Numerical study of a hybrid absorption-compressionhigh temperature heat pump for industrial waste heat recovery

Abstract

The present paper aims at exploring a hybrid absorption-compressionheat pump (HAC-HP) to upgrade and recover the industrial waste heatin the temperature range of 60°C–120°C. The new HAC-HPsystem proposed has a condenser, an evaporator, and one more solutionpump, compared to the conventional HAC-HP system, to allow flexibleutilization of energy sources of electricity and waste heat. In thesystem proposed, the pressure of ammonia-water vapor desorbed in thegenerator can be elevated by two routes; one is via the compressionof compressor while the other is via the condenser, the solution pump,and the evaporator. The results show that more ammonia-water vaporflowing through the compressor leads to a substantial higher energyefficiency due to the higher quality of electricity, however, onlya slight change on the system exergy efficiency is noticed. The temperaturelift increases with the increasing system recirculation flow ratio,however, the system energy and exergy efficiencies drop towards zero.The suitable operation ranges of HAC-HP are recommended for the wasteheat at 60°C, 80°C, 100°C, and 120°C. The recirculationflow ratio should be lower than 9, 6, 5, and 4 respectively for thesewaste heat, while the temperature lifts are in the range of 9.8°C–27.7°C, 14.9°C–44.1 °C, 24.4°C–64.1°C,and 40.7°C–85.7°C, respectively, and the system energyefficiency are 0.35–0.93, 0.32–0.90, 0.25–0.85,and 0.14–0.76.

Keywords

absorption compression / hightemperature heat pump / efficiency / industrial waste heat / thermodynamic analysis

Introduction

The industrial sector consumes about54% of the total world’s delivered energy, and what is worse is thatabout one sixth of the total energy consumed by industrial sectoris wasted as low-grade heat via radiation, exhausted gas, coolingfluid and so on . This energy can be recovered byvarious technologies, such as power generation via the organic Rankinecycle [1], refrigerationby absorption [2] andadsorption [3] technologies,and heat upgrade by heat pump [4], etc. Among these technologies, heat pump can efficiently pumpthe heat form low grade to high grade, and has been recognized asan efficient and practical solution to reduce greenhouse gas emission[5].

Compared to the conventional vapor-compressionheat pump, the hybrid absorption-compression heat pump (HAC-HP) isbelieved to be able to work for high temperature applications witha better system performance [6]. The HAC-HP system is a combination of the conventional vapor-compressionsystem and the absorption system, in which the use of binary mixtureas working fluid leads to the non-isothermal process of absorption/desorptionand relatively lower temperature difference between the heat sourceand the working fluid compared to the isothermal condensation/evaporationprocess in the conventional vapor-compression heat pump, thereby thesystem energy efficiency can increase due to the lower cycle irreversibilityand entropy generation. Some other benefits of using this hybrid systemare small swept volume of the compressor, higher heat transfer coefficient,environment-friendly refrigerant, more flexiblity in the changes oftemperature and capacity, and higher delivery temperature [7].

Brunin et al. [6] compared the working domains ofdifferent heat pump technologies, and concluded that the vapor-compressionsystem using hydrocarbon fluid and the hybrid system using ammonia-waterwere the only two technologies of high temperature heat pump as theabolishing of CFC and HCFC refrigerant. Minea and Chiriac [8] presented some design guidelinesand operation experiences of the ammonia-water based HAC-HP system,and tested a 4.5 MW prototype in field and demonstrated the heat upgradefrom 36°C industrial cooling water to 55°C useful hot waterwith a COP of 3.9. Kim et al. [9] experimentally tested a 10 kW HAC-HP system using ammonia-wateras working fluid, and elevated the heat from 50°C to over 90°Cwith ammonia mass fraction in the weak solution of 0.42. Jensen etal. [10] numerically investigatedthe influences of ammonia mass fraction and liquid circulation ratioon the system constrains, and the results showed that the maximumheat supply temperature was 111°C when using standard refrigerationcomponents without modifying the compressor (28 bar limitation); forhigh pressure ammonia based components, the maximum supply temperaturecould be 129°C (50 bar limitation) and for transcritical CO₂ based components, the maximum supply temperature couldbe 147°C (140 bar limitation). Jensen et al. [11] also thermodynamically analyzedthe HAC-HP system to recover waste heat from spray-drying facilities,and conducted exergoeconomic analysis to minimize the lifetime cost.The best solution they reported was an 895 kW heat pump with ammoniamass fraction at 0.82 and a circulation ratio of 0.43. This systemcould generate an economic saving of €146,000 and an annual CO₂ emission reduction of 227 ton. Bourouis et al. [12] studied a single stage HAC-HP systemusing ternary mixture of Trifluoroethanol-water-tetraethylenglycoldimethylether(TFE-H₂O-TEGDME) as working fluid, and theresults showed that the system could upgrade the waste heat form 80°Cto 120°C at a COP of 6.4.

The conventional HAC-HP systems mentionedabove are all based on a similar configuration with the main componentsof generator, absorber, solution pump, and compressor. Such systemuses electricity as the main driven force to upgrade the low-gradewaste heat. Though the system COP is high, the operation cost of theconventional HAC-HP system is still proportionally increased withthe amount of recovered waste heat. The current paper proposes toadd a condenser, an evaporator, and one more solution pump to thesystem to form a HAC-HP system using the compressor and the condenser-pump-evaporatortogether to elevate the pressure of the working fluid. There are twobenefits of employing such a dual energy source system. First, thesystem can be more flexible on the use of different energy source,i.e., electricity and/or waste heat, depending on their availabilities.Secondly, more waste heat can be recovered as the system requiredthe thermal inputs of both the generator and the evaporator whilethe required electricity can be reduced. The energy efficiency, exergyefficiency, temperature lift, and useful heat power are numericallycalculated and analyzed.

Working principle and analysis method

Ammonia-water was used as the workingfluid in the current study. The correlations developed by El-Sayedand Tribus [13] were usedto determine the equilibrium liquid and vapor states of ammonia-watermixture, while Gibbs free energy formulations reported by Zieglerand Trepp [14] were usedto calculate the enthalpies and entropies of ammonia-water mixture.The following hypothesises were used to simplify the numerical analysis:The cycle was operated in a steady-state; the solutions at the outletof the condenser and the evaporator were in a saturated state; thesolutions at the outlet of generator and absorber were in a saturatedstate; the vapor from the rectifier was in a saturated state; thethrottling does not change the solution enthalpy; after throttling,the solution entering the generator was at two-phase or in a saturatedliquid state; and the pressure drop and heat loss in the system wereboth negligible.

The schematic diagram of the HAC-HPsystem studied is shown in Fig. 1. The system consists of a columnshape generator, a rectifier on the top of the generator, a recuperator,an absorber, a condenser, an evaporator, a compressor, two solutionpumps, and a throttling valve.

There were two pressure levels inthe whole cycle, a low pressure P_L in the generator, rectifier and condenser,and a high pressure P_H in the evaporator, absorber and recuperator. In thecurrent study, the refrigerant ammonia mass fraction (after the rectifier),w_ref, was pre-definedat 0.9995 as suggested [15]. According to the predefined assumption, the ammonia-rich vaporat the outlet of the rectifier is in a saturated state; therefore, P_L could be determinedas the saturated vapor pressure at rectification temperature, T_rec and ammonia massfraction w_ref. Meanwhile, P_H could be determined by the saturated vapor at the outlet of theevaporator based on the waste heat temperature T_was and w_ref. Based on P_L and w_ref, the condensation temperature, T_con, could be determinedbased on the saturated liquid state at the outlet of the condenser.This condensation temperature should be slightly lower than the rectificationtemperature due to the temperature difference between the saturatedliquid and vapor of ammonia-water at the same pressure and ammoniamass fraction. Then, other thermodynamic states of saturated liquidat the outlet of condenser, saturated vapors at the outlet of evaporatorand rectifier could be determined.

The ammonia-water liquid at the outletof the generator was in a saturated state, where T₁ = T_was and P₁ = P_L, hence the ammonia mass fraction of the basic ammonia-waterliquid (weak solution) used, w_bas, could be determined by the thermodynamic state equation.Thereafter, enthalpy h₁ and entropy s₁ could be calculated. The recirculation flow ratio (FR) [15] was used to calculate the mass flow rate of the ammonia-water liquid. FR is defined in Eq. (1).

(1)FR={\displaystyle \frac{\mathrm{mass}\hspace{0.17em}\mathrm{flow}\hspace{0.17em}\mathrm{rate}\hspace{0.17em}\mathrm{of}\hspace{0.17em}\mathrm{liquid}\hspace{0.17em}\mathrm{leaving}\hspace{0.17em}\mathrm{generator}}{\mathrm{mass}\hspace{0.17em}\mathrm{flow}\hspace{0.17em}\mathrm{rate}\hspace{0.17em}\mathrm{of}\hspace{0.17em}\mathrm{vapor}\hspace{0.17em}\mathrm{leaving}\hspace{0.17em}\mathrm{generator}}}={\displaystyle \frac{{\dot{m}}_{\mathrm{1}}}{{\dot{m}}_{\mathrm{7}}}}.

Then mass flow rate of the refrigerant, ṁ₇, was pre-definedat 0.01 kg/s, so that the system scale was approximately 12 kW (refrigerationpower). Thereafter, the ammonia-water liquid mass flow rates ṁ₁, ṁ₆ and the ammonia-richliquid mass fraction w₆ could be calculated based on the mass balance equations.The temperature of the useful heat, T_use = T₄, was then determined by the saturated liquidat the outlet of the absorber at the pressure of P_H and the mass fraction of pump w₄ (w₄ =w₆). Besides, enthalpy h₄ could be calculated.

The ammonia-water liquid from thegenerator was pumped from P_L to P_H. The isentropic efficiency, h_pump1, of this solution pump isgiven as

(2){\eta }_{\mathrm{pump1}}={\displaystyle \frac{h_{\mathrm{2}\mathrm{s}}-h_{\mathrm{1}}}{h_{\mathrm{2}}-h_{\mathrm{1}}}}

where h_2s is the enthalpy of ammonia-water liquidat the outlet of the pump if the process is isentropic.h_pump1 was pre-defined to be 0.85 according to Ref. [16]. h_2s could be calculated by considering theammonia-water liquid at pressure P_H with an entropy value of s₁. The value of h₂ could be calculatedby Eq. (2). The pumping power consumed is then calculated by

(3){\dot{W}}_{\mathrm{pump1}}={\dot{m}}_{\mathrm{1}}(h_{\mathrm{2}}-h_{\mathrm{1}}).

The same method could be appliedto the second pump located between condenser and evaporator to obtain Ẇ_pump2. Meanwhile, the following two equations could beemployed to solve the compression process using the similar procedureas that for the pump.

(4){\eta }_{\mathrm{com}}={\displaystyle \frac{h_{\mathrm{9}\mathrm{s}}-h_{\mathrm{8}}}{h_{\mathrm{9}}-h_{\mathrm{8}}}},

(5){\dot{W}}_{\mathrm{com}}={\dot{m}}_{\mathrm{8}}(h_{\mathrm{9}}-h_{\mathrm{8}}),

where h_com was pre-definedto be 0.75 [16], and themass flow rate of the ammonia-water vapor entering the compressorcould be calculated based on the value of splitting ratio, R_{\mathrm{ref}}={\displaystyle \frac{{\dot{m}}_{\mathrm{8}}}{{\dot{m}}_{\mathrm{7}}}}.

For the recuperator, the thermalstates of the two inlet ammonia-water liquids (T₂, h₂, T₄, h₄) were determined. Then, the outlet temperature (T₃ and T₅) and the enthalpy (h₃ and h₅) of the two liquidscould be iteratively determined using heat balance equations, Eq.(6) and Eq. (7), and the logarithmic temperature difference ΔT_LMTD in Eq. (8).

(6){\dot{Q}}_{\mathrm{HE}}=\mathrm{\Delta }{T}_{\mathrm{LMTD}}UA,

(7){\dot{Q}}_{\mathrm{HE}}={\dot{m}}_{\mathrm{2}}(h_{\mathrm{3}}-h_{\mathrm{2}})={\dot{m}}_{\mathrm{4}}(h_{\mathrm{4}}-h_{\mathrm{5}}),

(8)\mathrm{\Delta }{T}_{\mathrm{LMTD}}={\displaystyle \frac{T_{\mathrm{4}}-T_{\mathrm{3}}-(T_{\mathrm{5}}-T_{\mathrm{2}})}{\ln \left({\displaystyle \frac{T_{\mathrm{4}}-T_{\mathrm{3}}}{T_{\mathrm{5}}-T_{\mathrm{2}}}}\right)}}.

The rectification reflux ratio, R_refl, could be usedto assist the calculation of the rectification process [17]. As depicted in Fig. 2, the limitedheight of the rectification column leads to a steeper operation linecompared to the ideal isothermal rectification line. The reflux ratiois defined as

(9)R_{\mathrm{refl}}={\displaystyle \frac{h_{\mathrm{pole}}-h_{\mathrm{7}}}{h_{\min }-h_{\mathrm{7}}}},

where h_min can be calculated by

(10)h_{\min }=h_{\mathrm{6v}}+(h_{\mathrm{6v}}-h_{\mathrm{6}\mathrm{l}}){\displaystyle \frac{w_{\mathrm{ref}}-w_{\mathrm{6v}}}{w_{\mathrm{6v}}-w_{\mathrm{6}\mathrm{l}}}}.

R_refl of 2 was used in the current study asrecommended [17]. h_pole can be calculatedby Eq. (9). The rectification heat is then

(11)Q_{\mathrm{rec}}={\dot{m}}_{\mathrm{7}}(h_{\mathrm{pole}}-h_{\mathrm{7}}).

The generation, evaporation and absorptionheat can be calculated by

(12){\dot{Q}}_{\mathrm{gen}}={\dot{m}}_{\mathrm{1}}{h}_{\mathrm{1}}+{\dot{m}}_{\mathrm{7}}{h}_{\mathrm{7}}+{\dot{Q}}_{\mathrm{rec}}-{\dot{m}}_{\mathrm{6}}{h}_{\mathrm{6}},

(13){\dot{Q}}_{\mathrm{eva}}={\dot{m}}_{\mathrm{12}}(h_{\mathrm{13}}-h_{\mathrm{12}}),

(14){\dot{Q}}_{\mathrm{abs}}={\dot{m}}_{\mathrm{3}}{h}_{\mathrm{3}}+{\dot{m}}_{\mathrm{9}}{h}_{\mathrm{9}}+{\dot{m}}_{\mathrm{13}}{h}_{\mathrm{13}}-{\dot{m}}_{\mathrm{4}}{h}_{\mathrm{4}}.

Finally, the energy efficiency andexergy efficiency of the HAC-HP cycle can be calculated by Eqs. (15)and (16), respectively.

(15){\eta }_{\mathrm{en}}={\displaystyle \frac{{\dot{Q}}_{\mathrm{abs}}}{({\dot{Q}}_{\mathrm{gen}}+{\dot{Q}}_{\mathrm{eva}}+{\dot{W}}_{\mathrm{com}}+{\dot{W}}_{\mathrm{pump1}}+{\dot{W}}_{\mathrm{pump2}}}},

(16){\eta }_{\mathrm{ex}}=\frac{{\dot{Q}}_{\mathrm{abs}}(\mathrm{1}-T_{\mathrm{amb}}/T_{\mathrm{use}})}{({\dot{Q}}_{\mathrm{gen}}+{\dot{Q}}_{\mathrm{eva}})(\mathrm{1}-T_{\mathrm{amb}}/T_{\mathrm{was}})+{\dot{W}}_{\mathrm{com}}+{\dot{W}}_{\mathrm{pump1}}+{\dot{W}}_{\mathrm{pump2}}}.

The parameters used in the calculationare summarized in Table 1.

Results and discussion

Figure 3 is an example of the operationstates of the HAC-HP system in the enthalpy-mass fraction chart, wherethe waste heat temperature is 80°C and FR = 7. The two pressure levels in the system are 11.1bar and 38.9 bar while the ammonia-lean and ammonia-rich solutionhave a mass fraction of 0.412 and 0.485 respectively, which givesa pre-defined refrigerant mass fraction of 0.9995.

Figure 4 demonstrates the energyefficiency of the HAC-HP system proposed at a waste heat temperatureof 80°C. The maximum energy efficiency obtained under the workingconditions studied is approximately 0.90. This maximum value occursunder the condition of FR = 1 and R_ref = 0. These curvesindicate that it is more efficient to use a compressor than a condenser-pump-evaporator,as can be seen from Fig. 4 that a lower refrigerant splitting ratioleads to a higher energy efficiency. For example, the highest h_en isabout 0.90 when using the compressor only under the condition of FR = 1 while the lowest h_en is only about0.43 when using the condenser-pump-evaporator only. The upgraded temperatureis from 95°C to approximately 132°C, which gives a temperaturelift in the range of 15°C–52°C. A higher FR leads to a smaller ammonia mass fractiondifference between the ammonia-rich and the ammonia-lean solution.The saturated ammonia-water liquid from the absorber has a relativelylower ammonia mass fraction. Therefore, the equilibrium temperaturein the absorber at a certain pressure (P_H has been determined by the evaporator) ishigher. A relatively flat energy efficiency curve can be noticed when FR is smaller than 6 for all values of R_ref, and the changeof energy efficiency is no more than 25%. Then, the energy efficienciesdrop dramatically towards zero as FR is higher than 7.

Figure 5 shows the exergy efficiencyof the system at a waste heat temperature of 80°C. Compared tothe energy efficiency curves in Fig.4, the gap between each exergyefficiency curve having different values of R_ref is smaller, e.g. less than8%. The exergy efficiency increases gently as FR increases to 3. Then, it decreases gently as FR increases to 6. And finally, it decreasesviolently towards zero. Based on these values of exergy efficiency,the optimal operating value of FR locates in the range of 1–6, where the exergy efficiency variesin the range of 0.50–0.61 while the corresponding energy efficiencyis 0.33–0.90 and the temperature lift is 14.9°C–44.1°C.

Figure 6 illustrates the upgradeduseful heat power at a waste heat temperature of 80°C. The powerof useful heat is only slightly reduced as FR increases from 1 to 4, e.g., from 11.97 kW to 11.20kW in the case of R_ref = 0 and from 9.6 kW to 8.9 kW in the case of R_ref = 1. Thereafter, the usefulheat power drops significantly towards zero as shown in Fig. 6. Toobtain these amounts of useful heat, the input heat and electricitypower required are exhibited in Fig.7. The required electricity consumedby the compressor and solution pumps is less than 3 kW, while therequired heat for the generator and evaporator can be as high as nearly24 kW. As shown in Fig. 7, when R_ref = 0, more electricity and less heat areconsumed compared to that in the case of R_ref = 1. However, the extent of variationof heat is more notable than that of electricity due to the lowerexergy possessed by thermal energy than electricity. More heat consumptionindicates lower energy efficiency. Nevertheless, if treating the wasteheat as totally free energy and only electricity is counted in Eq.(15), the system energy efficiency (or coefficient of performance)will be significantly higher and increase with the drop of R_ref.

Figure 8 displays the energy efficienciesof the HAC-HP system as a function of temperature lift at differentwaste heat temperatures from 60°C to 120°C. As shown in Fig.8, the energy efficiency curves shift to the lower right side as thewaste heat temperature increases, towards a higher temperature liftbut a lower energy efficiency. For example, for the waste heat temperatureof 120°C, the temperature lift is in the range of 40.7°C–101.0°C,the energy efficiency is lower than 0.76 when R_ref is 0 and is lower than 0.28when R_ref is 1. Table 2 presents the recommended operation conditions andcorresponding performances of the HAC-HP system before the violentdrop of system energy and exergy efficiencies and output useful heatpower.

Conclusions

The present study investigated ahybrid absorption-compression heat pump to upgrade and recover theindustrial waste heat in the temperature range of 60°C–120°C.The system proposed can use two routes to elevate the working fluidpressure, one is via the compressor and the other is via the condenser-pump-evaporator,so that the use of energy sources of electricity and waste heat canbe flexible. The major conclusions are summarized as follows.

(1) As the system recirculation flowratio increased, the temperature lift was improved; however, the systemenergy and exergy efficiencies dropped violently when the recirculationflow ratio was larger than a certain value.

(2) More ammonia-water vapor flowingthrough compressor led to a higher system energy efficiency; however,the exergy efficiency changed little with the vapor splitting ratio.

(3) The recirculation flow ratioshould be lower than 9, 6, 5, and 4 respectively as recommended fora waste heat temperature of 60°C, 80°C, 100°C, and 120°Crespectively, while the temperature lifts were 9.8°C–27.7°C,14.9°C–44.1°C, 24.4°C–64.1°C, and 40.7°C–85.7°Crespectively, and the system energy efficiency were 0.35–0.93,0.32–0.90, 0.25–0.85, and 0.14–0.76.

References

[1] Vélez F, Segovia  J J, Martin  M C, Antolin  G, Chejne F,  Quijano A. A technical, economical and market reviewof organic Rankine cycles for the conversion of low grade heat forpower generation. Renewable & SustainableEnergy Reviews, 2012, 16(6): 4175–4189

[2] Zhai X Q, Qu  M, Li Y,  Wang R Z. A review for research and new design options of solar absorption cooling systems. Renewable & Sustainable Energy Reviews, 2011, 15(9): 4416–4423

[3] Li T X, Wang  R Z, Li  H. Progress in the development of solid-gas sorption refrigerationthermodynamic cycle driven by low-grade thermal energy. Progress in Energy and Combustion Science, 2014, 40: 1–58

[4] Oluleye G, Smith  R, Jobson M. Modelling and screening heat pump optionsfor the exploitation of low grade waste heat in process site. Applied Energy, 2016, 169: 267–286

[5] Chua K J, Chou  S K, Yang  W M. Advances in heat pump systems: a review. Applied Energy, 2010, 87(12): 3611–3624

[6] Brunin O, Feidt  M, Hivet B. Comparison of the working domains ofsome compression heat pumps and a compression-absorption heat pump. International Journal of Refrigeration, 1997, 20(5): 308–318

[7] Hultén M, Berntsson  T. The compression/absorption cycle—influence of some major parameterson COP and a comparison with the compression cycle. International Journal of Refrigeration, 1999, 22(2): 91–106

[8] Minea V, Chiriac  F. Hybrid absorption heat pump with ammonia/water mixture—somedesign guidelines and district heating application. International Journal of Refrigeration, 2006, 29(7): 1080–1091

[9] Kim J, Park  S R, Baik  Y J, Chang  K C, Ra  H S, Kim  M, Kim C. Experimental study of operating characteristics of compression/absorption high-temperaturehybrid heat pump using waste heat. Renewable Energy, 2013, 54: 13–19

[10] Jensen J K, Markussen  W B, Reinholdt  L, Elmegaard B. On the development of high temperature ammonia-waterhybrid absorption-compression heat pumps. International Journal of Refrigeration, 2015, 58: 79–89

[11] Jensen J K, Markussen  W B, Reinholdt  L, Elmegaard B. Exergoeconomic optimization of an ammonia-water hybridabsorption-compression heat pump for heat supply in a spray-dryingfacility. International Journal of EnvironmentalEngineering, 2015, 6: 195–211

[12] Bourouis M, Nogues  M, Boer D,  Coronas A. Industrial heat recovery by absorption/compression heat pump usingTFE-H₂O-TEGDME working mixture. Applied Thermal Engineering, 2000, 20(4): 355–369

[13] El-sayed Y M, Tribus  M. Thermodynamic properties of water-ammonia mixtures theoretical implementation foruse in power cycles analysis.  American Society of Mechanical Engineers, Advanced Energy Systems Division(Publication), 1985, 1: 89–95

[14] Ziegler B, Trepp  C. Equation of state for ammonia-water mixtures. International Journal of Refrigeration, 1984, 7(2): 101–106

[15] Herold K E, Radermacher  R, Klein S A. Absorption Chillers and HeatPumps. Boca Raton: CRC Press, 1996

[16] Wu W, Wang  B L, Shi  W X, Li  X T. Performance improvement of ammonia/absorbent air source absorptionheat pump in cold regions. Building Services Engineering Research and Technology, 2014, 35(5): 451–464

[17] Kandlikar S. A new absorber heat recovery cycle to improve COP ofaqua-ammonia absorption refrigeration system. Ashrae Transactions, 1982, 88: 141–158

RIGHTS & PERMISSIONS

Higher Education Press and Springer-Verlag GmbHGermany