Small solar (thermal) water-pumping system

Solar Energy Vol. 57. No. I, pp. 69-76, 1996 Pergamon

PII: 0038~092X(96)80043-6 Copyright 0 1996 Elsevier Science Ltd

Printed in Great Britain. All rights reserved 0038-092X/96 1615.00+0.00

SMALL SOLAR (THERMAL) WATER-PUMPING SYSTEM

K. SPINDLER,* K. CHANDWALKER** and E. HAHNE* *Institut fur Thermodynamik und Warmetechnik, Universitat Stuttgart, Pfaffenwaldring 6, D-70550

Stuttgart, Germany and **Stiletto Engineers, Hyderabad, India

(Communicated by Peter Fraenkel)

Abstract-A small solar (thermal) water pump prototype was tested. The pump works on an organic Rankine cycle using refrigerant R113. The design of the pump is described. Detailed temperature and pressure measurements of the working fluid for different operating conditions are performed. The behavi- our of the cycle is analysed to get a clear picture of the thermodynamic process. Power-characteristic curves are obtained bv a svstematic variation of water temperature, pumping head and heat input. Copyright 0 1996 Elsevier Science Ltd.

1. INTRODUCTION

Water is the basic requirement for human sur- vival. A high percentage of the population that is without adequate water facilities is concen- trated in solar-abundant rural areas. Often these areas are dispersed and deprived of conventional sources of energy, either due to energy shortage or the availability of energy being restricted due to economic reasons.

Statistical data from India indicate that in this country alone 20 million small farmers (with holdings of less than 2 ha) exist on land where the water table is probably at a depth of 7 m or less. In India, about 40% of the ground water used for irrigation purposes is lifted by tradi- tional means (human and animal power) and the rest by diesel and electrical pumping systems. The situation in other parts of the developing world is similar. In view of the growing demands for irrigation water and the increasing cost of conventional fuels, the use of solar energy for irrigation purposes is highly desirable. Many millions of small farmers are likely to benefit from a small solar water pump.

A summary of feasible solar water-pumping systems is given by Bahadori (1978). The mechanical energy needed for pumping water may be produced by thermodynamic or direct-conversion methods. In thermodynamic conversion the high internal energy of a fluid may be utilized in Rankine-, Brayton- or Stirling-cycles. The direct conversion includes photovoltaic, thermoelectric and thermionic processes.

During recent years many photovoltaic water-pumping systems have been installed around the world. Technical details and effi-

ciencies are given by Rizvi and Sayigh (1988) and Whitfield (1988). Some Indian experiences are reported by Sangal et al. (1988). They reported that in some places trained staff were transferred and systems were being operated by people who were not very knowledgeable of these systems. In some locations the systems were a target of vandalism and parts were reported missing. The structure and results of a programme for photovoltaic water-pumping systems organized by the GTZ (Gesellschaft fur Technische Zusammenarbeit) are described by Posorski (1993). Auer and Vie1 (1993) give details of the measuring concept for these pro- jects. A marketing report and experiences, especially of the Sahel zone, is given by Hempel (1994).

A prototype of a solar thermal water pump working on a low-temperature organic Rankine cycle with refrigerant Rll is described by Kishore et al. (1986). The daily overall efficiency (hydraulic power to solar irradiation) was about 0.45%, with a peak of about 0.7%. The pump was lifting about 6.5 m3 of water per day from a depth of 11.2 m. The pump uses 7 m2 of single- glazed flat plate collectors with selectively coated absorbers. Solar powered diaphragm- pumps working on a Rankine cycle with R113 are described by Burton (1983) and by Hernandez and Murata (1988). The diaphragm pump, tested by Burton (1983) was pumping about 0.9 m3/h to a head of 3 m. The overall efficiency was 0.2%.

The prototype of a small solar water pump (SSP) described in this paper was developed by Mr Kiran Chandwalker at Stiletto Engineers, Hyderabad, India, under the following considerations.

IE 51:*-c 69

K. Spindler et al 70

(1)

(2)

(3)

(4)

The SSP should be small and inexpensive. It can be used in small farms for pumping water for drinking and irrigation purposes in India and other developing countries of semi-arid regions. The SSP should be simple in design. It should have least dependence on imported materials. It should require no specialized or expensive materials or production tech- niques. It should be manufactured in small workshops in developing countries. The SSP should be robust to withstand severe field conditions. It should be easy to operate and maintain. The SSP should be driven by solar energy.

The price of the SSP is estimated to be 2000 DM.

2. PRINCIPLE OF OPERATION

The solar water pump operates on a low- temperature organic Rankine cycle. The prototype presently uses R113 as working fluid. The boiling point at 1.013 bar is 47.6”C. The refrigerant passes the following changes of state in a closed loop: evaporation of liquid R113 using solar energy, vapour superheating using solar energy to produce high-pressure vapour, expansion of high-pressure vapour with output of mechanical energy to operate the water pump and the feed pump, condensation of R113

vapour at low pressure using pumped water and pressurization of the liquid R113 in the feed

pump. During rest periods the pressure remains

below the atmospheric pressure. In operation, the maximum pressure does not exceed 3 bar. This makes the system much safer than conventional high-pressure systems.

3. DESCRIPTION OF THE PUMP

The prototype of a small solar water pump was developed by Mr Kiran Chandwalker in Hyderabad, India. The pump was produced in India and brought to Germany. It was installed in the Institut fur Thermodynamik und Warmetechnik at Stuttgart University for trials, demonstration, data collection and other related research purposes.

The solar water pump operates on the thermodynamic Rankine cycle and presently uses R 113 (C,Cl,F,, trichlorotrifluoroethane) as working fluid. There is, however, the possibility of operating the system by alternative non- CFC-based organic fluids. The moderate temperature range allows various non-CFC fluids to be used, e.g. FC72 (C6Fi4). Some organic solvents have already been tried with reasonable success. A schematic layout of the water- pumping system is shown in Fig. 1.

The water-pumping system can be divided

Fig. 1. Schematic layout of the water-pumping system.

Small solar water-pumping system 71

into the refrigerant loop, the water loop and the heating device.

3.1. The refrigerant ~‘oop (a) The evaporator (low carbon steel, 113 mm

diameter, 125 mm long) with superheater (copper coil of 10 mm outer diameter) gener- ates high-pressure vapour.

(b) The expansion machine is a double-acting,

(c)

reciprocating piston-type expander (piston diameter 50 mm, operating stroke 40 mm, and 2.25 mm dead cavity on each side). The valve system consists of a main inlet shuttle type 3 port-valve, a shuttle type 3 port-exhaust valve and a switching system for the valves. The main inlet and exhaust valves are located in the cylinder block. They are of different configuration to allow their unidirectional operation. The valve- operating mechanism consists of spring- loaded levers, valve pusher plate and release system to operate the valves at the end of the forward and return stroke. The valve- operating mechanism is enclosed in a sealed extension of the main cylinder.

(d) The feed pump driven by the main piston is

(e)

(f-1

(8)

a reciprocating piston type single-operating pump. It is provided with non-return valves. The recuperator is a shell-and-tube single pass, counter-flow heat exchanger. It is designed to pre-heat the pumped feed back liquid using exhaust vapour from the expander. The condenser is a shell-and-tube single pass, counter-flow heat exchanger. It is cooled by the pumped water. The exhaust vapour is condensed at low pressure. The condenser is located in the system in such a way that the cooling water path is in the suction line of the main water pump. A non- return valve provided in the inlet side of the water path of the condenser ensures that the condenser always has water and thereby allows the cold start of the system. The refrigerant storage tank provides a removable flanged cover and connections for the inlet to the feed pump. A hand shut- off valve is also mounted on this storage tank for charging and discharging the system.

(h) The priming pump is a piston type single- operating pump operated by a hand lever through a hermetic seal. A non-return valve block is located in front of this priming pump. This pump is generally used for ini-

tiating the process after charging the system. Liquid refrigerant is pumped from the storage tank into the boiler.

(i) A sight glass mounted in the suction line of the feed pump allows visual observation of the working fluid.

The filling capacity of the refrigerant loop is about 3.5 kg.

3.2

(j)

(k)

The water loop A double-operating piston-type reciprocating water pump is connected to the top of the expansion machine and is driven through linkages by the output of the expander. The expander output is taken through a hermetic sealing system for connecting it to the water pump piston rod. The water pump has a 40 mm diameter piston with 38 mm stroke length. Two non-return valves are provided for inlet and outlet of water. The water pump suction (inlet) is connected to the pump through the condenser. So the condenser is cooled automatically when the pump is operating. A lower and upper water tank with a volume of 40 x 1O-3 m3 each. Both tanks are connected by flexible tubes with 19 mm inner diameter. An additional cooler has been used to keep the pumped water at a constant temperature.

3.3. The heating device The water pump is designed to be used with

a parabolic solar concentrator of about 2 m2 area. For data collection and indoor test the system is presently fitted with a bulb heating device. A reflector shell is placed around the boiler and superheater. In the heating device 10 bulbs (each 250 W) are installed. Two groups of five bulbs each can be controlled independently by voltage transformers to adjust the heat input. The optimum voltage range for operating the bulbs lies between 180 and 240 V.

4. EXPERIMENTAL SET-UP AND INSTRUMENTATION

A schematic flow diagram, including the main components of the water-pumping system, is shown in Fig. 2. Detailed measurements of temperature, pressure and flow rate are performed to evaluate the existing prototype and to provide a clear picture of the thermodynamic cycle.

All the measuring points are listed in Table 1 and indicated in Fig. 2.

72 K. Spindler et al.

valve 3 priming pump w evaporator

Fig. 2. Schematic flow diagram including measuring points.

Table 1. Temperature and pressure measuring points

No. Location Measurement

1 Evaporator PI> Tt 2 Superheater T2 3 Expansion machine inlet ~3. G 4 Expansion machine outlet T4 5 Recuperator inlet PS, Ts 6 Condensator inlet PC. T6 7 Refrigerant storage tank ~7, T7 8 Feed pump inlet 5 9 Recuperator return inlet ~93 Ts

10 Recuperator return outlet T 10 11 Evaporator inlet TlI 12 Water pump outlet &Z 13 Control Tl3 14 Lower water tank T4 IS Ambient T5

To measure the temperature of the working fluid, one-side closed capillary tubes (12, 1 x 0.1 mm, 70 mm length) are soldered in T- junctions of the flow loop. The capillary tubes are positioned in the centre line of the outer tube. Calibrated NiCr-Ni thermocouples (0 0.5 mm) are then inserted in the capillary tubes. The contact resistance between thermocouple and capillary tube has been reduced by adding a thermal conducting paste. The accuracy of the temperature measurement is + 0.1 K. The pressure is measured with a high-precision Bourdon- tube pressure gauge (k 10 mbar) and with an electronic pressure transducer. All the pressure taps can be separately connected to the pressure gauge by individual switching of the valves. The electrical power input of the heating device is

obtained by measuring the voltage and the current. The accuracy is about f2%. The mass flow rate of the pumped water was measured by collecting the water over a time interval of 30 s and by weighing. The accuracy of the mass flow rate is better than + 1%. The number of cycles per second (frequency of the pump) was obtained by counting the strokes within 60 s.

5. RESULTS

The operating conditions were varied by changing the power input to the boiler. This power value will be used as an independent variable in presenting the results.

The temperature and pressure distribution of the working fluid for different water temperatures at constant pumping head of 3.2 m and electrical input of 2570 W are shown in Figs 3 and 4. The temperature and pressure levels increase with rising water temperature. This increase is more significant in the condenser than in the boiler. The temperature and pressure difference of the refrigerant between boiler (measuring point 1 in Figs 3 and 4) and storage tank (measuring point 7 in Figs 3 and 4) decreases with increased water temperature.

The temperature and pressure distributions of the working fluid for different electrical input at a constant water temperature of 21°C and pumping head of 3.2 m are given in Figs 5 and 6. The temperature and pressure levels increase with increasing electrical input. This increase is


--I

6 7 6 9 10 11 meaming point ,0ca,lo”

Fig. 3. Temperature distribution of the working fluid for different water temperatures at constant pumping head of

3.2 m and electrical input of 2570 W.

IB’C I 0.6 _ ,~~~ ._. 1 r --

1 2 3 4 5 6 7 6 9 10 measunng point location

Fig. 4. Pressure distribution of the working fluid for different water temperatures at constant pumping head of

3.2 m and electrical input of 2570 W.

707 .:

Fig. 5. Temperature distribution of the working fluid for different electrical inputs at constant pumping head of 3.2 m

and water temperature of 21°C.

more significant in the boiler than in the condenser. The temperature and pressure difference of the refrigerant between boiler (measuring point 1 in Figs 5 and 6) and storage tank (measuring point 7 in Figs 5 and 6) increases with increasing electrical input.

For this prototype a vapour superheat is not found (T, x 7”). The addition of heat takes place only in the boiler. A temperature drop between

2.6

0.6 k-- I 1 2 3 4 5 6 7 6 9 IO

measuring point location

Fig. 6. Pressure distribution of the working fluid for different electrical inputs at constant pumping head of 3.2 m

and water temperature of 21°C.

the superheater exit and the expander inlet can be detected due to heat losses. The expansion of the refrigerant in the expansion machine (3-4) causes a large temperature and pressure drop. The temperature drop in the recuperator (5-+6) is smaller than in the condenser (6-7). Most of the heat is removed in the condenser. The preheating of the liquid refrigerant can be seen in the temperature increase T,+ T,,. Pressure losses due to friction between boiler outlet and expansion machine inlet ( l-+3), between expansion machine outlet and refrigerant storage tank (5+7) can be seen in Figs 4 and 6.

Figure 7 shows the hydraulic power as a function of the expander pressure difference (ps -p3). The hydraulic power increases with growing expander pressure difference, as expected. The dependence is approximately linear. The scatter is caused by mechanical friction in the pump. Figure 8 represents typical operating curves of the water pump at a constant water temperature of 21°C. The hydraulic

5 -. I

o L _ _._.. I

05 0.6 0.7 0.6 0.9 1 1.1 expander pressure difference [bar]

Fig. 7. Hydraulic power versus pressure difference in the expander for various operating conditions.


I 0. 1 -~-- - ---r-_--r !

0 025 05 075 1 1.25 I.5 1.75 2 frequency [l/S]

Fig. 8. Power-characteristic curves for different electrical Fig. 10. Volumetric flow rate versus frequency of the pump input at a constant water temperature of 21°C and pumping for different operating conditions.

heads of 3.2 and 5.0 m.

power rises linearly with increasing electrical input. An increase of pumping head from 3.2 to 5.0 m leads to an increase of hydraulic power by about 35%. The hydraulic power for various operating conditions is between the two curves of constant pumping head. For these experi- ments the maximum hydraulic power was found to be 5.2 W, and the minimum value was 2 W.

rate is obtained by the known cylinder geometry of the water pump. The average volumetric efficiency for various operating conditions is 72%, the maximum value is 84% and the minimum value is 52%.

Three power-characteristic curves for varying water temperature, pumping head and electrical input are given in Fig. 9. The hydraulic power is given as a function of the maximum temperature difference (Ti - T,,). If the water temperature is raised, the temperature level of the working fluid increases (Fig. 3) and the maximum temperature difference decreases. The effect on the hydraulic power can be neglected. The hydraulic power is raised by increasing pumping head or increasing electrical input. In both cases the maximum temperature difference is augmented.

The effect of the boiler temperature and the electrical input on the hydraulic power equivalent (volume flow rate x pumping head) is shown in Figs 11 and 12. A higher boiler temperature and a higher electrical input lead to an increase of the hydraulic power equivalent. The hydraulic

160 , . .

s . , .--

-I . I

. .

.

.

The variation of volumetric flow rate with the frequency of the pump is shown in Fig. 10. The line in Fig. 10 corresponds to a volumetric efficiency of 100%. This theoretical volume flow

0.w J- 60

--. I -~

65 70 75 boiler temperature [‘Cl

80

Fig. 11. Hydraulic power equivalent versus boiler tempcr- ature for different operating conditions.

5 1

i 35 40 45 50 55

maximum temperature difference [KI 60

Fig. 9. Power-characteristic curves for varying water peratures, pumping heads and electrical inputs.

2.00

0.00 ’ / I / I 1600 1800 2000 2200 24W 2600 2800 3000 3200

electrical input Iwl

Fig. 12. Hydraulic power equivalent versus electrical input for different operating conditions.


power equivalent lies between 1.2 and 1.6 m4/h for typical operating conditions with boiler temperatures between 75 and 80°C and an electrical input of about 2600 W.

The Rankine cycle due to measured pressures and temperatures of the working fluid is shown in the log p-h diagram for a typical operating condition (Fig. 13). The indicated points corre- spond to Table 1. The process parameters and the enthalpies are listed in Table 2 as number 1. It can be seen from Fig. 13 that the vapour superheat is low (points 1 and 3). The vapour expansion is close to the saturation line. The preheating of the liquid feed to the evaporator (9+10) is low.

The efficiency of the Rankine cycle q,ycle can be obtained by the ratio of expander work (ha - h4) and the heat of vaporization (h3 - h,,). For the calculations it is assumed that the enthalpy at the evaporator inlet (h,,) is nearly the same as the enthalpy at the recuperator

outlet (h,,). The work input of the feed pump is neglected. The efficiency of the given Rankine cycle is 6.8%, as shown in Fig. 13. The efficiency is 7.3% for a lower water temperature (2 in Table 2). An efficiency of 9% is achieved for a higher electrical input (3 in Table 2). The Carnot efficiency +, calculated with the highest fluid temperature T2 (superheater) and the lowest fluid temperature T, (refrigerant storage tank), is between 11 and 13%. The comparison of the cycle efficiency with the Carnot efficiency shows that about 30% of the heat is lost.

The overall system efficiency can be calculated

by

?l= qelectr.heating qheat utilization ~expander I?waterpump VC

The efficiency of the electrical heating

qelectr. heating was obtained by a calorimetric test. For this purpose the boiler was filled up with a certain amount of liquid refrigerant. The pres-

Table 2. Efficiency of the Rankine cycle

1 2 3

Pumping head (m) Electrical input (W) Water temperature (‘-C) Boiler pressure (bar) Condenser pressure (bar) Enthalpy of expander inlet h, (kJ/kg) Enthalpy of expander outlet h, z h, (kJ/kg) Enthalpy drop of expander h,-h, (kJ/kg) Enthalpy of evaporator inlet h,, Z/I,, (kJ/kg) Enthalpy of evaporator outlet h, z h, (kJ/kg) Heat of vaporization h3-h,, (kJ/kg) Efficiency qeycle = (h,-U/(M,,) (%) Carnot efficiency qc = (T,-T,)/T, (%)

3.2 3.2 8 2592 1926 2592

31.7 20.5 16.7 2.32 1.65 2.26 1.16 0.83 1.0

407 400 408 396 388 393

11 12 15 245 236 241 407 400 408 162 164 167

6.8 7.3 9 10.8 10.9 13

200 250 300 350 h-

400 kllkg

3. The Rankine cycle of the water pump in the log p-h diagram of R113 for typical operating conditions (pumping head 3.2 m, electrical input 2592 W, water temperature 31.7 C).

Fig. 1


sure and the temperature was observed during heating up with constant electrical input. The time was also measured when the liquid reached saturation and when the liquid was totally evap- orated. The efficiency of the electrical heating was found to be 0.3.

The expander efficiency could be calculated from the equations of state since both pressure and temperature are measured. But we do not know the vapour quality at the end of the expansion. The water-pump efficiency could be calculated from the expander work and the flow rate and the pumping head of the water. The mass flow rate of the refrigerant, which is neces- sary for the calculation of the expander work, was not measured for these test runs.

Due to lack of information, the efficiencies of the expander and the water pump are estimated to be ~,,pander =0.25 and ~waterpump = 0.3. The efficiencies of the expander and of the water pump are being improved in later models of the solar water-pumping system.

With qheat utilization=0.7 (30% heat 10~s) and r]c=O.13, the overall system efficiency is 0.2%, which is actually achieved. Relative to the electric power input, the hydraulic power represents an overall efficiency of about O.lLO.2% (from Fig. 12).

Trial runs with a parabolic solar concentrator (a 1.567 m) were made in November between 11 am and 1 pm at Hyderabad, India. At a solar irradiation of 983.5-1041 W/m2, the system showed a collection efficiency of 55.05%, resulting in a heat input to the boiler of about 1075 W. When the solar pump is exposed to the sun, the boiler is covered with a glass jacket to reduce convective heat losses and improve the collection efficiency. Typically boiler temperatures of 85-93°C are reached during sustained cycle at a solar irradiation of about 900 W/m’. Electrical heating results in boiler temperatures of about 63°C for similar input and hence is likely to be far less efficient.

The actual heat input to the boiler could be determined if the mass flow rate of the working fluid is measured. The water-pumping system uses a positive displacement piston type feed pump with 5.4 cm3 displacement per stroke. The volumetric efficiency of the feed pump is about 74%. Therefore the discharge per stroke is 4 cm3.

Thus the flow rate of the working fluid can be calculated if the number of strokes is known. Previous efforts with actual flow measurements were not successful due to fluctuating flow and the low values of the mass flow rate (z 2-7 g/s).

6. CONCLUSIONS

The experimental prototype of a small solar water pump demonstrates the possibility of using such a system for solar water pumping. The modest levels of operating conditions (max. pressure 3 bar, max. temperature 90°C) also provide possibilities of using alternative working fluids. With the collected thermodynamic data a detailed analysis of the system should now be taken up to improve the system efficiency.

Acknowledgements-The prototype of the SSP was developed and built up by Kiran Chandwalker, Hyderabad, India. The authors express their thanks to Kiran Chandwalker who made the prototype available for research studies. The authors gratefully acknowledge the financial support of the Eiselen Vermgchtnis. The authors would like to thank Peter Riiser who has contributed with his practical and experimental skills to this work.

REFERENCES

Auer F. and Vie1 L. (1993) PV-Trinkwasserpumpen-MeB- technikkonzept und Auswertebeispiele. Sonnenenergie 18(3), 17-19.

Bahadori M. N. (1978) Solar water pumping. Solur Energy 21, 307-316.

Burton R. (1983) A solar powered diaphragm pump. Solar Energy 31, 523-525.

Hempel C. (1994) Solarpumpen im Sahel-Markttendenzen und Erfahrungen aus dem Niger. Sonnenenergie 19(2), 4-11.

Hernandez 0. S. and Murata V. V. (1988) Development of an elastic diaphragm pump. Proc. 1988 Annual Meeting of the American Solar Energy Society, pp. 131- 134.

Kishore V. V. N.. Gandhi M. R., Pathak N., Rao K. S., Jaboyedoff P., Lehmann W. and Marquis, Ch. (1986) Development of a solar (thermal) water pump prototype-an Indo-Swiss experience. Solar Energy 36, 251-265.

Rizvi S. N. A. and Sayigh A. A. M. (1988) Photovoltaic powered water pumping-an appraisal. Conf. Proc. Int. Solar Energy Society, Reading. UK, pp. 21-27.

Posorski R. (1993) PV-Trinkwasserpumpen-Struktur und Ergebnisse des PVP-Programms. Sonnenenergie 18( 3). 11-16.

Sangal S. K., Ranjan V. and Gosain R. K. (1988) Photovol- taic water pumping-Indian experience. 20th IEEE Pho- rouolraic Specialists Conf 26-30 Sept., Las Vegas, NV, pp. 1188-1193.

Whitfield G. R. ( 1988) The efficiency of small solar photovoltaic water pumping systems. Conf. Proc. Int. Solar Energy Society, Reading, UK, pp. 42-49.

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Small solar (thermal) water-pumping system