- Email: [email protected]
A new open absorption heat pump for latent heat recovery from moist gas
Energy Conversion and Management 94 (2015) 438–446
Contents lists available at ScienceDirect
Energy Conversion and Management journal homepage: www.elsevier.com/locate/enconman
A new open absorption heat pump for latent heat recovery from moist gas Bicui Ye a, Jun Liu a, Xiangguo Xu a, Guangming Chen a,b,⇑, Jiao Zheng b a b
Institute of Refrigeration and Cryogenics, State Key Laboratory of Clean Energy Utilization, Zhejiang University, 38 Zheda Road, Hangzhou 310027, PR China Ningbo Institute of Technology, Zhejiang University, 1 Qianhu South Road, Ningbo 315100, PR China
a r t i c l e
i n f o
Article history: Received 10 November 2014 Accepted 1 February 2015
Keywords: Open absorption Double-stage mode Single-stage mode Latent heat recovery Thermodynamic analysis
a b s t r a c t Conventional drying processes discharge high humidity gas to the atmosphere. The exhaust gas contains large amount of energy. The direct discharging would result in relatively large energy waste. In order to improve the thermal efficiency of drying process, in this paper, a new open absorption heat pump system was proposed, which aims at recovering the latent heat from exhausted moist gas and producing steam for reutilization. The working principle was discussed in detail and thermodynamic models were established to analyze the performance of the new system. The new system can work under both singlestage and double-stage modes. Simulation results showed that the new system could utilize a heat source with lower generation temperature compared with that utilized by a traditional open absorption system. The temperature range of heat source for the double-stage mode is 130–160 °C, and that for the single-stage mode is 160–175 °C. The new system also eliminates the limitation of traditional close absorption system, whose evaporation temperature has to be lower than the dew point temperature of discharged moist gas to recover the latent heat of water steam. Simulation results also indicated an improved COPh of the new system compared with that of double-stage close absorption heat pump system. The COPh of the new system varied from 1.52 to 1.97 and the efficiency of heat recovery varied from 15.1% to 54.8% when the temperature of heat source varied from 135 °C to 175 °C and saturated steam of 100 °C was produced. Ó 2015 Elsevier Ltd. All rights reserved.
1. Introduction Increased energy consumption and greenhouse gas emissions have led to the concerns of the improvement of thermodynamic efficiency of integrated industrial processes. Drying process is widely applied in industries such as pulp and paper manufacturing, wood manufacturing and food processing industry and is one of the most energy-intensive unit operations. It was reported that the energy consumption of drying processes accounted for 9–25% of national energy consumption in the developed countries [1]. Furthermore, during the drying process, more than 70% of the energy consumed is exhausted to the environment in the form of
Abbreviations: AHP, absorption heat pump; A, absorber; A1, primary absorber; A2, secondary absorber; CAHP, close absorption heat pump; COPh, coefficient of performance of heat pump; CR, circulation ratio; G, generator; G1, primary generator; G2, secondary generator; NOAHP, new open absorption heat pump; OAHP, open cycle absorption heat pump; SHX, solution heat exchanger. ⇑ Corresponding author at: Institute of Refrigeration and Cryogenics, State Key Laboratory of Clean Energy Utilization, Zhejiang University, 38 Zheda Road, Hangzhou 310027, PR China. Tel.: +86 13857170531; fax: +86 0571 87951680. E-mail address: [email protected] (G. Chen). http://dx.doi.org/10.1016/j.enconman.2015.02.001 0196-8904/Ó 2015 Elsevier Ltd. All rights reserved.
moist gas [1–6]. This consequently results in a low energy efficiency of the drying process, which is typically in the range of 20–40% [5]. Because of the low thermodynamic-efficiency and high energyconsumption, there is a very large improving space for the drying process. In order to improve the thermodynamic efficiency, researchers have devoted cumulative efforts to recover the latent heat in the discharged moist gas by using the heat integration approaches. Heat exchanger is the most commonly used equipment for heat recovery, which has advantages of low investment and easy operation. In the research of Laurijssen et al. [7], waste heat in the exhaust gas was recovered by using heat exchanger to pre-heat the incoming air and the process water, leading to 15% decrease of the energy consumption in multi-cylinder paper drying sections. However, using heat exchanger to directly recovering the latent heat in the moist gas only could produce hot air with the temperature lower than the dew point of the moist gas. The traditional heat pump system, including both mechanical heat pump system and absorption heat pump system, has emerged as one of the widely accepted technologies integrated with the drying processes. Anderson et al. [8] estimated that the mechanical
B. Ye et al. / Energy Conversion and Management 94 (2015) 438–446
439
Nomenclature d h m p Q t x f k
absolute humidity (kg per kg dry air) enthalpy (kJ/kg) mass flow rate (kg/s) water vapor partial pressure (kPa) heat transfer rate (kW) temperature (°C) solution concentration (wt.) efficiency of heat recovery liquid–gas ratio
Subscripts a1 primary absorber a2 secondary absorber a absorber
heat pump system, which was coupled with common drying schemes used in Swedish sawmills, could decrease heat demand by 5.6 TW h/year at the cost of increased electricity demand by 1 TW h/year. Bakhtiari et al. [9] investigated the optimal integration of absorption heat pumps in a kraft pulping process using Pinch Analysis method. Ahmadi et al. [10,11] optimized an irreversible absorption heat pump system based on a new thermo-ecological criterion. Qu et al. [12] proposed three new configurations of absorption heat pump to recover latent heat in flue gas from natural gas boiler. The simulation results suggested 5–10% improvement in boiler efficiency. Lostec et al. [13] developed a model of wood chip drying process coupled with absorption heat pumps to conduct thermal and economic analysis. Researches [8–15] demonstrated that absorption heat pumps could effectively recover the latent heat of moist gas since the evaporation temperature could be much lower than the dew point temperature of the moist gas. However, the low evaporation temperature reduced the COPh of heat pump systems at the same time. To overcome the limitation of traditional close absorption system, whose evaporation temperature should be lower than the dew point temperature of discharged moist gas to recover the latent heat of water vapor, Westerlund et al. [16–18] proposed an open cycle absorption heat pump (OAHP) system to recover the latent heat from the moist gas. As shown in Fig. 1, the OAHP system consists mainly of three parts, absorber, generator and condenser. The moist gas contacts directly with the absorbent in the absorber. Water vapor in the moist gas is absorbed by the absorbent due to the water vapor partial pressure difference between moist gas and absorbents, giving off the latent heat of vapor to directly heat the moist gas, which could be reused by the drying process. Therefore, the latent heat of vapor could be recovered with
Fig. 1. Schematic diagram of traditional OAHP system [16].
c envir eva g1 g2 ge hg hp in max mg out recovered
condenser environment evaporation primary generator secondary generator generation high-pressure generator heat pump input maximum moist gas output recovered from the moist gas
the temperature much higher than the dew point temperature of moist gas. In the research of Wang and Zhang et al. [19,20], the discharged gas was heated from 70 °C to 125 °C in the absorber of the OAHP system and used to produce steam at 120 °C. Anderson et al. [8] demonstrated that the OAHP system could decrease energy consumption by 67.4% for sawmill drying process. Although the OAHP system performs well at recovering the latent heat of moist gas, to produce high temperature hot air or steam for drying process using the OAHP system would require an even higher generating temperature, which limits its application [16–21]. Consequently, a new OAHP system (NOAHP) is proposed in this paper, which could be operated at double-stage mode to utilize low temperature heat source for generator or single-stage mode to improve system energy efficiency when high temperature heat source for generator is available. In order to validate the feasibility of the NOAHP system, thermodynamic model was established to analyze the system performance based on the properties of chose absorbent and the mass and energy balance of each component. The performance of the NOAHP system was also compared to that of traditional open absorption and traditional close absorption heat pump system.
2. The description of the NOAHP system 2.1. The working process of the NOAHP system Fig. 2 shows the schematic of the new open absorption heat pump (NOAHP) system. It consists of two generators (primary and secondary generator G1, G2), two absorbers (primary and secondary absorber A1, A2), condenser and solution heat exchanger (SHX). The working process of the NOAHP system is described as follows. Firstly, the open absorption process, which occurs in A1, is a main feature of the new system. It is also the main difference between an open absorption cycle and a close absorption cycle. The water vapor flows into the absorber A1 from outside in terms of moisture in the moist gas and is discharged to outside in terms of condensate at condenser (or directly outputted from generator G1 as hot steam), instead of circulating in the system. This is a main reason why the system is called open cycle absorption. In A1, the moist gas contacts directly with strong solution and the moisture in the moist gas is absorbed when the water vapor partial pressure of the strong solution is lower than that of the moist gas. Latent heat given off during the absorption process is transported to water in the heat exchanger, which is installed in A1.
440
B. Ye et al. / Energy Conversion and Management 94 (2015) 438–446
Fig. 2. Schematic diagram of the NOAHP system.
Secondly, the water vapor absorbed by the weak solution in absorbers was heated and evaporated from the weak solution in G1 and G2. Water vapor generated in G1 could be directly outputted as high-temperature steam (in this case, condenser is not needed anymore), or flow to the condenser and give out condensing heat to produce steam/hot water. Water vapor generated in G2 flows into A2, where the vapor is absorbed by the weak solution from A1 and absorption heat is provided for producing hot water/air or steam. Finally, the solution circulates in the system. The strong solution is diluted in A1 by water vapor in the moist gas (3–4). Then the solution is pumped into A2, where it is further diluted by water vapor from G2 (4–5–6). Afterward, weak solution from A2 flows through SHX before it is regenerated in G1 and further regenerated in G2 (6–7–8–9–10–11). Energy from heat sources is supplied to G1 and G2 in order to regenerate weak solution. Strong solution from G2 also flows through SHX to exchange heat with weak solution before it flows into A1 (11–12–3). Fig. 3. p–T diagram of H2O/LiBr in the NOAHP system and OAHP system.
2.2. The fundamental principle of the NOAHP system H2O/LiBr serves as the working pair in this study. Fig. 3 is the p– T diagram of H2O/LiBr in the NOAHP system (the continuous line) and in the traditional OAHP system (the dash line). In Fig. 3, working conditions for the two systems are assumed as follows: the two systems are used to produce steam with the same pressure pout and the same temperature tout. tc is higher than tout to form a temperature difference for heat transfer and is also the same for two systems. Same condensation temperatures correspond to same condensation pressures, which approximately equal to the pressure of generator (for the NOAHP system, it is the pressure of G1, pg1). Therefore, pg1 in the NOAHP system equals to the pressure of generator in traditional OAHP system. As shown in Fig. 3, process of 8–9 is the generation process in G1 in the NOAHP system and process of 8–0 is the generation process in OAHP system. The concentration order of solution is the concentration of solution from G2, which is represented by xg2, and those of solutions from
G1, A1 and A2 are, xg1, xa1, and xa2, respectively. Solution from G1 for the NOAHP system and solution from generator for traditional OAHP system are of the same concentration to ensure a same absorption concentration of solution entering the open absorbers. The p–T lines of H2O/LiBr in Fig. 3 indicate that the NOAHP system can utilize a lower generation temperature than traditional OAHP system does. The generation temperature for OAHP system is approximately 162 °C in order to produce steam at pressure of pg1 and to regenerate solution at concentration of xg2 (as point 0 marked in Fig. 3). However, the regeneration process in the NOAHP system is conducted in two steps. G1 produces steam at pressure of pg1 and solution at concentration of xg1, which is lower than xg2 (as point 9 marked in Fig. 3). Consequently, the required generation temperature for G1 is only 146 °C. G2 further regenerates solution at concentration of xg2, and produce steam at pressure of pg2 (as
B. Ye et al. / Energy Conversion and Management 94 (2015) 438–446
point 11 marked in Fig. 3), which is lower than pg1. Thus, the heat source temperature for G2 is approximately 148 °C. Furthermore, the generation temperature could be even lower for G2 as pg2 decreases. On the other hand, the NOAHP system could work under two modes: single-stage mode and double-stage mode. When heat source temperature for G1 is high enough to produce solution with concentration of xg2, the heat supply for G2 could be turned off. In this case, G2 and A2 do not work and the NOAHP system works under single-stage mode. Otherwise, the NOAHP system works under double-stage mode when heat source temperature is not high enough to drive the single-stage mode. The switch between the single-stage mode and the double-stage mode can be simply realized by the automatic on–off control of the heat supply to G2. The practical operation of the NOAHP system would be further investigated in related studies. To optimize the performance of the NOAHP system, thermodynamic simulation is conducted in Sections 3 and 4. Section 5 compares the performance of NOAHP system with that of traditional close absorption heat pump system. 3. Thermodynamic model of the NOAHP system To simplify the model, the following assumptions are made in this study. (1) The system is performed in steady-state conditions; (2) Temperatures of heat sources for two generators G1 and G2 are the same; (3) The solution from G1, G2, A1 and A2 is saturated; (4) A temperature efficiency of SHX (the ratio between the temperature difference for the strong solution (t12–t11) and the theoretical available temperature difference (t11–t7)) is assumed at 80%; (5) A temperature difference of 10 °C is assumed between the heat source temperature and output temperature of generators; (6) Thermal leaks, solution loss and pressure drops are neglected; (7) The solution pump power and fan power are neglected. Based on above assumptions, mass and energy conservation equations are applied to the main components in the NOAHP system. (Numeric subscripts in following equations refer to numbers marked in Fig. 2.) The first law of thermodynamics of the new OAHP system is formulated as below:
Q g1 þ Q g2 þ Q a1;in ¼ Q a1;out þ Q a2 þ Q c
mdryair ðd2 d1 Þ ¼ m4 m3
Secondary absorber A2
mg2;watervapor ¼ m13 ¼ m6 m5
ð6Þ
m5 x5 ¼ m6 x6
ð7Þ
Q a2 þ m6 h6 ¼ mg2;watervapor h13 þ m5 h5
ð8Þ
Primary generator G1
mg1;watervapor ¼ m14 ¼ m8 m9
ð9Þ
m7 x7 ¼ m9 x9
ð10Þ
Q g1 þ m8 h8 ¼ m9 h9 þ mg1;watervapor h14
ð11Þ
Secondary generator G2
mg2;watervapor ¼ m13 ¼ m10 m11
ð12Þ
m10 x10 ¼ m11 x11
ð13Þ
Q g2 þ m10 h10 ¼ m11 h11 þ mg2;watervapor h13
ð14Þ
Condenser
m14 ¼ m15
ð15Þ
Q c ¼ mwatervapor ðh14 h15 Þ
ð16Þ
Since the mass flow rate of solution remains constant when flowing through pumps, valves and heat exchangers:
m3 ¼ m11 ¼ m12
ð17Þ
m6 ¼ m6 ¼ m8
ð18Þ
m9 ¼ m10
ð19Þ
m4 ¼ m5
ð20Þ
According to mass conservation, the amount of water vapor absorbed in A1 equals to that of water vapor generated at G1, therefore, it has:
mg1;watervapor ¼ m14 ¼ mdryair ðd2 d1 Þ
ð21Þ
Circulation ratio, CR, is defined as the mass of water vapor evaporated from G2 divided by the mass of water vapor evaporated from G1.
CR ¼
ð1Þ
where Qa1,in represents energy released by the moist gas, and Qa1,out represents energy obtained by the water in the heat exchanger, which is installed in A1. Qa2 represents absorption heat obtained by the water in the heat exchanger, which is installed in A2. Qc represents condensation heat obtained by the water in the heat exchanger, which is installed in the condenser. Qg1 and Qg2 represent the generation heat supplied to G1 and G2, respectively. The balance equations of each component are shown respectively as follows: Primary absorber A1
441
mg2;watervapor m13 ¼ mg1;watervapor m14
ð22Þ
The liquid–gas ratio k is defined as the mass of concentrated solution flowing into A1 divided by the mass of dry air flowing through A1.
k¼
m3 mdryair
ð23Þ
For thermodynamic performance evaluation, the coefficient of performance COPh is defined as:
COPh ¼
ð2Þ
Q a1;out þ Q a2 þ Q c Q g1 þ Q g2
ð24Þ
The efficiency of heat recovery from the moist gas f is defined as:
m3 x3 ¼ m4 x4
ð3Þ
Q a1;in ¼ mdryair ðh1 h2 Þ
ð4Þ
Q a1;in þ m3 h3 ¼ Q a1;out þ m4 h4
ð5Þ
f¼
mdryair ðh1 h2 Þ Q recovered ¼ Q max mdryair ðh1 henv ir Þ
ð25Þ
where Qrecovered represents energy recovered from the moist gas, Qmax represents the overall energy released when the moist gas is
442
B. Ye et al. / Energy Conversion and Management 94 (2015) 438–446
cooled to the environmental condition, and henvir represents enthalpy of gas under the environmental condition. Properties of the moist gas, including absolute humidity, relative humidity, water vapor partial pressure, temperature and enthalpy, are given by Hyland and Wexler [22,23]. Properties of H2O/LiBr, including concentration, temperature, water vapor partial pressure and enthalpy, are given by Pátek and Klomfar [24]. Thermodynamic calculations were conducted by EES (Engineering Equation Solver, Professional Version 8.889, 06/25/11). 4. Parametric analysis of the NOAHP system In order to provide reliable performance evaluation for the NOAHP system, drying process of paper manufacturing is used as the base-case condition in this study [12]. Energy of moist gas from the drying process of paper manufacturing is recovered by the NOAHP system, which produces required steam for the drying process. Base-case conditions are shown in Table 1. Simulation results indicate that COPh is 1.737 and efficiency of heat recovery f achieves 50.16% in the base-case condition. Since real operation conditions may vary over a wide range, parametric analysis is discussed below to optimize the performance of the NOAHP system. Main factors affecting the system performance are pressure of G2, pg2, temperature of heat source tge, absolute humidity of moist gas, dmg, and temperature of moist gas, tmg. The influence of each factor on COPh and f are given in Figs. 4–9. Operational parameters are kept constant at the basecase values (Table 1), except for parameters that marked in the figures.
Fig. 4. Effects of pg2 on COPh and f.
4.1. Effects of the pressure of G2 on COPh and f Fig. 5. Effects of pg2 on k and CR.
pg2 is an adjustable operational parameter, which could be controlled by the throttling valve between G1 and G2, in the NOAHP system under the double-stage mode. Fig. 4 shows the relationship between pg2 and performance parameters of the NOAHP system, which provides a basis for the optimization of pg2. It can be seen that with the increase of pg2, COPh initially rises and then peaks at a certain pg2. However, f decreases rapidly with the increase of pg2. COPh and f cannot be optimized simultaneously by adjusting
pg2. Considering energy from moist gas is recovered at the cost of even higher temperature energy, working conditions with a low COPh and a high f do not make sense from the energetic view. Therefore, the NOAHP system recovers energy from moist gas with the aim of consuming less high-temperature energy. Thus, the pg2, which corresponding to the maximum COPh, is defined as the optimized pg2.
Table 1 Operational parameters and performance of the NOAHP system at the base-case condition. Temperature (°C)
Water vapor partial pressure (kPa)
Mass flow rate (kg/s)
0.010 0.3031 0.1326
1.303 1.141
20 80 105
Steam Steam Steam Steam Steam Steam
100 100 150 150 100
Pressure (kPa) 101.3 101.3 112.7 64.97 101.3
0.1692 0.0519 0.1704 0.0501 0.1715
Heat source For G1 For G2
160 160
617.7 617.7
0.2116 0.0623
Absorbent (H2O/LiBr) Solution from G2 (11) Solution from A1 (4,5) Solution from A2 (6,7) Solution from G1 (9,10) Solution from SHX (12,3) Solution from SHX (8)
150 105 105 150 114 113
Water vapor partial pressure (kPa) 64.97 54.78 64.97 112.7 17.80 117.5
from from from from from
A1 A2 G1 (14) G2 (13) Conderser
Base-case performance COPh Efficiency of heat recovery f
1.737 0.5016
1.637 33.19 17.80
Absolute humidity (kg/kg)
Ambient conditions (0) Moist gas (1) Exhaust gas (2)
Concentration (wt.%) 0.6356 0.4460 0.41 0.5649 0.6356 0.41
0.4008 0.5712 0.6214 0.4509 0.4008 0.6214
B. Ye et al. / Energy Conversion and Management 94 (2015) 438–446
443
Fig. 6. Effects of pg2 on heat transfer rate of each component.
Fig. 7. Influence of tge on COPh and f.
f decreases rapidly with the increase of pg2. The concentration of solution entering A1, i.e., x3, decreases with the increase of pg2 according to the property of H2O/LiBr. Water vapor pressure of H2O/LiBr increases with the decrease of x3, leading to the drop of the mass transfer driving force between solution and moist gas. f thus decreases. The variation of COPh with pg2 can be explained as follows. Firstly, the decrease of f means the decrease of water vapor absorbed from the moist gas, i.e., mrecover, which equals to m14. The decrease of m14 leads to the decrease of Qa1,out, Qg1 and Qc. However, as shown in Fig. 5, CR decreases with the increase of pg2, which indicates the decrease of m13 is more dominant than the decrease of m14. The decrease of m13 leads to the decrease of Qg2 and Qa2. As shown in Fig. 6, Qa1,in, Qa1,out, Qg1 and Qc decrease almost linearly with the increase of pg2, whereas, Qg2 and Qa2 decreases along a parabola line. When pg2 is below 80 kPa, the decrease of Qg2 and Qa2 is faster than the decrease of Qa1,out, Qg1 and Qc, COPh increases. When pg2 increases to 80 kPa, Qg2 and Qa2 remains almost the same, the decrease of Qa1,out, Qg1 and Qc leads to the decrease of COPh. Secondly, the drop of liquid–gas ratio as shown in Fig. 5 leads to the increase of COPh since less generation heat is used to heat the solution. In sum, to get a maximum COPh for a given working condition, pg2 should be optimized. It is noteworthy that the value of COPh and f in the following analysis section are all calculated with optimized pg2.
to the decrease of x3. Therefore, the upward trend for f in the double-stage mode is smaller than that in the single-stage mode. The increase of f results in the increase of Qa1,out, Qg1 and Qc. On the other hand, the increase of pg2 leads to the decrease of concentration difference between solutions from G1 and G2, which indicates the decrease of m13. The decrease of m13 leads to the decrease of Qg2 and Qa2. The increase of Qa1,out and Qc is larger than the decrease of Qa2, leading to the increase of output energy from the NOAHP system. The increase of Qg1 and decrease of Qg2 leads to the slightly increase of input energy. As a result, COPh increases.
4.2. Effects of temperature of heat source on COPh and f The NOAHP system works under double-stage mode within the temperature range of 130–160 °C and works under single-stage mode within a higher temperature range of 160–175 °C. When tge is below 160 °C, the concentration of solution from G1 is not strong enough to absorb water vapor in the moist gas. The solution has to be further regenerated in G2. Thus, the NOAHP system works under the double-stage mode. When tge is above 160 °C, solution from G2 is close to the crystallization condition. In this case, the heat supplied to G2 should be closed. Single-stage mode therefore activates. As shown in Fig. 7, f increases with tge. Under the single-stage mode, concentration of solution from G1, i.e., x10, increases with the increase of tge. Under the double-stage mode, concentration of solution from G2, i.e., x3 also increases. The increases of x10 and x3 make the mass transfer driving force between the solution and moist gas in A1 increases. f thus increases with the increase of tge under both single-stage and double-stage modes. However, the optimized pg2, which corresponds to the maximum COPh under the double-stage mode, increases with the increase of tge, leading
4.3. Effects of absolute humidity of moist gas on COPh and f As shown in Fig. 8, both COPh and f increase with dmg. This is because the increase of dmg leads to the increase of water vapor partial pressure of moist gas, which increases the driving force between moist gas and solution in A1. Water vapor absorbed from the moist gas thus increases, which makes f and COPh increases. With the increase of mass transfer driving force between solution and moist gas, a lower absorption concentration of solution, i.e., x3, is allowed for the NOAHP system without decreasing the heat recovery efficiency. Therefore, the increase of the absolute humidity of moist gas can decrease the required temperature of the heat source for the NOAHP system. This finding confirms that the open absorption system is of particular interest when dealing with high humidity moist gas. 4.4. Effects of temperature of moist gas on COPh and f It can be seen from Fig. 9 that as tmg increases, f decreases slightly and COPh increases slightly. This is because the water
Fig. 8. Effects of dmg on COPh and f.
444
B. Ye et al. / Energy Conversion and Management 94 (2015) 438–446
Fig. 9. Effects of tmg on COPh and f.
vapor partial pressure of moist gas is only relevant to the absolute humidity, and has nothing to do with the gas temperature. Thus, the mass transfer driving force between moist gas and solution in A1 would remain constant with the increase of tmg. However, the increase of gas temperature decreases the heat transfer driving force between the moist gas and the solution, which impedes the heat transfer between the solution and heat exchanger installed in A1. Solution temperature thus increases, along with the increase of water vapor partial pressure of solution. This directly leads to the slight decrease of f. The slight decrease of f leads to the slight decrease of Qa1,out, Qg1 and Qc. On the other hand, corresponding pg2 increases with the increase of tmg, which decreases the mass of water vapor evaporated from G2, i.e., m13, leading to the decrease of Qg2 and Qa2. COPh is finally increased. Since the gas temperature has limited effect on f, the increase of COPh is mainly a result of the increase of pg2. In sum, simulation results above demonstrate that the NOAHP system performs well at recovering energy, especially the latent heat, from moist gas. The temperature of gas has limited effect on the performance of the NOAHP system while the absolute humidity of gas has significant effect. On the other hand, the NOAHP system can be operated with lower temperature heat source when the absolute humidity of gas is increased. Although efficiency of heat recovery f cannot be simultaneously optimized with the optimization of COPh, it can still achieve about 20–50%, even under a high absorption temperature (when 100 °C steam is produced). The optimized COPh can achieve from 1.19 to 1.97 when temperature of heat source varies from 130 to 175 °C.
sidering that efficiency of heat recovery f is strongly influenced by teva, the performance of double-stage CAHP is evaluated at different teva. Thirdly, pg2 of the double-stage CAHP system and pg2 of the NOAHP system have been optimized in the discussion conditions. f of the double-stage CAHP system and the NOAHP system are compared as shown in Fig. 10. It can be seen that f of the double-stage CAHP system is only influenced by peva. f is 12.5%, 26.3% and 43.9%, respectively when peva is 24.4 kPa, 21.3 kPa and 13.25 kPa and corresponding teva is 59.44 °C, 57.44 °C and 51.44 °C. On the other hand, f of the NOAHP system increases with tge as discussed in Section 4.2. It varies from 15.1% to 54.8% when tge increases from 135 °C to 175 °C. Fig. 11 shows that the NOAHP system has a much higher COPh than the double-stage CAHP system does. COPh of the NOAHP system increases from 1.52 to 1.97 as discussed in Section 4.2. COPh of the double-stage CAHP system varies from 1.30 to 1.43 with the increase of tge. In addition, COPh of the double-stage CAHP system also decreases slightly with the increase of peva. Table 2 shows the heat supplied (positive value) and rejected (negative value) by the components of the double-stage CAHP system and the NOAHP system at different temperatures of heat source. In the double-stage CHAP system, since the evaporation temperature has to be lower than the dew point temperature of moist gas, the corresponding evaporation pressure, i.e., peva, has to be low. This leads to a large pressure difference between the evaporator and high-pressure generator. Lots of energy has to be consumed to pressurize the water vapor circulating in the double-stage CAHP system from peva to the pressure of high-pressure generator phg. On the other hand, in the NOAHP system, due to the direct contact
Fig. 10. f of the double-stage CAHP system and the NOAHP system.
5. Thermodynamic performance comparison of the NOAHP system and traditional CAHP system As discussed in Sections 2 and 4, the NOAHP system can work under single-stage or double-stage mode. Correspondingly, both single-stage cycle traditional close absorption heat pump systems (CAHP) and double-stage cycle CAHP have been used to recover waste heat of moist gas [10–15]. Lostec et al. [13] found that under the given heat source (152 °C): (1) the single-stage CAHP was only effective when the output temperature was below 60 °C; (2) the double-stage CAHP should be used when the output temperature was from 60 to 100 °C. Since the NOAHP system also aims at a high output temperature, the performance of NOAHP system is compared with that of the double-stage CAHP system in this section. The discussions are based on the following facts. Firstly, doublestage CAHP system recovers waste heat by cooling and dehumidifying the moist gas, i.e., the evaporation temperature has to be lower than the dew point temperature teva < tdp. Secondly, con-
Fig. 11. COPh of the double-stage CAHP system and the NOAHP system.
445
B. Ye et al. / Energy Conversion and Management 94 (2015) 438–446 Table 2 Heat transfer rates of each components in the double-stage CAHP system and the NOAHP system. tge (°C)
Qeva (kW)
Qa1 (kW)
Qa2 (kW)
Qg1 (kW)
Qg2 (kW)
Qcon (kW)
Efficiency f
COPh
Double-stage CAHP system 140 232.8 150 232.8 160 232.8 170 232.8
271.3 261.4 252.6 244.6
306.2 299.2 301.8 299.3
275.1 264.5 253.8 244.8
310.1 303.9 308.3 306.8
240.6 240.6 240.6 240.6
0.2634 0.2634 0.2634 0.2634
1.398 1.410 1.414 1.422
tge (°C)
Qa1,out (kW)
Qa2 (kW)
Qg1 (kW)
Qg2 (kW)
Qcon (kW)
Efficiency f
COPh
286.1 378.5 357.7 528.2
140.6 65.95 4.34 0
326.9 425.8 413.1 595.4
146.7 71.15 4.783 0
346.9 461.5 464.7 660.9
0.208 0.280 0.404 0.548
1.633 1.823 1.979 1.997
Qa1,in (kW)
The NOAHP system 140 300 150 409 160 408.9 170 593.7
with moisture gas at environmental pressure, the pressure in A1 is much higher than peva of CHAP system. If phg in the double-stage CAHP system equals to pg1 in NOAHP system, the energy consumed to pressurize the water vapor absorbed in A1 to pg1 is relatively less, which is beneficial for the NOAHP system to achieve a higher COPh than that of the double-stage CAHP system. From the energetic viewpoint, in the double-stage CAHP system, the energy released in high-pressure absorber approximately equals to the energy supplied to low-pressure generator because water vapor evaporated from low-pressure generator is completely absorbed in the high-pressure absorber. Meanwhile, the energy released in high-pressure absorber also approximately equals to the energy supplied to high-pressure generator because water vapor absorbed in high-pressure absorber is totally evaporated in the high-pressure generator. For the same reason, energy released from low-pressure absorber approximately equals to energy supplied to low-pressure generator and energy supplied to high-pressure generator approximately equals to energy released in the condenser. Therefore, if neglecting the energy loss of the whole cycle, the maximum COPh of the double-stage CAHP system would be 1.5. On the other hand, the maximum COPh of the NOAHP system is 2 under the single-stage mode since energy supplied to G1 is used to produce energy released both in A1 and the condenser. And the theoretical minimum COPh of 1.5 would be achieved when the NOAHP system works under the double-stage mode. Thus, COPh of the NOAHP system is higher than that of the double-stage CAHP system. 6. Conclusion A new open absorption heat pump (NOAHP) system has been proposed to be applied in the heat recovery of moist gas in this study. The new system can utilize a wider temperature range of heat source by the innovation of system structure, which allows the new system to work under two operational modes. When temperature of heat source is high enough (160–175 °C), NOAHP system works under single-stage mode to ensure a high COP, and when temperature of heat source is not high enough to drive the single-stage mode, double-stage mode is operated to utilize heat source with a lower temperature range (130–160 °C). Simulation results show that under the base-case condition, COPh can achieve 1.737 and efficiency of heat recovery f achieves 50.16% when 100 °C saturated steam is produced. The thermodynamic parameter analysis of the new system shows that: (1) pressure of G2 pg2 could be optimized to attain the maximum COPh; (2) increasing temperature of heat source improve COPh and efficiency of heat recovery;
(3) temperature of the moist gas has little impact on performance of the new system; (4) increasing absolute humidity of the moist gas improve performance of the new system. The parameter analysis suggests that the NOAHP system is feasible for energy recovery of moist gas with high absolute humidity and even a low temperature. The performance comparison between the NOAHP system and traditional double-stage CAHP system shows that COPh of the new system is much higher than that of the double-stage CAHP system. COPh of the new system achieves from 1.52 to 1.97 while COPh of the double-stage CAHP system is only from 1.30 to 1.43. The NOAHP system can be applied to any industrial applications where moist gas is discharged and wasted. It recovers energy of moist gas and produces high-temperature steam for re-utilization. The application of the NOAHP system can thus improve energy efficiency for many industrial unit operations. Further work of the new OAHP system will mainly cover: (1) exploring less corrosive absorbent, which is suitable for the open absorption structure; (2) improving structure of open absorber to improve absorption efficiency; (3) carrying out experimental work of the new system, including the control of single-stage mode and double-stage mode and the use of anticorrosive material. Acknowledgements This work was financially supported by the major program of the National Natural Science Foundation of China (50890184) and major program of the Industrial Technology Innovation Program of Ningbo (2013B10029). References [1] Misha S, Mat S, Ruslan MH, Sopian K. Review of solid/liquid desiccant in the drying applications and its regeneration methods. Renew Sustain Energy Rev 2012;16:4686–707. [2] Goh LJ, Othaman MY, Mat S, Ruslan H, Sopian K. Review of heat pump systems for drying applications. Renew Sustain Energy Rev 2011;15:4788–96. [3] Jana AK. Advances in heat pump assisted distillation column: a review. Energy Convers Manage 2014;77:287–97. [4] Law R, Harvey A, Reay D. Opportunities for low-grade heat recovery in the UK food processing industry. Appl Therm Eng 2013;53:188–96. [5] Minea V. Heat pump assisted drying: recent technological advances and R&D needs. Drying Technol 2013;31:1177–89. [6] Hepbasli A, Biyik E, Ekren O, Gunerhan H, Araz M. A key review of wastewater source heat pump (WWSHP) systems. Energy Convers Manage 2014;88:700–22. [7] Laurijssen J, Gram FJ, Worrell E, Faaij A. Optimizing the energy efficiency of conventional multi-cylinder dryers in the paper industry. Energy 2010;35:3738–50. [8] Anderson JO, Westerlund L. Improved energy efficiency in sawmill drying system. Appl Energy 2014;113:891–901. [9] Bakhtiari B, Fradette L, Legros R, Paris J. Opportunities for the integration of absorption heat pumps in the pulp and paper process. Energy 2010;35:4600–6.
446
B. Ye et al. / Energy Conversion and Management 94 (2015) 438–446
[10] Ahmadi MH, Ahmadi MA, Mehrpooya M, Feidt M. Thermodynamic optimization of Stirling heat pump based on multiple criteria. Energy Convers Manage 2014;80:319–28. [11] Ahmadi MH, Ahmadi MA, Mehrpooya M, Sameti M. Thermo-ecological analysis and optimization performance of an irreversible three-heat-source absorption heat pump. Energy Convers Manage 2015;90:175–83. [12] Qu M, Abdelaziz O, Yin HX. New configurations of a heat recovery absorption heat pump integrated with a natural gas boiler efficiency improvement. Energy Convers Manage 2014;87:175–84. [13] Lostec BL, Galanis N, Baribeault J, Millette J. Wood chip drying with an absorption heat pump. Energy 2008;33:500–12. [14] Hamdan MA, Rossides SD, Khalil RH. Thermal energy storage using thermochemical heat pump. Energy Convers Manage 2013;65:721–4. [15] Chaturvedi SK, Gagrani VD, Abdel-Salam TM. Solar-assisted heat pump – a sustainable system for low-temperature water heating applications. Energy Convers Manage 2014;77:550–7. [16] Westerlund L, Hermansson R, Fagerstrom J. Flue gas purification and heat recovery: a biomass fired boiler supplied with an open absorption system. Appl Energy 2012;96:444–50. [17] Johansson L, Westerlund L. Description of pilot plant used for timber and biofuel drying. Energy 2000;25:1067–79.
[18] Johansson L, Westerlund L. Energy efficient bio fuel drying with an open absorption system parameter study in order to reduce investment costs. Appl Energy 2000;67:231–44. [19] Wang YQ, Lior N. Thermoeconomic analysis of a low-temperature multi-effect thermal desalination system coupled with an absorption heat pump. Energy 2011;36:3878–87. [20] Zhang XD, Hu DP, Li ZY. Performance analysis on a new multi-effect distillation combined with an open absorption heat transformer driven by waste heat. Appl Therm Eng 2014;62:239–44. [21] Wei F, Xiao YH, Zhang SJ. The off-design performance of open cycle absorption heat transformer. J. Eng. Therm. 2009;30:1081–5 [Chinese]. [22] Hyland RW, Wexler A. Formulations for the thermodynamic properties of the saturated phases of H2O from 173.15 to 473.15 K. ASHRAE Trans 1983;89:500–19. [23] Hyland RW, Wexler A. Formulations for the thermodynamic properties of dry air from 173.15 to 372.15 K at pressures to 5 MPa. ASHRAE Trans 1983;89:520–35. [24] Pátek J, Klomfar J. A computationally effective formulation of the thermodynamic properties of LiBr–H2O solutions from 273 to 500 K over full composition range. Int J Refrig 2006;29:566–78.
Contents lists available at ScienceDirect
Energy Conversion and Management journal homepage: www.elsevier.com/locate/enconman
A new open absorption heat pump for latent heat recovery from moist gas Bicui Ye a, Jun Liu a, Xiangguo Xu a, Guangming Chen a,b,⇑, Jiao Zheng b a b
Institute of Refrigeration and Cryogenics, State Key Laboratory of Clean Energy Utilization, Zhejiang University, 38 Zheda Road, Hangzhou 310027, PR China Ningbo Institute of Technology, Zhejiang University, 1 Qianhu South Road, Ningbo 315100, PR China
a r t i c l e
i n f o
Article history: Received 10 November 2014 Accepted 1 February 2015
Keywords: Open absorption Double-stage mode Single-stage mode Latent heat recovery Thermodynamic analysis
a b s t r a c t Conventional drying processes discharge high humidity gas to the atmosphere. The exhaust gas contains large amount of energy. The direct discharging would result in relatively large energy waste. In order to improve the thermal efficiency of drying process, in this paper, a new open absorption heat pump system was proposed, which aims at recovering the latent heat from exhausted moist gas and producing steam for reutilization. The working principle was discussed in detail and thermodynamic models were established to analyze the performance of the new system. The new system can work under both singlestage and double-stage modes. Simulation results showed that the new system could utilize a heat source with lower generation temperature compared with that utilized by a traditional open absorption system. The temperature range of heat source for the double-stage mode is 130–160 °C, and that for the single-stage mode is 160–175 °C. The new system also eliminates the limitation of traditional close absorption system, whose evaporation temperature has to be lower than the dew point temperature of discharged moist gas to recover the latent heat of water steam. Simulation results also indicated an improved COPh of the new system compared with that of double-stage close absorption heat pump system. The COPh of the new system varied from 1.52 to 1.97 and the efficiency of heat recovery varied from 15.1% to 54.8% when the temperature of heat source varied from 135 °C to 175 °C and saturated steam of 100 °C was produced. Ó 2015 Elsevier Ltd. All rights reserved.
1. Introduction Increased energy consumption and greenhouse gas emissions have led to the concerns of the improvement of thermodynamic efficiency of integrated industrial processes. Drying process is widely applied in industries such as pulp and paper manufacturing, wood manufacturing and food processing industry and is one of the most energy-intensive unit operations. It was reported that the energy consumption of drying processes accounted for 9–25% of national energy consumption in the developed countries [1]. Furthermore, during the drying process, more than 70% of the energy consumed is exhausted to the environment in the form of
Abbreviations: AHP, absorption heat pump; A, absorber; A1, primary absorber; A2, secondary absorber; CAHP, close absorption heat pump; COPh, coefficient of performance of heat pump; CR, circulation ratio; G, generator; G1, primary generator; G2, secondary generator; NOAHP, new open absorption heat pump; OAHP, open cycle absorption heat pump; SHX, solution heat exchanger. ⇑ Corresponding author at: Institute of Refrigeration and Cryogenics, State Key Laboratory of Clean Energy Utilization, Zhejiang University, 38 Zheda Road, Hangzhou 310027, PR China. Tel.: +86 13857170531; fax: +86 0571 87951680. E-mail address: [email protected] (G. Chen). http://dx.doi.org/10.1016/j.enconman.2015.02.001 0196-8904/Ó 2015 Elsevier Ltd. All rights reserved.
moist gas [1–6]. This consequently results in a low energy efficiency of the drying process, which is typically in the range of 20–40% [5]. Because of the low thermodynamic-efficiency and high energyconsumption, there is a very large improving space for the drying process. In order to improve the thermodynamic efficiency, researchers have devoted cumulative efforts to recover the latent heat in the discharged moist gas by using the heat integration approaches. Heat exchanger is the most commonly used equipment for heat recovery, which has advantages of low investment and easy operation. In the research of Laurijssen et al. [7], waste heat in the exhaust gas was recovered by using heat exchanger to pre-heat the incoming air and the process water, leading to 15% decrease of the energy consumption in multi-cylinder paper drying sections. However, using heat exchanger to directly recovering the latent heat in the moist gas only could produce hot air with the temperature lower than the dew point of the moist gas. The traditional heat pump system, including both mechanical heat pump system and absorption heat pump system, has emerged as one of the widely accepted technologies integrated with the drying processes. Anderson et al. [8] estimated that the mechanical
B. Ye et al. / Energy Conversion and Management 94 (2015) 438–446
439
Nomenclature d h m p Q t x f k
absolute humidity (kg per kg dry air) enthalpy (kJ/kg) mass flow rate (kg/s) water vapor partial pressure (kPa) heat transfer rate (kW) temperature (°C) solution concentration (wt.) efficiency of heat recovery liquid–gas ratio
Subscripts a1 primary absorber a2 secondary absorber a absorber
heat pump system, which was coupled with common drying schemes used in Swedish sawmills, could decrease heat demand by 5.6 TW h/year at the cost of increased electricity demand by 1 TW h/year. Bakhtiari et al. [9] investigated the optimal integration of absorption heat pumps in a kraft pulping process using Pinch Analysis method. Ahmadi et al. [10,11] optimized an irreversible absorption heat pump system based on a new thermo-ecological criterion. Qu et al. [12] proposed three new configurations of absorption heat pump to recover latent heat in flue gas from natural gas boiler. The simulation results suggested 5–10% improvement in boiler efficiency. Lostec et al. [13] developed a model of wood chip drying process coupled with absorption heat pumps to conduct thermal and economic analysis. Researches [8–15] demonstrated that absorption heat pumps could effectively recover the latent heat of moist gas since the evaporation temperature could be much lower than the dew point temperature of the moist gas. However, the low evaporation temperature reduced the COPh of heat pump systems at the same time. To overcome the limitation of traditional close absorption system, whose evaporation temperature should be lower than the dew point temperature of discharged moist gas to recover the latent heat of water vapor, Westerlund et al. [16–18] proposed an open cycle absorption heat pump (OAHP) system to recover the latent heat from the moist gas. As shown in Fig. 1, the OAHP system consists mainly of three parts, absorber, generator and condenser. The moist gas contacts directly with the absorbent in the absorber. Water vapor in the moist gas is absorbed by the absorbent due to the water vapor partial pressure difference between moist gas and absorbents, giving off the latent heat of vapor to directly heat the moist gas, which could be reused by the drying process. Therefore, the latent heat of vapor could be recovered with
Fig. 1. Schematic diagram of traditional OAHP system [16].
c envir eva g1 g2 ge hg hp in max mg out recovered
condenser environment evaporation primary generator secondary generator generation high-pressure generator heat pump input maximum moist gas output recovered from the moist gas
the temperature much higher than the dew point temperature of moist gas. In the research of Wang and Zhang et al. [19,20], the discharged gas was heated from 70 °C to 125 °C in the absorber of the OAHP system and used to produce steam at 120 °C. Anderson et al. [8] demonstrated that the OAHP system could decrease energy consumption by 67.4% for sawmill drying process. Although the OAHP system performs well at recovering the latent heat of moist gas, to produce high temperature hot air or steam for drying process using the OAHP system would require an even higher generating temperature, which limits its application [16–21]. Consequently, a new OAHP system (NOAHP) is proposed in this paper, which could be operated at double-stage mode to utilize low temperature heat source for generator or single-stage mode to improve system energy efficiency when high temperature heat source for generator is available. In order to validate the feasibility of the NOAHP system, thermodynamic model was established to analyze the system performance based on the properties of chose absorbent and the mass and energy balance of each component. The performance of the NOAHP system was also compared to that of traditional open absorption and traditional close absorption heat pump system.
2. The description of the NOAHP system 2.1. The working process of the NOAHP system Fig. 2 shows the schematic of the new open absorption heat pump (NOAHP) system. It consists of two generators (primary and secondary generator G1, G2), two absorbers (primary and secondary absorber A1, A2), condenser and solution heat exchanger (SHX). The working process of the NOAHP system is described as follows. Firstly, the open absorption process, which occurs in A1, is a main feature of the new system. It is also the main difference between an open absorption cycle and a close absorption cycle. The water vapor flows into the absorber A1 from outside in terms of moisture in the moist gas and is discharged to outside in terms of condensate at condenser (or directly outputted from generator G1 as hot steam), instead of circulating in the system. This is a main reason why the system is called open cycle absorption. In A1, the moist gas contacts directly with strong solution and the moisture in the moist gas is absorbed when the water vapor partial pressure of the strong solution is lower than that of the moist gas. Latent heat given off during the absorption process is transported to water in the heat exchanger, which is installed in A1.
440
B. Ye et al. / Energy Conversion and Management 94 (2015) 438–446
Fig. 2. Schematic diagram of the NOAHP system.
Secondly, the water vapor absorbed by the weak solution in absorbers was heated and evaporated from the weak solution in G1 and G2. Water vapor generated in G1 could be directly outputted as high-temperature steam (in this case, condenser is not needed anymore), or flow to the condenser and give out condensing heat to produce steam/hot water. Water vapor generated in G2 flows into A2, where the vapor is absorbed by the weak solution from A1 and absorption heat is provided for producing hot water/air or steam. Finally, the solution circulates in the system. The strong solution is diluted in A1 by water vapor in the moist gas (3–4). Then the solution is pumped into A2, where it is further diluted by water vapor from G2 (4–5–6). Afterward, weak solution from A2 flows through SHX before it is regenerated in G1 and further regenerated in G2 (6–7–8–9–10–11). Energy from heat sources is supplied to G1 and G2 in order to regenerate weak solution. Strong solution from G2 also flows through SHX to exchange heat with weak solution before it flows into A1 (11–12–3). Fig. 3. p–T diagram of H2O/LiBr in the NOAHP system and OAHP system.
2.2. The fundamental principle of the NOAHP system H2O/LiBr serves as the working pair in this study. Fig. 3 is the p– T diagram of H2O/LiBr in the NOAHP system (the continuous line) and in the traditional OAHP system (the dash line). In Fig. 3, working conditions for the two systems are assumed as follows: the two systems are used to produce steam with the same pressure pout and the same temperature tout. tc is higher than tout to form a temperature difference for heat transfer and is also the same for two systems. Same condensation temperatures correspond to same condensation pressures, which approximately equal to the pressure of generator (for the NOAHP system, it is the pressure of G1, pg1). Therefore, pg1 in the NOAHP system equals to the pressure of generator in traditional OAHP system. As shown in Fig. 3, process of 8–9 is the generation process in G1 in the NOAHP system and process of 8–0 is the generation process in OAHP system. The concentration order of solution is the concentration of solution from G2, which is represented by xg2, and those of solutions from
G1, A1 and A2 are, xg1, xa1, and xa2, respectively. Solution from G1 for the NOAHP system and solution from generator for traditional OAHP system are of the same concentration to ensure a same absorption concentration of solution entering the open absorbers. The p–T lines of H2O/LiBr in Fig. 3 indicate that the NOAHP system can utilize a lower generation temperature than traditional OAHP system does. The generation temperature for OAHP system is approximately 162 °C in order to produce steam at pressure of pg1 and to regenerate solution at concentration of xg2 (as point 0 marked in Fig. 3). However, the regeneration process in the NOAHP system is conducted in two steps. G1 produces steam at pressure of pg1 and solution at concentration of xg1, which is lower than xg2 (as point 9 marked in Fig. 3). Consequently, the required generation temperature for G1 is only 146 °C. G2 further regenerates solution at concentration of xg2, and produce steam at pressure of pg2 (as
B. Ye et al. / Energy Conversion and Management 94 (2015) 438–446
point 11 marked in Fig. 3), which is lower than pg1. Thus, the heat source temperature for G2 is approximately 148 °C. Furthermore, the generation temperature could be even lower for G2 as pg2 decreases. On the other hand, the NOAHP system could work under two modes: single-stage mode and double-stage mode. When heat source temperature for G1 is high enough to produce solution with concentration of xg2, the heat supply for G2 could be turned off. In this case, G2 and A2 do not work and the NOAHP system works under single-stage mode. Otherwise, the NOAHP system works under double-stage mode when heat source temperature is not high enough to drive the single-stage mode. The switch between the single-stage mode and the double-stage mode can be simply realized by the automatic on–off control of the heat supply to G2. The practical operation of the NOAHP system would be further investigated in related studies. To optimize the performance of the NOAHP system, thermodynamic simulation is conducted in Sections 3 and 4. Section 5 compares the performance of NOAHP system with that of traditional close absorption heat pump system. 3. Thermodynamic model of the NOAHP system To simplify the model, the following assumptions are made in this study. (1) The system is performed in steady-state conditions; (2) Temperatures of heat sources for two generators G1 and G2 are the same; (3) The solution from G1, G2, A1 and A2 is saturated; (4) A temperature efficiency of SHX (the ratio between the temperature difference for the strong solution (t12–t11) and the theoretical available temperature difference (t11–t7)) is assumed at 80%; (5) A temperature difference of 10 °C is assumed between the heat source temperature and output temperature of generators; (6) Thermal leaks, solution loss and pressure drops are neglected; (7) The solution pump power and fan power are neglected. Based on above assumptions, mass and energy conservation equations are applied to the main components in the NOAHP system. (Numeric subscripts in following equations refer to numbers marked in Fig. 2.) The first law of thermodynamics of the new OAHP system is formulated as below:
Q g1 þ Q g2 þ Q a1;in ¼ Q a1;out þ Q a2 þ Q c
mdryair ðd2 d1 Þ ¼ m4 m3
Secondary absorber A2
mg2;watervapor ¼ m13 ¼ m6 m5
ð6Þ
m5 x5 ¼ m6 x6
ð7Þ
Q a2 þ m6 h6 ¼ mg2;watervapor h13 þ m5 h5
ð8Þ
Primary generator G1
mg1;watervapor ¼ m14 ¼ m8 m9
ð9Þ
m7 x7 ¼ m9 x9
ð10Þ
Q g1 þ m8 h8 ¼ m9 h9 þ mg1;watervapor h14
ð11Þ
Secondary generator G2
mg2;watervapor ¼ m13 ¼ m10 m11
ð12Þ
m10 x10 ¼ m11 x11
ð13Þ
Q g2 þ m10 h10 ¼ m11 h11 þ mg2;watervapor h13
ð14Þ
Condenser
m14 ¼ m15
ð15Þ
Q c ¼ mwatervapor ðh14 h15 Þ
ð16Þ
Since the mass flow rate of solution remains constant when flowing through pumps, valves and heat exchangers:
m3 ¼ m11 ¼ m12
ð17Þ
m6 ¼ m6 ¼ m8
ð18Þ
m9 ¼ m10
ð19Þ
m4 ¼ m5
ð20Þ
According to mass conservation, the amount of water vapor absorbed in A1 equals to that of water vapor generated at G1, therefore, it has:
mg1;watervapor ¼ m14 ¼ mdryair ðd2 d1 Þ
ð21Þ
Circulation ratio, CR, is defined as the mass of water vapor evaporated from G2 divided by the mass of water vapor evaporated from G1.
CR ¼
ð1Þ
where Qa1,in represents energy released by the moist gas, and Qa1,out represents energy obtained by the water in the heat exchanger, which is installed in A1. Qa2 represents absorption heat obtained by the water in the heat exchanger, which is installed in A2. Qc represents condensation heat obtained by the water in the heat exchanger, which is installed in the condenser. Qg1 and Qg2 represent the generation heat supplied to G1 and G2, respectively. The balance equations of each component are shown respectively as follows: Primary absorber A1
441
mg2;watervapor m13 ¼ mg1;watervapor m14
ð22Þ
The liquid–gas ratio k is defined as the mass of concentrated solution flowing into A1 divided by the mass of dry air flowing through A1.
k¼
m3 mdryair
ð23Þ
For thermodynamic performance evaluation, the coefficient of performance COPh is defined as:
COPh ¼
ð2Þ
Q a1;out þ Q a2 þ Q c Q g1 þ Q g2
ð24Þ
The efficiency of heat recovery from the moist gas f is defined as:
m3 x3 ¼ m4 x4
ð3Þ
Q a1;in ¼ mdryair ðh1 h2 Þ
ð4Þ
Q a1;in þ m3 h3 ¼ Q a1;out þ m4 h4
ð5Þ
f¼
mdryair ðh1 h2 Þ Q recovered ¼ Q max mdryair ðh1 henv ir Þ
ð25Þ
where Qrecovered represents energy recovered from the moist gas, Qmax represents the overall energy released when the moist gas is
442
B. Ye et al. / Energy Conversion and Management 94 (2015) 438–446
cooled to the environmental condition, and henvir represents enthalpy of gas under the environmental condition. Properties of the moist gas, including absolute humidity, relative humidity, water vapor partial pressure, temperature and enthalpy, are given by Hyland and Wexler [22,23]. Properties of H2O/LiBr, including concentration, temperature, water vapor partial pressure and enthalpy, are given by Pátek and Klomfar [24]. Thermodynamic calculations were conducted by EES (Engineering Equation Solver, Professional Version 8.889, 06/25/11). 4. Parametric analysis of the NOAHP system In order to provide reliable performance evaluation for the NOAHP system, drying process of paper manufacturing is used as the base-case condition in this study [12]. Energy of moist gas from the drying process of paper manufacturing is recovered by the NOAHP system, which produces required steam for the drying process. Base-case conditions are shown in Table 1. Simulation results indicate that COPh is 1.737 and efficiency of heat recovery f achieves 50.16% in the base-case condition. Since real operation conditions may vary over a wide range, parametric analysis is discussed below to optimize the performance of the NOAHP system. Main factors affecting the system performance are pressure of G2, pg2, temperature of heat source tge, absolute humidity of moist gas, dmg, and temperature of moist gas, tmg. The influence of each factor on COPh and f are given in Figs. 4–9. Operational parameters are kept constant at the basecase values (Table 1), except for parameters that marked in the figures.
Fig. 4. Effects of pg2 on COPh and f.
4.1. Effects of the pressure of G2 on COPh and f Fig. 5. Effects of pg2 on k and CR.
pg2 is an adjustable operational parameter, which could be controlled by the throttling valve between G1 and G2, in the NOAHP system under the double-stage mode. Fig. 4 shows the relationship between pg2 and performance parameters of the NOAHP system, which provides a basis for the optimization of pg2. It can be seen that with the increase of pg2, COPh initially rises and then peaks at a certain pg2. However, f decreases rapidly with the increase of pg2. COPh and f cannot be optimized simultaneously by adjusting
pg2. Considering energy from moist gas is recovered at the cost of even higher temperature energy, working conditions with a low COPh and a high f do not make sense from the energetic view. Therefore, the NOAHP system recovers energy from moist gas with the aim of consuming less high-temperature energy. Thus, the pg2, which corresponding to the maximum COPh, is defined as the optimized pg2.
Table 1 Operational parameters and performance of the NOAHP system at the base-case condition. Temperature (°C)
Water vapor partial pressure (kPa)
Mass flow rate (kg/s)
0.010 0.3031 0.1326
1.303 1.141
20 80 105
Steam Steam Steam Steam Steam Steam
100 100 150 150 100
Pressure (kPa) 101.3 101.3 112.7 64.97 101.3
0.1692 0.0519 0.1704 0.0501 0.1715
Heat source For G1 For G2
160 160
617.7 617.7
0.2116 0.0623
Absorbent (H2O/LiBr) Solution from G2 (11) Solution from A1 (4,5) Solution from A2 (6,7) Solution from G1 (9,10) Solution from SHX (12,3) Solution from SHX (8)
150 105 105 150 114 113
Water vapor partial pressure (kPa) 64.97 54.78 64.97 112.7 17.80 117.5
from from from from from
A1 A2 G1 (14) G2 (13) Conderser
Base-case performance COPh Efficiency of heat recovery f
1.737 0.5016
1.637 33.19 17.80
Absolute humidity (kg/kg)
Ambient conditions (0) Moist gas (1) Exhaust gas (2)
Concentration (wt.%) 0.6356 0.4460 0.41 0.5649 0.6356 0.41
0.4008 0.5712 0.6214 0.4509 0.4008 0.6214
B. Ye et al. / Energy Conversion and Management 94 (2015) 438–446
443
Fig. 6. Effects of pg2 on heat transfer rate of each component.
Fig. 7. Influence of tge on COPh and f.
f decreases rapidly with the increase of pg2. The concentration of solution entering A1, i.e., x3, decreases with the increase of pg2 according to the property of H2O/LiBr. Water vapor pressure of H2O/LiBr increases with the decrease of x3, leading to the drop of the mass transfer driving force between solution and moist gas. f thus decreases. The variation of COPh with pg2 can be explained as follows. Firstly, the decrease of f means the decrease of water vapor absorbed from the moist gas, i.e., mrecover, which equals to m14. The decrease of m14 leads to the decrease of Qa1,out, Qg1 and Qc. However, as shown in Fig. 5, CR decreases with the increase of pg2, which indicates the decrease of m13 is more dominant than the decrease of m14. The decrease of m13 leads to the decrease of Qg2 and Qa2. As shown in Fig. 6, Qa1,in, Qa1,out, Qg1 and Qc decrease almost linearly with the increase of pg2, whereas, Qg2 and Qa2 decreases along a parabola line. When pg2 is below 80 kPa, the decrease of Qg2 and Qa2 is faster than the decrease of Qa1,out, Qg1 and Qc, COPh increases. When pg2 increases to 80 kPa, Qg2 and Qa2 remains almost the same, the decrease of Qa1,out, Qg1 and Qc leads to the decrease of COPh. Secondly, the drop of liquid–gas ratio as shown in Fig. 5 leads to the increase of COPh since less generation heat is used to heat the solution. In sum, to get a maximum COPh for a given working condition, pg2 should be optimized. It is noteworthy that the value of COPh and f in the following analysis section are all calculated with optimized pg2.
to the decrease of x3. Therefore, the upward trend for f in the double-stage mode is smaller than that in the single-stage mode. The increase of f results in the increase of Qa1,out, Qg1 and Qc. On the other hand, the increase of pg2 leads to the decrease of concentration difference between solutions from G1 and G2, which indicates the decrease of m13. The decrease of m13 leads to the decrease of Qg2 and Qa2. The increase of Qa1,out and Qc is larger than the decrease of Qa2, leading to the increase of output energy from the NOAHP system. The increase of Qg1 and decrease of Qg2 leads to the slightly increase of input energy. As a result, COPh increases.
4.2. Effects of temperature of heat source on COPh and f The NOAHP system works under double-stage mode within the temperature range of 130–160 °C and works under single-stage mode within a higher temperature range of 160–175 °C. When tge is below 160 °C, the concentration of solution from G1 is not strong enough to absorb water vapor in the moist gas. The solution has to be further regenerated in G2. Thus, the NOAHP system works under the double-stage mode. When tge is above 160 °C, solution from G2 is close to the crystallization condition. In this case, the heat supplied to G2 should be closed. Single-stage mode therefore activates. As shown in Fig. 7, f increases with tge. Under the single-stage mode, concentration of solution from G1, i.e., x10, increases with the increase of tge. Under the double-stage mode, concentration of solution from G2, i.e., x3 also increases. The increases of x10 and x3 make the mass transfer driving force between the solution and moist gas in A1 increases. f thus increases with the increase of tge under both single-stage and double-stage modes. However, the optimized pg2, which corresponds to the maximum COPh under the double-stage mode, increases with the increase of tge, leading
4.3. Effects of absolute humidity of moist gas on COPh and f As shown in Fig. 8, both COPh and f increase with dmg. This is because the increase of dmg leads to the increase of water vapor partial pressure of moist gas, which increases the driving force between moist gas and solution in A1. Water vapor absorbed from the moist gas thus increases, which makes f and COPh increases. With the increase of mass transfer driving force between solution and moist gas, a lower absorption concentration of solution, i.e., x3, is allowed for the NOAHP system without decreasing the heat recovery efficiency. Therefore, the increase of the absolute humidity of moist gas can decrease the required temperature of the heat source for the NOAHP system. This finding confirms that the open absorption system is of particular interest when dealing with high humidity moist gas. 4.4. Effects of temperature of moist gas on COPh and f It can be seen from Fig. 9 that as tmg increases, f decreases slightly and COPh increases slightly. This is because the water
Fig. 8. Effects of dmg on COPh and f.
444
B. Ye et al. / Energy Conversion and Management 94 (2015) 438–446
Fig. 9. Effects of tmg on COPh and f.
vapor partial pressure of moist gas is only relevant to the absolute humidity, and has nothing to do with the gas temperature. Thus, the mass transfer driving force between moist gas and solution in A1 would remain constant with the increase of tmg. However, the increase of gas temperature decreases the heat transfer driving force between the moist gas and the solution, which impedes the heat transfer between the solution and heat exchanger installed in A1. Solution temperature thus increases, along with the increase of water vapor partial pressure of solution. This directly leads to the slight decrease of f. The slight decrease of f leads to the slight decrease of Qa1,out, Qg1 and Qc. On the other hand, corresponding pg2 increases with the increase of tmg, which decreases the mass of water vapor evaporated from G2, i.e., m13, leading to the decrease of Qg2 and Qa2. COPh is finally increased. Since the gas temperature has limited effect on f, the increase of COPh is mainly a result of the increase of pg2. In sum, simulation results above demonstrate that the NOAHP system performs well at recovering energy, especially the latent heat, from moist gas. The temperature of gas has limited effect on the performance of the NOAHP system while the absolute humidity of gas has significant effect. On the other hand, the NOAHP system can be operated with lower temperature heat source when the absolute humidity of gas is increased. Although efficiency of heat recovery f cannot be simultaneously optimized with the optimization of COPh, it can still achieve about 20–50%, even under a high absorption temperature (when 100 °C steam is produced). The optimized COPh can achieve from 1.19 to 1.97 when temperature of heat source varies from 130 to 175 °C.
sidering that efficiency of heat recovery f is strongly influenced by teva, the performance of double-stage CAHP is evaluated at different teva. Thirdly, pg2 of the double-stage CAHP system and pg2 of the NOAHP system have been optimized in the discussion conditions. f of the double-stage CAHP system and the NOAHP system are compared as shown in Fig. 10. It can be seen that f of the double-stage CAHP system is only influenced by peva. f is 12.5%, 26.3% and 43.9%, respectively when peva is 24.4 kPa, 21.3 kPa and 13.25 kPa and corresponding teva is 59.44 °C, 57.44 °C and 51.44 °C. On the other hand, f of the NOAHP system increases with tge as discussed in Section 4.2. It varies from 15.1% to 54.8% when tge increases from 135 °C to 175 °C. Fig. 11 shows that the NOAHP system has a much higher COPh than the double-stage CAHP system does. COPh of the NOAHP system increases from 1.52 to 1.97 as discussed in Section 4.2. COPh of the double-stage CAHP system varies from 1.30 to 1.43 with the increase of tge. In addition, COPh of the double-stage CAHP system also decreases slightly with the increase of peva. Table 2 shows the heat supplied (positive value) and rejected (negative value) by the components of the double-stage CAHP system and the NOAHP system at different temperatures of heat source. In the double-stage CHAP system, since the evaporation temperature has to be lower than the dew point temperature of moist gas, the corresponding evaporation pressure, i.e., peva, has to be low. This leads to a large pressure difference between the evaporator and high-pressure generator. Lots of energy has to be consumed to pressurize the water vapor circulating in the double-stage CAHP system from peva to the pressure of high-pressure generator phg. On the other hand, in the NOAHP system, due to the direct contact
Fig. 10. f of the double-stage CAHP system and the NOAHP system.
5. Thermodynamic performance comparison of the NOAHP system and traditional CAHP system As discussed in Sections 2 and 4, the NOAHP system can work under single-stage or double-stage mode. Correspondingly, both single-stage cycle traditional close absorption heat pump systems (CAHP) and double-stage cycle CAHP have been used to recover waste heat of moist gas [10–15]. Lostec et al. [13] found that under the given heat source (152 °C): (1) the single-stage CAHP was only effective when the output temperature was below 60 °C; (2) the double-stage CAHP should be used when the output temperature was from 60 to 100 °C. Since the NOAHP system also aims at a high output temperature, the performance of NOAHP system is compared with that of the double-stage CAHP system in this section. The discussions are based on the following facts. Firstly, doublestage CAHP system recovers waste heat by cooling and dehumidifying the moist gas, i.e., the evaporation temperature has to be lower than the dew point temperature teva < tdp. Secondly, con-
Fig. 11. COPh of the double-stage CAHP system and the NOAHP system.
445
B. Ye et al. / Energy Conversion and Management 94 (2015) 438–446 Table 2 Heat transfer rates of each components in the double-stage CAHP system and the NOAHP system. tge (°C)
Qeva (kW)
Qa1 (kW)
Qa2 (kW)
Qg1 (kW)
Qg2 (kW)
Qcon (kW)
Efficiency f
COPh
Double-stage CAHP system 140 232.8 150 232.8 160 232.8 170 232.8
271.3 261.4 252.6 244.6
306.2 299.2 301.8 299.3
275.1 264.5 253.8 244.8
310.1 303.9 308.3 306.8
240.6 240.6 240.6 240.6
0.2634 0.2634 0.2634 0.2634
1.398 1.410 1.414 1.422
tge (°C)
Qa1,out (kW)
Qa2 (kW)
Qg1 (kW)
Qg2 (kW)
Qcon (kW)
Efficiency f
COPh
286.1 378.5 357.7 528.2
140.6 65.95 4.34 0
326.9 425.8 413.1 595.4
146.7 71.15 4.783 0
346.9 461.5 464.7 660.9
0.208 0.280 0.404 0.548
1.633 1.823 1.979 1.997
Qa1,in (kW)
The NOAHP system 140 300 150 409 160 408.9 170 593.7
with moisture gas at environmental pressure, the pressure in A1 is much higher than peva of CHAP system. If phg in the double-stage CAHP system equals to pg1 in NOAHP system, the energy consumed to pressurize the water vapor absorbed in A1 to pg1 is relatively less, which is beneficial for the NOAHP system to achieve a higher COPh than that of the double-stage CAHP system. From the energetic viewpoint, in the double-stage CAHP system, the energy released in high-pressure absorber approximately equals to the energy supplied to low-pressure generator because water vapor evaporated from low-pressure generator is completely absorbed in the high-pressure absorber. Meanwhile, the energy released in high-pressure absorber also approximately equals to the energy supplied to high-pressure generator because water vapor absorbed in high-pressure absorber is totally evaporated in the high-pressure generator. For the same reason, energy released from low-pressure absorber approximately equals to energy supplied to low-pressure generator and energy supplied to high-pressure generator approximately equals to energy released in the condenser. Therefore, if neglecting the energy loss of the whole cycle, the maximum COPh of the double-stage CAHP system would be 1.5. On the other hand, the maximum COPh of the NOAHP system is 2 under the single-stage mode since energy supplied to G1 is used to produce energy released both in A1 and the condenser. And the theoretical minimum COPh of 1.5 would be achieved when the NOAHP system works under the double-stage mode. Thus, COPh of the NOAHP system is higher than that of the double-stage CAHP system. 6. Conclusion A new open absorption heat pump (NOAHP) system has been proposed to be applied in the heat recovery of moist gas in this study. The new system can utilize a wider temperature range of heat source by the innovation of system structure, which allows the new system to work under two operational modes. When temperature of heat source is high enough (160–175 °C), NOAHP system works under single-stage mode to ensure a high COP, and when temperature of heat source is not high enough to drive the single-stage mode, double-stage mode is operated to utilize heat source with a lower temperature range (130–160 °C). Simulation results show that under the base-case condition, COPh can achieve 1.737 and efficiency of heat recovery f achieves 50.16% when 100 °C saturated steam is produced. The thermodynamic parameter analysis of the new system shows that: (1) pressure of G2 pg2 could be optimized to attain the maximum COPh; (2) increasing temperature of heat source improve COPh and efficiency of heat recovery;
(3) temperature of the moist gas has little impact on performance of the new system; (4) increasing absolute humidity of the moist gas improve performance of the new system. The parameter analysis suggests that the NOAHP system is feasible for energy recovery of moist gas with high absolute humidity and even a low temperature. The performance comparison between the NOAHP system and traditional double-stage CAHP system shows that COPh of the new system is much higher than that of the double-stage CAHP system. COPh of the new system achieves from 1.52 to 1.97 while COPh of the double-stage CAHP system is only from 1.30 to 1.43. The NOAHP system can be applied to any industrial applications where moist gas is discharged and wasted. It recovers energy of moist gas and produces high-temperature steam for re-utilization. The application of the NOAHP system can thus improve energy efficiency for many industrial unit operations. Further work of the new OAHP system will mainly cover: (1) exploring less corrosive absorbent, which is suitable for the open absorption structure; (2) improving structure of open absorber to improve absorption efficiency; (3) carrying out experimental work of the new system, including the control of single-stage mode and double-stage mode and the use of anticorrosive material. Acknowledgements This work was financially supported by the major program of the National Natural Science Foundation of China (50890184) and major program of the Industrial Technology Innovation Program of Ningbo (2013B10029). References [1] Misha S, Mat S, Ruslan MH, Sopian K. Review of solid/liquid desiccant in the drying applications and its regeneration methods. Renew Sustain Energy Rev 2012;16:4686–707. [2] Goh LJ, Othaman MY, Mat S, Ruslan H, Sopian K. Review of heat pump systems for drying applications. Renew Sustain Energy Rev 2011;15:4788–96. [3] Jana AK. Advances in heat pump assisted distillation column: a review. Energy Convers Manage 2014;77:287–97. [4] Law R, Harvey A, Reay D. Opportunities for low-grade heat recovery in the UK food processing industry. Appl Therm Eng 2013;53:188–96. [5] Minea V. Heat pump assisted drying: recent technological advances and R&D needs. Drying Technol 2013;31:1177–89. [6] Hepbasli A, Biyik E, Ekren O, Gunerhan H, Araz M. A key review of wastewater source heat pump (WWSHP) systems. Energy Convers Manage 2014;88:700–22. [7] Laurijssen J, Gram FJ, Worrell E, Faaij A. Optimizing the energy efficiency of conventional multi-cylinder dryers in the paper industry. Energy 2010;35:3738–50. [8] Anderson JO, Westerlund L. Improved energy efficiency in sawmill drying system. Appl Energy 2014;113:891–901. [9] Bakhtiari B, Fradette L, Legros R, Paris J. Opportunities for the integration of absorption heat pumps in the pulp and paper process. Energy 2010;35:4600–6.
446
B. Ye et al. / Energy Conversion and Management 94 (2015) 438–446
[10] Ahmadi MH, Ahmadi MA, Mehrpooya M, Feidt M. Thermodynamic optimization of Stirling heat pump based on multiple criteria. Energy Convers Manage 2014;80:319–28. [11] Ahmadi MH, Ahmadi MA, Mehrpooya M, Sameti M. Thermo-ecological analysis and optimization performance of an irreversible three-heat-source absorption heat pump. Energy Convers Manage 2015;90:175–83. [12] Qu M, Abdelaziz O, Yin HX. New configurations of a heat recovery absorption heat pump integrated with a natural gas boiler efficiency improvement. Energy Convers Manage 2014;87:175–84. [13] Lostec BL, Galanis N, Baribeault J, Millette J. Wood chip drying with an absorption heat pump. Energy 2008;33:500–12. [14] Hamdan MA, Rossides SD, Khalil RH. Thermal energy storage using thermochemical heat pump. Energy Convers Manage 2013;65:721–4. [15] Chaturvedi SK, Gagrani VD, Abdel-Salam TM. Solar-assisted heat pump – a sustainable system for low-temperature water heating applications. Energy Convers Manage 2014;77:550–7. [16] Westerlund L, Hermansson R, Fagerstrom J. Flue gas purification and heat recovery: a biomass fired boiler supplied with an open absorption system. Appl Energy 2012;96:444–50. [17] Johansson L, Westerlund L. Description of pilot plant used for timber and biofuel drying. Energy 2000;25:1067–79.
[18] Johansson L, Westerlund L. Energy efficient bio fuel drying with an open absorption system parameter study in order to reduce investment costs. Appl Energy 2000;67:231–44. [19] Wang YQ, Lior N. Thermoeconomic analysis of a low-temperature multi-effect thermal desalination system coupled with an absorption heat pump. Energy 2011;36:3878–87. [20] Zhang XD, Hu DP, Li ZY. Performance analysis on a new multi-effect distillation combined with an open absorption heat transformer driven by waste heat. Appl Therm Eng 2014;62:239–44. [21] Wei F, Xiao YH, Zhang SJ. The off-design performance of open cycle absorption heat transformer. J. Eng. Therm. 2009;30:1081–5 [Chinese]. [22] Hyland RW, Wexler A. Formulations for the thermodynamic properties of the saturated phases of H2O from 173.15 to 473.15 K. ASHRAE Trans 1983;89:500–19. [23] Hyland RW, Wexler A. Formulations for the thermodynamic properties of dry air from 173.15 to 372.15 K at pressures to 5 MPa. ASHRAE Trans 1983;89:520–35. [24] Pátek J, Klomfar J. A computationally effective formulation of the thermodynamic properties of LiBr–H2O solutions from 273 to 500 K over full composition range. Int J Refrig 2006;29:566–78.