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Thermodynamic Modelling and Parametric Study of a Two Stage Compression-Absorption Refrigeration System for Ice Cream Hardening Plant
Available online at www.sciencedirect.com
ScienceDirect Energy Procedia 109 (2017) 190 – 202
International Conference on Recent Advancement in Air Conditioning and Refrigeration, RAAR 2016, 10-12 November 2016, Bhubaneswar, India
Thermodynamic Modelling and Parametric Study of a Two Stage Compression-Absorption Refrigeration System for Ice Cream Hardening Plant Bhavesh Patela,*, Surendra Singh Kachhwahaa, Bhaumik Modia a
Pandit Deendayal Petroleum University, Raisan, Gandhinagar 382007, Gujarat, India
Abstract The present communication introduces a new concept of Two Stage Vapor Compression-Absorption Cascade Refrigeration System (TSVCACRS) for achieving low temperature industrial cooling. The system comprises of Two Stage Vapor Compression Refrigeration System (TSVCRS) having flash intercooler integrated with single stage Vapor Absorption Refrigeration System (VARS); thermally coupled by means of cascade condenser heat exchanger. The cascade condenser heat exchanger works as an evaporator for VARS and the condenser for TSVCRS. The proposed system has been designed and simulated for Havmor Ice cream Limited located at GIDC, Naroda, Ahmedabad; to check the thermodynamic performance feasibility with their existing installed TSVCRS based ice cream hardening refrigeration plant of 525 TR (1850 kW). Ammonia and LiBr-H2O have been considered as working fluid pair in proposed TSVCACRS. The results show that proposed TSVCACRS system would minimize the compressor work up to 28%, compared to an existing installed TSVCRS. The exergetic efficiency of the VAR, VCR subsystems and integrated TSVCACRS is found to be 32.78%, 60.29% and 53.59%, respectively. Moreover, the optimum generator temperature for the proposed system is found to be 85°C from the energetic and exergetic point of view. 2017The TheAuthors. Authors. Published by Elsevier Ltd. is an open access article under the CC BY-NC-ND license ©2017 © Published by Elsevier Ltd. This Peer-review under responsibility of the organizing committee of RAAR 2016. (http://creativecommons.org/licenses/by-nc-nd/4.0/). Peer-review under responsibility of the organizing committee of RAAR 2016. Keywords: Ice cream Plant, Energy, Exergy, Irreversibility, Two stage vapor compression refrigeration system (TSVCRS), Two stage vapor compression absorption refrigeration system (TSVCACRS).
_________________ * Corresponding author. Tel.: +91-98-24-339868; E-mail address: [email protected] (B. Patel)
1876-6102 © 2017 The Authors. Published by Elsevier Ltd. This is an open access article under the CC BY-NC-ND license (http://creativecommons.org/licenses/by-nc-nd/4.0/). Peer-review under responsibility of the organizing committee of RAAR 2016. doi:10.1016/j.egypro.2017.03.091
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191
Nomenclature ܿ e EES h ܫሶ P r T
specific heat at constant pressure specific Exergy engineering Equation Solver enthalpy Irreversibility pressure relative irreversibility Temperature
COP ܧሶ EV HP ݉ሶ ܳሶ ݏሶ TSVCACRS
TSVCRS
Two Stage Vapor Compression Refrigeration System
ܹሶ
coefficient of performance exergy flow rate Expansion Valve High pressure mass flow rate Heat Entropy Two Stage Vapor Compression Absorption Cascade Refrigeration System Work
θ
carnot factor
1, 2… c comp e g gen k m prv s T
state points condenser compressor evaporator generator generation component mechanical pressure reducing valve isentropic total
Greek Symbols Ƞ
efficiency
Superscript/ Subscript 0 a cc D EV HPC i LPC P R SHE
1.
environmental state Absorber cascade condenser Desorber expansion valve high pressure compressor state point low pressure compressor Pump rational solution heat exchanger
Introduction
Research by the Netherlands Environmental Assessment Agency has predicts that by around 2060, the amount of energy consumed worldwide in cooling will overtake that consumed in heating. According to the forecast, the world’s populations will reach 9 billion by 2050 in which the role of cold in food security will be pivotal. According to the UN Food and Agriculture Organization, global food demand is set to grow by 50% in that time. The cooling and refrigeration sectors are one of the major consumers of energy demand in India. India is the world’s leading producer of milk and the second-largest producer of fruits and vegetables, but the country watches nearly 20 percent of that yield, $10 billion worth of food go to waste for lack of a food-supply infrastructure that can keep its food fresh from farm to table. India faces a more daunting challenge with its cold chain than other parts of the world. Most of the country experiences extreme weather, so cooling technologies are more energy-intensive than that of in milder climates. Peak power outages are also routine, often forcing cold rooms to operate off costly backup generators. Further, it is interesting to note that the India today faces a greater demand for non-electricity applications, i.e., where the primary energy source could have been applied in a more efficient way than use of electricity-driven applications. All of this is to say that energy efficiency is paramount. To address that issue if we just replicate the old technologies, we’re heading for environmental disaster. Moreover, with increasing demand, the size of refrigeration and air conditioning unit is also increasing. Unfortunately, while the demand for more energy continues to grow, its scarcity increases and consumption of fossil fuel increases as well. In order to meet the future worldwide energy needs and to slow down the pace of global warming, the improvement of energy efficiency and the creation of sustainable energy sources have to be addressed simultaneously. Therefore the present study analyses the two stage vapor compression absorption system (TSVCACRS) which significantly reduce energy (electricity) consumption and dependency on high grade energy for low temperature industrial cooling applications, food preservation and storage. There is great potential for reducing high grade energy consumption by combining the Vapor Absorption Refrigeration System (VARS) and Vapor Compression Refrigeration System (VCRS) in series which is known as Vapor Compression Absorption Cascaded Refrigeration System (VCACRS). The application of cascaded
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refrigeration system maintains the advantages of both vapor compression and vapor absorption systems while minimizing the limitations of both simultaneously [1]. According to Jain et al. [1], the electrical energy consumption in vapor compression-absorption cascade refrigeration system is reduced by 61% and COP of compression section is improved by 155% compared to conventional VCRS. Researchers [2-4] have found 31 to 51% of reduction in energy consumption of cascade refrigeration system compare to conventional VCRS for the same cooling capacity and also pointed out the simultaneous energy need of electricity and heat. Further, the exhaustive thermodynamic analysis of such a single stage compression-absorption cascade system is available in the literature but work remains left for two stage vapor compression-absorption system which can be used for achieving low temperature cooling up to (-60°C). In the present study, two stage vapor compression absorption cascade refrigeration system (TSVCACRS) consisting of VCR subsystem (Fig.2) having ammonia as working fluid integrated with VAR subsystem having H2O–LiBr as fluid pair is proposed as an alternative to installed ammonia based two stage reciprocating vapor compression refrigeration ice cream plant. 2.
System analysis
2.1 Cycle description
Fig. 1. Existing (TSVCRS) for Ice cream plant
Bhavesh Patel et al. / Energy Procedia 109 (2017) 190 – 202
Fig. 2. Proposed (TSVCACRS) for Ice-cream Plant The schematic diagram of installed two stage vapor compression system (TSVCRS) based ice cream plant at Havmor Ice cream Limited, situated at G.I.D.C, Naroda, Ahmedabad, Gujarat is depicted in Fig. 1. Ammonia is used as a refrigerant in the system. Fig. 2 shows the line diagram of proposed TSVCACRS in which TSVCRS is thermally coupled with single effect VARS by means of cascade condenser. The cascade condenser serves as evaporator of conventional VAR subsystem and condenser of two stage VCR subsystem. The condenser of the compression section rejects the thermal load (the heat absorbed by the evaporator plus the work supplied to the compressor) to absorption section through cascade condenser, which leads to reduction of the pressure ratio across the compressor due to lower condenser temperature. The proposed system simultaneously requires low grade energy (heat in generator of VAR subsystem) and high grade energy (electricity in HP and LP compressor of VCR subsystem). Consequently, the discharge pressure of the compressor is very high in TSVCRS (Fig. 1) which would decrease to a greater extent for
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TSVCACRS (shown in Fig. 2) due to intermittent temperature achieved in cascade condenser. Thus, the pressure ratio across the compressor is reduce. As a consequences of the reduced pressure ratio, the electricity consumption and size of the compressor gets reduce. 2.2 Assumptions Mathematical modeling has been done based on following assumptions [1]: x x x x
All the system components operate under steady state. All the pressure and heat losses/gains in different components and pipelines of the system are neglected. The process occurring in all the expansion valves are isenthalpic. At given temperature and pressure, water-LiBr solutions in the generator and the absorber are assumed to be in equilibrium. x Only physical exergy is considered for exergy analysis. Moreover, kinetic, potential, nuclear, magnetic and chemical energies are also neglected. x The ambient temperature (ܶ ) and pressure (ܲ ) are taken as 25°C and 0.1MPa respectively. 3.
Mathematical modeling and Validation
3.1 Energetic and exergetic analyses x Energetic analysis Energetic analysis of the system is based on the first law of thermodynamic. The following set of fundamental governing equations (Mass balance, Energy balance and Entropy generation) is used for modelling individual components. (i) Mass balance,
σ ݉ሶ ൌ Ͳ;
(1)
(ii) Material Balance,
σ ݉ݔሶ ൌ Ͳ
(2)
(iii) Energy balance,
σ ܳሶ σ ܹሶ σ ݉ሶ ݄ െ σ ݉ሶ௨௧ ݄௨௧ ൌ Ͳ ;
(3)
x Exergetic Analysis Exergetic analysis is based on the second law of thermodynamic. Second law helps the designers to identify a particular component causing significant irreversible losses due to entropy generation in the system. ሶ
ொ ሶ ܵ ൌ σ ݉ሶ ݏ௨௧ െ σ ݉ሶ ݏ െ σ ቀ ቁ Ͳ
(4)
(ii) Specific exergy at point,
݁ ൌ ሺ݄ െ ݄ ሻ െ ܶ ሺݏ െ ݏ ሻ
(5)
(iii) Exergy flow rate at point,
ܧሶ ൌ ݉ ሾሺ݄ െ ݄ ሻ െ ܶ ሺݏ െ ݏ ሻሿ
(6)
(i) Entropy generation,
்
3.2 Modeling of integrated two stage vapor compression-absorption cascade refrigeration system Governing equations used for each component of TSVCACRS system are formulated by applying the fundamental equations (1) to (6). The modeled thermodynamic equations for individual components are presented in Appendix A. The state points mentioned in Appendix A correspond to Fig. 2. A computer program has been written and simulated
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in Engineering Equation Solver (EES) for thermodynamic model of proposed system. Moreover, thermophysical properties of various states of fluids are taken from the built in functions of EES. x Performance parameters (i) As per the entropy principle, all the real processes are irreversible and generate entropy. The associated irreversibility with the process can be determined by means of Gouy-Stodola law [5], ሶ ܫሶ ൌ ܶ ܵǡ
(7)
where, is real environmental temperature, ሶୣ୬ is entropy generation rate and k is the component of system. (ii) The total irreversibility rate of the subsystems (TSVCRS and VARS) and integrated TSVCACRS is given by, ்ܫሶ ൌ σ ܫሶ
(8)
where, k is the components of the individual sub system or trigeneration system. (iii) The energetic efficiency (COP) of various systems are given by, ்ܱܲܥௌோௌ ൌ ܱܲܥோௌ ൌ
ொሶ ௐሶ
(9)
ொሶ ொሶ ାௐሶು
்ܱܲܥௌோௌ ൌ
(10) ொሶ
(11)
ሺ ା୕ሶౝା౦ ሻ
(iv) The rational / exergetic efficiency (ߟோ ) of various systems are given by, ூ
ቁ ߟோǡ்ௌோௌ ൌ ͳ െ ቀ ೄೇೃೄ ሶ
(12)
ௐ
ߟோǡோௌ ൌ ͳ െ ൬
ூೇಲೃೄ
ொ כఏǡೌೝ
(13)
൰
where, carnot factor ሺߠǡ௧ ሻ ൌ ͳ െ ൬
்బ
்ǡೞೠೝ
ߟோǡ்ௌோௌ ൌ ͳ െ ൬
ூೄೇಲೃೄ ൰ ௐሶ ାௐሶ ାொ כఏǡೌೝ
൰ and ܶǡௌ௨ ൌ
்మభ ି்మమ ୪୬ቀ మభ ቁ మమ
(14)
(v)The relative irreversibility in each component of the cycle is a convenient non dimensional form for examining the relative contributions of different plant components to total cycle irreversibility defined by the relative exergy destruction ratio: ݎ ൌ
ூೖ ூೄೇಲೃೄ
(15)
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3.3 Model Validation Individual models are developed for TSVCRS and single stage VCRS using the built-in thermodynamic properties and functions of EES. The thermodynamic model equations are highly nonlinear in nature and are solved using Engineering Equation Solver (EES). Individual validation for both the model was carried out and comparison of performance data obtained from the current model and the reference model are presented in this section. (i) Validation of two stage vapor compression refrigeration system (TSVCRS) The model of TSVCRS is validated with the theoretical results obtained by Nikolaidis and Probert [6] with R22 as the refrigerant for the following set of input variables: T e = -35°C, Tcond = 25°C, Ten = 20°C, Qe = 100 kW, and ηisen = 0.73. The comparison of values of the present and reference model is summarized in Table 1. Table 1. Validation of present model of TSVCRS with Nikolaidis et al. [6] Sr. No.
Components
Parameters
Nikolaidis Model [6]
Present Model
Difference (%)
1
LPC
WLPC
23.23
23.2
0.13
2
Evaporator
Qe
100
100
0.00
3
HPC
WHPC
33.04
33.05
-0.03
4
Condenser
QC
140.5
140.5
0.00
System
்ܫሶ
35.7
35.9
-0.56
COP
1.78
1.778
0.11
WT
56.27
56.25
0.04
5
Table 1 shows the comparison of first law and second law performance parameters of proposed system and reference system. It shows that all the calculated and operating parameters are predicted within the error range of ± 0.60 %. (ii) Validation of single stage vapor compression-absorption cascade refrigeration system (VCACRS) In order to validate the present model of VCACRS, the results have been compared with the numerical results presented by Jain et. al. [1] for the input parameters: Qe = 66.67 kW, Te = -4.1oC, Tc = 43.6oC, Ta = 40oC, Tg = 90oC and Tcc = 18oC. Table 2. Validation of present model of VCACRS with Jain et al. [1] Sr. No.
Components
Parameters
Jain Model [1]
Present Model
Difference (%)
1
Compressor
Wcomp
9.414
9.329
0.90
2
Evaporator
Qe
66.67
66.67
0.00
3
Cascade Condenser
QCC
76.08
76
0.11
4
Absorber
Qa
100.1
99.95
0.15
5
Generator
Qg
104.9
104.8
0.10
6
Pump
Wp
0.0019
0.00194
-2.11
7
Condenser
Qc
80.96
80.85
0.14
8
VCRS
COP
7.082
7.146
-0.90
9
VARS
COP
0.725
0.7252
-0.03
10
System
COP
0.583
0.5841
-0.19
IT
22.81
22.59
0.96
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Table 2 shows the good coherence between the results of the current thermodynamic model and the model described by Jain et al. [1]. It can be seen that the maximum error in irreversibility calculations for different components is found to be within ± 2.15%. 3.4 Design conditions and operating parameters The subsystems of the proposed thermal system are TSVCRS and VARS. The main objective of integrating TSVCRS with VARS is to achieve low temperature cooling (around -45 to -55°C) for food preservation applications and industrial cooling applications. The complete integrated cascaded system (TSVCACRS) depicted in Fig. 2 is modeled in EES by applying the mass, energy and exergy balance equations discussed earlier. The input design data are given in Table 3. The overlap temperature (ܶ െ ܶଽ ) across the cascaded condenser was taken to be 5°C. In the present study, Ammonia is considered as a refrigerant for two stage VCR subsystem and H2O-LiBr is considered in VAR subsystem of integrated TSVCACRS. In industrial refrigeration, Ammonia is widely used as refrigerant due to its easy availability and lower cost though it is highly toxic in nature. The proposed system can achieve low temperature around (-50 to -60°C) in the evaporator of two stage VCRS due to lower condensing temperature in cascade condenser. Table 3. Input Parameters for proposed TSVCACRS Parameter
Value
Condenser coolant inlet temperature (ܶଵଽ in ºC) [7] Condenser coolant outlet temperature (ܶଶ in ºC)
28
[7]
30
Generator coolant inlet temperature (ܶଶଵ in ºC)[1]
100
Generator coolant outlet temperature (ܶଶଶ in ºC) [1]
95
Generator temperature (ܶଵ in ºC)
[1]
Absorber coolant inlet temperature (ܶଶଷ in ºC)
90 [1]
Absorber coolant outlet temperature (ܶଶସ in ºC)
35
[1]
38
Absorber Temperature (ܶଵ in ºC) [1]
40
Rate of heat absorbed by evaporator (ܳሶ in kW) [7]
1850
Evaporator temperature (ܶସᇱ in ºC) [7] Evaporation temperature in cascade condenser (ܶଽ in ºC)
10
Effectiveness of solution heat exchanger (ߝ௦௫ ) [1]
0.6
Isentropic efficiency of compressor (ߟ௦ ) [2]
0.85
Electrical efficiency of pump (ߟ )
4.
-45 [1]
[7]
0.9
Degree of overlap in cascade condenser (ܶ௩ in ºC) [5]
5
Environment temperature (ܶ in ºC)
25
Atmospheric pressure (ܲ in kPa)
101.3
Results and analysis
In this section, the results of computer simulation carried out for the proposed two stage vapor compressionabsorption cascade refrigeration system are presented and discussed. The comparison of performance parameters of TSVCRS (Fig.1) and TSVCACRS (Fig.2) based on the energetic and exergetic analysis of the process (heat and work, COP of various system, irreversibility and second law analysis) are calculated and tabulated in Table 4. For the same cooling capacity of 525 TR, in comparison to standalone TSVCRS, the combined LPC and HPC compressors of TSVCACRS requires 372 kW less work due to integration of VARS which minimizes the pressure ratio across the evaporator and condenser as mentioned in Table 4. Moreover, the compressors capacities and sizes also reduce due to lower discharge temperature of HPC compressor maintained in cascade condenser by integration
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of VARS. The pressure ratio of TSVCACRS is reduced to 13.38 from which is to be 24.8 for standalone TSVCRS. The COP of compression section of TSVCACRS is obtained 2.26 which is 1.61 for standalone TSVCRS. The overall COP of the proposed system is found to be 0.45 which is low value due to integration of VARS since it works on the low grade heat. The COP of the VARS in proposed system is found to be 0.76. From the point of view of second law (the irreversibility and rational efficiency as mentioned in Table 4), the standalone TSVCRS is far better in comparison to TSVCACRS since it utilizes the high grade energy (electricity). Moreover, the exergetic efficiency of the TSVCACRS is found to be 53.59%. The property plot for both systems are shown in Fig. 3 and Fig. 4 which clearly shows that the integration of VARS reduces the condensing temperature (35°C) as shown in Fig. 3 to intermediate temperature (15°C) which leads to less compressor work to achieve same cooling capacity at -45°C. Table 4. Results of energetic performance parameters Sr. No. 1
Performance parameters ܳሶௗ (kW)
Low grade energy
ܳሶ (kW) ܳሶ (kW) ܳሶ௩ (kW) ܳሶ௦ௗ (kW) 2
High grade energy
TSVCACRS
(Fig. 1)
(Fig. 2)
2914.0
2671.0
-
3110.0
-
3269.0
1850.0
1850.0
-
2512.0
ܳሶௌு (kW) ܹሶ (kW)
454.7
345.4
ܹሶு (kW)
690
472.3
315.6
ܹሶ (kW)
-
0.03
3
Operating parameter
Pressure ratio
24.8
13.38
4
First law parameters
்ܱܲܥௌோௌ
1.61
2.26
ܱܲܥோௌ
-
0.76
்ܱܲܥௌோௌ
-
0.45
I (kW)
420.5
675.8
િ܀
63.26
53.59
5
Second law parameters
Ammonia
Ammonia kg -K 5. 20 9
TSVCRS 104 119.4°C
TSVCACRS
1 78 5.
4 43 5.
104
9 33 6.
P [kPa]
74.93°C
103
105
kJ /
105
P [kPa]
TSVCRS
35°C
15°C -18.94°C
102
0
500
9 33 6.
49.03°C
103
-45°C
-45°C
101 -500
-K
8 88 5.
82.15°C
-11.7°C
102
g /k kJ
1000
1500
h [kJ/kg]
Fig.3. Property chart for TSVCRS
2000
2500
101 -500
0
500
1000
1500
h [kJ/kg]
Fig.4. Property chart for TSVCACRS
2000
2500
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The exergetic performance parameters of the proposed TSVCACRS are calculated and presented in Table 5. The irreversibility and exergetic efficiency of the proposed system and sub system along with the individual components are tabulated and mentioned is Table 5. The generator, absorber, evaporator, cascade condenser, condenser contributes the 20.69%, 19.44%, 15.03%,11.54% and 8.21% irreversibility relative to system irreversibility, respectively. The maximum exergetic efficiency of flash cooler is obtained as 92.02%. The exergetic efficiency of the generator, evaporator, HPC compressor, LPC compressor, solution heat exchanger, absorber, condenser and cascade condenser is found to be 78.15%, 73.77%, 70.97%, 69.77%, 56.05%, 46.77%, 41.69% and 41.42% respectively. The exergetic efficiency of the VAR and VCR sub systems are obtained 32.78% and 60.29%, respectively. The exergetic efficiency of the proposed TSVCACRS is found to be 53.59%. The highest exergetic efficiency is obtained for VCR subsystem as it utilizes the high grade energy (electricity) which produces the maximum useful work. The value obtained in case of VAR subsystem is lower as it utilises the low grade heat. The value of exergetic efficiency of the system lies between the values of VCR and VAR subsystem as it utilises both high grade and low grade energy. Table 5. Results of exergetic performance parameters of TSVCACRS (Fig. 2)
Second law Analysis Component/ System
۷ሶ
ܓܚ
િ܀
[kW]
[%]
[%]
Condenser
55.50
8.21
41.69
Absorber
131.4
19.44
46.77
Generator
139.8
20.69
78.15
Evaporator
101.6
15.03
73.77
Cascade condenser
77.96
11.54
41.42
Low pressure compressor
38.78
5.74
69.77
High pressure compressor Pump
47.35 0.00
7.00 0.00
70.97 0.00
Pressure reducing valve
0.00
0.00
0.00
Solution heat exchanger
19.56
2.89
56.05
Expansion valve 1
12.05
1.78
0.00
Expansion valve 2
17.59
2.60
0.00
Expansion valve 3
4.87
0.72
0.00
Flash cooler
29.36
4.34
92.02
VAR subsystem
429.1
-
32.78
VCR subsystem
324.7
-
60.29
TSVCACRS
675.82
-
53.59
4.1 Influence of generator temperature According to Fig. 5, with the hike in the generator temperature, the thermal capacity needed for the generator reduces for the same cooling capacity. Further, the coefficients of performance of the absorption section COP VARS and two stage vapor compression absorption cascade refrigeration system COP TSVCACRS increases up to 90°C. It can be also observed that the exergetic efficiency of the TSVCACRS also improves significantly up to 90°C and then it decreases marginally. The irreversibility of the system and the thermal capacity requirement in generator decreases as the generator temperature increases up to 90°C and then it starts increasing. Thus, the optimum temperature for the generator lies between 85-90°C.
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4.2 Influence of condenser and absorber temperature In this section, the performance of the system is predicted when the absorber and condenser temperature is increased from 35°C to 45°C whilst the other input parameters and conditions of the system are kept constant. For the absorber saturation temperature, the COP and exergetic efficiency of the TSVCACRS decreases by 4% and 0.66%, respectively. Further, the change in condenser saturation from 35-45°C causes reduction in the system COP and exergetic efficiency by 6% and 1.2%, respectively. 0.46
54
54 0.46 Ta
0.44
COPTSVCACRS
R,TSVCACRS (%)
COPTSVCACRS
53
53.5
Tc
Ta
0.44
Tc
53
0.43
R,TSVCACRS (%)
0.45
53.5 0.45
Ta
52.5
52.5 0.42
COPTSVCACRS R,TSVCACRS 0.43 75
80
85
90
95
Tc
COPTSVCACRS R,TSVCACRS 0.41
52 100
52 35
40
45
Tc and Ta [°C]
Tg [°C]
Fig. 5. COP and exergetic efficiency Versus Effect of generator saturation temperature (Tg).
Fig. 6. COP and exergetic efficiency Versus Effect of absorber and condenser saturation temperature (Ta, Tc)
4.3 Influence of cascade overlap temperature The overlap temperature (T 7-T9) of the cascade condenser is an important design parameter as it influences the performance of both the subsystems. It is observed in Fig. 7 that the COP and exergetic efficiency of the TSVCACRS decreases approx. by 4% and 2%, respectively. 0.46
54
0.5
15
20
25
30
35
40
45
58
T0 [C] 0.48
52
0.46 50 0.44
0.43
COPTSVCACRS
R,TSVCACRS 0.42
51 5
6
7
46
0.42
COPTSVCACRS
R,TSVCACRS (%)
0.44
COPTSVCACRS
54 53
R,TSVCACRS
COPTSVCACRS
0.45
8
9
10
Toverlap [°C]
Fig. 7. COP and exergetic efficiency Versus Effect of cascade condenser overlap temperature
0.4 650
R,TSVCACRS 670
690
710
42 730
Itotal [kW]
Fig. 8. COP and exergetic efficiency Versus Effect of ambient temperature
Bhavesh Patel et al. / Energy Procedia 109 (2017) 190 – 202
4.4 Influence of ambient temperature The effect of the ambient temperature on the system performance is investigated over the range of 15-45°C and presented in Fig. 8. The COP of proposed system remains unaffected with change in ambient temperature as the COP of individual systems remains unchanged. The exergetic efficiency of VCR subsystem of TSVCACRS reduces by 4%. Additionally the exergetic efficiency of the VAR subsystem decreases drastically by 41% as the difference between the ambient and the generator temperature decreases which leads to less useful work output from the supplied heat. Thus, according to Fig. 8 the exergetic efficiency of the proposed TSVCACRS is reduced by 14% as the irreversibility of the system increases 10% for the change of ambient temperature from 15°C to 45°C. 5.
Conclusion
In this communication, the thermodynamic performance and feasibility study of the two stage vapor compressionabsorption refrigeration system (TSVCACRS) has been carried out. The proposed system has been modeled and simulated for the existing installed 525 TR (1850 kW) two stage vapor compression refrigeration system (TSVCRS) for ice cream hardening and storage unit for -50 to -35°C evaporator temperature at Havmor ice cream limited. The comparison study of the TSVCRS and proposed TSVCACRS has been carried out. The main results obtained from the thermodynamic simulation and parametric study are concluded below: x The results show that proposed TSVCACRS system would minimize the compressor work up to 28%, compared to an existing installed TSVCRS. x The pressure ratio of compression section for proposed TSVCACRS is predicted to be 13.8 due to intermediate temperature achieved in cascade condenser, which is 24.8 in case of standalone TSVCRS. Thus the capacity of the HPE and LPE compressors would minimize. x The overall COP of the proposed system is obtained 0.45 which is low value in comparison to TSVCRS having COP of 1.61 since the proposed system utilizes the low grade energy in absorption section. x The exergetic efficiency of the proposed TSVCACRS is found to be 10% less compared to existing installed TSVCRS since 255 kW more irreversibility is associated to TSVCACRS. x The exergetic efficiency of the VAR, VCR subsystems and integrated TSVCACRS is found to be 32.78%, 60.29% and 53.59%, respectively. The lower value of exergetic efficiency in case of VAR subsystem is observed due to usage of low grade heat as a source of energy. x The highest irreversibility occurs in generator with 20% followed by absorber with 19.44%, evaporator 15.03%, cascade condenser with 11.54% and condenser with 8.21%. x The optimum generator temperature is found to be 85°C as COP and exergetic efficiencies are increase up to that temperature and then it remains constant up to 90°C and afterwards it starts decreasing. x The increase in condenser and absorber saturation temperature decreases the COP and exergetic efficiencies of the system. Moreover, in the case of condenser the rate of decrease of COP and exergetic efficiency is higher as temperature goes on increases. x The overlap temperature of the cascade condenser affects the system performance at great extent so it is considered as a critical design parameter for the system. The best performance of the system is obtained at 5°C overlap temperature. x The COP of the system remains unaffected as the ambient temperature increases but the exergetic efficiency decreases 14% for the temperature range of 15 to 45°C. Acknowledgements The authors also thank Havmor Ice cream Limited located at GIDC, Naroda, Ahmedabad; for their support and allowing us to their premises to get the realistic data.
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Appendix A. An example appendix Table A.1 Thermodynamic equations for proposed TSVCACRS Sr. No
Components
First law Equations
Second Law Equations
1
Evaporator
Qe = m4 * (h1 – h2)
Ie = T0* ((S1 - S4) - Qe/Tcs)
LP Compressor
WLPC = m1 * (h2-h1) / (ηm * ηe) h2 = h1 + (h2isen – h1) / ηisen pFlash = (pH /pL)0.5 mLT (h3 – h2) = mHT (h8 – h5) h3 = h4 WHPC = mHT * (h6-h5) / (ηm * ηe) h6 = h5 + (h6isen – h5) / ηisen mw * (h9 - h18) = mHT * (h6 - h7) h7 = h8 m9 + m15 = m10 Qa = mw*h9 + m15*h15 – m10*h10 Qa = ma * (h23 - h24) Wp = mss *(P16 - P9)/( ρss * ηp) Wp = m11*h11 – m10*h10 m13*h13 -m14*h14 = m12*h12 – m11*h11 ε = (T13 -T14)/(T13-T11) m12 = m13 + mw m16*h16 + m13*h13 = Qg + m12*h12 Qg = mg* (h21 – h22) h14 = h15
2 3 4 5 6 7
Flash Cooler Ex. V/v 1 HP Compressor Cascade Cond. Ex V/v 2 Absorber
8 9 10
Pump SHE Generator
11
13
PRV Condenser
14
Ex V/v 3
12
ILPC = T0* (S2 - S1) IFlash = T0* ((S3 - S2) + (S5 - S8)) Iev1 = T0* (S4 - S3) IHPC = T0* (S6 - S5) ICC = T0* ((S7 - S6) + (S9 - S18)) Iev2 = T0* (S8 - S7) Ia = T0* (S10 - S9 - S15 + Qa/Ta) Ip = T0* (S11 - S10) ISHE = T0* (S12 - S11) + (S14 -S13) Ig = T0* (S16 + S13 - mss * S12 ) - (Qg/Tg) Iprv = T0* (S15 - S14)
Qc = mw* (h16 – h17) = mc* (h20-h19)
Ic = T0* (S17 - S16) + (S20 - S19)
h17 = h18
Iev3 = T0* (S18 - S17)
References [1] Vaibhav Jain, S.S. Kachhwaha, Gulshan Sachdeva. Thermodynamic performance analysis of a vapor compression-absorption cascaded refrigeration system. Energy Conversion and Management 2013; 75: 685–700. [2] Cimsit C, and Ozturk I. Analysis of compression-absorption cascade refrigeration cycles. Appl Therm Eng 2012; 40:311–7. [3] Wang L, Ma A, Tan Y, Cui X, and Cui H. Study on solar assisted cascade refrigeration system. Energy Procedia 2012; 16(C):1503–9. [4] Garimella S, Brown AM, and Nagavarapu AK. Waste heat driven absorption/vapor-compression cascade refrigeration system for megawatt scale, high-flux, low-temperature cooling. Int J Refrig 2011; 34:1776–85. [5] Colorado D, Velazquez VM. Exergy analysis of a compression–absorption cascade system for refrigeration. Int J Energy Res 2013; 37(14):1851-65. [6] Nikolaidis C and Probert D. Exergy-method analysis of two-stage vapor compression refrigeration-plants performance, Appl Energy. 1998; 60:241–56. [7] Havmor Ice cream Limited. GIDC, Naroda, Ahmedabad. [Visited on 12.01.2016].
ScienceDirect Energy Procedia 109 (2017) 190 – 202
International Conference on Recent Advancement in Air Conditioning and Refrigeration, RAAR 2016, 10-12 November 2016, Bhubaneswar, India
Thermodynamic Modelling and Parametric Study of a Two Stage Compression-Absorption Refrigeration System for Ice Cream Hardening Plant Bhavesh Patela,*, Surendra Singh Kachhwahaa, Bhaumik Modia a
Pandit Deendayal Petroleum University, Raisan, Gandhinagar 382007, Gujarat, India
Abstract The present communication introduces a new concept of Two Stage Vapor Compression-Absorption Cascade Refrigeration System (TSVCACRS) for achieving low temperature industrial cooling. The system comprises of Two Stage Vapor Compression Refrigeration System (TSVCRS) having flash intercooler integrated with single stage Vapor Absorption Refrigeration System (VARS); thermally coupled by means of cascade condenser heat exchanger. The cascade condenser heat exchanger works as an evaporator for VARS and the condenser for TSVCRS. The proposed system has been designed and simulated for Havmor Ice cream Limited located at GIDC, Naroda, Ahmedabad; to check the thermodynamic performance feasibility with their existing installed TSVCRS based ice cream hardening refrigeration plant of 525 TR (1850 kW). Ammonia and LiBr-H2O have been considered as working fluid pair in proposed TSVCACRS. The results show that proposed TSVCACRS system would minimize the compressor work up to 28%, compared to an existing installed TSVCRS. The exergetic efficiency of the VAR, VCR subsystems and integrated TSVCACRS is found to be 32.78%, 60.29% and 53.59%, respectively. Moreover, the optimum generator temperature for the proposed system is found to be 85°C from the energetic and exergetic point of view. 2017The TheAuthors. Authors. Published by Elsevier Ltd. is an open access article under the CC BY-NC-ND license ©2017 © Published by Elsevier Ltd. This Peer-review under responsibility of the organizing committee of RAAR 2016. (http://creativecommons.org/licenses/by-nc-nd/4.0/). Peer-review under responsibility of the organizing committee of RAAR 2016. Keywords: Ice cream Plant, Energy, Exergy, Irreversibility, Two stage vapor compression refrigeration system (TSVCRS), Two stage vapor compression absorption refrigeration system (TSVCACRS).
_________________ * Corresponding author. Tel.: +91-98-24-339868; E-mail address: [email protected] (B. Patel)
1876-6102 © 2017 The Authors. Published by Elsevier Ltd. This is an open access article under the CC BY-NC-ND license (http://creativecommons.org/licenses/by-nc-nd/4.0/). Peer-review under responsibility of the organizing committee of RAAR 2016. doi:10.1016/j.egypro.2017.03.091
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191
Nomenclature ܿ e EES h ܫሶ P r T
specific heat at constant pressure specific Exergy engineering Equation Solver enthalpy Irreversibility pressure relative irreversibility Temperature
COP ܧሶ EV HP ݉ሶ ܳሶ ݏሶ TSVCACRS
TSVCRS
Two Stage Vapor Compression Refrigeration System
ܹሶ
coefficient of performance exergy flow rate Expansion Valve High pressure mass flow rate Heat Entropy Two Stage Vapor Compression Absorption Cascade Refrigeration System Work
θ
carnot factor
1, 2… c comp e g gen k m prv s T
state points condenser compressor evaporator generator generation component mechanical pressure reducing valve isentropic total
Greek Symbols Ƞ
efficiency
Superscript/ Subscript 0 a cc D EV HPC i LPC P R SHE
1.
environmental state Absorber cascade condenser Desorber expansion valve high pressure compressor state point low pressure compressor Pump rational solution heat exchanger
Introduction
Research by the Netherlands Environmental Assessment Agency has predicts that by around 2060, the amount of energy consumed worldwide in cooling will overtake that consumed in heating. According to the forecast, the world’s populations will reach 9 billion by 2050 in which the role of cold in food security will be pivotal. According to the UN Food and Agriculture Organization, global food demand is set to grow by 50% in that time. The cooling and refrigeration sectors are one of the major consumers of energy demand in India. India is the world’s leading producer of milk and the second-largest producer of fruits and vegetables, but the country watches nearly 20 percent of that yield, $10 billion worth of food go to waste for lack of a food-supply infrastructure that can keep its food fresh from farm to table. India faces a more daunting challenge with its cold chain than other parts of the world. Most of the country experiences extreme weather, so cooling technologies are more energy-intensive than that of in milder climates. Peak power outages are also routine, often forcing cold rooms to operate off costly backup generators. Further, it is interesting to note that the India today faces a greater demand for non-electricity applications, i.e., where the primary energy source could have been applied in a more efficient way than use of electricity-driven applications. All of this is to say that energy efficiency is paramount. To address that issue if we just replicate the old technologies, we’re heading for environmental disaster. Moreover, with increasing demand, the size of refrigeration and air conditioning unit is also increasing. Unfortunately, while the demand for more energy continues to grow, its scarcity increases and consumption of fossil fuel increases as well. In order to meet the future worldwide energy needs and to slow down the pace of global warming, the improvement of energy efficiency and the creation of sustainable energy sources have to be addressed simultaneously. Therefore the present study analyses the two stage vapor compression absorption system (TSVCACRS) which significantly reduce energy (electricity) consumption and dependency on high grade energy for low temperature industrial cooling applications, food preservation and storage. There is great potential for reducing high grade energy consumption by combining the Vapor Absorption Refrigeration System (VARS) and Vapor Compression Refrigeration System (VCRS) in series which is known as Vapor Compression Absorption Cascaded Refrigeration System (VCACRS). The application of cascaded
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refrigeration system maintains the advantages of both vapor compression and vapor absorption systems while minimizing the limitations of both simultaneously [1]. According to Jain et al. [1], the electrical energy consumption in vapor compression-absorption cascade refrigeration system is reduced by 61% and COP of compression section is improved by 155% compared to conventional VCRS. Researchers [2-4] have found 31 to 51% of reduction in energy consumption of cascade refrigeration system compare to conventional VCRS for the same cooling capacity and also pointed out the simultaneous energy need of electricity and heat. Further, the exhaustive thermodynamic analysis of such a single stage compression-absorption cascade system is available in the literature but work remains left for two stage vapor compression-absorption system which can be used for achieving low temperature cooling up to (-60°C). In the present study, two stage vapor compression absorption cascade refrigeration system (TSVCACRS) consisting of VCR subsystem (Fig.2) having ammonia as working fluid integrated with VAR subsystem having H2O–LiBr as fluid pair is proposed as an alternative to installed ammonia based two stage reciprocating vapor compression refrigeration ice cream plant. 2.
System analysis
2.1 Cycle description
Fig. 1. Existing (TSVCRS) for Ice cream plant
Bhavesh Patel et al. / Energy Procedia 109 (2017) 190 – 202
Fig. 2. Proposed (TSVCACRS) for Ice-cream Plant The schematic diagram of installed two stage vapor compression system (TSVCRS) based ice cream plant at Havmor Ice cream Limited, situated at G.I.D.C, Naroda, Ahmedabad, Gujarat is depicted in Fig. 1. Ammonia is used as a refrigerant in the system. Fig. 2 shows the line diagram of proposed TSVCACRS in which TSVCRS is thermally coupled with single effect VARS by means of cascade condenser. The cascade condenser serves as evaporator of conventional VAR subsystem and condenser of two stage VCR subsystem. The condenser of the compression section rejects the thermal load (the heat absorbed by the evaporator plus the work supplied to the compressor) to absorption section through cascade condenser, which leads to reduction of the pressure ratio across the compressor due to lower condenser temperature. The proposed system simultaneously requires low grade energy (heat in generator of VAR subsystem) and high grade energy (electricity in HP and LP compressor of VCR subsystem). Consequently, the discharge pressure of the compressor is very high in TSVCRS (Fig. 1) which would decrease to a greater extent for
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TSVCACRS (shown in Fig. 2) due to intermittent temperature achieved in cascade condenser. Thus, the pressure ratio across the compressor is reduce. As a consequences of the reduced pressure ratio, the electricity consumption and size of the compressor gets reduce. 2.2 Assumptions Mathematical modeling has been done based on following assumptions [1]: x x x x
All the system components operate under steady state. All the pressure and heat losses/gains in different components and pipelines of the system are neglected. The process occurring in all the expansion valves are isenthalpic. At given temperature and pressure, water-LiBr solutions in the generator and the absorber are assumed to be in equilibrium. x Only physical exergy is considered for exergy analysis. Moreover, kinetic, potential, nuclear, magnetic and chemical energies are also neglected. x The ambient temperature (ܶ ) and pressure (ܲ ) are taken as 25°C and 0.1MPa respectively. 3.
Mathematical modeling and Validation
3.1 Energetic and exergetic analyses x Energetic analysis Energetic analysis of the system is based on the first law of thermodynamic. The following set of fundamental governing equations (Mass balance, Energy balance and Entropy generation) is used for modelling individual components. (i) Mass balance,
σ ݉ሶ ൌ Ͳ;
(1)
(ii) Material Balance,
σ ݉ݔሶ ൌ Ͳ
(2)
(iii) Energy balance,
σ ܳሶ σ ܹሶ σ ݉ሶ ݄ െ σ ݉ሶ௨௧ ݄௨௧ ൌ Ͳ ;
(3)
x Exergetic Analysis Exergetic analysis is based on the second law of thermodynamic. Second law helps the designers to identify a particular component causing significant irreversible losses due to entropy generation in the system. ሶ
ொ ሶ ܵ ൌ σ ݉ሶ ݏ௨௧ െ σ ݉ሶ ݏ െ σ ቀ ቁ Ͳ
(4)
(ii) Specific exergy at point,
݁ ൌ ሺ݄ െ ݄ ሻ െ ܶ ሺݏ െ ݏ ሻ
(5)
(iii) Exergy flow rate at point,
ܧሶ ൌ ݉ ሾሺ݄ െ ݄ ሻ െ ܶ ሺݏ െ ݏ ሻሿ
(6)
(i) Entropy generation,
்
3.2 Modeling of integrated two stage vapor compression-absorption cascade refrigeration system Governing equations used for each component of TSVCACRS system are formulated by applying the fundamental equations (1) to (6). The modeled thermodynamic equations for individual components are presented in Appendix A. The state points mentioned in Appendix A correspond to Fig. 2. A computer program has been written and simulated
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in Engineering Equation Solver (EES) for thermodynamic model of proposed system. Moreover, thermophysical properties of various states of fluids are taken from the built in functions of EES. x Performance parameters (i) As per the entropy principle, all the real processes are irreversible and generate entropy. The associated irreversibility with the process can be determined by means of Gouy-Stodola law [5], ሶ ܫሶ ൌ ܶ ܵǡ
(7)
where, is real environmental temperature, ሶୣ୬ is entropy generation rate and k is the component of system. (ii) The total irreversibility rate of the subsystems (TSVCRS and VARS) and integrated TSVCACRS is given by, ்ܫሶ ൌ σ ܫሶ
(8)
where, k is the components of the individual sub system or trigeneration system. (iii) The energetic efficiency (COP) of various systems are given by, ்ܱܲܥௌோௌ ൌ ܱܲܥோௌ ൌ
ொሶ ௐሶ
(9)
ொሶ ொሶ ାௐሶು
்ܱܲܥௌோௌ ൌ
(10) ொሶ
(11)
ሺ ା୕ሶౝା౦ ሻ
(iv) The rational / exergetic efficiency (ߟோ ) of various systems are given by, ூ
ቁ ߟோǡ்ௌோௌ ൌ ͳ െ ቀ ೄೇೃೄ ሶ
(12)
ௐ
ߟோǡோௌ ൌ ͳ െ ൬
ூೇಲೃೄ
ொ כఏǡೌೝ
(13)
൰
where, carnot factor ሺߠǡ௧ ሻ ൌ ͳ െ ൬
்బ
்ǡೞೠೝ
ߟோǡ்ௌோௌ ൌ ͳ െ ൬
ூೄೇಲೃೄ ൰ ௐሶ ାௐሶ ାொ כఏǡೌೝ
൰ and ܶǡௌ௨ ൌ
்మభ ି்మమ ୪୬ቀ మభ ቁ మమ
(14)
(v)The relative irreversibility in each component of the cycle is a convenient non dimensional form for examining the relative contributions of different plant components to total cycle irreversibility defined by the relative exergy destruction ratio: ݎ ൌ
ூೖ ூೄೇಲೃೄ
(15)
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3.3 Model Validation Individual models are developed for TSVCRS and single stage VCRS using the built-in thermodynamic properties and functions of EES. The thermodynamic model equations are highly nonlinear in nature and are solved using Engineering Equation Solver (EES). Individual validation for both the model was carried out and comparison of performance data obtained from the current model and the reference model are presented in this section. (i) Validation of two stage vapor compression refrigeration system (TSVCRS) The model of TSVCRS is validated with the theoretical results obtained by Nikolaidis and Probert [6] with R22 as the refrigerant for the following set of input variables: T e = -35°C, Tcond = 25°C, Ten = 20°C, Qe = 100 kW, and ηisen = 0.73. The comparison of values of the present and reference model is summarized in Table 1. Table 1. Validation of present model of TSVCRS with Nikolaidis et al. [6] Sr. No.
Components
Parameters
Nikolaidis Model [6]
Present Model
Difference (%)
1
LPC
WLPC
23.23
23.2
0.13
2
Evaporator
Qe
100
100
0.00
3
HPC
WHPC
33.04
33.05
-0.03
4
Condenser
QC
140.5
140.5
0.00
System
்ܫሶ
35.7
35.9
-0.56
COP
1.78
1.778
0.11
WT
56.27
56.25
0.04
5
Table 1 shows the comparison of first law and second law performance parameters of proposed system and reference system. It shows that all the calculated and operating parameters are predicted within the error range of ± 0.60 %. (ii) Validation of single stage vapor compression-absorption cascade refrigeration system (VCACRS) In order to validate the present model of VCACRS, the results have been compared with the numerical results presented by Jain et. al. [1] for the input parameters: Qe = 66.67 kW, Te = -4.1oC, Tc = 43.6oC, Ta = 40oC, Tg = 90oC and Tcc = 18oC. Table 2. Validation of present model of VCACRS with Jain et al. [1] Sr. No.
Components
Parameters
Jain Model [1]
Present Model
Difference (%)
1
Compressor
Wcomp
9.414
9.329
0.90
2
Evaporator
Qe
66.67
66.67
0.00
3
Cascade Condenser
QCC
76.08
76
0.11
4
Absorber
Qa
100.1
99.95
0.15
5
Generator
Qg
104.9
104.8
0.10
6
Pump
Wp
0.0019
0.00194
-2.11
7
Condenser
Qc
80.96
80.85
0.14
8
VCRS
COP
7.082
7.146
-0.90
9
VARS
COP
0.725
0.7252
-0.03
10
System
COP
0.583
0.5841
-0.19
IT
22.81
22.59
0.96
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Table 2 shows the good coherence between the results of the current thermodynamic model and the model described by Jain et al. [1]. It can be seen that the maximum error in irreversibility calculations for different components is found to be within ± 2.15%. 3.4 Design conditions and operating parameters The subsystems of the proposed thermal system are TSVCRS and VARS. The main objective of integrating TSVCRS with VARS is to achieve low temperature cooling (around -45 to -55°C) for food preservation applications and industrial cooling applications. The complete integrated cascaded system (TSVCACRS) depicted in Fig. 2 is modeled in EES by applying the mass, energy and exergy balance equations discussed earlier. The input design data are given in Table 3. The overlap temperature (ܶ െ ܶଽ ) across the cascaded condenser was taken to be 5°C. In the present study, Ammonia is considered as a refrigerant for two stage VCR subsystem and H2O-LiBr is considered in VAR subsystem of integrated TSVCACRS. In industrial refrigeration, Ammonia is widely used as refrigerant due to its easy availability and lower cost though it is highly toxic in nature. The proposed system can achieve low temperature around (-50 to -60°C) in the evaporator of two stage VCRS due to lower condensing temperature in cascade condenser. Table 3. Input Parameters for proposed TSVCACRS Parameter
Value
Condenser coolant inlet temperature (ܶଵଽ in ºC) [7] Condenser coolant outlet temperature (ܶଶ in ºC)
28
[7]
30
Generator coolant inlet temperature (ܶଶଵ in ºC)[1]
100
Generator coolant outlet temperature (ܶଶଶ in ºC) [1]
95
Generator temperature (ܶଵ in ºC)
[1]
Absorber coolant inlet temperature (ܶଶଷ in ºC)
90 [1]
Absorber coolant outlet temperature (ܶଶସ in ºC)
35
[1]
38
Absorber Temperature (ܶଵ in ºC) [1]
40
Rate of heat absorbed by evaporator (ܳሶ in kW) [7]
1850
Evaporator temperature (ܶସᇱ in ºC) [7] Evaporation temperature in cascade condenser (ܶଽ in ºC)
10
Effectiveness of solution heat exchanger (ߝ௦௫ ) [1]
0.6
Isentropic efficiency of compressor (ߟ௦ ) [2]
0.85
Electrical efficiency of pump (ߟ )
4.
-45 [1]
[7]
0.9
Degree of overlap in cascade condenser (ܶ௩ in ºC) [5]
5
Environment temperature (ܶ in ºC)
25
Atmospheric pressure (ܲ in kPa)
101.3
Results and analysis
In this section, the results of computer simulation carried out for the proposed two stage vapor compressionabsorption cascade refrigeration system are presented and discussed. The comparison of performance parameters of TSVCRS (Fig.1) and TSVCACRS (Fig.2) based on the energetic and exergetic analysis of the process (heat and work, COP of various system, irreversibility and second law analysis) are calculated and tabulated in Table 4. For the same cooling capacity of 525 TR, in comparison to standalone TSVCRS, the combined LPC and HPC compressors of TSVCACRS requires 372 kW less work due to integration of VARS which minimizes the pressure ratio across the evaporator and condenser as mentioned in Table 4. Moreover, the compressors capacities and sizes also reduce due to lower discharge temperature of HPC compressor maintained in cascade condenser by integration
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of VARS. The pressure ratio of TSVCACRS is reduced to 13.38 from which is to be 24.8 for standalone TSVCRS. The COP of compression section of TSVCACRS is obtained 2.26 which is 1.61 for standalone TSVCRS. The overall COP of the proposed system is found to be 0.45 which is low value due to integration of VARS since it works on the low grade heat. The COP of the VARS in proposed system is found to be 0.76. From the point of view of second law (the irreversibility and rational efficiency as mentioned in Table 4), the standalone TSVCRS is far better in comparison to TSVCACRS since it utilizes the high grade energy (electricity). Moreover, the exergetic efficiency of the TSVCACRS is found to be 53.59%. The property plot for both systems are shown in Fig. 3 and Fig. 4 which clearly shows that the integration of VARS reduces the condensing temperature (35°C) as shown in Fig. 3 to intermediate temperature (15°C) which leads to less compressor work to achieve same cooling capacity at -45°C. Table 4. Results of energetic performance parameters Sr. No. 1
Performance parameters ܳሶௗ (kW)
Low grade energy
ܳሶ (kW) ܳሶ (kW) ܳሶ௩ (kW) ܳሶ௦ௗ (kW) 2
High grade energy
TSVCACRS
(Fig. 1)
(Fig. 2)
2914.0
2671.0
-
3110.0
-
3269.0
1850.0
1850.0
-
2512.0
ܳሶௌு (kW) ܹሶ (kW)
454.7
345.4
ܹሶு (kW)
690
472.3
315.6
ܹሶ (kW)
-
0.03
3
Operating parameter
Pressure ratio
24.8
13.38
4
First law parameters
்ܱܲܥௌோௌ
1.61
2.26
ܱܲܥோௌ
-
0.76
்ܱܲܥௌோௌ
-
0.45
I (kW)
420.5
675.8
િ܀
63.26
53.59
5
Second law parameters
Ammonia
Ammonia kg -K 5. 20 9
TSVCRS 104 119.4°C
TSVCACRS
1 78 5.
4 43 5.
104
9 33 6.
P [kPa]
74.93°C
103
105
kJ /
105
P [kPa]
TSVCRS
35°C
15°C -18.94°C
102
0
500
9 33 6.
49.03°C
103
-45°C
-45°C
101 -500
-K
8 88 5.
82.15°C
-11.7°C
102
g /k kJ
1000
1500
h [kJ/kg]
Fig.3. Property chart for TSVCRS
2000
2500
101 -500
0
500
1000
1500
h [kJ/kg]
Fig.4. Property chart for TSVCACRS
2000
2500
Bhavesh Patel et al. / Energy Procedia 109 (2017) 190 – 202
The exergetic performance parameters of the proposed TSVCACRS are calculated and presented in Table 5. The irreversibility and exergetic efficiency of the proposed system and sub system along with the individual components are tabulated and mentioned is Table 5. The generator, absorber, evaporator, cascade condenser, condenser contributes the 20.69%, 19.44%, 15.03%,11.54% and 8.21% irreversibility relative to system irreversibility, respectively. The maximum exergetic efficiency of flash cooler is obtained as 92.02%. The exergetic efficiency of the generator, evaporator, HPC compressor, LPC compressor, solution heat exchanger, absorber, condenser and cascade condenser is found to be 78.15%, 73.77%, 70.97%, 69.77%, 56.05%, 46.77%, 41.69% and 41.42% respectively. The exergetic efficiency of the VAR and VCR sub systems are obtained 32.78% and 60.29%, respectively. The exergetic efficiency of the proposed TSVCACRS is found to be 53.59%. The highest exergetic efficiency is obtained for VCR subsystem as it utilizes the high grade energy (electricity) which produces the maximum useful work. The value obtained in case of VAR subsystem is lower as it utilises the low grade heat. The value of exergetic efficiency of the system lies between the values of VCR and VAR subsystem as it utilises both high grade and low grade energy. Table 5. Results of exergetic performance parameters of TSVCACRS (Fig. 2)
Second law Analysis Component/ System
۷ሶ
ܓܚ
િ܀
[kW]
[%]
[%]
Condenser
55.50
8.21
41.69
Absorber
131.4
19.44
46.77
Generator
139.8
20.69
78.15
Evaporator
101.6
15.03
73.77
Cascade condenser
77.96
11.54
41.42
Low pressure compressor
38.78
5.74
69.77
High pressure compressor Pump
47.35 0.00
7.00 0.00
70.97 0.00
Pressure reducing valve
0.00
0.00
0.00
Solution heat exchanger
19.56
2.89
56.05
Expansion valve 1
12.05
1.78
0.00
Expansion valve 2
17.59
2.60
0.00
Expansion valve 3
4.87
0.72
0.00
Flash cooler
29.36
4.34
92.02
VAR subsystem
429.1
-
32.78
VCR subsystem
324.7
-
60.29
TSVCACRS
675.82
-
53.59
4.1 Influence of generator temperature According to Fig. 5, with the hike in the generator temperature, the thermal capacity needed for the generator reduces for the same cooling capacity. Further, the coefficients of performance of the absorption section COP VARS and two stage vapor compression absorption cascade refrigeration system COP TSVCACRS increases up to 90°C. It can be also observed that the exergetic efficiency of the TSVCACRS also improves significantly up to 90°C and then it decreases marginally. The irreversibility of the system and the thermal capacity requirement in generator decreases as the generator temperature increases up to 90°C and then it starts increasing. Thus, the optimum temperature for the generator lies between 85-90°C.
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4.2 Influence of condenser and absorber temperature In this section, the performance of the system is predicted when the absorber and condenser temperature is increased from 35°C to 45°C whilst the other input parameters and conditions of the system are kept constant. For the absorber saturation temperature, the COP and exergetic efficiency of the TSVCACRS decreases by 4% and 0.66%, respectively. Further, the change in condenser saturation from 35-45°C causes reduction in the system COP and exergetic efficiency by 6% and 1.2%, respectively. 0.46
54
54 0.46 Ta
0.44
COPTSVCACRS
R,TSVCACRS (%)
COPTSVCACRS
53
53.5
Tc
Ta
0.44
Tc
53
0.43
R,TSVCACRS (%)
0.45
53.5 0.45
Ta
52.5
52.5 0.42
COPTSVCACRS R,TSVCACRS 0.43 75
80
85
90
95
Tc
COPTSVCACRS R,TSVCACRS 0.41
52 100
52 35
40
45
Tc and Ta [°C]
Tg [°C]
Fig. 5. COP and exergetic efficiency Versus Effect of generator saturation temperature (Tg).
Fig. 6. COP and exergetic efficiency Versus Effect of absorber and condenser saturation temperature (Ta, Tc)
4.3 Influence of cascade overlap temperature The overlap temperature (T 7-T9) of the cascade condenser is an important design parameter as it influences the performance of both the subsystems. It is observed in Fig. 7 that the COP and exergetic efficiency of the TSVCACRS decreases approx. by 4% and 2%, respectively. 0.46
54
0.5
15
20
25
30
35
40
45
58
T0 [C] 0.48
52
0.46 50 0.44
0.43
COPTSVCACRS
R,TSVCACRS 0.42
51 5
6
7
46
0.42
COPTSVCACRS
R,TSVCACRS (%)
0.44
COPTSVCACRS
54 53
R,TSVCACRS
COPTSVCACRS
0.45
8
9
10
Toverlap [°C]
Fig. 7. COP and exergetic efficiency Versus Effect of cascade condenser overlap temperature
0.4 650
R,TSVCACRS 670
690
710
42 730
Itotal [kW]
Fig. 8. COP and exergetic efficiency Versus Effect of ambient temperature
Bhavesh Patel et al. / Energy Procedia 109 (2017) 190 – 202
4.4 Influence of ambient temperature The effect of the ambient temperature on the system performance is investigated over the range of 15-45°C and presented in Fig. 8. The COP of proposed system remains unaffected with change in ambient temperature as the COP of individual systems remains unchanged. The exergetic efficiency of VCR subsystem of TSVCACRS reduces by 4%. Additionally the exergetic efficiency of the VAR subsystem decreases drastically by 41% as the difference between the ambient and the generator temperature decreases which leads to less useful work output from the supplied heat. Thus, according to Fig. 8 the exergetic efficiency of the proposed TSVCACRS is reduced by 14% as the irreversibility of the system increases 10% for the change of ambient temperature from 15°C to 45°C. 5.
Conclusion
In this communication, the thermodynamic performance and feasibility study of the two stage vapor compressionabsorption refrigeration system (TSVCACRS) has been carried out. The proposed system has been modeled and simulated for the existing installed 525 TR (1850 kW) two stage vapor compression refrigeration system (TSVCRS) for ice cream hardening and storage unit for -50 to -35°C evaporator temperature at Havmor ice cream limited. The comparison study of the TSVCRS and proposed TSVCACRS has been carried out. The main results obtained from the thermodynamic simulation and parametric study are concluded below: x The results show that proposed TSVCACRS system would minimize the compressor work up to 28%, compared to an existing installed TSVCRS. x The pressure ratio of compression section for proposed TSVCACRS is predicted to be 13.8 due to intermediate temperature achieved in cascade condenser, which is 24.8 in case of standalone TSVCRS. Thus the capacity of the HPE and LPE compressors would minimize. x The overall COP of the proposed system is obtained 0.45 which is low value in comparison to TSVCRS having COP of 1.61 since the proposed system utilizes the low grade energy in absorption section. x The exergetic efficiency of the proposed TSVCACRS is found to be 10% less compared to existing installed TSVCRS since 255 kW more irreversibility is associated to TSVCACRS. x The exergetic efficiency of the VAR, VCR subsystems and integrated TSVCACRS is found to be 32.78%, 60.29% and 53.59%, respectively. The lower value of exergetic efficiency in case of VAR subsystem is observed due to usage of low grade heat as a source of energy. x The highest irreversibility occurs in generator with 20% followed by absorber with 19.44%, evaporator 15.03%, cascade condenser with 11.54% and condenser with 8.21%. x The optimum generator temperature is found to be 85°C as COP and exergetic efficiencies are increase up to that temperature and then it remains constant up to 90°C and afterwards it starts decreasing. x The increase in condenser and absorber saturation temperature decreases the COP and exergetic efficiencies of the system. Moreover, in the case of condenser the rate of decrease of COP and exergetic efficiency is higher as temperature goes on increases. x The overlap temperature of the cascade condenser affects the system performance at great extent so it is considered as a critical design parameter for the system. The best performance of the system is obtained at 5°C overlap temperature. x The COP of the system remains unaffected as the ambient temperature increases but the exergetic efficiency decreases 14% for the temperature range of 15 to 45°C. Acknowledgements The authors also thank Havmor Ice cream Limited located at GIDC, Naroda, Ahmedabad; for their support and allowing us to their premises to get the realistic data.
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Appendix A. An example appendix Table A.1 Thermodynamic equations for proposed TSVCACRS Sr. No
Components
First law Equations
Second Law Equations
1
Evaporator
Qe = m4 * (h1 – h2)
Ie = T0* ((S1 - S4) - Qe/Tcs)
LP Compressor
WLPC = m1 * (h2-h1) / (ηm * ηe) h2 = h1 + (h2isen – h1) / ηisen pFlash = (pH /pL)0.5 mLT (h3 – h2) = mHT (h8 – h5) h3 = h4 WHPC = mHT * (h6-h5) / (ηm * ηe) h6 = h5 + (h6isen – h5) / ηisen mw * (h9 - h18) = mHT * (h6 - h7) h7 = h8 m9 + m15 = m10 Qa = mw*h9 + m15*h15 – m10*h10 Qa = ma * (h23 - h24) Wp = mss *(P16 - P9)/( ρss * ηp) Wp = m11*h11 – m10*h10 m13*h13 -m14*h14 = m12*h12 – m11*h11 ε = (T13 -T14)/(T13-T11) m12 = m13 + mw m16*h16 + m13*h13 = Qg + m12*h12 Qg = mg* (h21 – h22) h14 = h15
2 3 4 5 6 7
Flash Cooler Ex. V/v 1 HP Compressor Cascade Cond. Ex V/v 2 Absorber
8 9 10
Pump SHE Generator
11
13
PRV Condenser
14
Ex V/v 3
12
ILPC = T0* (S2 - S1) IFlash = T0* ((S3 - S2) + (S5 - S8)) Iev1 = T0* (S4 - S3) IHPC = T0* (S6 - S5) ICC = T0* ((S7 - S6) + (S9 - S18)) Iev2 = T0* (S8 - S7) Ia = T0* (S10 - S9 - S15 + Qa/Ta) Ip = T0* (S11 - S10) ISHE = T0* (S12 - S11) + (S14 -S13) Ig = T0* (S16 + S13 - mss * S12 ) - (Qg/Tg) Iprv = T0* (S15 - S14)
Qc = mw* (h16 – h17) = mc* (h20-h19)
Ic = T0* (S17 - S16) + (S20 - S19)
h17 = h18
Iev3 = T0* (S18 - S17)
References [1] Vaibhav Jain, S.S. Kachhwaha, Gulshan Sachdeva. Thermodynamic performance analysis of a vapor compression-absorption cascaded refrigeration system. Energy Conversion and Management 2013; 75: 685–700. [2] Cimsit C, and Ozturk I. Analysis of compression-absorption cascade refrigeration cycles. Appl Therm Eng 2012; 40:311–7. [3] Wang L, Ma A, Tan Y, Cui X, and Cui H. Study on solar assisted cascade refrigeration system. Energy Procedia 2012; 16(C):1503–9. [4] Garimella S, Brown AM, and Nagavarapu AK. Waste heat driven absorption/vapor-compression cascade refrigeration system for megawatt scale, high-flux, low-temperature cooling. Int J Refrig 2011; 34:1776–85. [5] Colorado D, Velazquez VM. Exergy analysis of a compression–absorption cascade system for refrigeration. Int J Energy Res 2013; 37(14):1851-65. [6] Nikolaidis C and Probert D. Exergy-method analysis of two-stage vapor compression refrigeration-plants performance, Appl Energy. 1998; 60:241–56. [7] Havmor Ice cream Limited. GIDC, Naroda, Ahmedabad. [Visited on 12.01.2016].