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Heat pump operated freeze concentration system with tubular heat exchanger for seawater desalination
Energy for Sustainable Development 15 (2011) 184–191
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Energy for Sustainable Development
Heat pump operated freeze concentration system with tubular heat exchanger for seawater desalination M.V. Rane a,⁎, Y.S. Padiya b a b
Mechanical Engineering Department, Indian Institute of Technology, Bombay, Mumbai, India Mechanical Engineering Department, Konkan Gyanpeeth College of Engineering, Karjat, Dist. Raigad, Maharashtra, India
a r t i c l e
i n f o
Article history: Received 15 July 2010 Revised 6 March 2011 Accepted 6 March 2011 Available online 9 April 2011 Keywords: Desalination Heat pump Layer freezing Energy efficient
a b s t r a c t Desalination of seawater can provide an almost inexhaustible source of freshwater if it can be made affordable. Distillation and filtration, the commonly used processes, have high operating and maintenance cost. Low operating temperature and low latent heat of fusion of water makes freezing technology worth considering for further development. Here, water is selectively frozen, in the form of ice from sea water, and melted after isolating it from the concentrated seawater to generate pure water which can be made potable. However, desalination processes based on freezing have not been exploited because of the operational difficulties in ice–water separation, high cost of equipment and high parasitic power requirement. This paper discusses a patented layer freezing based technology which has competitive initial and operating cost and eliminates operational difficulties of conventional freezing system. It is scalable and is coupled with a heat pump which selectively freezes water from seawater in the evaporator and melts the ice in the subsequent phase when it serves as a condenser. The condenser optimally utilises the latent heat of melting of ice to partially condense the refrigerant and the excess heat is rejected to ambient. It avoids the need of ice scraper/separation mechanisms. Use of vented-double-wall tube–tube heat exchanger, TT_HE, enables refrigerant and seawater/potable water to exchange heat without the use of intermediate fluids, while keeping the inclusion low. TT_HE is a reliable heat exchanger which ensures that refrigerant and seawater/water do not mix. Operating COP of the heat pump is in the range of 8 to 12, which results in specific energy consumption in the range of 9 to 11 kWhe/m3 of water produced. Comparison of features with other freezing desalination processes is also presented. © 2011 International Energy Initiative. Published by Elsevier Inc. All rights reserved.
Introduction Desalination is the process of separation of salt from water. It is mainly carried out using evaporation, membrane separation and freezing technologies. In evaporation technology, seawater is heated above its boiling point (normally 100.8 °C and 101.7 °C at 3.5% and 7% salt concentrations) and the generated vapour is condensed to get water. Multistage flashing (MSF), multiple effect evaporation (MEE), mechanical vapour compression (MVC), humidification and dehumidification (HDH), membrane distillation (MD), pervaporization and solar distillation are the major techniques which come under evaporation category. MSF, MEE and MVC are suitable for 100 to 500,000 m3/d fresh water production capacity plant. The specific energy consumption of MSF and MEE varies from 150 to 300 kJ/kg with additional electrical input of 2 to 4 kWhe/m3 for pumping liquor and water, while energy consumption of MVC is 8 to 15 kWhe/m3 (Cipollina et al., 2005). These technologies require pre-treatment of seawater, corrosion resistant
⁎ Corresponding author: Tel.: +91 22 2576 7514, +91 22 2576 4527, +91 22 2576 4593, +91 22 2576 8514; fax: +91 22 2572 4544, +91 22 2572 6875. E-mail address: [email protected] (M.V. Rane).
material and permanent qualified maintenance staff (Al-Enezi et al., 2006). HDH, MD, pervaporization and solar distillation are low operating temperature technologies and thus minimise corrosion related issues. They are mostly used for small capacity water generation, e.g. 100 l/d. HDH technology is simple in operation, but availability of low mass transfer coefficient in existing humidifier, 1 to 3 kg/s·m3·(kgw/kgda), (Amara et al., 2004) and in dehumidifier, increases the plant size and energy required for air circulation. In MD and pervaporization transport of water vapour from seawater occurs through microporous hydrophobic membranes. The specific energy consumption of MD can be low in the range of 73 to 75 kJ/kg when operated with low grade heat between 50 and 100 °C. It has potential to generate potable water with lower water cost, less than 0.5 $/m3. However, approximately after 800 h of use membrane may get damaged due to biofouling (Meindersma et al., 2006). Solar desalination has been reported by many researchers. It can be used directly or coupled with other technologies such as MSF, MEE, vapour compression, RO, MD and electro dialysis, where solar collectors are used for heat generation. Direct solar desalination method includes use of solar still, which has no moving parts and is simple in operation. It is competitive to indirect desalination plant in small scale production.
0973-0826/$ – see front matter © 2011 International Energy Initiative. Published by Elsevier Inc. All rights reserved. doi:10.1016/j.esd.2011.03.001
M.V. Rane, Y.S. Padiya / Energy for Sustainable Development 15 (2011) 184–191
Nomenclature COP coefficient of performance latent heat of fusion, kJ/kg hlat_f hsw specific enthalpy of seawater, kJ/kg kWhh/c/e/m3 heating/cooling/electrical energy per unit volume of water mf mass flow rate, kg/s Q heat flow rate, W t temperature volf volume flow rate W power η efficiency φheat.leak percentage of heat leakage
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seawater to reduce fouling rate of membrane. In case fouling is excessive, membrane can no longer be cleaned. This increases membrane replacement and pre-treatment costs. The pre-treatment cost for 100 l per day capacity plant was found to be 10€/m3 (700 Rs/m3) while it is 0.3 to 0.9 €/m3 (21 to 63 Rs/m3) in large capacity plant (Elfil et al., 2007). Thus, for small production capacities of the range of 1 to 100 l per day of production using sea water as the feed, RO process may not be an optimum choice. Distillation and membrane separation technologies thus have difficulties mainly due to scaling of equipment, plant size, space requirement and reliability. The major problems in the systems discussed are scaling of the equipment and related operational cost. Freezing technologies being operated at lower temperature obviates the scaling related issues and increases equipment life. Freezing technologies
Subscript 2stage two stage compression c.LHE latent heat exchanger condenser car Carnot e evaporator EC external condenser FCS freeze concentration system HPC high pressure compressor HPC.llar high pressure compressor in parallel compression system LHE latent heat exchanger llar parallel llar.cmp.FCS parallel compression freeze concentration system LPC low pressure compressor LPC.llar low pressure compressor in parallel compression system sw seawater tot total w water
However, its productivity is low, 2 to 5 l/d·m2 compared to indirect technologies. Indirect solar technology operating with multiple effects HDH techniques has resulted in water generation of 12 l/d·m2 (Farid and Al-Hajaj, 1996). Although this is three times higher than simple solar still, the cost associated with air circulation is high due to large air side pressure drop in humidifier and dehumidifier. Solar collectors used for seawater desalination could be flat plate collector, evacuated tube collector or parabolic trough collector (Qiblawey and Banat, 2008). In membrane separation technology, salt is separated from seawater through semi permeable membrane by reverse osmosis (RO) process. This process is most extensively used due to ease of operation, and is applicable to a wide range of capacities, from household level to entire cities. However requirement of high pumping pressure of seawater through membrane, 55 to 70 bar, increases specific energy consumption and maintenance cost (Avlonitis et al., 2003). The high pumping pressure of seawater is due to its high osmotic pressure (22.7 bar at 3.5% salt concentration), which increases further with increased salt concentration (Spiegler, 1966). The specific energy consumption for large capacity plant without energy recovering device is 8 to 12 kWhe/m3. This specific energy consumption increases over a period of time due to fouling/ageing of the membrane, which also affects product water quality. The specific energy consumption of RO can be reduced by installing brine energy recovery devices such as turbo, Pelton turbine or pressure exchanger. The use of turbo and Pelton turbine have resulted in specific energy consumption of 3.7 to 5.3 kWhe/m3 (Stover, 2007), while a pressure exchanger reported specific energy consumption of 3.02 kWhe/m3 in 14,888 m3/d capacity plant (Avlonitis et al., 2003). However, high cost of energy recovery device may restrict their use in small capacity plants. Further RO requires high pre-treatment of
In freezing technology, seawater is cooled below its freezing temperature, which is a function of its salinity. As salt concentration in normal seawater is much less than its eutectic composition as 23.3 wt. % NaCl, ice crystals of pure water are formed when it starts freezing. Depending on equipment design, ice could form in suspension with the mother liquor or as a frozen layer on the heat transfer surface. Common methods of freezing include direct, indirect or vacuum freezing and their classification is as shown in Fig. 1. Direct freezing here relates to process where immiscible refrigerant and seawater directly comes in contact with each other. This lowers the driving potential for heat and mass transfer and eliminates cost of heat transfer surface. Evaporation of refrigerant cools seawater below its freezing point and form ice crystals. Ice crystals formed are in suspension with mother liquor and are separated and collected in container, while converted refrigerant vapour is condensed in an ice container and is reused. The refrigerants used are iso-butane/n-butane or Freon114. The advantage of this system is its lower energy consumption, but part retention of refrigerant traces with ice formed makes the generated water non potable. Reverse Osmosis (RO) combined with direct contact freezing, DCF, could reduce the overall energy consumption by about 13% and 17% when compared with separate RO and DCF plants respectively, along with reduction in brine outflow. Here rejected brine from RO is used as feed to freezing system (Madani and Aly, 1989). In vacuum freezing technique, the seawater surface is maintained at a temperature below its triple point. The triple point of seawater with 3.5% salt by weight is at −2.1 °C and 0.0051 bar. Evaporation of water generates ice crystals in suspension with mother liquor, with lower triple point of −3.11 °C and 0.0046 bar after 33% evaporation of water. The ice formed is collected in container, while generated water vapour is compressed and condensed in the frozen ice chamber. The combined liquid due to condensation of vapour and melted ice is used as potable water. The main energy required in the process is for compressing the vapour. Here compressor design is difficult, owing to large specific volume of the water vapour, about 20,600 m3 per tonne of water. Further, due care is needed to avoid the presence of noncondensable gases in the system, thus placing rigid requirement of system components. Leak proof joints are necessary to maintain vacuum in the process, with additional rigid container designs. The system also requires extra refrigeration system to remove heat gain from atmosphere, compressor, pumps and other mechanical devices (Spiegler and Laird, 1980). Vacuum absorption system eliminates the need of compressor. Low vapour pressure of absorbent are able to absorb water vapour from seawater and generate vacuum over seawater surface. Heating the absorbent regenerates water vapour which can be condensed in ice chamber to generate water. However the disadvantages of vacuum operation are still present in this system. Further, selection of proper absorbent is required to avoid toxicity and solubility in water, which
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Fig. 1. Freezing methods for seawater desalination.
will otherwise contaminate the product water. The system also requires large thermal energy due to high boiling point elevation of desiccant and pumping power to circulate concentrated/diluted absorbent. Cerci (2003) proposed freezing of water by passing low water vapour pressure air over it. This avoids the problem of vacuum operation and refrigerant mixing. Low water vapour pressure air can be obtained by passing air over desiccants, which can be later regenerated using external heat. The system claimed to generate a maximum of 28.2 g of ice per kg of dry air flow, at 100% effectiveness of heat exchanger and with use of zero water vapour pressure in air. This system therefore consumes minimum thermal energy of 713.7 kJ/kg of water. Practical heat exchanger effectiveness as 0.7 will further increase energy requirement to 1374 kJ/kg of water. This energy requirement is much higher compared to multistage flashing or multiple effect evaporation techniques. Further, high air circulation power and similar ice–brine separation techniques make the proposed concept energy intensive and complicated. Several freeze desalination techniques are compared in Table 1. Indirect freezing is a method where refrigerant and seawater are separated in crystallizer by heat transfer surface (HTS) and allows operation of seawater at atmospheric pressure. Therefore, the major problems in direct and vacuum freezing method can be avoided by indirect methods. Crystallizers mainly used are of tubular heat exchanger type. Here, ice layer formed on HTS is easier to manage and can be removed either by scraping in scraped surface heat exchanger (SSHE) or with fluidized particles in fluidized bed heat exchanger (FBHE). SSHE has rotating scraper which continuously removes ice from inside the crystallizer wall. However, the system is required to operate under low tolerance and high frictional power between scraper and ice layer, adding extra load on the cooling unit. Further, newborn ice particles (or patches) on cooling surface areas are at freezing point. Thus ice gets scraped off before being sub cooled, depending on the temperature of the cooling surface and speed of the scraper. In order to use this technology for large capacity, either higher length or larger diameter crystalliser is required, and both these can affect scraper power. FBHE consists of hard particles suspended in a vertical column by the upward flowing liquid. Its gentle scouring action prevents the build up of ice layer on crystallizer wall and maintains the surface constantly clean. This gives high heat transfer rates and lower initial cost over SSHE. However, its ice removal rate is lower than SSHE (Habib and Farid,
2006). Further, separation of ice crystals from suspended particles is also difficult. Thus both the process of ice formation in SSHE and FBHE suffers from ice brine separation. Ice crystals generated in the discussed method are either in suspension with mother liquor or later mixed with brine. Crystals formed in suspension are flat in shape, small in size (as large as 0.4 mm and as small as 0.02 mm) and have high surface-to-weight ratio. Further, these crystals are prone to be contaminated by wetting with mother liquor and sizable proportion of the mother liquor is also held in the many interstices between crystals (Spiegler, 1966). Separation of ice and brine is carried out in decanter or wash column. Here, slurry mixture is fed from bottom of decanter. Since ice has lower specific gravity, 0.92, compared to brine, 1.05 at −4.5 °C, it floats over brine. The crystals at top of decanter are washed with part of generated water and are collected in ice chamber. The process thus requires large amount of water to wash adhered regulated brine which is a water loss. The removal of trapped salt solution in the formed ice is difficult and this makes the water non potable. Thus these processes have major problems of ice seawater separation and also require separate refrigeration system to maintain lower temperature in crystallizer. A freeze concentration system (FCS) that incorporates a heat pump was developed by Rane and Jabade (2002). This system obviates the problem of ice brine separation. It has heat transfer surface (HTS) made of stainless steel, plastic or metal coated with plastic, over which aqueous solution flows due to gravity and the surface is open to atmosphere. Refrigerant passages were bonded on HTS by suitable thermal bonding material and freeze the aqueous solution during refrigerant evaporation or melts water during refrigerant condensation. The system has two latent heat exchangers (LHE), which alternately work as an evaporator and condenser of heat pump, depending on refrigerant flow direction. Seawater passes to evaporator side of heat pump, where water from it is selectively frozen on HTS of LHE and concentrated brine comes out. The deposition of ice layer on HTS after prolonged operation reduces the evaporator pressure below preset value. This activates the controller of system and reverses the direction of refrigerant and seawater flow and evaporator and condenser exchange their duties. The condensation of refrigerant melts the frozen ice which was formed in previous cycles and produce potable water as seawater is now supplied to evaporator section. Thus in one LHE freezing takes place while in other melting of previously
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Table 1 Comparison of freeze desalination methods. Method
Advantages
Disadvantages
Energy consumption
Reference
Direct freezing with n-butane
a. Less energy consumption
9.42 kWhe/m3
Madani (1992)
Direct freezing with air
a. Atmospheric pressure operation b. No water contamination a. Eliminates need of secondary refrigerant a. Eliminates refrigerant and compressor
a. Retention of refrigerant traces b. Difficult ice brine separation a. High energy consumption b. Difficult ice brine separation a. Vacuum operation b. Difficult ice brine separation a. High energy consumption for desiccant regeneration b. Vacuum operation c. Less potable d. Difficult ice brine separation a. Required large thermal energy b. Vacuum operation c. Difficult ice brine separation a. Vacuum operation b. Difficult ice brine separation a. High scrapper power b. Restricted capacity per crystalliser c. Difficult ice brine separation
380 kWhth/m3
Cerci (2003)
Vacuum freezing with vapour compression Vacuum absorption vapour compression
Vacuum freezing ejector absorption
a. Eliminates compressor b. Eliminates use of refrigerant
Vacuum freezing multi phase transformation a. High heat transfer rate b. Avoid low pressure compressor Indirect freezing with scrapped surface a. Atmospheric pressure operation heat exchanger, SSHE b. Avoid mixing of refrigerant and seawater c. High ice removal rate Indirect freezing with fluidised bed a. Atmospheric pressure operation heat exchanger b. Simple operation c. High overall heat transfer coefficient than in SSHE Freeze concentration system with a. Eliminate difficulty of ice brine heat pump, FCSwHP separation b. Atmospheric pressure operation c. Simple operation d. Modularity e. Low cost material can be used Freeze concentration system with two stage a. Maintain the advantage of FCSwHP compression using tubular heat exchanger b. Lower inclusion rate due to high fluid velocity and high ice layer thickness c. Lower ice melting loss d. Heat exchanger wall thickness is less than in SSHE e. Higher mass transfer coefficient f. Lower energy consumption
frozen ice takes place. The process of freezing-melting continues in cyclic manner. Seawater flow is always directed to evaporator side of respective cycle operation and this forms solution free water. The cycle also has the advantages of high COP operation due to low temperature lift and maximum utilisation of latent heat, which lowers the specific energy consumption to 6 kWhe/m3 of water production. However as melting of ice cannot reject all heat gain in the system, extra heat has to be rejected to ambient temperature heat sink through an additional compressor. This increases specific energy consumption of the system. The two compressors of the system are arranged in parallel and each one picks up part of heat from evaporator and rejects it to respective pressures (Rane and Jabade, 2005). The HPC thus operates at high pressure ratio, which lowers its volumetric efficiency as well as isentropic efficiency. Further, heat exchangers used in such freeze concentration system are open and exposed to atmosphere. Due to this open structure of the heat exchanger, a significant amount of inclusion/solute of the feed stream occurs in the layer if ice is formed on the heat transfer surface. In an open structure heat exchanger the flow of solution is due to gravity, which limits the feed velocity and the mixing/turbulence. This heat exchanger therefore has lower heat and mass transfer coefficients. Further, lower velocity results in increase of solute inclusion in the frozen layer. Here velocity of the solution depends on the flow rate per unit wetted surface width. Increasing the velocity required increase in solution flow rate. This would increase the film thickness, resulting in inclusion of solute in the frozen layer. Although inclusion
11.9 kWhe/m
3
11.4 kWhe/m3
50 to 70 kWhth/m3
El-Nashar (1984)
9.3 to 10.6 kWhe/m3 Cheng et al. (1987) Habib and Farid (2006)
a. Less ice removal rate b. Difficult ice brine separation
a. Batch process results in thermal cycling loss b. Lower mass transfer coefficient c. Required larger heat transfer area for latent heat exchanger d. High specific energy consumption
Cheng et al. (1987) Cheng et al. (1987)
Habib and Farid (2006)
9 to 12 kWhe/m3
a. Batch process results in thermal cycling loss, 8 to 12 kWhe/m3 which further increases due to large thermal mass of tube b. Need metal tube heat exchanger
Rane and Jabade (2005)
Present method
obtained was within acceptable limits, required surface area and cost increases significantly. Two stage freeze concentration system with tubular heat exchanger The method proposed here suggests the use of a patented vented double wall Tube–Tube Heat Exchanger (TT_HE) as a heat transfer surface instead of open-to-atmosphere HTS (Rane and Tandale, 2002; Rane and Padiya, 2010). The double wall feature of TT_HE addresses the potability issues and avoids accidental mixing of water and refrigerant. TT_HE has already proven its reliability in various applications such as recovery of heat from refrigeration and air conditioning system, potable water chilling and heating (Rane, 2010). TT_HE with copper tubes on the refrigerant sides and Cu–Ni on seawater sides is more economical than the conventional tube-intube heat exchanger (TiT_HE) which requires larger diameter refrigerant tube encasing the Cu–Ni seawater tube. Since the overall heat transfer coefficient is governed by heat transfer coefficient on the seawater/ice side, TT_HE will also offer similar overall heat transfer coefficient like those of TiT_HE. Evaporating/condensing refrigerant passes through one set of tubes in the TT_HE and the solution passes through the other set of tubes as shown in Fig. 2. Since, the solution passes through tubes and not over an open surface, the velocity of the solution can be increased as desired by optimising pumping power. Moderate velocity of flow reduces the inclusion rate and fouling rate, along with increase
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Working principle of two-stage compression based freeze concentration system
Fig. 2. Representation of tube-tube heat exchanger for latent heat exchanger.
in heat and mass transfer coefficient, which reduces size of system (Rane and Jabade, 2005). Bends in the TT_HE introduce secondary flows in the solution and help to further increase heat and mass transfer rate. Effective wetting of the heat transfer surface is assured as the flow is inside the tubes. Further, as only one side of ice layer in tube is exposed to seawater, related melting and inclusion losses are lower than in conventional scraped surface systems. The SW tubes can also be cleaned in place (CIP) using chemical descaling agents, thus reducing the down time and labour involved in descaling. In order to increase system COP with minimum heat rejection from external condenser, a two stage compression system was designed as shown in Fig. 3 The low pressure compressor (LPC) compresses refrigerant from evaporator pressure to intermediate pressure which corresponds to melting temperature of ice. As complete condensation of refrigerant is not possible in LHE condenser, uncondensed refrigerant vapour from it is further compressed to a higher pressure by high pressure compressor (HPC) and is completely condensed in an external condenser. HPC compressor thus operates between intermediate pressure and external condenser pressure. The flash chamber judiciously used at the intermediate pressure reduces flashing at the inlet of evaporator. This reduces the size and work required by the low pressure compressor. It also reduces the superheat of the vapour at inlet of high pressure compressor and helps to reduce its power consumption. Low superheat in various heat exchangers also improves the heat transfer coefficients of respective exchangers.
As shown in Fig. 3, the system has two latent heat exchangers LHE1 and LHE2, an external condenser EC, two compressors LPC and HPC, a heat recovery liquid–liquid heat exchanger LL_HE, liquid separator cum flash chamber (LS) and valves. The opening and closing of four-way valves V1, V2 and V3 are as per required flow direction. Two 4-way valves, V1 and V2, are used to switch the refrigerant flow between LHE1 and LHE2, while 3-way valve V4 and a 4-way valve V3 are used to direct the seawater, water and brine flow. The cycle operates with intermediate reversing of refrigerant and seawater flow directions. Both latent heat exchangers (LHEs) alternately work as evaporator and condenser of heat pump. Low pressure compressor works between latent heat exchangers, while high pressure compressor works between latent heat exchanger working as a condenser and an external condenser. Refrigerant flow circuit was divided into two sections; flow between latent heat exchangers and flow between high pressure compressor and external condenser. The operation of first phase (half cycle) where LHE1 acts as an evaporator and LHE2 acts as a condenser is explained in the following sections. As shown in Fig. 3, incoming seawater is pre-cooled from 1 to 2 in three stream liquid–liquid heat exchanger (LL_HE) and is fed to LHE1 at 3 through valve V4. The evaporation of refrigerant from ‘c’ to ‘d’, selectively freezes the water from seawater in the form of an ice layer on inner periphery of seawater carrying tube. Concentrated brine from LHE1 is let out at 4 and rejected to atmosphere through LL_HE and EC as 4-6-8-10. In LL_HE cold brine stream 6 pre-cools the incoming seawater feed and heated to ‘8’. The seawater path thus is 1-2-3-4-6-8-10. Evaporated refrigerant from ‘d’ is fed to low pressure compressor through valve V1 and is compresses to intermediate pressure from ‘a’ to ‘b’. The compressed refrigerant is then fed to LHE2 at ‘c’ through valve V2. The refrigerant at LHE2 is partially condensed to ‘d’ due to melting of ice, which was formed in previous cycle. The melted ice at ‘5’ which is water, is passed through LL_HE through valve V3 and this also helps in pre-cooling of incoming seawater feed. The water is collected as potable water at 9. The water flow path thus is 5-7-9. The LL_HE is a three-fluid heat exchanger as shown in Fig. 4. The partially condensed
Fig. 3. Schematic of two stage heat pump based freeze desalination system.
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LL_HE may reduce its temperature to −2 °C, as entering concentrated brine temperature is at −4 °C. This reduces the heat duty on LHE. The total cooling load on evaporator of freeze concentrated system FCS, is the sum of sensible duty (Qsesible, FCS) for pre-cooling of seawater from −2 °C to −4 °C and latent duty (Qlatent, FCS) of water as: Q sensible:FCS = ðmf sw;3:5% × hsw;3:5%;2C Þ ðmf sw;7:0% × hsw;7:0%;4C Þ Fig. 4. Configuration of liquid-liquid heat exchanger, LL_HE.
= 6:28 kWhc = m3
refrigerant from LHE2 is fed to liquid separator, LS, through valve V1. In LS, vapour and liquid part of refrigerant are separated. The vapour ‘g’ which is a combination of uncondensed vapour coming from V1 and flash vapour coming out of E2, is compressed to ‘h’ in the high pressure compressor, HPC, and fed to an external condenser, EC. In EC, brine from LL_HE, completely condenses the refrigerant vapour to ‘i’. The condensed refrigerant from EC is throttled through high pressure throttle valve EV2 to intermediate pressure and is fed to LS. The liquid phase from LS was throttled to evaporator pressure through throttle valve, EV1, from f to k and is supplied to LHE1 at ‘c’ through valve V2. Further, evaporation of refrigerant in LHE1 brings the state to ‘d’ and the cycle is completed. The routing of refrigerant from EC to LS reduces the vapour fraction circulation in system. The flow path of refrigerant cycle through LHEs thus is a-b-c-d-e-f-k-c'-d'-a, while refrigerant flows through HPC will be g-h-i-j. The cycle continues to operate till the evaporator pressure falls below the set pressure and reverses the refrigerant flow and changes the seawater flow path. Evaporator pressure reduces as ice layer builds in the tubes carrying seawater. This increases the thermal resistance. After reduction of evaporator pressure below preset value, the reversing of the refrigerant valves V1 and V2 occurs which changes the flow directions of seawater and refrigerant. This enables the interchange of function of the two LHEs. LHE2 now becomes evaporator and LHE1 is the condenser. The seawater flow is also directed to LHE2. The refrigerant flow across LS, EC and HPC, however, remains the same. As the LPC feeds refrigerant to LHE1, the ice which was formed in previous cycle melts and seawater flowing through LHE2 partially freezes. Heat of freezing of water is picked up by the evaporating refrigerant. Thus, both the processes of freezing of water and meting of ice occurs concurrently in the two LHEs. This cycle further continues to operate till evaporator pressure, that is, refrigerant pressure in LHE2 falls below the set pressure. Refrigerant flow path becomes a-b-c'-d'-e-f-k-c-d-a, while the flow through EC-HPC-LS remains same. Similarly, the flow path of seawater and water becomes 1-2-3'-5-6-8-10 and 4-7-8 respectively. The seawater flow is thus always directed to evaporator of respective cycle, where freezing occurs. Calculation of specific power consumption for 1000 kg/h water production Assuming that seawater concentration is doubled as it passes over freeze concentrating surface, a feed flow rate, mfsw, of 2000 kg/h is required for generating 1000 kg/h of potable water, mfw. Assuming a minimum approach of 3 °C in heat exchangers and neglecting variation of specific heat of seawater with concentration and temperature, the following calculations are presented. As the freezing point of sea water at 7.0% salt concentration is −4 °C, the operating temperatures of refrigerant in evaporator will be −7 °C, while melting of ice at 0 °C required condensing temperature of 3 °C. Carnot COP of 26.7 is calculated for −7 °C evaporator and 3 °C condensing temperature. The performance of commercially available refrigeration system for condensing temperature higher than 30 °C revealed that its Carnot efficiency with R22 as the refrigerant varies from 0.4 to 0.6 depending on lift and capacity. Carnot efficiency is defined as the ratio of actual COP to Carnot COP of refrigeration cycle. Assuming Carnot efficiency of 0.6, actual COP of 15.92 is possible for compressor operating between −7 °C and 3 °C. The pre-cooling of seawater in
Q latent:FCS ¼mf w × hlatf = 93:06 kWhc = m3
ð1Þ
ð2Þ
where, hsw enthalpy of seawater at respective concentration, kJ/kg hlat_f latent heat of fusion of water, kJ/kg. Thus, heat load on evaporator is 99.34 kWhc/m3. Assuming heat gain from ambient to FCS system is 5% of total heat load, total cooling load on the evaporator will be 104.3 kWhc/m3. Taking COP of cycle as 15.92, specific power consumption is 6.55 kWhe/m3 of water, if LHE condenser alone rejects all heat gained in the system. But in actual, LHE condenser can reject only part of total heat load equivalent to sensible heating and melting heat load of ice. The excess heat contributed by compressor power input and heat gain from ambient, has to reject to atmosphere through an external condenser, EC. Assuming condensing temperature of EC as 36 °C, the net COP of freeze concentration system COPnet.FCS, has been analysed in following section. The Sankey diagram representing heat distribution in two-stage freeze concentration system for 1 m3/h of water production is shown in Fig. 5. As per the calculation in preceding section total heat to be rejected in condenser is 110.8 kWhh/m3. The latent heat condenser can reject only 96.2 kWhh/m3 heat, which is equivalent to sensible cooling of ice from −4 °C to 0 °C and the melting heat load. The remaining heat 14.66 kWhh/m3 is rejected in external condenser, which is operating at 36 °C. Assuming COP of 3.6 for pumping the excess heat from intermediate temperature to the ambient heat sink, high pressure compressor power will be 4.07 kWhe/m3. The total heat rejected in external condenser, QEC, is 14.66 + 4.07 = 18.73 kWhh/m3. This results in total power consumption by both compressors, Wtot.cmp, of 10.62 kWhe/m3 and net cycle COP, COPnet.FCS, of 9.82. The effect of variation in Carnot efficiency of low pressure compressor, ηcar.LPC, on various system parameters is given in Fig. 6. Comparing parallel compression (Rane and Jabade, 2005) and two stage compression systems, it was found that overall specific energy consumption of both systems was almost same at ηcar.LPC of 0.45 to 0.5. However, possibility of higher Carnot efficiency for LPC, ηcar.LPC, above 0.5 will reduce overall energy consumption of proposed system. This is because, in parallel compression system LPC picks up only part of evaporator heat load, equivalent to rejected heat for melting of ice in LHE condenser and thus has lower rating. While in two-stage system, LPC takes total evaporator heat load irrespective of LHE condenser capacity, and thus has higher power consumption. Thus a lower Carnot efficiency of LPC has a greater impact in a two-stage system. On the other hand HPC in parallel compression system picks up balance heat from evaporator and rejects it to EC, and thus has higher lift and power consumption. While in two stage system, HPC operates between intermediate pressure corresponding to LHE condenser temperature and EC pressure and thus has lower lift and power consumption. The overall power consumption of both systems is thus almost the same. However, proposed two-stage system has advantages of lower heat rejection rate at external condenser and is compact. The system thus has potential to provide water at specific energy consumption, 10.62 kWhe/m3, and without additional cost of pretreatment as in distillation or membrane processes. Heat exchanger
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tubes could be of suitable material such as Cu–Ni, aluminium alloy or epoxy coated steel. Diameter and thickness of the tubes could be in the range of 12.7 to 25.4 mm and the wall thickness of the order of 1 to 2 mm. Small system using 1 TR capacity compressor has been conceived and will be developed in near future.
Conclusions
Fig. 5. Sankey diagram: Two stages heat pump operated freeze concentration system for 1000 kg/h of water production.
Desalination of seawater as a source of potable water is likely to continue in the future as available pure water sources are reducing. Conventional desalination technologies such as reverse osmosis, multiple effect evaporator, multistage flashing and mechanical vapour compression have equipment scaling related issues which increase the operational and maintenance cost. Freeze desalination technique operates at low temperature and has less scaling problems. Conventional freezing methods include direct, indirect or vacuum freezing techniques. Each of these methods has operational difficulties like ice– seawater separation and loss of pure water. Further as direct and vacuum technologies are operated at lower pressure, their operational difficulties are more. Indirect freezing has the disadvantage of more power consumption due to scraper operation in scraped surface heat exchanger, while in fluidised bed heat exchanger ice removal rate is low. Novel freeze desalination systems coupled with a heat pump and using patented vented-double-wall tube–tube heat exchanger (TT_HE) was developed. System utilises two stage compression processes and optimally utilises latent heat of ice melting for refrigerant condensation. Since the melting of ice occurs in absence of seawater, the need for separation of ice–seawater is eliminated. The use of TT_HE eliminates accidental mixing of seawater, refrigerant and pure water. It allows moderate velocity of seawater flow, which reduces salt inclusion rate, fouling rate and improves compactness of system. Washing requirement and melting loss are further reduced because of increased ice layer thickness and selective flow of seawater. Two-stage compression with automated control valve enables this system to be deployed for a wide range of capacities. System offers significant initial and operating cost saving when compared with conventional freeze desalination systems. The expected energy consumption for the two stage compression system is 9 to 11 kWhe/m3 of water produced with a cycle COP ranging from 8 to 12.
Fig. 6. Variation in heat pump operated freeze concentration system parameters with Carnot efficiency of low pressure compressor.
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References Al-Enezi G, Ettouney H, Fawzy N. Low temperature humidification dehumidification desalination process. Energy Convers Manage 2006;47:470–84. Amara MB, Houcine I, Guizani A, Maalej M. Theoretical and experimental study of a pad humidifier used in a seawater desalination process. Desalination 2004;168: 1-12. Avlonitis SA, Kouroumbas K, Vlachakis N. Energy consumption and membrane replacement cost for seawater RO desalination plants. Desalination 2003;157: 151–8. Cerci Y. A new ideal evaporative freezing cycle. Int J Heat Mass Transfer 2003;46: 2967–74. Cheng CY, Cheng WC, Yang MD. The vacuum freezing multiple phase transformation process. Desalination 1987;67:139–53. Cipollina A, Micale G, Rizzut L. A critical assessment of desalination operations in Sicily. Desalination 2005;182:1-12. Elfil H, Hamed A, Hannachi A. Technical evaluation of a small-scale reverse osmosis desalination unit for domestic water. Desalination 2007;203:319–26. El-Nashar AM. Solar desalination using the vacuum freezing ejector absorption (VFEA) process. Desalination 1984;49:293–319. Farid M, Al-Hajaj AW. Solar desalination with a humidification–dehumidification cycle. Desalination 1996;106:427–9.
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Habib B, Farid M. Heat transfer and operating conditions for freeze concentration in a liquid–solid fluidized bed heat exchanger. Chem Eng Process 2006;45:698–710. Madani A. Zero-discharge direct-contact freezing/solar evaporator desalination complex. Desalination 1992;85:179–95. Madani AA, Aly SE. A combined RO/freezing system to reduce inland rejected brine. Desalination 1989;75:241–58. Meindersma GW, Guijt CM, de Haan AB. Desalination and water recycling by air gap membrane distillation. Desalination 2006;187:291–301. Qiblawey HM, Banat F. Solar thermal desalination technologies. Desalination 2008;220: 633–44. Rane MV. Energy conservation using superheat recovery and heat pumps. ISHRAE 2010;13(2):40–60. Rane, M.V, Jabade, S.K., 2002. Freeze concentration system. Indian Patent. 204 956. Rane MV, Jabade SK. Freeze concentration of sugarcane juice in a jaggery making process. Appl Therm Eng 2005;25:2122–37. Rane, M.V., Padiya, Y.S., 2010. Freeze concentration system with tubular heat exchangers. Indian Patent Application # 1915/MUM/2010. Rane, M.V., Tandale, M.S., 2002. Tube-tube heat exchangers. Indian Patent. 205 362. Spiegler KS. Principles of desalination. New York: Academic Press; 1966. Spiegler KS, Laird ADK. Principles of desalination. New York: Academic Press; 1980. Stover Richard L. Seawater reverse osmosis with isobaric energy recovery devices. Desalination 2007;203:168–75.
Contents lists available at ScienceDirect
Energy for Sustainable Development
Heat pump operated freeze concentration system with tubular heat exchanger for seawater desalination M.V. Rane a,⁎, Y.S. Padiya b a b
Mechanical Engineering Department, Indian Institute of Technology, Bombay, Mumbai, India Mechanical Engineering Department, Konkan Gyanpeeth College of Engineering, Karjat, Dist. Raigad, Maharashtra, India
a r t i c l e
i n f o
Article history: Received 15 July 2010 Revised 6 March 2011 Accepted 6 March 2011 Available online 9 April 2011 Keywords: Desalination Heat pump Layer freezing Energy efficient
a b s t r a c t Desalination of seawater can provide an almost inexhaustible source of freshwater if it can be made affordable. Distillation and filtration, the commonly used processes, have high operating and maintenance cost. Low operating temperature and low latent heat of fusion of water makes freezing technology worth considering for further development. Here, water is selectively frozen, in the form of ice from sea water, and melted after isolating it from the concentrated seawater to generate pure water which can be made potable. However, desalination processes based on freezing have not been exploited because of the operational difficulties in ice–water separation, high cost of equipment and high parasitic power requirement. This paper discusses a patented layer freezing based technology which has competitive initial and operating cost and eliminates operational difficulties of conventional freezing system. It is scalable and is coupled with a heat pump which selectively freezes water from seawater in the evaporator and melts the ice in the subsequent phase when it serves as a condenser. The condenser optimally utilises the latent heat of melting of ice to partially condense the refrigerant and the excess heat is rejected to ambient. It avoids the need of ice scraper/separation mechanisms. Use of vented-double-wall tube–tube heat exchanger, TT_HE, enables refrigerant and seawater/potable water to exchange heat without the use of intermediate fluids, while keeping the inclusion low. TT_HE is a reliable heat exchanger which ensures that refrigerant and seawater/water do not mix. Operating COP of the heat pump is in the range of 8 to 12, which results in specific energy consumption in the range of 9 to 11 kWhe/m3 of water produced. Comparison of features with other freezing desalination processes is also presented. © 2011 International Energy Initiative. Published by Elsevier Inc. All rights reserved.
Introduction Desalination is the process of separation of salt from water. It is mainly carried out using evaporation, membrane separation and freezing technologies. In evaporation technology, seawater is heated above its boiling point (normally 100.8 °C and 101.7 °C at 3.5% and 7% salt concentrations) and the generated vapour is condensed to get water. Multistage flashing (MSF), multiple effect evaporation (MEE), mechanical vapour compression (MVC), humidification and dehumidification (HDH), membrane distillation (MD), pervaporization and solar distillation are the major techniques which come under evaporation category. MSF, MEE and MVC are suitable for 100 to 500,000 m3/d fresh water production capacity plant. The specific energy consumption of MSF and MEE varies from 150 to 300 kJ/kg with additional electrical input of 2 to 4 kWhe/m3 for pumping liquor and water, while energy consumption of MVC is 8 to 15 kWhe/m3 (Cipollina et al., 2005). These technologies require pre-treatment of seawater, corrosion resistant
⁎ Corresponding author: Tel.: +91 22 2576 7514, +91 22 2576 4527, +91 22 2576 4593, +91 22 2576 8514; fax: +91 22 2572 4544, +91 22 2572 6875. E-mail address: [email protected] (M.V. Rane).
material and permanent qualified maintenance staff (Al-Enezi et al., 2006). HDH, MD, pervaporization and solar distillation are low operating temperature technologies and thus minimise corrosion related issues. They are mostly used for small capacity water generation, e.g. 100 l/d. HDH technology is simple in operation, but availability of low mass transfer coefficient in existing humidifier, 1 to 3 kg/s·m3·(kgw/kgda), (Amara et al., 2004) and in dehumidifier, increases the plant size and energy required for air circulation. In MD and pervaporization transport of water vapour from seawater occurs through microporous hydrophobic membranes. The specific energy consumption of MD can be low in the range of 73 to 75 kJ/kg when operated with low grade heat between 50 and 100 °C. It has potential to generate potable water with lower water cost, less than 0.5 $/m3. However, approximately after 800 h of use membrane may get damaged due to biofouling (Meindersma et al., 2006). Solar desalination has been reported by many researchers. It can be used directly or coupled with other technologies such as MSF, MEE, vapour compression, RO, MD and electro dialysis, where solar collectors are used for heat generation. Direct solar desalination method includes use of solar still, which has no moving parts and is simple in operation. It is competitive to indirect desalination plant in small scale production.
0973-0826/$ – see front matter © 2011 International Energy Initiative. Published by Elsevier Inc. All rights reserved. doi:10.1016/j.esd.2011.03.001
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Nomenclature COP coefficient of performance latent heat of fusion, kJ/kg hlat_f hsw specific enthalpy of seawater, kJ/kg kWhh/c/e/m3 heating/cooling/electrical energy per unit volume of water mf mass flow rate, kg/s Q heat flow rate, W t temperature volf volume flow rate W power η efficiency φheat.leak percentage of heat leakage
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seawater to reduce fouling rate of membrane. In case fouling is excessive, membrane can no longer be cleaned. This increases membrane replacement and pre-treatment costs. The pre-treatment cost for 100 l per day capacity plant was found to be 10€/m3 (700 Rs/m3) while it is 0.3 to 0.9 €/m3 (21 to 63 Rs/m3) in large capacity plant (Elfil et al., 2007). Thus, for small production capacities of the range of 1 to 100 l per day of production using sea water as the feed, RO process may not be an optimum choice. Distillation and membrane separation technologies thus have difficulties mainly due to scaling of equipment, plant size, space requirement and reliability. The major problems in the systems discussed are scaling of the equipment and related operational cost. Freezing technologies being operated at lower temperature obviates the scaling related issues and increases equipment life. Freezing technologies
Subscript 2stage two stage compression c.LHE latent heat exchanger condenser car Carnot e evaporator EC external condenser FCS freeze concentration system HPC high pressure compressor HPC.llar high pressure compressor in parallel compression system LHE latent heat exchanger llar parallel llar.cmp.FCS parallel compression freeze concentration system LPC low pressure compressor LPC.llar low pressure compressor in parallel compression system sw seawater tot total w water
However, its productivity is low, 2 to 5 l/d·m2 compared to indirect technologies. Indirect solar technology operating with multiple effects HDH techniques has resulted in water generation of 12 l/d·m2 (Farid and Al-Hajaj, 1996). Although this is three times higher than simple solar still, the cost associated with air circulation is high due to large air side pressure drop in humidifier and dehumidifier. Solar collectors used for seawater desalination could be flat plate collector, evacuated tube collector or parabolic trough collector (Qiblawey and Banat, 2008). In membrane separation technology, salt is separated from seawater through semi permeable membrane by reverse osmosis (RO) process. This process is most extensively used due to ease of operation, and is applicable to a wide range of capacities, from household level to entire cities. However requirement of high pumping pressure of seawater through membrane, 55 to 70 bar, increases specific energy consumption and maintenance cost (Avlonitis et al., 2003). The high pumping pressure of seawater is due to its high osmotic pressure (22.7 bar at 3.5% salt concentration), which increases further with increased salt concentration (Spiegler, 1966). The specific energy consumption for large capacity plant without energy recovering device is 8 to 12 kWhe/m3. This specific energy consumption increases over a period of time due to fouling/ageing of the membrane, which also affects product water quality. The specific energy consumption of RO can be reduced by installing brine energy recovery devices such as turbo, Pelton turbine or pressure exchanger. The use of turbo and Pelton turbine have resulted in specific energy consumption of 3.7 to 5.3 kWhe/m3 (Stover, 2007), while a pressure exchanger reported specific energy consumption of 3.02 kWhe/m3 in 14,888 m3/d capacity plant (Avlonitis et al., 2003). However, high cost of energy recovery device may restrict their use in small capacity plants. Further RO requires high pre-treatment of
In freezing technology, seawater is cooled below its freezing temperature, which is a function of its salinity. As salt concentration in normal seawater is much less than its eutectic composition as 23.3 wt. % NaCl, ice crystals of pure water are formed when it starts freezing. Depending on equipment design, ice could form in suspension with the mother liquor or as a frozen layer on the heat transfer surface. Common methods of freezing include direct, indirect or vacuum freezing and their classification is as shown in Fig. 1. Direct freezing here relates to process where immiscible refrigerant and seawater directly comes in contact with each other. This lowers the driving potential for heat and mass transfer and eliminates cost of heat transfer surface. Evaporation of refrigerant cools seawater below its freezing point and form ice crystals. Ice crystals formed are in suspension with mother liquor and are separated and collected in container, while converted refrigerant vapour is condensed in an ice container and is reused. The refrigerants used are iso-butane/n-butane or Freon114. The advantage of this system is its lower energy consumption, but part retention of refrigerant traces with ice formed makes the generated water non potable. Reverse Osmosis (RO) combined with direct contact freezing, DCF, could reduce the overall energy consumption by about 13% and 17% when compared with separate RO and DCF plants respectively, along with reduction in brine outflow. Here rejected brine from RO is used as feed to freezing system (Madani and Aly, 1989). In vacuum freezing technique, the seawater surface is maintained at a temperature below its triple point. The triple point of seawater with 3.5% salt by weight is at −2.1 °C and 0.0051 bar. Evaporation of water generates ice crystals in suspension with mother liquor, with lower triple point of −3.11 °C and 0.0046 bar after 33% evaporation of water. The ice formed is collected in container, while generated water vapour is compressed and condensed in the frozen ice chamber. The combined liquid due to condensation of vapour and melted ice is used as potable water. The main energy required in the process is for compressing the vapour. Here compressor design is difficult, owing to large specific volume of the water vapour, about 20,600 m3 per tonne of water. Further, due care is needed to avoid the presence of noncondensable gases in the system, thus placing rigid requirement of system components. Leak proof joints are necessary to maintain vacuum in the process, with additional rigid container designs. The system also requires extra refrigeration system to remove heat gain from atmosphere, compressor, pumps and other mechanical devices (Spiegler and Laird, 1980). Vacuum absorption system eliminates the need of compressor. Low vapour pressure of absorbent are able to absorb water vapour from seawater and generate vacuum over seawater surface. Heating the absorbent regenerates water vapour which can be condensed in ice chamber to generate water. However the disadvantages of vacuum operation are still present in this system. Further, selection of proper absorbent is required to avoid toxicity and solubility in water, which
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Fig. 1. Freezing methods for seawater desalination.
will otherwise contaminate the product water. The system also requires large thermal energy due to high boiling point elevation of desiccant and pumping power to circulate concentrated/diluted absorbent. Cerci (2003) proposed freezing of water by passing low water vapour pressure air over it. This avoids the problem of vacuum operation and refrigerant mixing. Low water vapour pressure air can be obtained by passing air over desiccants, which can be later regenerated using external heat. The system claimed to generate a maximum of 28.2 g of ice per kg of dry air flow, at 100% effectiveness of heat exchanger and with use of zero water vapour pressure in air. This system therefore consumes minimum thermal energy of 713.7 kJ/kg of water. Practical heat exchanger effectiveness as 0.7 will further increase energy requirement to 1374 kJ/kg of water. This energy requirement is much higher compared to multistage flashing or multiple effect evaporation techniques. Further, high air circulation power and similar ice–brine separation techniques make the proposed concept energy intensive and complicated. Several freeze desalination techniques are compared in Table 1. Indirect freezing is a method where refrigerant and seawater are separated in crystallizer by heat transfer surface (HTS) and allows operation of seawater at atmospheric pressure. Therefore, the major problems in direct and vacuum freezing method can be avoided by indirect methods. Crystallizers mainly used are of tubular heat exchanger type. Here, ice layer formed on HTS is easier to manage and can be removed either by scraping in scraped surface heat exchanger (SSHE) or with fluidized particles in fluidized bed heat exchanger (FBHE). SSHE has rotating scraper which continuously removes ice from inside the crystallizer wall. However, the system is required to operate under low tolerance and high frictional power between scraper and ice layer, adding extra load on the cooling unit. Further, newborn ice particles (or patches) on cooling surface areas are at freezing point. Thus ice gets scraped off before being sub cooled, depending on the temperature of the cooling surface and speed of the scraper. In order to use this technology for large capacity, either higher length or larger diameter crystalliser is required, and both these can affect scraper power. FBHE consists of hard particles suspended in a vertical column by the upward flowing liquid. Its gentle scouring action prevents the build up of ice layer on crystallizer wall and maintains the surface constantly clean. This gives high heat transfer rates and lower initial cost over SSHE. However, its ice removal rate is lower than SSHE (Habib and Farid,
2006). Further, separation of ice crystals from suspended particles is also difficult. Thus both the process of ice formation in SSHE and FBHE suffers from ice brine separation. Ice crystals generated in the discussed method are either in suspension with mother liquor or later mixed with brine. Crystals formed in suspension are flat in shape, small in size (as large as 0.4 mm and as small as 0.02 mm) and have high surface-to-weight ratio. Further, these crystals are prone to be contaminated by wetting with mother liquor and sizable proportion of the mother liquor is also held in the many interstices between crystals (Spiegler, 1966). Separation of ice and brine is carried out in decanter or wash column. Here, slurry mixture is fed from bottom of decanter. Since ice has lower specific gravity, 0.92, compared to brine, 1.05 at −4.5 °C, it floats over brine. The crystals at top of decanter are washed with part of generated water and are collected in ice chamber. The process thus requires large amount of water to wash adhered regulated brine which is a water loss. The removal of trapped salt solution in the formed ice is difficult and this makes the water non potable. Thus these processes have major problems of ice seawater separation and also require separate refrigeration system to maintain lower temperature in crystallizer. A freeze concentration system (FCS) that incorporates a heat pump was developed by Rane and Jabade (2002). This system obviates the problem of ice brine separation. It has heat transfer surface (HTS) made of stainless steel, plastic or metal coated with plastic, over which aqueous solution flows due to gravity and the surface is open to atmosphere. Refrigerant passages were bonded on HTS by suitable thermal bonding material and freeze the aqueous solution during refrigerant evaporation or melts water during refrigerant condensation. The system has two latent heat exchangers (LHE), which alternately work as an evaporator and condenser of heat pump, depending on refrigerant flow direction. Seawater passes to evaporator side of heat pump, where water from it is selectively frozen on HTS of LHE and concentrated brine comes out. The deposition of ice layer on HTS after prolonged operation reduces the evaporator pressure below preset value. This activates the controller of system and reverses the direction of refrigerant and seawater flow and evaporator and condenser exchange their duties. The condensation of refrigerant melts the frozen ice which was formed in previous cycles and produce potable water as seawater is now supplied to evaporator section. Thus in one LHE freezing takes place while in other melting of previously
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Table 1 Comparison of freeze desalination methods. Method
Advantages
Disadvantages
Energy consumption
Reference
Direct freezing with n-butane
a. Less energy consumption
9.42 kWhe/m3
Madani (1992)
Direct freezing with air
a. Atmospheric pressure operation b. No water contamination a. Eliminates need of secondary refrigerant a. Eliminates refrigerant and compressor
a. Retention of refrigerant traces b. Difficult ice brine separation a. High energy consumption b. Difficult ice brine separation a. Vacuum operation b. Difficult ice brine separation a. High energy consumption for desiccant regeneration b. Vacuum operation c. Less potable d. Difficult ice brine separation a. Required large thermal energy b. Vacuum operation c. Difficult ice brine separation a. Vacuum operation b. Difficult ice brine separation a. High scrapper power b. Restricted capacity per crystalliser c. Difficult ice brine separation
380 kWhth/m3
Cerci (2003)
Vacuum freezing with vapour compression Vacuum absorption vapour compression
Vacuum freezing ejector absorption
a. Eliminates compressor b. Eliminates use of refrigerant
Vacuum freezing multi phase transformation a. High heat transfer rate b. Avoid low pressure compressor Indirect freezing with scrapped surface a. Atmospheric pressure operation heat exchanger, SSHE b. Avoid mixing of refrigerant and seawater c. High ice removal rate Indirect freezing with fluidised bed a. Atmospheric pressure operation heat exchanger b. Simple operation c. High overall heat transfer coefficient than in SSHE Freeze concentration system with a. Eliminate difficulty of ice brine heat pump, FCSwHP separation b. Atmospheric pressure operation c. Simple operation d. Modularity e. Low cost material can be used Freeze concentration system with two stage a. Maintain the advantage of FCSwHP compression using tubular heat exchanger b. Lower inclusion rate due to high fluid velocity and high ice layer thickness c. Lower ice melting loss d. Heat exchanger wall thickness is less than in SSHE e. Higher mass transfer coefficient f. Lower energy consumption
frozen ice takes place. The process of freezing-melting continues in cyclic manner. Seawater flow is always directed to evaporator side of respective cycle operation and this forms solution free water. The cycle also has the advantages of high COP operation due to low temperature lift and maximum utilisation of latent heat, which lowers the specific energy consumption to 6 kWhe/m3 of water production. However as melting of ice cannot reject all heat gain in the system, extra heat has to be rejected to ambient temperature heat sink through an additional compressor. This increases specific energy consumption of the system. The two compressors of the system are arranged in parallel and each one picks up part of heat from evaporator and rejects it to respective pressures (Rane and Jabade, 2005). The HPC thus operates at high pressure ratio, which lowers its volumetric efficiency as well as isentropic efficiency. Further, heat exchangers used in such freeze concentration system are open and exposed to atmosphere. Due to this open structure of the heat exchanger, a significant amount of inclusion/solute of the feed stream occurs in the layer if ice is formed on the heat transfer surface. In an open structure heat exchanger the flow of solution is due to gravity, which limits the feed velocity and the mixing/turbulence. This heat exchanger therefore has lower heat and mass transfer coefficients. Further, lower velocity results in increase of solute inclusion in the frozen layer. Here velocity of the solution depends on the flow rate per unit wetted surface width. Increasing the velocity required increase in solution flow rate. This would increase the film thickness, resulting in inclusion of solute in the frozen layer. Although inclusion
11.9 kWhe/m
3
11.4 kWhe/m3
50 to 70 kWhth/m3
El-Nashar (1984)
9.3 to 10.6 kWhe/m3 Cheng et al. (1987) Habib and Farid (2006)
a. Less ice removal rate b. Difficult ice brine separation
a. Batch process results in thermal cycling loss b. Lower mass transfer coefficient c. Required larger heat transfer area for latent heat exchanger d. High specific energy consumption
Cheng et al. (1987) Cheng et al. (1987)
Habib and Farid (2006)
9 to 12 kWhe/m3
a. Batch process results in thermal cycling loss, 8 to 12 kWhe/m3 which further increases due to large thermal mass of tube b. Need metal tube heat exchanger
Rane and Jabade (2005)
Present method
obtained was within acceptable limits, required surface area and cost increases significantly. Two stage freeze concentration system with tubular heat exchanger The method proposed here suggests the use of a patented vented double wall Tube–Tube Heat Exchanger (TT_HE) as a heat transfer surface instead of open-to-atmosphere HTS (Rane and Tandale, 2002; Rane and Padiya, 2010). The double wall feature of TT_HE addresses the potability issues and avoids accidental mixing of water and refrigerant. TT_HE has already proven its reliability in various applications such as recovery of heat from refrigeration and air conditioning system, potable water chilling and heating (Rane, 2010). TT_HE with copper tubes on the refrigerant sides and Cu–Ni on seawater sides is more economical than the conventional tube-intube heat exchanger (TiT_HE) which requires larger diameter refrigerant tube encasing the Cu–Ni seawater tube. Since the overall heat transfer coefficient is governed by heat transfer coefficient on the seawater/ice side, TT_HE will also offer similar overall heat transfer coefficient like those of TiT_HE. Evaporating/condensing refrigerant passes through one set of tubes in the TT_HE and the solution passes through the other set of tubes as shown in Fig. 2. Since, the solution passes through tubes and not over an open surface, the velocity of the solution can be increased as desired by optimising pumping power. Moderate velocity of flow reduces the inclusion rate and fouling rate, along with increase
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Working principle of two-stage compression based freeze concentration system
Fig. 2. Representation of tube-tube heat exchanger for latent heat exchanger.
in heat and mass transfer coefficient, which reduces size of system (Rane and Jabade, 2005). Bends in the TT_HE introduce secondary flows in the solution and help to further increase heat and mass transfer rate. Effective wetting of the heat transfer surface is assured as the flow is inside the tubes. Further, as only one side of ice layer in tube is exposed to seawater, related melting and inclusion losses are lower than in conventional scraped surface systems. The SW tubes can also be cleaned in place (CIP) using chemical descaling agents, thus reducing the down time and labour involved in descaling. In order to increase system COP with minimum heat rejection from external condenser, a two stage compression system was designed as shown in Fig. 3 The low pressure compressor (LPC) compresses refrigerant from evaporator pressure to intermediate pressure which corresponds to melting temperature of ice. As complete condensation of refrigerant is not possible in LHE condenser, uncondensed refrigerant vapour from it is further compressed to a higher pressure by high pressure compressor (HPC) and is completely condensed in an external condenser. HPC compressor thus operates between intermediate pressure and external condenser pressure. The flash chamber judiciously used at the intermediate pressure reduces flashing at the inlet of evaporator. This reduces the size and work required by the low pressure compressor. It also reduces the superheat of the vapour at inlet of high pressure compressor and helps to reduce its power consumption. Low superheat in various heat exchangers also improves the heat transfer coefficients of respective exchangers.
As shown in Fig. 3, the system has two latent heat exchangers LHE1 and LHE2, an external condenser EC, two compressors LPC and HPC, a heat recovery liquid–liquid heat exchanger LL_HE, liquid separator cum flash chamber (LS) and valves. The opening and closing of four-way valves V1, V2 and V3 are as per required flow direction. Two 4-way valves, V1 and V2, are used to switch the refrigerant flow between LHE1 and LHE2, while 3-way valve V4 and a 4-way valve V3 are used to direct the seawater, water and brine flow. The cycle operates with intermediate reversing of refrigerant and seawater flow directions. Both latent heat exchangers (LHEs) alternately work as evaporator and condenser of heat pump. Low pressure compressor works between latent heat exchangers, while high pressure compressor works between latent heat exchanger working as a condenser and an external condenser. Refrigerant flow circuit was divided into two sections; flow between latent heat exchangers and flow between high pressure compressor and external condenser. The operation of first phase (half cycle) where LHE1 acts as an evaporator and LHE2 acts as a condenser is explained in the following sections. As shown in Fig. 3, incoming seawater is pre-cooled from 1 to 2 in three stream liquid–liquid heat exchanger (LL_HE) and is fed to LHE1 at 3 through valve V4. The evaporation of refrigerant from ‘c’ to ‘d’, selectively freezes the water from seawater in the form of an ice layer on inner periphery of seawater carrying tube. Concentrated brine from LHE1 is let out at 4 and rejected to atmosphere through LL_HE and EC as 4-6-8-10. In LL_HE cold brine stream 6 pre-cools the incoming seawater feed and heated to ‘8’. The seawater path thus is 1-2-3-4-6-8-10. Evaporated refrigerant from ‘d’ is fed to low pressure compressor through valve V1 and is compresses to intermediate pressure from ‘a’ to ‘b’. The compressed refrigerant is then fed to LHE2 at ‘c’ through valve V2. The refrigerant at LHE2 is partially condensed to ‘d’ due to melting of ice, which was formed in previous cycle. The melted ice at ‘5’ which is water, is passed through LL_HE through valve V3 and this also helps in pre-cooling of incoming seawater feed. The water is collected as potable water at 9. The water flow path thus is 5-7-9. The LL_HE is a three-fluid heat exchanger as shown in Fig. 4. The partially condensed
Fig. 3. Schematic of two stage heat pump based freeze desalination system.
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LL_HE may reduce its temperature to −2 °C, as entering concentrated brine temperature is at −4 °C. This reduces the heat duty on LHE. The total cooling load on evaporator of freeze concentrated system FCS, is the sum of sensible duty (Qsesible, FCS) for pre-cooling of seawater from −2 °C to −4 °C and latent duty (Qlatent, FCS) of water as: Q sensible:FCS = ðmf sw;3:5% × hsw;3:5%;2C Þ ðmf sw;7:0% × hsw;7:0%;4C Þ Fig. 4. Configuration of liquid-liquid heat exchanger, LL_HE.
= 6:28 kWhc = m3
refrigerant from LHE2 is fed to liquid separator, LS, through valve V1. In LS, vapour and liquid part of refrigerant are separated. The vapour ‘g’ which is a combination of uncondensed vapour coming from V1 and flash vapour coming out of E2, is compressed to ‘h’ in the high pressure compressor, HPC, and fed to an external condenser, EC. In EC, brine from LL_HE, completely condenses the refrigerant vapour to ‘i’. The condensed refrigerant from EC is throttled through high pressure throttle valve EV2 to intermediate pressure and is fed to LS. The liquid phase from LS was throttled to evaporator pressure through throttle valve, EV1, from f to k and is supplied to LHE1 at ‘c’ through valve V2. Further, evaporation of refrigerant in LHE1 brings the state to ‘d’ and the cycle is completed. The routing of refrigerant from EC to LS reduces the vapour fraction circulation in system. The flow path of refrigerant cycle through LHEs thus is a-b-c-d-e-f-k-c'-d'-a, while refrigerant flows through HPC will be g-h-i-j. The cycle continues to operate till the evaporator pressure falls below the set pressure and reverses the refrigerant flow and changes the seawater flow path. Evaporator pressure reduces as ice layer builds in the tubes carrying seawater. This increases the thermal resistance. After reduction of evaporator pressure below preset value, the reversing of the refrigerant valves V1 and V2 occurs which changes the flow directions of seawater and refrigerant. This enables the interchange of function of the two LHEs. LHE2 now becomes evaporator and LHE1 is the condenser. The seawater flow is also directed to LHE2. The refrigerant flow across LS, EC and HPC, however, remains the same. As the LPC feeds refrigerant to LHE1, the ice which was formed in previous cycle melts and seawater flowing through LHE2 partially freezes. Heat of freezing of water is picked up by the evaporating refrigerant. Thus, both the processes of freezing of water and meting of ice occurs concurrently in the two LHEs. This cycle further continues to operate till evaporator pressure, that is, refrigerant pressure in LHE2 falls below the set pressure. Refrigerant flow path becomes a-b-c'-d'-e-f-k-c-d-a, while the flow through EC-HPC-LS remains same. Similarly, the flow path of seawater and water becomes 1-2-3'-5-6-8-10 and 4-7-8 respectively. The seawater flow is thus always directed to evaporator of respective cycle, where freezing occurs. Calculation of specific power consumption for 1000 kg/h water production Assuming that seawater concentration is doubled as it passes over freeze concentrating surface, a feed flow rate, mfsw, of 2000 kg/h is required for generating 1000 kg/h of potable water, mfw. Assuming a minimum approach of 3 °C in heat exchangers and neglecting variation of specific heat of seawater with concentration and temperature, the following calculations are presented. As the freezing point of sea water at 7.0% salt concentration is −4 °C, the operating temperatures of refrigerant in evaporator will be −7 °C, while melting of ice at 0 °C required condensing temperature of 3 °C. Carnot COP of 26.7 is calculated for −7 °C evaporator and 3 °C condensing temperature. The performance of commercially available refrigeration system for condensing temperature higher than 30 °C revealed that its Carnot efficiency with R22 as the refrigerant varies from 0.4 to 0.6 depending on lift and capacity. Carnot efficiency is defined as the ratio of actual COP to Carnot COP of refrigeration cycle. Assuming Carnot efficiency of 0.6, actual COP of 15.92 is possible for compressor operating between −7 °C and 3 °C. The pre-cooling of seawater in
Q latent:FCS ¼mf w × hlatf = 93:06 kWhc = m3
ð1Þ
ð2Þ
where, hsw enthalpy of seawater at respective concentration, kJ/kg hlat_f latent heat of fusion of water, kJ/kg. Thus, heat load on evaporator is 99.34 kWhc/m3. Assuming heat gain from ambient to FCS system is 5% of total heat load, total cooling load on the evaporator will be 104.3 kWhc/m3. Taking COP of cycle as 15.92, specific power consumption is 6.55 kWhe/m3 of water, if LHE condenser alone rejects all heat gained in the system. But in actual, LHE condenser can reject only part of total heat load equivalent to sensible heating and melting heat load of ice. The excess heat contributed by compressor power input and heat gain from ambient, has to reject to atmosphere through an external condenser, EC. Assuming condensing temperature of EC as 36 °C, the net COP of freeze concentration system COPnet.FCS, has been analysed in following section. The Sankey diagram representing heat distribution in two-stage freeze concentration system for 1 m3/h of water production is shown in Fig. 5. As per the calculation in preceding section total heat to be rejected in condenser is 110.8 kWhh/m3. The latent heat condenser can reject only 96.2 kWhh/m3 heat, which is equivalent to sensible cooling of ice from −4 °C to 0 °C and the melting heat load. The remaining heat 14.66 kWhh/m3 is rejected in external condenser, which is operating at 36 °C. Assuming COP of 3.6 for pumping the excess heat from intermediate temperature to the ambient heat sink, high pressure compressor power will be 4.07 kWhe/m3. The total heat rejected in external condenser, QEC, is 14.66 + 4.07 = 18.73 kWhh/m3. This results in total power consumption by both compressors, Wtot.cmp, of 10.62 kWhe/m3 and net cycle COP, COPnet.FCS, of 9.82. The effect of variation in Carnot efficiency of low pressure compressor, ηcar.LPC, on various system parameters is given in Fig. 6. Comparing parallel compression (Rane and Jabade, 2005) and two stage compression systems, it was found that overall specific energy consumption of both systems was almost same at ηcar.LPC of 0.45 to 0.5. However, possibility of higher Carnot efficiency for LPC, ηcar.LPC, above 0.5 will reduce overall energy consumption of proposed system. This is because, in parallel compression system LPC picks up only part of evaporator heat load, equivalent to rejected heat for melting of ice in LHE condenser and thus has lower rating. While in two-stage system, LPC takes total evaporator heat load irrespective of LHE condenser capacity, and thus has higher power consumption. Thus a lower Carnot efficiency of LPC has a greater impact in a two-stage system. On the other hand HPC in parallel compression system picks up balance heat from evaporator and rejects it to EC, and thus has higher lift and power consumption. While in two stage system, HPC operates between intermediate pressure corresponding to LHE condenser temperature and EC pressure and thus has lower lift and power consumption. The overall power consumption of both systems is thus almost the same. However, proposed two-stage system has advantages of lower heat rejection rate at external condenser and is compact. The system thus has potential to provide water at specific energy consumption, 10.62 kWhe/m3, and without additional cost of pretreatment as in distillation or membrane processes. Heat exchanger
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tubes could be of suitable material such as Cu–Ni, aluminium alloy or epoxy coated steel. Diameter and thickness of the tubes could be in the range of 12.7 to 25.4 mm and the wall thickness of the order of 1 to 2 mm. Small system using 1 TR capacity compressor has been conceived and will be developed in near future.
Conclusions
Fig. 5. Sankey diagram: Two stages heat pump operated freeze concentration system for 1000 kg/h of water production.
Desalination of seawater as a source of potable water is likely to continue in the future as available pure water sources are reducing. Conventional desalination technologies such as reverse osmosis, multiple effect evaporator, multistage flashing and mechanical vapour compression have equipment scaling related issues which increase the operational and maintenance cost. Freeze desalination technique operates at low temperature and has less scaling problems. Conventional freezing methods include direct, indirect or vacuum freezing techniques. Each of these methods has operational difficulties like ice– seawater separation and loss of pure water. Further as direct and vacuum technologies are operated at lower pressure, their operational difficulties are more. Indirect freezing has the disadvantage of more power consumption due to scraper operation in scraped surface heat exchanger, while in fluidised bed heat exchanger ice removal rate is low. Novel freeze desalination systems coupled with a heat pump and using patented vented-double-wall tube–tube heat exchanger (TT_HE) was developed. System utilises two stage compression processes and optimally utilises latent heat of ice melting for refrigerant condensation. Since the melting of ice occurs in absence of seawater, the need for separation of ice–seawater is eliminated. The use of TT_HE eliminates accidental mixing of seawater, refrigerant and pure water. It allows moderate velocity of seawater flow, which reduces salt inclusion rate, fouling rate and improves compactness of system. Washing requirement and melting loss are further reduced because of increased ice layer thickness and selective flow of seawater. Two-stage compression with automated control valve enables this system to be deployed for a wide range of capacities. System offers significant initial and operating cost saving when compared with conventional freeze desalination systems. The expected energy consumption for the two stage compression system is 9 to 11 kWhe/m3 of water produced with a cycle COP ranging from 8 to 12.
Fig. 6. Variation in heat pump operated freeze concentration system parameters with Carnot efficiency of low pressure compressor.
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