Exergy analysis of R1234ze(Z) as high temperatureheat pump working fluid with multi-stage compression

Abstract

In this paper, the simulation approach and exergy analysis ofmulti-stage compression high temperature heat pump (HTHP) systemswith R1234ze(Z) working fluid are conducted. Both the single-stageand multi-stage compression cycles are analyzed to compare the systemperformance with 120°C pressurized hot water supply based uponwaste heat recovery. The exergy destruction ratios of each componentfor different stage compression systems are compared. The resultsshow that the exergy loss ratios of the compressor are bigger thanthat of the evaporator and the condenser for the single-stage compressionsystem. The multi-stage compression system has better energy and exergyefficiencies with the increase of compression stage number. Comparedwith the single-stage compression system, the coefficient of performance(COP) improvements of the two-stage and three-stage compression systemare 9.1% and 14.6%, respectively. When the waste heat source temperatureis 60°C, the exergy efficiencies increase about 6.9% and 11.8%for the two-stage and three-stage compression system respectively.

Keywords

multi-stage compression / hightemperature heat pump / heat recovery / exergy destruction / R1234ze(Z) workingfluid

Introduction

Water-source heat pump (WSHP) technologiessuch as geothermal water-source heat pumps and seawater water-sourceheat pumps are energy-saving methods that have been used in commercialand residential heating applications [1]. Besides, the WSHP with waste heat recovery has betterprospects in industrial applications for its high efficiency and hightemperature output. Survey results of temperature range requirementsin some industries revealed that most of the temperature requirementsin industry are above 60°C, for example, wood drying, food andbeverage production, dyeing process, district heating and crude oilheating [2,3]. Watanabe [4] summarized the temperatures of heatdemand in different types of industrial processes and divided theminto air, water and steam applications. In most industrial fields,coal-fired boilers, gas-fired boilers, oil-fired boilers and electricboilers are often used for hot water supply and process heating. However,these heating equipment waste more primary energy and cause environmentalpollution problems because of their low efficiencies and combustionemissions. Meanwhile, large amounts of low grade energy can be recoveredto satisfy the requirements of production and living for useful purposes.High temperature heat pumps (HTHP) offer a most practical solutionto those problems. They recover the heat from industrial waste andproduce high temperature hot water. The most important of all, hightemperature heat pumps have proven to be an effective way to reducegreenhouse gases emission.

Li et al. [5] studied the system performance ofa water-to-water HTHP unit with the R22/R141b mixture to provide hotwater with a temperature of 80°C. The highest hot water temperaturecould reach 85°C. Liu et al. [6] developed a water-to-water heat pump system with the HTR01 refrigerantto recover waste heat with temperatures of 30°C–60°Cand provide moderately high-temperature hot water with temperaturesof 70°C–90°C. Brown et al. [7] proposed fluorinated olefin R-1234ze(Z) as a possiblehigh-temperature refrigerant for HTHP and estimated the refrigerantthermal properties and potential system performances. Wang et al.[8] developed a heat pumpsystem with heat recovery in building heating and experimentally studiedthe system performance. The hot water temperature is reported to be85°C in most of the test conditions. Pan et al. [9] evaluated several moderately high-temperaturerefrigerants and found that a zeotropic mixture refrigerant HC600/HFC245fawas a good option for the HTHP. The system achieved a COP of 3.83at an evaporating temperature of 55°C and condensation temperatureof 100°C. Chamoun et al. [10] developed a dynamic model for an industrial heat pump using wateras refrigerant and the reported temperature range is around 120°C–130°C.To achieve a large temperature lift with a higher temperature output,the exergy loss in the compression and expansion process for the single-stagecompression system is very high, which lowers the compression efficiencyand degrades the system performance. As a result, a multi-stage compressiondesign, for example, multi-stage heat pump system is developed toovercome those problems.

Compared to a single-stage compressionheat pump system, the temperature lift of a multi-stage heat pumpsystem is larger and the compression efficiency is higher. Therefore,it is possible to achieve a higher COP. Some researchers have alreadycarried out both theoretical and experimental studies of the thermodynamicperformance of multi-stage compression heat pump systems. Redónet al. [11] analyzed theinfluence of design parameters and injection conditions for two differentconfigurations of two stage cycles. The results show that the twostage systems could reach improvements of 30% in terms of COP. Kondouand Koyama [12] presentedan exploratory assessment of multiple-stage heat pump systems withheat recovery using environmentally friendly refrigerants. The resultsindicate that a multiple-stage “extraction” cycle drasticallyreduces the throttling loss and exergy loss in the condensers, resultingin the highest COP among four kinds of proposed cycle configurations.It is also demonstrated the potential use of multiple-stage high-temperatureheat pumps is a promising way to reduce the primary energy consumptionfor industrial applications. Lee et al. [13] proposed a saturation cycle consisting of saturationcompression and saturation expansion in order to improve system performanceby reducing the thermodynamic loss associated with single phase gascompression and isenthalpic expansion. The simulation results indicatethat it is more beneficial if the multi-stage cycle is used for thecycle at a higher pressure ratio. When the four-stage cycle is applied,the COP improvements of D2Y60 (mixture of R32 and R1234yf), CO₂ and propane are 46.9%, 43.2% and 38.2%, respectivelyunder an extreme heating condition. Arpagaus et al. [14] investigated and evaluated eightkinds of heat pump cycles based on the first and second law efficiency.Thermodynamic simulations reveal that multi-stage compressor cycleshave the highest COPs and second law efficiencies, followed by cascade,ejector, and expansion valve cycles.

Although there are some investigationson multi-stage compression heat pump systems, there is still a lackof studies on waste heat recovery industrial heat pumps with a hightemperature lift, especially with low GWP refrigerants. The systemperformance and heat recovery efficiency for industrial heat pumpscan be further improved. In this paper, multi-stage compression heatpump systems with R1234ze(Z) refrigerant are investigated to improvethe thermal performance and system exergy efficiency. The thermodynamicsimulation approach of multi-stage compression heat pumps is builtto evaluate the system performance. Based on exergy analysis, theenergy destruction at each separated component is developed to getthe better performance and improvement potential of the proposed systems.The thermodynamic analysis is conducted based on the variation ofwaste heat source temperature. The power consumption, system COP,and exergy efficiency under different working conditions are analyzed.The exergy destruction ratio of each process is compared and validatedwith reference results.

Multi-stage compression heat pump systems

Figure 1(a) shows the schematic diagramsof the two-stage vapor compression heat pump system. The working principleis as follows: The refrigerant vapor is compressed to an intermediatetemperature and pressure gas first, after being mixed with the intermediatepressure vapor from the flash tank. The mixed vapor is compressedby the higher-stage compressor to high temperature and pressure gas.It exchanges heat with the water in the condenser and becomes liquidrefrigerant. Then, the saturated liquid refrigerant is further cooleddown in the subccoler. The high pressure refrigerant from the subcooleris throttled by the upper-stage expansion valve and becomes a liquid-gasmixture of intermediate pressure. The refrigerant mixture is thenseparated into liquid phase and vapor phase in the flash tank withintermediate pressure. The vapor refrigerant of intermediate pressureis mixed with the discharge gas from the lower-stage compression.The liquid refrigerant is further throttled by the lower-stage expansionvalve and becomes a liquid-gas mixture of low pressure and low temperature.The low pressure refrigerant mixture flows back to the evaporatorwhere it absorbs heat from waste water and vaporizes for the nextcycle.

From Fig. 1(b) it can be seen thatdue to the two-phase separation in the flash tank, the liquid enteringthe evaporator has a lower enthalpy compared to that of a single-stagecycle. Thus the enthalpy difference across the evaporator is greaterthan that of a single-stage cycle. The saturated vapor from the flashtank also has a lower temperature than that of the vapor in the compressor,which helps to reduce the compressor discharge temperature. The reducedcompressor power consumption leads to a higher system COP.

Figure 2 shows the schematic anddiagrams of the three-stage compression heat pump system. The workingprinciple is as follows. The refrigerant from the evaporator is compressedin the first-stage and then mixed with the vapor refrigerant of thefirst-stage pressure. The mixed vapor enters the compressor for thesecond-stage compression. After mixing with the vapor refrigerantof the second-stage pressure, the mixed refrigerant is further compressedin the third-stage. Then, the refrigerant is cooled down in the condenserand flows through the third-stage expansion valve. The refrigerantis separated into liquid phase and vapor phase in the flash tank П.The vapor refrigerant of the second-stage pressure is mixed with thedischarge gas from the second-stage compression. The liquid refrigerantis further cooled down in the subccoler and enters the second-stageexpansion valve. The refrigerant is separated again in the flash tankІ. The vapor refrigerant of the first-stage pressure is mixedwith the discharge gas from the first-stage compression. The liquidrefrigerant enters the first-stage expansion valve and then is heatedby the waste heat in the evaporator.

From Fig. 1(b) and Fig. 2 (b) itcan be seen that the pressure difference of each compression stageand expansion process become small for the multi-stage compressionheat pump. The irreversible loss can be greatly reduced to improvethe exergy efficiency. Taking into account the pressure ratio anddischarge temperature, the temperature difference of evaporating andcondensing can be even larger based on the increase of the compressionstage. For the low grade industrial waste heat, it is a promisingapproach to develop high-temperature heat pumps with 120°C pressurizedhot water supply.

Table 1 presents a comparison ofthe characteristics and properties of selected refrigerants for industrialhigh temperature heat pumps. R600a, R717, and R744 are the conventionalrefrigerants. As a refrigerant, R600a has an explosion risk in additionto the hazards associated with non-flammable CFC refrigerants. Theuse of R600a as a refrigerant in industrial high temperature heatpumps is potentially dangerous. The natural refrigerant R717, i.e.,ammonia, exhibits excellent thermodynamic properties, as mentionedby many forerunners, but it also exhibits quite strong toxicity. Thecritical temperature of R744 is low and the corresponding operatingpressure is much higher than other refrigerants. Therefore, it isonly considered for transcritical refrigeration cycle in this paper.Beside, these low GWP refrigerants, R134a and R245fa are also comparedas the reference, which are used for a high-temperature heat pumpor organic Rankine cycle (ORC) recently. Because they both have ahigh-GWP value, hydrofluoroolefins (HFOs) such as R1234yf, R1234ze(E), R1234ze (Z), and R1233zd (E) were introduced. R1234ze (E) andR1234ze (Z) are the newly recognized substances as refrigerants. Theyhave the similar thermal properties and features for high temperatureheat pump applications. Specifically, the low-GWP refrigerant R1234ze(E) and the isomer R1234ze (Z) have been vigorously investigated inthis decade as alternatives to R134a and R245fa.

Simulation approach and exergy analysis

Simulation approach

Multi-stage compression heat pumpsystems with R1234ze(Z) working fluid are modeled by using EngineeringEquation Solver (EES) (Software, F-chart, 2012). The multi-stage compression(two-stage and three-stage) is compared with the single-stage compressioncycle. The expansion process is assumed to be an isenthalpic process,and the pressure drops through the heat exchangers are neglected.Moreover, the degrees of superheating and subcooling are assumed tobe constant. The waste heat source temperature is assumed to be between50°C and 90°C in the simulation. R1234ze (Z) is selected ascandidate working fluid. The refrigerant-side is assumed as follows:

Degree of subcooling (SC): 20°C;

Degree of superheating (SH): 5°C;

Isentropic efficiency is calculatedas a function of the pressure ratio; (h_is = 0.8014–0.0484(P_d/P_s)) [15].

Condensing temperature (T_c) is 5°C higherthan the hot water outlet temperature;

Evaporating temperature is 10°Clower than the waste water inlet temperature (T_ww);

The water mass flow rate of wasteheat source is set as 20 kg/s, and the temperature difference betweenwaste water inlet and outlet (DT_e) is5°C;

The initial intermediate pressures(P_in,i) are selected toresult in equal pressure ratios across the compression stages to minimizethe compressor power.

The vapor refrigerant is injectedto the next compression process, and the injection amount is controlledto maintain the discharge temperature after compression.

Figure 3 demonstrates the flowchartof multi-stage compression heat pump modeling with R1234ze(Z) refrigerant.The total mass flow rate is determined with the guessed value of intermediatepressures. The injected mass flow rate of each stage is decided toachieve the designed degree of superheat.

Exergy analysis

In a vapor compression heat pumpsystem, there are usually four major processes: evaporation, compression,condensation, and expansion. External energy (power) is supplied tothe compressor and outside heat is added to the system by evaporator,whereas in the condenser heat rejection is occurred for the system.For different refrigerants, the system cooling capacity and heatingcapacity are not the same, which cause a change in energy efficiencyfor the systems. Exergy losses in various components of the systemare not the same. For the multi-stage compression heat pump, the heattransfer process also exists in the subcooler. Exergy is consumedor destroyed due to the entropy created depending on the associatedprocesses. To specify the exergy losses or destructions of the multi-stagecompression heat pump, exergy analysis is necessary.

According to the second law of thermodynamics,a practical process is always irreversible. In a heat pump system,irreversible losses are caused by different factors, such as no-isentropiccompression and temperature difference of heat transfer. Accordingto the second law of thermodynamics, exergy analysis equations areexpressed as

(1)E_{\mathrm{heat,in}}+E_{\mathrm{mass,in}}+E_{\mathrm{work}}=E_{\mathrm{heat,out}}+E_{\mathrm{mass,out}}+I_{\mathrm{rr}},

(2)\Psi =(h-h_{\mathrm{0}})-T_{\mathrm{0}}(s-s_{\mathrm{0}}),

where I_rr is irreversibleloss; E_heat, E_mass and E_work are exergies during heat transfer, mass transfer, and the workingprocess; Y is specific exergy inany state; T₀ is the surrounding temperature; h₀ and s₀ are the enthalpy and entropy of workingfluid at the temperature of T₀ and pressure of 0.1 MPa. There is no mass transferbetween hybrid source heat pump and the surrounding environment, hence E_mass = 0.

Theoretical exergy losses in differentcomponents are calculated according to Refs. [16–18].

For the compression process:

Compressor work,

(3)W_{\mathrm{c}}=m(h_{\mathrm{out}}-h_{\mathrm{in}}).

Power consumption,

(4)W_{\mathrm{el}}=W_{\mathrm{c}}/{\eta }_{\mathrm{m}}\cdot {\eta }_{\mathrm{e}}.

So, the exergy loss,

(5)I_{\mathrm{comp}}=m({\psi }_{\mathrm{in}}-{\psi }_{\mathrm{out}})+W_{\mathrm{el}}=m[(h_{\mathrm{in}}-h_{\mathrm{out}})-T_{\mathrm{0}}(s_{\mathrm{in}}-s_{\mathrm{out}})]+W_{\mathrm{el}},

where I_comp is the theoretical exergy destruction of the compressor, m is the mass flow rate of working fluid, s_in and s_out are specificentropy of working fluid at inlet and outlet of the compressor, h_m isthe mechanical efficiency of the compressor, and h_e is the electricalefficiency of the motor. With reference to Ref. [15], it is assumed that mechanicalefficiency of the compressor is 95% and the electrical efficiencyof the motor is 98%.

For condensation process:

(6)I_{\mathrm{cond}}=m_{\mathrm{R1234ze}}({\psi }_{\mathrm{R1234ze,in}}-{\psi }_{\mathrm{R1234ze,out}})+m_{\mathrm{hw}}({\psi }_{\mathrm{hw,in}}-{\psi }_{\mathrm{hw,out}})=m_{\mathrm{R1234ze}}[(h_{\mathrm{R1234ze,in}}-h_{\mathrm{R1234ze,out}})-T_{\mathrm{0}}(s_{\mathrm{R1234ze,in}}-s_{\mathrm{R1234ze,out}})]+m_{\mathrm{hw}}[(h_{\mathrm{hw,in}}-h_{\mathrm{hw,out}})-T_{\mathrm{0}}(s_{\mathrm{hw,in}}-s_{\mathrm{hw,out}})].

For evaporation process:

I_{\mathrm{e}\mathrm{v}\mathrm{a}\mathrm{p}}=m_{\mathrm{R1234ze}}({\psi }_{\mathrm{R1234ze,in}}-{\psi }_{\mathrm{R1234ze,out}})+m_{\mathrm{w}\mathrm{w}}({\psi }_{\mathrm{w}\mathrm{w,in}}-{\psi }_{\mathrm{w}\mathrm{w,out}})

(7)=m_{\mathrm{R1234ze}}[(h_{\mathrm{R1234ze,in}}-h_{\mathrm{R1234ze,out}})-T_{\mathrm{0}}(s_{\mathrm{R1234ze,in}}-s_{\mathrm{R1234ze,out}})]+m_{\mathrm{ww}}[(h_{\mathrm{ww,in}}-h_{\mathrm{ww,out}})-T_{\mathrm{0}}(s_{\mathrm{ww,in}}-s_{\mathrm{ww,out}})].

For subcooling process:

(8)=m_{\mathrm{R1234ze}}[(h_{\mathrm{R1234ze,in}}-h_{\mathrm{R1234ze,out}})-T_{\mathrm{0}}(s_{\mathrm{R1234ze,in}}-s_{\mathrm{R1234ze,out}})]+m_{\mathrm{hw}}[(h_{\mathrm{hw,in}}-h_{\mathrm{hw,out}})-T_{\mathrm{0}}(s_{\mathrm{hw,in}}-s_{\mathrm{hw,out}})].I_{\mathrm{sub}}=m_{\mathrm{R1234ze}}({\psi }_{\mathrm{R1234ze,in}}-{\psi }_{\mathrm{R1234ze,out}})+m_{\mathrm{hw}}({\psi }_{\mathrm{hw,in}}-{\psi }_{\mathrm{hw,out}})

For expansion process:

(9)I_{\exp }=m_{\mathrm{R1234ze}}({\psi }_{\mathrm{in}}-{\psi }_{\mathrm{out}})=m_{\mathrm{R1234ze}}{T}_{\mathrm{0}}(s_{\mathrm{out}}-s_{\mathrm{in}})[\mathrm{T}\mathrm{htottling},{{\displaystyle h}}_{\exp ,\mathrm{in}}={{\displaystyle h}}_{\exp ,\mathrm{out}}].

Total exergy destruction,

(10)I_{\mathrm{total}}=I_{\mathrm{comp}}+I_{\mathrm{cond}}+I_{\mathrm{evap}}+I_{\exp }+I_{\mathrm{sub}}.

Compared to the conventional energyanalysis, exergy analysis can quantitatively characterize the thermodynamicimperfection of the heat transfer process and the possibility of thermodynamicdevelopment for the heat pump system. Exergy efficiency is definedas the ratio of the total exergy increasement to the total power consumptionof the multi-stage compression heat pump.

(11){\eta }_x=(E_{\mathrm{h}\mathrm{w},\mathrm{o}\mathrm{u}\mathrm{t}}-E_{\mathrm{w}\mathrm{w},\mathrm{i}\mathrm{n}})/W_{\mathrm{e}\mathrm{l}},

where the total exergy output E_hw,out and the exergy input E_ww,in of the multi-stagecompression heat pump can be given by

(12)E_{\mathrm{hw,out}}=m_{\mathrm{hw}}({\psi }_{\mathrm{hw,out}}-{\psi }_{\mathrm{hw,in}})=m_{\mathrm{hw}}[(h_{\mathrm{hw,out}}-h_{\mathrm{hw,in}})-T_{\mathrm{0}}(s_{\mathrm{hw,out}}-s_{\mathrm{hw,in}})],

(13)E_{\mathrm{ww,in}}=m_{\mathrm{ww}}({\psi }_{\mathrm{ww,in}}-{\psi }_{\mathrm{ww,out}})=m_{\mathrm{ww}}[(h_{\mathrm{ww,in}}-h_{\mathrm{ww,out}})-T_{\mathrm{0}}(s_{\mathrm{ww,in}}-s_{\mathrm{ww,out}})].

Results and discussion

The system performances of multi-stagecompression heat pump systems are evaluated under the operating conditionsfor 120°C pressurized hot water supply. The waste heat recoveredin the evaporator is supposed as a constant of 420 kW.

Variation of system performance

The variation of total power consumptionwith waste heat source temperature for different stage compressionsis depicted in Fig. 4. As the waste heat source temperature increasedfrom 50°C to 90°C, the total power consumption was found todecrease from 180 kW to 70.5 kW for the single-stage compression system.For the two-stage and three-stage compression system, the total powerconsumption decrease from 170 kW to 68.4 kW and from 163 kW to 65.7kW, respectively. This was mainly due to the decrease of the compressionratio caused by the evaporating temperature rising. It also can beseen that the three-stage compression system always has the smallestcompressor power consumption. With the waste heat source temperatureincreasing, the power saving of multi-stage compression became lessand less significant.

Based on the simulation results oftotal power consumption and heating capacity, the variation of systemCOP with waste heat source temperature is analyzed in Fig. 5. It canbe seen that system COP increased with the waste heat source temperaturerising. As the waste heat source temperature increased from 50°Cto 90°C, the system COP increased from 3.1 to 6.7 for the single-stagecompression system and from 3.5 to 6.98 for the two-stage compressionsystem, respectively. The system COP of the three-stage compressionsystem was 3.74 for 50°C waste heat source temperature and 7.14for 90°C waste heat source temperature. Under the same waste heatsource temperature condition, the COP improvement of the multi-stagecompression system was significant. Compared with the single-stagecompression system, the COP improvements of the two-stage compressionsystem are 12.2% and 6.3% for 50°C and 80°C waste heat sourcetemperature, respectively. The COP improvements of the three-stagecompression system are 19.8% for 50°C waste heat source temperatureand 9.4% for 80°C waste heat source temperature.

Figure 6 reveals the variation ofexergy efficiency with waste heat source temperature for differentstage compression systems. As the waste heat source temperature increased,the exergy efficiency decreased gradually. For the single-stage andtwo-stage compression systems, the exergy efficiencies decreased by18.3% and 19.2% when the waste heat source temperature increased from50°C to 90°C. For the three-stage compression system, theexergy efficiencies decreased by 20.2% when the waste heat sourcetemperature increased from 50°C to 90°C. With the evaporatingtemperature increasing, the exergy destruction of compression processand expansion process became less and less serious. The total exergydestruction of the system is decreased while the exergy loss of condensationand evaporation process remains a constant because of the fixed heattransfer approach temperature. Taking all of the above factors intoconsideration, the relatively irreversible loss increased with thewaste heat source temperature rising. So the exergy efficiencies ofall the single-stage and multi-stage compression systems decreased.For the waste heat source temperature of 60°C, the improvementsof the exergy efficiency were 6.9% and 11.8% for the two-stage andthree-stage compression systems when compared with the single-stagecompression system.

Table 2 lists the simulation resultsof R1234ze (Z) heat pumps evaluated under the 60°C waste heatsource temperature condition. As the stage number is increased, thecompressor work is reduced and heating capacity is increased, whichresults in the increase of COP. Table 2 also lists the mass flow rateand pressure of each stage in the compression process. The reducedpressure ratio results in a reduction of the compressor work consumption,and finally results in a COP improvement. When the three-stage compressionheat pump is applied, the COP is improved by 16.4% under the 60°Cwaste heat source temperature condition.

Variation of exergy destruction ratio

The exergy destruction ratio reflectsthe percentage of exergy loss for each process in heat pump systems.Figure 7 shows the variation of exergy destruction ratio of the compressionprocess with waste heat source temperature. It can be seen that theexergy destruction ratio of the compression process decreased withthe waste heat source temperature rising. As the waste heat sourcetemperature increased from 50°C to 90°C, the exergy destructionratio of the compression process was found to decrease from 44% to36% for the single-stage compression system. For the two-stage andthree-stage compression system, the destruction ratio decreased from42% to 31% and from 40% to 28%, respectively. The main reason forthis was that the exergy destruction in the compressors decreasedwith the rise of evaporating pressure.

Figure 8 displays the variation ofthe exergy destruction ratio of the condensation processes with wasteheat source temperature. As the waste heat source temperature increasedfrom 50°C to 90°C, the exergy destruction ratio of the condensationprocesses increased from 29% to 43% for the two-stage compressionsystem and from 32% to 45% for the three-stage compression system.For the single-stage compression system, the exergy destruction ratioof the condensation processes had the same increment (about 14%) withthe multi-stage compression system. This is mainly due to the decreaseof the total exergy destruction and the constant exergy loss of condensationwith waste heat source temperature increasing.

Figure 9 reveals the variation ofthe exergy destruction ratio of the expansion process with waste heatsource temperature. As the waste heat source temperature increasedfrom 50°C to 90°C, the exergy destruction ratio of the expansionprocess was found to decrease from 21.4% to 12.3% for the single-stagecompression system. For the two-stage and three-stage compressionsystems, the destruction ratio decreased from 17.3% to 10.1% and from14% to 9%, respectively. The reason for this is that the increaseof waste heat source temperature can effectively reduce the pressuredifference between evaporating and condensing, and then reduce theexergy loss of the expansion process.

Figure 10 shows the variation ofthe exergy destruction ratio of the evaporation processes with wasteheat source temperature. It can be seen that the exergy destructionratio of the evaporation processes increased by 5.2% and 3.7% forthe two-stage and three-stage compression system, respectively. Withthe waste heat source temperature increasing from 50°C to 90°C,the exergy destruction ratio of the evaporation processes increasedfrom 8.4% to 15% for the single-stage compression system. The mainreason for evaporation processes was similar to that for the condensationprocess. The decrease of the total exergy destruction may take animportant part in the rise of the exergy destruction ratio for theevaporation processes.

Comparison with reference results

The exergy destruction ratio of thesingle-stage compression heat pump is compared to the calculated resultsof Fukuda et al. [19].The comparison is performed assuming the R1234ze(Z) condensing temperatureof 125°C is maintained. For different heat pump process, the exergydestruction ratio of the simulation and reference results are comparedin Fig. 11. It can be seen that the exergy destruction ratios of thecompression and condensation processes take up a large proportion.While those of the expansion and evaporation processes account for13% and 12%, respectively. The relative deviation of the simulationand reference results is less than 5%. This reveals that the simulationmethod is able to reflect the system performance and predict the exergydestruction for different heat pump process.

Conclusions

Multi-stage compression heat pumpsystems with R1234ze(Z) refrigerant are investigated to recover thewaste heat in industrial processes. The pressurized water is graduallyheated by the subcooler and the condenser to a high-temperature of120°C. Simulation approach is conducted to analyze the performanceof multi-stage compression heat pump systems under different wasteheat source temperature conditions. Besides, exergy analysis is alsoconducted to investigate the exergy destruction and exergy loss ratio.The main conclusions are as follows.

The waste heat source temperaturehas a great influence on the total power consumption and system COPof the multi-stage compression heat pumps. As the waste heat sourcetemperature increases from 50°C to 90°C, system COP increasesfrom 3.5 to 6.98 for the two-stage compression system and from 3.74to 7.14 for the two-stage compression system, respectively. As thestage number increases for the same waste heat recovered, the multi-stagecompression heat pump has less power consumption.

The multi-stage compression heatpump significantly improves the system COP and exergy efficiency.When the three-stage compression heat pump is applied, the COP isimproved by 16.4% under the 60°C waste heat source temperatureconditions. The exergy efficiency is improved by 6.9% and 11.8% forthe two-stage and three-stage compression systems when compared withthe single-stage compression system.

The exergy destruction ratio of eachprocess is investigated for multi-stage compression heat pumps. Withthe waste heat source temperature increasing, the exergy destructionratio of the compression and expansion processes decreases, whilethat of the condensation and evaporation processes increases. Underthe same operating conditions, the three-stage compression heat pumphas the minimum exergy destruction of the compression and expansionprocess. The results indicate that the three-stage compression heatpump system has obvious advantage of exergy efficiency with heat recovery.

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