


Preheater 1 and Heat Exchanger 2

Heat duty of Preheater 1 and 2, temperature, enthalpy and entropy of State 2, H3, 16 and H4 are calculated.

Evaporator

Heat Exchanger

Heat duty of Evaporator, temperature, enthalpy and entropy of State 3 and H2 are calculated.

Recuperator 1, Heat Exchanger 2 and 3

Heat duty of Recuperator, temperature, enthalpy and entropy of State 10 and 1 are calculated.

Condenser and 2

Cooler

Heat duty of Condenser, temperature, enthalpy and entropy of State 7 and 11 are calculated.

Separator


Turbine

Expander

Power output of Turbine, temperature, pressure, enthalpy and entropy of State 4 are calculated.

Mixer 1 and 2

Mixer


Splitter

Tee


RCY

Recycle

Temperature, enthalpy and entropy of State 21-1 are calculated.

Separator

1


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Table A3. Simulation model of each component of the absorption refrigeration cycle Component

Block


Pump

Pump

Power requirement, temperature, enthalpy and entropy of State 2 are calculated.

SOLV

Valve


REFV

Valve


Solution Heat Heater+ Heater Exchanger

Temperature of State 6, heat duty of Solution Heat Exchanger, enthalpy and entropy of State 3 are calculated.

Condenser

Heater +Heater

Heat duty of Condenser, temperature, enthalpy and entropy of State 9 and CW3 are calculated.

Evaporator

Heater +Heater

Heat duty of Evaporator, temperature, enthalpy and entropy of State 11 and Coolout are calculated.

Absorber

Mass flow, enthalpy and entropy of Mixer+ (Heater+ State 12 are calculated. Heat duty of Absorber, temperature, enthalpy and Heater) entropy of CW2 are calculated.

Generator

HeatX+ Flash2

Heat duty of Generator, temperature, pressure, enthalpy and entropy of State 5 and 8 are calculated.


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Table A4. Simulation conditions for Kalina cycle (a)

Mass Heat source flow/ temperature/℃ kg·s−1

Mass fraction (Ammonia water)

Turbine inlet Turbine outlet Separation pressure/kPa pressure/kPa temperature/℃

100

0.8539

0.90

1700

910

90

110

0.7682

0.90

1900

910

100

120

0.7015

0.90

2100

910

110

130

0.6465

0.90

2300

910

120

140

0.5980

0.90

2500

910

130

150

0.5491

0.90

2900

910

140

Table A5. Simulation conditions for Kalina cycle (b)

Mass Heat source flow/ temperature/℃ kg·s−1

Mass fraction (Ammonia water)

Turbine inlet Turbine outlet Separation Splitter pressure/kPa pressure/kPa temperature/℃ ratio

140

0.4760

0.90

2000

700

70

0.2

150

0.4760

0.90

2700

700

70

0.2

160

0.5150

0.90

3500

770

69

0.3

170

0.4805

0.90

4500

740

65

0.24

180

0.4300

0.90

6500

650

80

0.15

190

0.4395

0.90

7000

670

60

0.16

200

0.4376

0.90

8000

670

60

0.16


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Table A6. Simulation conditions for the absorption refrigeration cycle

Mass flow/ kg·s−1

Mass fraction (LiBr)

Effectiveness of solution High side Low side Evaporation Condensation heat pressure/Pa pressure/Pa temperature/℃ temperature/℃ exchanger

5.4

0.58

0.64

7381.39

872

5

40

Table A7. Simulation conditions for the vapor compression refrigeration cycle Mass Heat source flow/ temperature/℃ kg·s−1

Compressor inlet pressure/Pa

Compressor outlet pressure/Pa

Evaporation Condensation temperature/℃ temperature/℃

100

0.1495

872

8000

5

40

110

0.1742

872

8000

5

40

120

0.1918

872

8000

5

40

130

0.2050

872

8000

5

40

140

0.2154

872

8000

5

40

150

0.2239

872

8000

5

40

160

0.2310

872

8000

5

40

170

0.2365

872

8000

5

40

180

0.2415

872

8000

5

40

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For Table of Contents Use Only TABLE OF CONTENTS (TOC) GRAPHIC

SYNOPSIS A novel indicator is proposed to compare the energy performance between power and absorption refrigeration cycles considering different waste heat temperature improving the sustainability of energy systems.


37

Energy Performance Comparison between Power and Absorption

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