A small capacity steam-ejector refrigerator: experimental investigation of a system using ejector with movable primary nozzle

ELSEVIER

Int J. Refrig. Vol. 20, No. 5, pp. 352-358, 1997 © 1997 Elsevier Science Ltd and IIR All rights reserved. Printed in Great Britain PII:SO140-7007(97)00008-X 0140-7007/97/$17.00 + .00

A small capacity steam-ejector refrigerator: experimental investigation of a system using ejector with movable primary nozzle Satha Aphornratana Department of Mechanical Engineering, Sirinthorn International I n s t i t u t e o f T e c h n o l o g y , T h a m m a s a t U n i v e r s i t y , P.O. B o x 22 T h a m m a s a t R a n g s i t P o s t Office, P a t u m t h a n i 12121, T h a i l a n d

lan W. Eames I n s t i t u t e o f Building T e c h n o l o g y , D e p a r t m e n t o f A r c h i t e c t u r e a n d Building T e c h n o l o g y , T h e U n i v e r s i t y o f N o t t i n g h a m , U n i v e r s i t y P a r k , Nottingham NG7 2RD, UK Received 22 July 1996; revised 6 January 1997; accepted 6 February 1997 This paper describes an experimental study of a steam-ejector refrigerator using an ejector with a primary nozzle that could be moved axially within the mixing chamber section. The effects on coefficient of performance and cooling capacity produced by adjusting the position of the nozzle were studied. The experimental rig and method are described and results are presented which clearly show the benefit of using such a primary nozzle. Copyright © Elsevier Science Ltd and IIR.

(Keywords:refrigeratingmachine;ejector;design;performance)

R6frig6rateur de petite puissance 5. ejection de vapeur: Etude exp6rimentale d'un syst6me utilisant un 6jecteur buse primaire mobile Cet article ddcrit l'~tude exp~rimentale d'un r~frig~rateur it ~jection de vapeur utilisant un ~jecteur it buse primaire qui peut se dkplacer dans l'axe de la chambroe de m~lange. On a ~tudi~ les effets de la position de la buse sur le COP et la puissance frigorifique. On d~crit le banc d'essai et la m~thode ; on pr~sente les r~sultats qui mettent en ~vidence les avantages de l'utilisation d'un tel systbme. © Elsevier Science Ltd et IIR. (Mots cles: machine frigorifique;6jecteur; conception; performance)

its capital cost and maintenance should make it become a serious competitor with any other cycle 1. Figure 1 shows a schematic diagram of an ejector refrigeration cycle. High pressure and high temperature refrigerant vapour is evolved in a boiler to produce the primary (motive) fluid. This enters the primary nozzle of the ejector, where it expands to produce a supersonic flow that creates a low pressure region within the mixing chamber. This region of low pressure draws vapour from the evaporator (secondary flow). The primary and secondary fluids are combined in the mixing chamber of the ejector. At the mixing chamber's throat, a transverse shock-wave

An ejector refrigerator is similar to an absorption refrigerator, since both can be powered by low grade heat energy, with the addition of a small quantity of work input required to circulate the working fluids. Therefore, both systems can convert waste heat to useful refrigeration and may be cheaper to operate than conventional vapour compression cycles whose energy input is entirely in the form of mechanical work. When compared with an absorption system, an ejector system is relatively simple to construct, operate and control; it also uses only single component working fluid (refrigerant only). Even its coefficient of performance (COP) is relatively low but

352

A small capacity steam-ejector refrigerator boiler

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Figure l A steam ejector refrigeration cycle Figure 1 Cvcle JHgor(fique g~~jection de vapeur

is induced to create a compression effect. The mixed stream is then discharged, via a diffuser, to a condenser, where the vapour is condensed. The liquid refrigerant accumulated in the condenser is returned to the boiler via a pump whilst the remainder is expanded through the throttling valve to the evaporator, thus completing the cycle. As the work input required to circulate the fluid is typically < 1% of the heat supplied to the boiler, the COP may be given as: COP -

heat input at the evaporator heat input at the boiler

Theoretical and experimental studies of a small capacity steam-ejector refrigerator have been presented previously 1. These studies showed that COP and cooling capacity are both dependent on the operating temperatures. The area ratio of the primary nozzle throat to the diffuser throat was also important. For fixed boiler and evaporator temperatures, when the condenser pressure is allowed to be decreased below a certain critical value the COP and cooling capacity were found to be constant and the ejector entrained the same amount of the secondary fluid. This phenomenon may occur due to the flow choking within the mixing chamber. Therefore, an increase in boiler pressure or a reduction in condenser pressure may not always increase the COP and cooling capacity. Furthermore, if condenser pressure is increased to a value greater than a certain critical value, the ejector m a r lose its function completely, causing the COP and cooling capacity to drop sharply to zero. In a previous paper ~, it was concluded that, for a given condenser pressure (limited by the cooling water temperature) and evaporator temperature (limited by the cooling application heat source temperature), a steam ejector refrigerator will provide its maximum performance when the boiler pressure is adjusted in order to allow the ejector to operate precisely at its critical condenser pressure condition. Also, when the geometry of the ejector is fixed, cooling capacity can only be increased by reducing the boiler temperature as the condenser pressure falls or I JR 20-5-£

353

by allowing the evaporator temperature to rise, which may not be possible. From the literature, the performance of an ejector has been shown to be dependent on the position of the primary nozzle2'3. The effect of the nozzle position on ejector performance has not been clearly explained. In practice, ejectors are usually designed with a fixed primary nozzle position. The optimum position of the nozzle within the mixing chamber being experimentally determined. The ESDU design guide2 suggests that the nozzle should be placed at a distance of 0.5 1.0 length of the mixing chamber's throat diameter upstream of the mixing chamber inlet. However, because of the complex nature of the flow structure, it is difficult to give precise recommendations for the optimum nozzle position. In this current paper, the ejector used was designed so that the primary nozzle exit position, relative to the mixing chamber, could be adjusted in order to maximize the system performance when the operating conditions were difference from the design point. Tests were conducted with various boiler, evaporator and condenser saturation temperatures.

An experimental refrigerator An experimental refrigerator with a cooling capacity of 2kW was constructed. Water was used as the refrigerant. Figure 2 shows a schematic diagram of the system. The boiler design was based on the thermosyphon principle. Its maximum heating capacity was

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evaporator cooling load 685 W ¢"1 Tboiler = 120°C, Peon = 30 mbar • Tboiler = 120°C, Peon = 35 mbar O Tboiler= 130°C, Peon = 35 mbar • Tboiler = I30°C, Pcon = 40 mbar

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Figure 4 Variations of the evaporator temperatures with the NXP Figure 4 Variations de temperatures h l'~vaporateur en fonction de la position de la buse (NXP)

7 kW, provided by two 3.5 kW electric heaters. The evaporator was based on a flash type design. A single 3.25kW electric heater was used to simulate the evaporator heat load. All the electric heaters were controlled using variable transformers. A shell and coil condenser was used, cooled by water taken from the laboratory's cooling tower. The boiler was covered with a 30 mm thickness of glass fibre wool with aluminium foil backing. The evaporator was covered with a 20mm thickness of neoprene foam rubber. The average boiler heat loss was estimated to be 25% of the electrical power input and the average evaporator unwanted heat gain was found to be ca 12% 1. Two circulation pumps were employed in the system; a pneumatic diaphragm pump was used to return the liquid water collected in the condenser to the boiler and evaporator, and a magnetically coupled centrifugal pump was used to circulate water through the evaporator. Figure 3 shows a sectional drawing of the test

ejector. It was designed based on methods provided in the literature I-3. The nozzle was mounted on a threaded shaft which allowed the distance between the nozzle exit and the mixing chamber inlet to be adjusted in order to determine the influence of the nozzle position on the performance of the ejector. The nozzle exit position (NXP) was defined as the distance between the nozzle exit and the mixing chamber inlet planes as shown in Figure 3. The NXP has a positive value when the nozzle is placed inside the mixing chamber, and is negative when outside the mixing chamber. The boiler, condenser and evaporator were charged with deionized water. The performance of the experimental refrigerator was obtained by measuring the time averaged electric power input to the evaporator and generator heaters over a steady state running time of 30-60 min.

Optimum nozzle position These tests were conducted by setting the evaporator

A small capacity steam-ejector refrigerator

355

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heat load constant at 685 W (110 4-0.5 V across the heater). At each setting of the boiler and condenser conditions, the primary nozzle position was varied and the evaporator temperature was measured. If the compression ratio (condenser to the evaporator pressures) increased while the entrained fluid flow was fixed (fixed cooling capacity), it was indicated that the ejector performance was improved. The optimum nozzle position was thought to be at the point where the minimum evaporator temperature was achieved. The results of these experiments are shown in Figure 4 and 5. It was found that: For fixed boiler and condenser pressures, the evaporator temperature decreased as the primary nozzle position was moved into the mixing chamber. The temperature can be reduced to a minimum level at some nozzle positions. Further movement of the nozzle into the mixing chamber caused the evaporator temperature to rise. -- For a fixed nozzle position, the evaporator temperature dropped when the condenser pressure was reduced or when the boiler pressure was increased. However, the effect of the condenser and boiler pressures were reduced when the NXP was large. The optimum nozzle position was shown to depend on the boiler and condenser pressures. Increasing the condenser pressure or reducing the boiler pressure moved the optimum nozzle position into the mixing chamber, and vice versa. According to the tests' results, a single optimum nozzle position cannot be defined to meet all operating conditions. Each operating condition required a particular nozzle position. Similar experiments were conducted by Hamner 4 using RII as the working fluid and with zero secondary flow

(the evaporator was isolated from the ejector). In Hamner's case, the optimum nozzle position was determined by measuring the minimum pressure in the mixing chamber. However, the pressure varied only very slightly with the nozzle position. According to Zeren 5, the pressure in the mixing chamber was not only a function of the boiler and condenser pressures but also the mass flow rate of the secondary fluid. From Figures 4 and 5, the optimum NXP was found in the range of 0-15ram within the mixing chamber inlet section. This is in contrast to the recommendation from ESDU 2 which suggested placing the nozzle exit 0.5-1.0 length of the mixing chamber's throat diameter upstream of the start of the mixing chamber (equivalent to an NXP of - 9 to - 1 8 m m ) . This may be due to the fact that the primary steam pressure used is relatively low compared with the pressure that is commonly used in industrial applications (2-3.6 bar compared with 5-20 bar).

Effect of the nozzle position on the system performance These tests were conducted by setting the evaporator's thermostat at 5°C and the boiler's thermostat at 130°C. For each NXP, the condenser pressure was varied from below to above the critical value. During the tests, the electric power input to the boiler and the evaporator were measured. Refrigerator COP based on electric power input, to both the boiler and evaporator, is a measure of overall performance and includes all the unwanted heat losses and gains to the system. Figure 6 (see also Table 1) shows the effect of the nozzle position on COP. For fixed boiler and evaporator temperatures, the COP and cooling capacity can be varied as much as 100% by changing on the nozzle position. Moving the nozzle into the

S. Aphornratana and I. W. Eames

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Figure 6 Variations of experimental performance (COP) with N X P and condenser pressure for the evaporator temperature of 5°C and the boiler temperature of 130°C Figure 6 Variations de la performance exp~rimentale (COP) on fonction de la position de la buso ( N X P ) et de la pression au condenseur, pour une temperature h lYvaporateur de 5 deg C et une tempOrature au bouilleur de 130 deg C

Table 1 Performance of the experimental refrigerator at

m i x i n g c h a m b e r c a u s e d t h e C O P to fall a n d c o o l i n g c a p a c i t y to d e c r e a s e w h e n t h e boiler i n p u t w a s m a i n t a i n e d c o n s t a n t . H o w e v e r , t h e cycle c o u l d be o p e r a t e d a t a h i g h e r critical c o n d e n s e r p r e s s u r e . By retracting the nozzle from the mixing chamber, the C O P a n d c o o l i n g c a p a c i t y c a n be i n c r e a s e d at t h e e x p e n s e o f t h e critical c o n d e n s e r p r e s s u r e .

critical condenser pressure operation with various NXPs Tableau 1 Performance du r~frig~rateur expdrimental fonctionnant en pression critique au condenseur, avec diff~rentes positions de la buse Condenser N X P (mm)

Temp. (°C)

Pressure (mbar)

Qevap (W)

COP

-4 11 26 41 56

29.5 29.8 30.8 31.2 31.9

41 42 44 45 47

754 650 528 458 372

0.2238 0.1929 0.1567 0.1360 0.1104

Operation and control of a steam ejector refrigerator T h i s s e c t i o n gives m e t h o d s o f o p e r a t i n g a n d c o n t r o l ling a s m a l l scale s t e a m ejector r e f r i g e r a t o r u s i n g a n ejector w i t h a m o v e a b l e p r i m a r y n o z z l e p o s i t i o n . P e r f o r m a n c e m a p s were c o n s t r u c t e d so t h a t t h e e j e c t o r c o u l d be t u n e d b y v a r y i n g t h e boiler t e m p e r a -

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Figore 7 Performance characteristic (COP) o f the experimental steam ejector refrigerator Figure 7 Caract~;ristique de performance (COP) du r~frig~rateur exp~;rimental ?t ~!jection de vapeur

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Figure 8 Performance characteristic (cooling capacity) of the experimental steam ejector refrigerator Figure 8 Caractkristique de performance (puissance frigorifique) du r~frig~rateur experimental h ~jection de vapeur Table 2 Performance at off design operations of the steam ejector refrigerator as shown in Figures 7 and 8* Tableau 2 Performance du rOfrigOrateur ~ djection de vapeur en cas de fonctionnement hors des normes de conception, selon les figures 7 et 8 Heat input (W)

Temperature (°C)

Operating point on the figures

NXP (mm)

COP

Evaporator

Boiler

Evaporator

Boiler

Condenser pressure (mbar)

a b c d e f g h i

26 26 26 26 26 26 26 I1 11

0.210 0.210 0.210 0.225 0.240 0.278 0.197 0.210 0.261

710 710 710 710 752 936 690 710 850

3369 3369 3369 3160 3138 3369 3510 3369 3257

7.5 7.5 7.5 7.0 7.5 10.0 7.5 6.0 7.5

130.0 130•0 130.0 127.2 126.9 130.0 131.7 130.0 128.5

45.7 42.5 25.0 42.5 42.5 47.6 47.6 42.5 42.5

* The data provided in this table are obtained graphically from the figures•

ture a n d pressure or N X P in order to o b t a i n the m a x i m u m p e r f o r m a n c e w h e n the c o n d e n s e r pressure is changed due to the variation of the e n v i r o n m e n t temperature• The experiments showed that, for given e v a p o r a t o r a n d c o n d e n s e r pressures, m a x i m u m C O P a n d cooling capacity were o b t a i n e d when the cycle was o p e r a t e d at a boiler t e m p e r a t u r e t h a t allowed the ejector to operate at its critical c o n d e n s e r pressure• The p e r f o r m a n c e m a p s in Figures 7 a n d 8 were constructed from d a t a t a k e n at critical c o n d e n s e r pressure operation for the N X P s o f 11 a n d 26 m m . The boiler a n d e v a p o r a t o r isotherm lines s h o w n in the figures can be used only when the ejector was o p e r a t e d at its critical condenser pressure. A n example o f using these figures is now given. Referring to Figures 7 a n d 8 (data for each operating point on these figures are p r o v i d e d in Table 2); It is assumed that, the cycle is normally designed to operate at p o i n t a (referenced to solid

line curves) with a N X P o f 2 6 m m a n d critical c o n d e n s e r pressure of 4 5 . 7 m b a r . If the c o n d e n s e r pressure falls to 42.5 m b a r , due to a reduction o f its cooling water temperature, while the boiler a n d e v a p o r a t o r temperatures are held constant, the cycle will then operate at point b with the C O P with the cooling capacity essentially remaining the same*. As the boiler t e m p e r a t u r e is fixed, for any condenser pressure < 45.7 m b a r , the C O P a n d cooling capacity essentially remain c o n s t a n t as shown by the horizontal line a - b - c , b u t the ejector is no longer operating at its critical c o n d e n s e r pressure. In order to improve cycle p e r f o r m a n c e when the c o n d e n s e r pressure is * For this case, the evaporator and boiler isotherm lines can not be used to indicate the evaporator and boiler temperatures as it does not operate with critical condenser pressure (see Table 2 for operating conditions). It must be noted that the boiler and evaporator temperatures given by the iostherm lines can be used only when the ejector is operated with critical condition.

3,58

S. Aphornratana and I. W. Eames

reduced from 45.7 to 42.5mbar, the ejector must be allowed to operate at a new critical condenser pressure (42.5mbar) by using one of the following methods (see Figures 7 and 8): Constant cooling capacity, lower evaporator temperature Point d with NXP of 26 mm (referenced to solid line curves). If the boiler temperature is reduced to 127.2°C when the condenser pressure is decreased to 42.5 mbar, the cycle operating point will move to point d with the critical condenser pressure of 42.5 mbar. This produces a constant cooling capacity at a lower evaporator temperature (7.0°C). The COP is also increased as the primary nozzle is always choked, reducing the boiler temperature automatically reduces heat input to the boiler. Point h with NXP of 11 mm (referenced to dotted line curves): If the boiler temperature is maintained at 130.0°C and the nozzle is retracted to NXP of l l m m when the condenser pressure falls to 42.5mbar, the cycle operating point will be moved to point h with the critical condenser pressure of 42.5 mbar (from the figures, point h and b are shown to be the same, however, point h is operated with critical condition and referenced to the dotted line curves while point b is not operated with critical condition). This causes the COP and the cooling capacity to remain constant, but at a lower evaporator temperature (6°C). Constant evaporator temperature, higher cooling capacity Point e with N X P of 26mm (referenced to solid line curves): If the cycle is already operating at point d, and the boiler temperature is slightly further reduced from 127.2 to 126.9°C, the evaporator temperature will return to 7.5°C and the cycle operating point will move to point e with the critical condenser pressure of 42.5 mbar. The COP and the cooling capacity will rise. Point i with N X P of l l mm (referenced to dotted line curves): If the cycle is already operating at point h with an NXP of 11 mm and the boiler temperature is reduced from 130.0 to 128.5°C, the evaporator temperature will increase back to 7,5°C and the cycle can be operated at point i with the critical condenser pressure at 42.5 mbar. This causes the COP and cooling capacity to be increase. If the condenser pressure is increased higher than the design point such as on a hot day points f and g show the possible operating conditions when the

condenser pressure is increased to 47.6mbar. As the condenser pressure is increased higher than the critical value, in order to establish a new critical condenser pressure, the cycle must operate with a higher boiler pressure and fixed evaporator temperature (which reduces cooling capacity and COP) or a higher evaporator temperature and fixed boiler temperature (which increases cooling capacity and

coP). Conclusions

This paper describes the experimental studies of a small scale steam ejector refrigerator using ejector with adjustable primary nozzle position. The test showed that a single optimum primary nozzle position cannot be defined to meet all operating conditions. Each operating condition requires a particular optimum nozzle position. The COP and cooling capacity can be varied as much as 100% by changing the nozzle position. Moving the nozzle into the mixing chamber caused the COP and cooling capacity to decrease when the boiler input and temperature was maintained constant. However, the cycle could be operated at a higher critical condenser pressure. When the nozzle was retracted from the mixing chamber, the COP and cooling capacity was found to increase, but the critical condenser pressure was reduced. The use of an ejector with movable primary nozzle provides a more flexible operation than a totally fixed geometry unit. An increase in the cooling capacity can be achieved by retracting the nozzle from the mixing chamber as the condenser pressure falls without changing either the evaporator or boiler temperatures. In practice, the nozzle position may be automatically controlled by monitoring the saturated temperatures and pressures in the boiler, evaporator and condenser.

References 1

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4

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Eames, I. W., Aphornratana, S., Haider, H. A theoretical and experimental study of a small scale steam jet refrigerator. Int. J. Refrig. (1995) 18 378-386 ESDU, Ejector and jet pump, data item 86030, ESDU International Ltd, London, UK, (1985) Keenan, J. H., Neumann, E. P., Lustwerk, F. An investigation of ejector design by analysis and experiment. ASME J. Appl. Mech., (1950), Sept. 299-309 Hamner, R. M. An investigation of an ejector-compression refrigeration cycle and its applications to heating, cooling, and energy conservation. Ph.D. thesis, The University of Alabama, Birmingham, USA (1978) Zeren, F. Freon-12 vapor compression jet pump solar cooling system. Ph.D. thesis, Texas A&M University, USA (1982)