Performance enhancement of a two-stage vapour compression heat pump with solution circuits by eliminating the rectifier

Performance enhancement of a two-stage vapour compression heat pump with solution circuits by eliminating the rectifier

Milind V. Rane Energy Concepts Co, 627 Ridgely Avenue, Annapo l i s , M D 21401, U S A

Karim Amrane I C F Inc, G l o b a l Change Divis ion, 1850 K Street N.W. , Suite 1000, Wash ing ton , D.C.

20006, U S A Reinhard Radermacher

D e p a r t m e n t of Mechanica l Engineering, Center for Env i ronmenta l Energy Engineering, The Univers i ty o f Mary land , College Park, M D 20742, U S A

Received 5 February 1992; revised 30 November 1992

Performances of four versions of a two-stage vapour compression heat pump with solution circuits (VCHSC) are compared. The VCHSC represents a cascade system that requires only one compressor. Performance curves for the best cycle, the modified version with a bleed line and a desuperheater are obtained and studied in detail. The four cycles are compared for the same total UA value (product of overall heat transfer coefficient and area) including all heat exchangers while pumping heat from an average desorber temperature of -2.5°C to an average absorber temperature of 105°C. Simulation results show that the cycle with a bleed line and a desuperheater has 40 to 50% higher cooling COPs and 30 to 40% higher cooling loads as compared to the cycle with a rectifier. Both the rectifier and the bleed line cycles show an increase of up to 20% in the cooling COP and capacity by adding a desuperheater. Performance curves are obtained by simulating the best cycle at various weak solution concentrations in the low temperature loop and different solution pump flow rates. The parameters studied are the cooling COP, the solution heat exchanger effectiveness, the pressure ratio, the solution temperature glides in the absorbers and desorbers, the low temperature desorber load and the distribution of the UA value. Cooling COPs up to 1.05 and pressure ratios as low as 7.05 are obtained for the modified VCHSC. The cooling COPs are twice as much and the pressure ratios are two thirds less of what would be found in a conventional single-stage vapour compression system. Changing the weak solution concentration in the low temperature loop from 40 to 80 wt% ammonia increased the cooling load by four times. The results indicate that the two-stage VCHSC with a bleed line and a desuperheater can work at temperatures above 100°C and achieve temperature lifts of more than 100 K with well known and environmentally safe refrigerants. (Keywords: heat pump; absorption system; compression system; hybrid cycle; simulation; water; ammonia; performance)

Am61ioration de la performance d'une pompe fi chaleur bi6tag6e fi compression de vapeur avec des circuits de

solution, grfice/t l'61imination du rectificateur On a compar{ les petT[brmances de quatre configurations d'une pompe h chah, ur bibtag& h compression de vapeur avec des circuits de solution. Cette pompe g, chaleur eonstitue un syst~me en cascade qui ne nbcessite qu'un seul compresseur. On a obtenu et btudib avec soin les eourbes de la perjormance pour le meilleur ~Tcle et la con[iguration mo&'~be eomportant un dispositif de soutirage et un dbsurehauffeur. On compare les quatre cycles pour la m6me valeur totale UA (produit du coefficient d'bchange global et de la surface de tran,~,rt ), incluant tous les bchangeurs de chaleur, en pompant la chaleur h tree temp&ature moyenne du d~sorbeur de - 2,5 deg C et une temp&ature moyenne de I'absorbeur de 105 deg C. Les r&ultats de la simulation montrent que le cycle comportant un di.sposit(f de soutirage et un d&urehauffeur pr&ente des COP de rejkoidissement sup&ieurs de 40 h 50% et des charges de rr!h'oidissement sup&ieures de 30 h 40% h ceux d'un cTcle comportant un rectificateur. A vec I'ineorporation d'un d&'urchaz¢ffeur, h, COP et la capacitb de rejroidissement du cycle h rectificateur et du o'cle h disposit~lde soutirage ont enregistrb une augmentation allantjusqu'a 20%. On obtient les courbes de perJbrmanee en simulant le meilleur r3'cle h plusieurs concentrations de la solution pauvre clans la boucle basse temp&ature et h d(If&ents debits de la solution. Les param&res &udibs sont le COP de refi'oidissement, I'effieacitb de I'g'changeur de ehaleur, le taux de compression, les glissements de temp&ature de la solution clans les absorbeurs et d&orbeurs, la charge du d~;sorbeur h basse temp&ature et la distribution de la valeur UA. On obtient des COP de r~'/?oidissement allantjusqu 'e't 1,05 et des taux de compression de 7,05 minirnum avec la configuration modifibe de la pompe h chaleur bi&agr;e h compression de vapeur et h circuits. Les COP de rc{lroidissement sont deux fois plus blev& et les taux de compression sont inJg'rieurs de deux tiers h ceux d'une pompe h chaleur classique, mono&agbe et h compression de vapeur. En mod(tfant la concentration de solution pauvre dans la boucle basse temp&ature, de 40 h 80% en ma~se d'ammoniac, on a multiplib par 4 la charge de reJroidissement. Les rbsultats indiquent que la pompe a chaleur &udibe, avec w7 disposit(f de soutirage et an d&urchauffbur, peut travailler fi des tempOratures sup&ieures h 100 deg C et atteindre des ~;h;vations de temperature de plus de 100 K avec des J?igorig~nes courants et st~rs du point de vue de l'en vironnement. (Mots clOs: pompe fi chaleur; syst6me fi absorption; syst6me fi compression; cycle hybride; simulation; cam ammoniac; performance)

It is economica l ly des i rable to provide heat ing and cooling duties with one system such as a heat pump act ing as both a heater and a refr igerator . But this task becomes

difficult to accompl ish in the v a p o u r compress ion heat pump as the t empera tu re difference between the heat sink and the heat source, i.e., the t empera tu re lift,

0140 7007/93/040247 11 ~:~ 1993 Butterworth Hcinemann Ltd and IIR Iqev. Int. Froid 1993 Vol 16 No 4 2 4 7

Performance enhancement of a two-stage heat pump: M. V. Rane et al.

Nomenclature

A COP EV F H HT IT LMTD

LT Q U

Heat transfer area (m 2) Coefficient of performance Expansion valve Mass flow rate (kg s l) Specific enthalpy (kJ kg ~) High temperature Intermediate temperature Logarithmic mean temperature difference (K) Low temperature Heat transferred (kW) Overall heat transfer coefficient

(kW m -2 K ~) SHX Solution heat exchanger SP Solution pump STG Solution temperature glide VCHSC Vapour compression heat pump with

solution circuits W Compressor or pump(s) work (kW) wt Weight X Overall composition

Subscripts

in Inlet out Outlet

increases, because of increasingly large pressure ratios. It is well known from thermodynamics that the relation between the temperature lift and the pressure ratio in the heat pump cycle is dictated by the slope of the vapour pressure curve of the working fluid. Therefore, if large temperature lifts with low pressure ratios are desired, one would look for a fluid with a very flat vapour pressure curve. Unfortunately, most of today's refrigerants exhi- bit similar vapour pressure curves with relatively steep slopes making it practically impossible with the present technology to pump heat across temperature lifts of 100 K or more without resorting to cascaded heat pumps with two compressors and two different working fluids. Such heat pumps use halogenated refrigerants such as R114 and R22 which are known to be harmful to the ozone layer of the atmosphere and will be banned in the near future. As an alternative, a heat pump that works along a moderately sloped pseudo vapour pressure curve although employing common and well-known working fluids has been proposed ~-3. This cycle, i.e., a two-stage vapour compression heat pump with solution circuits (VCHSC), belongs to the large family of absorption/ compression heat pumps. It employs mixtures as the working fluid and is capable of simultaneously produc- ing cooling capacity at for example 0°C and heat output at 100°C for very moderate pressure ratios, making poss- ible the use of single-stage compressor units. With the mixture ammonia-water as the working fluid, temperatures as high as 200°C can be achieved. The heat pump has been studied previously 4,5 and two experimental rigs demonstrating its feasibility have been constructed and tested 68.

In this study, a steady state computer simulation model of a two-stage VCHSC has been developed. The model is based on heat and mass balances and heat transfer relationships for each component of the cycle and is applied to a prototype unit. The heat exchangers are described by their UA values. The objective of this work is to predict the performance of the heat pump. First, the conventional two-stage VCHSC with rectifier is introduced. Then, a modified version of the cycle that eliminates the need for rectification and improves the system's performance is presented. The effects of mixture composition, heat exchanger size and pump volumetric flow rate on the performance of the modified two-stage VCHSC are investigated.

Description of the cycles

Figure 1 represents the two-stage vapour compression heat pump with two solution circuits, a bleed line and a desuperheater on a pressure-temperature diagram. This representation helps visualize the pressure and temperature values of the heat exchangers accommodating a phase change (namely the absorbers and the desorbers).

The low temperature (LT) stage of the two-stage VCHSC, as shown in Figure 1, pumps heat from the lower temperature in desorber A to an intermediate temperature (IT) in absorber B. Heat is exchanged internally between absorber B and desorber C. The high temperature (HT) stage pumps heat from the intermediate temperature in desorber C to the higher temperature in absorber D. Thus, heat is pumped from the lower temperature in desorber A to the higher temperature in absorber D. As heat is pumped through a smaller tern, perature difference in each stage the pressure ratio in a two-stage VCHSC can be 40 to 65% lower compared to that in a single fluid Rankine cycle heat pump. However~ the capacity of the two-stage system is about 45% less than the single stage system. This is because in addition to compressing the LT desorber vapour, the compressor has to compress vapour from the IT desorber, which does not contribute to the cooling capacity. The four versions of the two-stage VCHSC which are simulated are as follows:

Case l: Cycle with a rectifier Case 2: Cycle with a rectifier and a desuperheater

(Figure 2) Case 3: Modified cycle with a bleed line, and Case 4: Modified cycle with a bleed line and a desuper-

heater (Figure 3).

Cases 1 and 2 are similar; the only additional component in case 2 is the desuperheater. Figure 2 shows the layout of a two-stage VCHSC with a rectifier and a desuperheater (Case 2). A rectifier is used in this cycle to prevent the transfer of absorbent from the H T loop to the LT loop. Although resulting in a loss of efficiency as explained in the subsequent paragraphs, the rectification process is necessary to keep the system operational: The cycle is composed of 12 main components: seven heat exchangers, two solution pumps, two expansion valves and a compressor. The seven heat exchangers are the LT desorber, the absorber/desorber heat exchanger which

248 Int. J. Refrig. 1993 Vo116 No 4

Performance enhancement of a two-stage heat pump. M.V. Rane et a l.

In P

Desorber A Desorber C Qin De~ber A ~J {Desorber ~

LT SP HT SP

I I T source T sink

LT IT HT

/ /

/

Desuperheater / /

I ?~esup. / z

Figure I Two-stage vapour compression heat pump with ammonia water solution circuits: modified cycle with a bleed line and a desuperheater Figure 1 Pornpe ~'1 chaleur bi[~tag~;e h compression de vapeur avec des circuits de solution ammoniac eau: cycle mod(]id avec un dispositif de soutirage et un d~;surchauff~'ur

Figure 2 Figure 2

9

34 33

3

~ 2"1

LTSHX HTSHX

_ 1,4,

f- Two-stage VCHSC: cycle with a rectifier and a desuperheater Pompe • chaleur bi~;tag~e ~i compression de vapeur et circuits: cycle avec un rectfficateur et un (#surchauff'eur

constitutes the IT absorber and the IT desorber heat exchanger, the HT absorber, the low and the high temperature solution heat exchangers (SHX), the rectifier and the desuperheater. All heat exchangers are counter current.

Heat from the LT heat source is picked up in the LT desorber by the working fluid. The evaporating mixture is then separated in a phase separator where the vapour (rich in refrigerant) exiting at the top is fed to the compressor and compressed to a higher pressure level and the liquid weak in refrigerant coming out at the bottom is pumped to the same high pressure in a solution pump. The weak solution absorbs the high pressure vapour delivered by the compressor in the IT absorber. The resulting strong solution exchanges heat with the weak

solution in the LT SHX and is finally expanded down to the low pressure in an expansion valve.

The HT stage is set up in the same way as the LT stage except that the IT desorber receives heat in its totality from the IT absorber and the HT absorber delivers heat to the HT sink. Since the mixture composition varies during evaporation and condensation, the constant pressure desorption and absorption processes are associated with a change in temperature of the working fluid. This variation of solution temperature is referred to in this text as solution temperature glide (STG). The STG can be changed by varying the pump flow rate.

It should be noted that when a mixture with boiling points not too far apart is used such as ammonia-water, as is the case in this study, the vapour desorbed in the IT

Rev. Int. Froid 1993 Vo116 No 4 249

Performance enhancement of a two-stage heat pump. M. V. Rane et al.

Figure 3 Figure 3

9

32 3 '

LT S H X

5

,3 °I , =

1

~8

.... I 3

22 H T S H X

e l +

<. <

<

Two-stage VCHSC: modified cycle with a desuperheater and without a rectifier Pompe h chaleur bi~tag~e h compression de vapeur et circuits." cycle modifi~ avec un d~surchauffeur et sans rect!ticateur

desorber is weak in refrigerant (ammonia) as compared to that desorbed in the LT desorber. This is due to the solution concentration in the IT desorber being lower than that in the LT desorber. Since the two vapour streams are compressed together in one compressor, the compressed vapour absorbed in the intermediate and high temperature absorbers is of the same concentration (somewhere between the concentrations of the vapours desorbed in the low and intermediate temperature desorbers). As a result, water from the HT loop is transferred to the LT loop. This shifts the average solution concentrations in the low and the high temperature loops closer together and the heat pump will fail to operate in steady state condition.

Conventionally, water balance is achieved by rectify- ing the vapour emerging from the IT desorber so that its concentration is the same as that of the vapour emerging from the LT desorber. In the rectifier, the vapour is purified by exchanging heat with a cold sink. During this process, a small portion of the vapour stream condenses and is then used as a reflux stream. Since the vapour streams at both the rectifier and the LT desorber outlets are ideally in equilibrium at the same pressure and concentration, they are consequently at the same temperature. Therefore, the rectification heat can only be picked up by the LT desorber heat transfer fluid. This reduces the cooling capacity of the heat pump and adversely affects the cooling COP. One could argue that the weak solution entering the LT SHX can pick up the rectifier heat. However, this will increase the temperature of the strong solution leaving the LT SHX and consequently will also lead to reduced LT desorber cooling capacity. The penalty on COP is the same for the two cases discussed above.

Compressed vapour at the outlet of the compressor is superheated and its temperature is much higher than the sink temperature. Desuperheating the compressed vapour reduces the heat rejected in the absorbers. A reduced IT absorber load means that less heat needs to be picked up by the IT desorber which in turn decreases the amount of vapour desorbed. This reduces the compressor work and hence increases the cooling COP. For a given compressor, desuperheating would also lead to

increased cooling capacity because of reduced vapour flow rate to produce the cooling effect.

Cases 3 and 4 are similar; the only additional component in Case 4 is the desuperheater while the rectifier is eliminated. Figure 3 shows the layout of the heat pump with a bleed line and a desuperheater (Case 4) similar to the laboratory prototype at the University of Maryland. As can be seen from Figure 3, Case 2 can be modified to Case 4 by removing the rectifier and incorporating a bleed line.

A new scheme to eliminate the need for the rectification in the two-stage VCHSC i s proposed by incorporating a bleed line. Water balance in the high and the low temperature circuits is now maintained by bleeding 2 to 5% of the weak solution flow from the outlet of the LT SHX to the inlet of the HT expansion valve: Eliminating the rectifier and introducing the bleed line reduces the losses in the system. The rectification heat no longer needs to be picked up by the LT desorber. This increases the cooling capacity of the heat pump and results in a higher cooling COP. Now the rectifier loss is traded for the mixing loss due to the bleed line.

The HT and IT absorbers could be put in series and so can the LT and IT desorbers. This will lead to a DAHX (desorber absorber heat exchange) cycle which does not need a separate rectifier and desuperheater. ThiS cycle is being investigated at the University of Maryland.

Computer simulation model

In order to formulate the necessary governing equations and to simplify the system modelling, the following assumptions were made:

1. Pressure drops in the system are negligible (except for the two expansion valves).

2. Heat losses to the surroundings are negligible: 3, The state of the strong solutions leaving the interme-

diate and high temperature absorbers is saturated. 4. The efficiency of the two l iquid pumps is 7 0 % . 5. L M T D is based on the inlet and exit temperatures of

each heat exchanger. LMTD calculated in this fash- ion is a fair representation of actual L M T D (based on changing heat capacity) because of small temper-


Performance enhancement of a two-stage heat pump: M.V. Rane et al.

ature glide (less than 15 K) which results in small change in concentration (less than 20%) in the absorber and desorber.

The steady state simulation model is based on mass and energy conservation equations along with heat transfer relations and thermodynamic properties of the mixture. The following equations may be written at every component of the heat pump:

Mass conservation of refrigerant E ( F * X ) i n = Y',(F*X)om

Overall mass balance

Energy balance

absorber~tesorber for which a UA is calculated and compared with the input UA. If the two UAs are not equal, the weak solution concentration of the HT loop is adjusted and the following steps repeated. Next, the enthalpy at HT expansion valve inlet (state 19) is compared to the enthalpy at stage 17. If the two enthalpies are not equal, the volumetric flow rate at 10 is adjusted and all subsequent phases redone. Once the enthalpy balance is established, the state at the LT desorber inlet is sought by simulating the LT expansion valve. Finally, the LT desorber is simulated and its UA calculated and compared to the UA given as input data. If the two UAs are not equal, the temperature at the LT desorber outlet (state 9) is adjusted and all calculations repeated from the beginning.

Heat exchangers

Compressor and pumps

E(F*H)in = ~(F*H)om

E(F*H)in + W

Results

Simulation conditions

= ~ ( F * H ) ....

Heat transfer Q = UA(LMTD)

Given as input data to the computer program are the glycol-water inlet and outlet temperatures in the high temperature absorber and low temperature desorber, the weak solution concentration in the first stage solution circuit (low temperature stage), the volumetric flow rates of the compressor and the two solution pumps and the products of the overall heat transfer coefficient and heat transfer area of all heat exchangers. Isentropic and volumetric efficiencies of the compressor are obtained by polynomial functions fitted to experimental data of a screw-type compressor. The thermodynamic properties of the working fluid mixture, i.e. ammonia water, used for the computations, are evaluated from computer subroutines described in Ziegler and Trepp 9. The quanti- ties calculated by the simulation model consist of temperature, pressure and mass flow rate at all state points corresponding to Figure 3 as well as heat flows, effectiveness and gliding temperature differences at every heat exchanger in the cycle. Also calculated are the cooling coefficient of performance of the heat pump defined as the ratio of the LT desorber load to the sum of the compressor and pump work inputs.

The model's logic is shown in the flowchart in Figure 4. The simulation process begins first by guessing the mix- lure temperature at the LT desorber outlet (state 9 in Figure 3). With this temperature and the given circulat- ing mixture composition, the low side pressure is calculated. At this point, the volumetric flow rate at state 10 as well as the weak solution of the HT stage are guessed. The temperature at the HT absorber outlet (state 15) is also guessed and the high side pressure is found. Using these data, the compressor, the desuperheater, the two pumps and the solution heat exchanger of the HT stage are simulated. Next the HT absorber is simulated and its UA determined. If the UA calculated is not equal to the desired UA (given as input data), the temperature at the HT absorber outlet is adjusted, a new high pressure level is obtained and the simulation of the compressor, the desuperheater, the two pumps and the HT SHX is redone. Once the compared UAs are equal (within a given tolerance), the solution heat exchanger of the LT stage is simulated followed by the intermediate

The performance of the four versions of the two-stage VCHSC was simulated at the following conditions:

1. (a) Heat source temperature: Inlet T(28) = 0°C and outlet T(29) = - 5°C

(b) Heat sink temperature: Inlet T(26) = 100°C and outlet T(27) = l l0°C

The glycol water temperatures at the inlet and outlet of the heat source and heat sink were specified as above. The mass flow rate of glycol water was determined by the LT desorber and HT absorber loads.

2. The compressor volume flow rate was fixed at 0.01 m 3 s-l for all cases.

3. The sum of all UA values (products of overall heat transfer coefficients and the heat transfer areas) of all heat exchangers was assumed to be 7 kW K ~. This value was chosen because the LMTDs in the heat exchangers (except the desuperheater) are about 5 K for a LT weak solution concentration of 60 wt% ammonia, and the LT and the HT pump flow rates are2.5 x 1 0 - S m 3 s - l a n d 3 . 5 x 10 5m3s ~ respectively for Case 1. The values for the LT weak solution concentration and the low and the high temperature pump flow rates were selected in the middle of the range of concentrations and flow rates investigated.

4. The optimal distribution of UA values among the different heat exchangers was obtained for each of the four cases with the help of an optimization program based on Hooke and Jeeve's direct search ~°. The objective function, i.e., the cooling COP, was maximized by changing the UA values indepen- dently during the optimization subject to the con- straint that their sum equalled 7 kW K-I . The opti- mum allocations of UA values in the four cases are listed in Table 1. This distribution was then maintained for all other performance evaluations at other weak solution concentrations. The sum of UA values for cases 1 and 2 exceeded the total value of 7 kW K ~ by an amount equal to the rectifier's UA. This is due to the fact that the rectifier UA is calculated (assuming ideal rectification) and is not an input variable (that can be changed by the optimization program) as are the other UA values.

5. The weak ammonia water solution concentration in the LT loop was varied between 40 and 80 wt%



GUESS TEMPERATURE @ g T(g~T(28~DT

FLOW RAIE @ I0,

IN SECOND SOLU.CIRC

GUESS TEMPERATURE @ IS T(15)'T(26)*OT

| SIMULATE COPlPRESSOR |

i SIMULATE OESUPERHEATER l

| SIMULATE LT PUMP ]

I S , ~ A . E . r ~u.~ I

I S,.ULATE Hr SHX I

SliIULA [E LT SHX

ADJUST T(15)

AOJUST X(23)

ABSORBER-DESORBER

ADJUSTFV(IO) I " ' -

T RESU~.TS

Figure 4 Flowchart of the two-stage vapour compression heat pump with solution circuits Figure 4

YES

Logiciel pour une pompe h chaleur bidtagke h compression de vapeur avec des circuits de solution

Table 1 Distribution of UA values for different heat exchangers after optimization of the cycles Tableau 1 Distribution des valeurs UA pour diff~rents dchangeurs de ehaleur apr~s optimisation des cycles

Case l Case 2 Case 3 Case 4 UA values (kW K ~) (kW K ~) (kW K ~) (kW K ~)

LT desorber 1.35 1.50 1.35 1.50 LT SHX 0.91 0.83 0.91 0.83 IT absorber/desorber 1.84 1.64 1.84 1.64 HT SHX 1.19 t.26 1.19 1.26 HT absorber 1.71 1.62 1.71 1.62 Rectifier 0.03 0.03 N/A N/A Desuperheater N/A 0. l 5 N/A 0.15

Total UA 7.03 7.03 7.00 7.00

a m m o n i a . T h e m i n i m u m w e a k s o l u t i o n c o n c e n t r a - t ions inves t iga ted in this s tudy are d i f ferent fo r d i f ferent s i m u l a t i o n c o n d i t i o n s because the a b s o r b e r a n d or the d e s o r b e r a p p e a r overs ized at ve ry low c o n c e n t r a t i o n s wh ich leads to ve ry [ow capac i t i e s and a fa i lu re o f the m o d e l to c o n v e r g e .

P e r f o r m a n c e cu rves fo r the best cycle, the case wi th h ighes t C O P , were then o b t a i n e d b y s i m u l a t i n g the cycle at v a r i o u s L T so lu t i on p u m p f low rates . L T so lu t ion p u m p flow ra tes o f 1.0 x 10-5. 1.5 x 10 5. 2.5 x 10-5 a n d 3.5 x 10-5 m 3 s -1 were s tudied. T h e H T so lu t i on p u m p flow ra tes were 1.5 x 1 0 - 5 2,5 x 10-5. 3.5 x 10-5 a n d 4.5 x 10-5 m 3 s - l . S l ight ly h igher f low ra tes in the H T l o o p lead to be t t e r m a t c h in the S T G s m the a b s o r b e r / d e s o r b e r hea t exchange r . T h e p u m p f low ra tes were c h o s e n as i n d e p e n d e n t va r i ab l e s because they



~ D 0 8 0 (D

• - - 0 7 0

o o O

£

~J

0 60 LT Stage Purrp Flow Rate 2 5E-5 m3,'s

k H T S'age :um~ Flaw Rate 3 5E 5 m 3 / s

, modified c~,cle with bleed I~ne cod desuperheafer 0.50 modified *ycle with bleed qne

cycle wth rectifier and desuperheater c~cle w 'h reef fieF

oao L . . . . . . . . t , l l . - - k k ~ l , t t L l ~ . . . .

35 4 5 55 65 75 8 5

Weak Solution Concentrat ion (wry ammon ia )

Figure 5 Figure 5 c voles

Comparison of cooling COPs for the four cycles Comparaison des COP de rejroidissement pour les quatre

800

?

6.00 ~g O

E03 400

o i

k 2O0

0 O0 35 . . . .

Figure 6 Figure 6 c.rcles

,i,~T,, , ~ " --TTTr ' ' : T ' " ' r T T , r r~q i T T T rTTT]

I

LT Sloge Pump Flow Role 4 25E-5 m3/s 2

H.T. Sloge Pump Flow Role 55[ 5 m S / s

• modified cycle with bleed line and desuPerhealer modified cycle with bleed line

• cycle with rec~Vier and desuperheater * cycle with reef;fieF

I ± , ~ l ± t l t t . . . . . . . : _ * . I L l . . . . , L ~ I , 45 55 65 75 85

Weak Solution Concentration (wt' ammonia)

Comparison of cooling loads for the four cycles Comparaison des charges de re/roidis'sement pour les quatre

influence the solution temperature glides and hence the performance of the VCHSC.

Results and discussion

Figure 5 represents the variation of COP versus weak solution concentration for the four cases. The cooling COP of the cycle with the rectifier (Case 1) is the lowest among the four cases. The maximum cooling COP for Case 1 is 0.72. Among the four cases, the best performance is obtained for the cycle with the bleed line and the desuperheater (Case 4), with a maximum cooling COP of 1.05. Over the entire range of weak solution concentration, the cooling COP for Case 4 is about 50% higher than that for Case 1.

The advantage of desuperheating the vapour entering the absorber is very clear, and can be seen by comparing the COP curves for Cases 1 and 3 with those for Cases 2 and 4 respectively. The desuperheater reduces the temperature and enthalpy of the vapour entering the absorbers by transferring the superheat to the high temperature sink. This reduces the amount of heat exchanged in the absorbers. A reduction in the intermediate temperature (IT) absorber duty reduces the IT desorber duty. This leads to reduction in the amount of vapour desorbed in the IT desorber, thereby reducing the mass flow rate through the compressor for a fixed LT desorber load. In other words the compressor work per unit cooling capacity reduces and the cooling COP increases. For the operating conditions considered here, this leads to an increase in cooling COP of as much as 20%.

The reduction in the HT absorber duty does not contribute to the improvement in COP. However, desuperheating the vapour entering the HT absorber might make sense from a practical standpoint (cost consideration). By exchanging the heat in the desuperbeater instead of the HT absorber, smaller heat exchange areas may be required because the LMTD in the desuperheater would be much higher than those encountered in the HT absorber. On the other hand the heat transfer coefficients for exchanging heat with the vapour in the desuperheater may be much lower than those for an absorbing solution in the HT absorber. Thus, based on these trade-offs, of

the two factors mentioned above, it might be desirable to desuperheat only the vapour being absorbed in the IT absorber without sacrificing the improvement in COP.

Replacing the rectifier at the outlet of the IT desorber with the bleed line between the LT SHX outlet and the HT expansion valve inlet also leads to significant improvement in COP. Comparing the COP curve for Case 1 with Case 3 and that for Case 2 with Case 4 indicates that cooling COP enhancement of the order of 25 to 30% is achievable. By incorporating the bleed line and eliminating the rectifier, the losses associated with the rectifier are traded for the losses associated with the mixing of the relatively strong solution, at the outlet of the LT SHX, with the solution at the inlet of the HT expansion valve.

The rectifier needs to be cooled, and the only low temperature sink available is the LT desorber (unless the cycle is opened to a third temperature level, in which case the difference between the cooling and the heating COPs will no longer be 1). By dumping the heat of rectification in the LT desorber, the net cooling capacity of the cycles with the rectifier is about 30 to 40% lower than the LT desorber load. This reduces the cooling COP by about the same percentages.

On the other hand, in the cycles with the bleed line there is no direct penalty on the net cooling load. How- ever, due to the mixing of two solution streams with different concentrations, the IT desorber load increases slightly. This is to return the excess ammonia delivered in the HT loop back to the LT loop. It turns out that this loss is much smaller than the rectification loss and explains the higher COPs obtained for Cases 3 and 4,

Figure 6 shows the variation of cooling load with respect to the LT weak solution concentration. The cooling load increased four-fold when the weak solution concentration in the LT loop was changed from 40 to 80 wt% ammonia. The increase in cooling load is due to the increase in the desorber pressure, which results from increased solution concentration for a fixed desorber temperature. Increased desorber pressure means increased vapour density at the inlet of the compressor. Since the compressor displacement is constant, the vapour mass flow rate is directly proportional to the density of the vapour at the compressor inlet•

Rev. Int. Fro id 1 9 9 3 V o 1 1 6 N o 4 2 5 3

Performance enhancement of a two-stage heat pump." M. V• Rane et a l.

1 I0 ::::::::::::::::::::::::::::::: : : : . . . .

! i [

10C i

~Z) 0 9 0

oO 0 8 0 (.D

7

/

LT Sioge Pump FIo~ Rote 2 5 E - 5 m S / s

H T Stoge Pump Flow Rote • I 5E-5 m 3 / s

2 5 E - 5 mS/s u 35E-5 m3/s

4 5 E 5 m 3 / s

118 : :- :!::: . . . . . . . . . . . . . . . .

ca_ 0 } 90 C)

S}Gge F( }q F) [}igw Rqt,-t 0 3 5 E - 4 m S / s

[ T Stage Pump Flow Pate • I OE-5 m S / s : !5E-5 m 3 / s

2 5 E 5 rn3}/s SF -5 m3/~

0 60 : • i J ; ~ 2 ~J i L , [ [ £ ~ L I [ J I ~ [ , ; J 1 t [ 35 45 55 65 75

Weak Solution Concentrat ion (wt:, ammon ia ) _V, 45 c?,

v V e a k S o l u t i o n C o n e e n ! r c t i o r ' ( w t a m m o n i a )

85

Figure 7 Variation of cooling COP with the low temperature loop weak solution concentration and high temperature pump flow rate Figure 7 Variation du COP de refoidissement avec la concentration en solution pauvre dans la boucle basse temperature pour d(ff'~rents d~bits masse h haute tempdrature

Figure 8 Variation of cooling COP with the low temperature loop weak solution concentration and flow rate Figure 8 Variation du COP de refroidissement arecla concentration en solution pauvre et le ddbit masse dans la boucle basse temperature

The cooling loads for Cases 3 and 4 are about 10% higher than those for Cases 1 and 2 respectively. This is due to reduced mass flow rate of vapour per unit cooling capacity for the cycles with the bleed line as compared to those with the rectifier. The reasons for the reduction in mass flow rate of vapour per unit cooling capacity for Cases 3 and 4 have been discussed in the previous paragraphs.

Another important observation that can be made from Figure 6 is the increase in cooling capacity for Cases 2 and 4, cases with a desuperheater, as compared to Cases 1 and 3, the cases without a desuperheater. The increase in cooling capacities is in the range of 25 to 30%. There- fore, the total increase in the cooling capacity for Case 4 as compared to that for Case 1 is in the range of 30 to 40%. This means that Case 4 not only has about 50% higher cooling COPs, but simultaneously has about 30% higher cooling capacities.

The cycle with a bleed line and a desuperheater, Case 4, which has the best performance was then selected for further investigation. Case 4 was simulated at various solution pump flow rates in the high and low temperature loops and weak solution concentrations in the LT loop. The simulation results are analysed in the following paragraphs•

Figure 7 shows the variation of cooling COP with the LT loop weak solution concentration and the HT loop solution pump flow rate. For a fixed LT pump flow rate of 2.5 x 10 5 m 3 s-~ the cooling COP curves peak at the LT weak solution concentrations of 45, 50, 55 and 60 wt% ammonia for the HT pump flow rates of 1.5 x 10 -5, 2.5 x 10-5, 3.5 x 10 5 and 4.5 x 10-5 m 3 s : respectively. The location of the peak in the cooling COP curve is influenced by the match in the temperature glides in the three main heat exchangers, namely the LT desorber, the IT absorber/desorber and the HT absorber. A good match in the temperature glides in these heat exchangers leads to low pressure ratios across the compressor and lower compression work. This results in higher COP.

The variation in COP with changing HT pump flow rate at any particular LT loop weak solution concent-

ration can also be explained based on the matching of the temperature glides.

Similar trends in COP curves were observed at LT pump flow rates of 1.0 x 10 5, 1.5 x 10-5 mad 3.5 x 10 5 m 3 s ~ (the graphs for these flow rates are not included here to restrict the length of this paper). It was observed that the cooling COPs were highest for the HT pump flow rate of 3.5 x l0 Sm 3s t. This again is the effect of better matching of the temperature glides in the case of an HT pump flow rate of3.5 x 10-5 m 3 s-~: This is probably due to the fact that the COP of the system was optimized (with respect to UA distribution) at this HT pump flow rate.

Variation of cooling COP with the LT loop weak solution concentration and pump flow rate is shown in Figure 8. Here again there is only one peak in the COP curve for each of the LT pump flow rates. The location of the COP peak along the concentration axis depends on two main parameters: the pressure ratio across the compressor, and the SHX effectiveness• The pressure ratio depends on the match in the temperature glides in the three heat exchangers: the LT desorber, the IT absorber/ desorber and the H T absorber. The influence of these parameters will be analysed in detail in the following paragraphs.

Figure 9 shows the variation of pressure ratio (the ratio of the absorber to desorber pressure) with the LT loop weak solution concentration and pump flow rate. As one would expect, the curves for pressure ratio show the inverse behaviour of the COP curves. Thus, the gain in COP is primarily a consequence of the reduction in pressure ratio. As the pressure ratio increases, the work increases, and consequently the COP decreases, The variation in the pressure ratio is dependent on the match in the temperature glides on the working fluid side and the heat transfer fluid side of the L T desorber and the HT absorber, and the match in the temperature glides of the solution on the absorber and the desorber sides of the IT absorber/desorber heat exchanger.

The variation of the solution temperature glide in the LT desorber is shown in Figure ! O. The maximum glides for all the LT pump flow rates are observed at a weak solution concentration of about 50 wt% ammonia and


Performance enhancement of a two-stage heat pump: M.V. Rane et a l.

f H.T Pure Flow Ra~e

f L .T Stage Pump Flow Rate * 1.0E 5 m~/~m s ]

15E 5 ! 825 - 2 g E - 5

5 5 [ - 5 m g / s

o 800 [

° f GE

~ z s L

759

EL

7 25

7 O0 ,

i

4

"~ f#.. •

675 : . . . . . . . . . ii, :~: tttl . . . . . .

55 45 55 65 75 85

Weak Solulion Concentrafion (wf': amrnonial

Figure 9 Variation of pressure ratio with the low temperature loop weak solution concentration and flow rate Figure 9 I/ariation du laux de compression avecla eoncenlration de la solution pauvre et lu d~;hit masse chins la boucle hasse temp~;rature

<

.C

c~ q~

25 ~ - ~ - r y T l " r ' r , ~ - - - ~ T : T T T " r [ I [ r i " ~ ' r r , -

H I 0 55EStage,& P u r ¢~:)sm3 Flow Rate

L T Stage P u m p g lo~ Rate

¢*

J

5

- 1 0 E - 5 m 3 / s I 5 E - 5 m S / s 2 5 k - 5 rT'5 / s 3 5 [ 5 m 3 ~ s

0 1 i l l , : . L , 1 . . . . . . . . ;

55 45 5 5 6 5 75 85

Weak Solution Concentration ( w t ornmonic)

Figure I 1 Variation of absorber glide with the low' temperature loop weak solution concentration and flow rate Figure 11 Variation du glissement de l'absorheur avec la concentration de la solution pauvre et le &;bit masse dans la boucle hasse temperature

2

'2)

o oJ

C2:1 15

E

o

5.

]D

H T Stage Pump Flow Ra'e 0 . 3 5 E - 4 m 3 / s

LT . S lage P u m p Flow Ra'e * 1 . O E - 5 mS/'s

2 . 5 E - 5 55E 5 mSIs

, 335 J , I , , . 4 5 " 1 . 1 9 5 ~ , i , . . . . 6 5 1 1 , ~ 1 . ~ 7 5

Weak Solution Concentration ( w f ammonia ~5

Figure 10 Variation of desorber glide with the low temperature loop weak solution concentration and flow rate Figure 10 Variation du glissement du dksorbeur avee la concentration de ht solution pauvre el le dkbit masse duns la boucle basse temp~'rature

2 5 : T : T . . . . I I T . . . . T : ' r " ' " ~ 1 - , r r l r : -

2O

~D x:]

0

q~ £:5

.E

- - 5 0 e ~

I

e

t~ T S lage P u m p Flow Ra'e 0 5 5 E 4 m3/s

L * S tage P u m p Flaw Rate • I 0 E - 5 m S / s , I 5 [ - 5 r n 3 / s

2 5 E 5 m3 , / s 5 5 [ 5 rT/3 'S

i [ L 1 1 . . . . . . i [ [ ] _ i i

4 5 5 5 6 5 7 5 ~ 5

Weak Solution Concentration ( w t ' ammonia)

Figure 12 Variation of absorber glide with the low temperature loop weak solution concentration and flow rate Figure 12 Variation du glissement du d6~orbeur avecla concentration de la solution pauvre et le d~;bit masse dans la bouch' hasse temp~;rature

the glides taper off at the high and the low end of the solution concentration. This is not surprising, because by moving towards the high or the low end of the solution concentration we are in fact approaching the cases of pure fluids where the glides will essentially be zero. The glides decrease with the increase in the LT pump flow rate because the desorbing solution has to undergo smaller changes in concentration to cater for the same cooling load. At any particular weak solution concentration, the cooling load remains almost unchanged because the compressor volumetric flow rate is constant. The horizontal line on the figure represents the heat transfer fluid temperature glide of the LT desorber which was fixed at 5 K.

The solution temperature glides in the IT absorber have similar trends (Figure 11). However, the temperature glides in the IT absorber are higher than those observed in the LT desorber. This is explained by the

temperature of the vapour being absorbed in the IT absorber which is much higher than the saturation temperature of the solution. This in turn causes the heat duty of the IT absorber to be higher than that of the LT desorber.

The solution temperature glides for the IT desorber (Figure 12) and the HT absorber (Figure 13) increase as the weak solution concentration in the LT loop varies from 40 to 80 wt% ammonia because for these cases the weak solution concentration in the HT loop is always below 50 wt% ammonia. This means that these curves represent the first half (low concentration side) of the overall profile shown in Figures 10 and 11. Also, note that the solution temperature glides in the IT desorber and the HT absorber are of the same order of magnitude. This is because the vapour being absorbed in the HT absorber is desuperheated and is at the maximum temperature in the absorber. The temperature glide on the

Rev. Int. Froid 1 9 9 3 V o 1 1 6 No 4 2 5 5


25 ~ i b i I r T-I F : 7 ;TTTT- - - - T r T : 1 r f ! ' " T~ q T T " : f - "

L 2 O -

L

/D

E

=,o - - -./::::: ......

T L ~ : : ~ HT Stage Pure Flow Rote

E [ /z / "'- 0.35E-4 rn3~s ' - - ~ / / / - / ~ " L>T S tage Pump Flow Rate ~2 5 [ ~(/ .- , . IOE-5 m3/s

7:3 i[ ~ 15E-5 mS/s I ~" ~ 2 .5E -5 m3 /s

35E 5 mS/s

[

35 45 55 65 75 8fi

Weak Solution Concentrat ion (wt~ ammon ia )

Figure 13 Variation of absorber glide with the low temperature loop weak solution concentration and flow rate Figure 13 Variation du glissement de l'absorbeur avec la concentration de la solution pauvre et le dkbit masse dans la boucle basse tempkrature

10000

c~ g " 50

% 9500

L~

92 5C

L~

E o., 90 0O

$ I 8 " 0 0

. . . . t ~ - . :<':

~r s,oge P~mp n,~ Rc,~ 0 35 [ - 4 mS /s

L T Stage Pump [ o~ Rate ' ! 0[--5 m3, / s

1 . 5E .5 m f / s 2.5[-5 m3/s 35E 5 ,~31s

~5 45 55 65 75 !

Weak Solution Concentration (wt~ amman]o)

Figure 15 Variation of HT SHX effectiveness with the low temperature loop weak solution concentration and flow rate Figure 15 Variation de l'efficacitk de l'@changeur de chaleur de la solution h haute temp@rature avec la concentration de la solution pauvre et le d#bit masse dans la boucle basse tempbratare

100 .00

9750 Co .>_ Z

9500

UJ

TX 92 50 La

&

90 .00

j a 87 .50

q ~ T T - ~ I F i w + l q r ~ - -T r r T . - - - nT v r T-r ~ T ] r [ * : T : r ~ r ]

. . . . - - _ , ~ : : ~: ;=== :=~ [

/./

o"

/ ,~w/Z

~.j//1

.49

<,

H f Stage Pum Flow Rate 0 ,35E -4 rm 3 >Ps

LT Stage Pump Flow Rote * IOE-5 r n3 / s

15E -5 mS /s 2 .5 [ - 5 m3/s 3.5E-5 m3/s i

3 5 4 5 5 5 65 7 5 0 5

Weak Solution Concentration (wt;~, ammonia)

Figure 14 Variation of LT SHX effectiveness with the low temperature loop weak solution concentration and flow rate Figure 14 Variation de l'efficacit~ de l'bchangeur de chaleur de la solution ~ basse temperature avec la concentration de la solution pauvre et le dbbit masse dans la boucle basse tempbrature

heat transfer fluid side of the HT absorber was fixed at 10 K as shown on the figure by the horizontal line.

Now, by looking more closely at Figures 10 to 13, one can see that the pressure rating curves go through a minimum at a weak solution concentration where there is a best match in the temperature glides. That is to say that the LT desorber glide is close to 5 K, the glides in the IT absorber and the IT desorber match each other and the HT absorber glide is close to 10 K.

The other important parameter influencing the COP is the SHX effectiveness. The variation o f the LT and HT SHX effectiveness with the LT loop pump flow rate and weak solution concentration is represented in Figures 14 and 15. The LT SHX effectiveness decreases with an increase in the LT pump flow rate because the heat duty consistently increases with the solution flow rate, which in turn results in higher L M T D s (log mean temperature differences) for a fixed U A value. On the other hand, the

SHX effectiveness increases with higher weak solution concentratton because of a reduction in the SHX heat duty (the solution density and consequently the solution mass flow rate decreases with increases in concentration). This results in lower L M T D s and higher SHX effectiveness since the heat exchanger UA is constant. The HT

I SHX effectiveness is not significantly altered because the HT pump flow rate remains the same for the four LT pump flow rates.

The influence of SHX effectiveness on the COP of the cycle is significant. To best illustrate this let us compare the COPs. pressure ratios and the SHX effectiveness al the LT loop weak solution concentratton of 60 wt% ammonia. Figure 9 shows that the pressure ratios for the four flow rates are almost the same. However. the COPs in Figure 8 under the same conditions are significantly different. This variation in the COP is very similar to the variation of the LT SHX effectiveness in Figure 14. The minor differences can be accounted for by considering the variation o f the HT SHX effectiveness.

Since the SHX effectiveness changes with pump flow rate. the optimization program should have included pump flow rates as additional variables to obtain an overall opt imum distribution of UA values among the heat exchangers. However. due to limited computational power this scheme was not pursued.

Figure 16 shows the variation of cooling load with respect to the LT weak solution concentration. The cooling load increases four times when the weak solution concentration in the LT loop is changed from 40 to 80 wt% ammonia. The desorber load decreases with increases in the LT pump flow rate. As discussed earlier. an increase in the solution flow rate decreases the SHX effectiveness. This increases the temperature and the enthalpy o f the solution entering the LT desorber and thus leads to the decrease in desorber capacity. The curves for the desorber capacity are flatter at the higher concentrations because the vapour pressure o f the solution does not change significantly.

The simulation model was validated against experimental data TM. In general, simulated COPs were within 5-10% of the experimental values.



~ 7 O

6 0

© 5 @

._E 4 .O

z ©

(D 5 C

: ~ r T ~ - i i ~ 1 T - - ~ l l , , = T I , I 4 q 2

::::2a

J

H T S tage Pure Flow Rate . . . . . . . . 37~ 1 J L T Stage Pump Flaw Rote

1 0 E - 5 m 3 / s i ~ * 1 5 E - 5 m S / s

2 C ~ 25E 5 mS/s 35E 5 mS,Is i

35 45 55 65 75 85

Weak Solution Concentration (wt~' ammonia)

Figure 16 Variation of cooling load with the low temperature loop weak solution concentration and flow rate Figure 16 Variation de la charge de refroidissement avec la concentration de la solution pauvre et le d~'bit masse clans la boucle basse temperature

concentrations and decrease with increases in solution pump flow rates.

A change in cooling capacity by up to a factor of four can be achieved by varying the LT weak solution concentration from 40 to 80 wt% ammonia.

Compared to conventional single-stage ammonia vapour compression cycles, the two-stage VCHSC provides cooling COPs twice as high and pressure ratios of about two thirds less.

Finally, this analysis has shown that the two-stage VCHSC with a bleed line and a desuperheater can operate at temperatures above 100°C and reach temperature lifts beyond 100 K with well known and environmentally safe refrigerants.

Acknowledgements

The support for this study by the National Science Foun- dation, the US Department of Energy, the American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. and the University of Maryland, Engi- neering Research Center is gratefully acknowledged.

Conclusions

Based on a computer simulation model, the steady state performance of a two-stage vapour compression heat pump with solution circuits has been predicted and analysed in detail. The results indicate that the modified cycle with a bleed line and a desuperheater has the best overall performance. Compared to the basic two-stage VCHSC with a rectifier, improvements in cooling COPs and cooling loads by as much as 50 and 40% respectively are achieved. Both cycles, i.e., the basic cycle with a rectifier and the modified one with a bleed line, showed up to a 20% increase in cooling COP and capacity by incorporating a desuperheater.

The pressure ratio curves show the inverse behaviour of the COP curves, with values ranging from 7 to 8.5 when pumping heat across a temperature lift of 100 K. The pressure ratio is found to depend on the match in the temperature glides on the working and heat transfer fluid sides of the LT desorber and HT absorber as well as on the match in the temperature glides of the solution in the IT absorber/desorber heat exchanger. The glides decrease with an increase in the LT solution pump flow rate.

The SHX effectivenesses are sensitive to the variation in the solution concentrations and pump flow rates. Both effectivenesses are found to increase with higher solution

References

1 Alefeld, G. Rules for the design of multi-stage absorption and compression heat pumps (in German) Brennstoff: Warme-KraJ? (1982) 34 142--162

2 Alefeld, G. Multi-stage apparatus having working fluid and absorption cycles and methods of operation thereof US patent No 4 531 374 (1985)

3 Radermaeher, R. An example of the manipulation of effective vapor pressure curves by thermodynamic cycles Trans ASME; J Eng Gas Turb Power (1988) 10

4 Radermaeher, R. Advanced versions of heat pumps with zeotro- pic refrigerant mixtures A S H R A E Trans (1986) 92

5 Pourreza-Djourshari, S., Radermaeher, R. Calculation of the performance of vapor compression heat pumps with solution circuits using the mixture R22-DEGDME Int J Re[?~ (1986) 9

6 Radermaeher, R., Zheng, J., Herald, K.E. Vapor compression heat pump with two-stage solution circuit: Proof-of-concept- unit International Workshop on Absorption Heat Pumps, Lon- don, UK (1988)

7 Raue, M.V., Radermaeher, R. Feasibility study of a two-stage vapour compression heat pump with ammonia water solution circuits: experimental results lnt J R¢fft'ig (1993) 16 (4) 258 264

8 Ziegler, F., Hammer, G. Experimental results of a double-lift compression-absorption heat pump BHRA Confi'rence, Munich, Germany (1990)

9 Ziegler, B. and Trepp, C. Equation of state for ammonia water mixtures lnt J Re]Pig (1984) 7 10l 107

10 Hooke, R., Jeeves, T.A. 'Direct Search' solution of numerical and statistical problems J Assoe Cornput Machin (1961) 8 212 229

11 Amrane, K. Simulation of vapor compression heat pumps with solution circuits PhD thesis University of Maryland, USA (1991)


Documents

Performance enhancement of a two-stage vapour compression heat pump with solution circuits by eliminating the rectifier