The Journal of Macrotrends in Energy and Sustainability Thermodynamic Vapor Cycles for Converting Low- to Medium-Grade Heat To

Bahram Saadatfar, Reza Fakhrai and Torsten Fransson, JMES Vol 2 Issue 1 2014

The Journal of MacroTrends in Energy and Sustainability

MACROJOURNALS

Thermodynamic Vapor Cycles for Converting Low- to Medium-grade Heat to Power: A State-of-the- art Review and Future Research Pathways

Bahram Saadatfar, Reza Fakhrai, Torsten Fransson Department of Energy Technology, Royal Institute of Technology KTH, Stockholm, Sweden

Abstract The moderate temperature of renewable energy sources and vast amounts of industrial waste heat demands implementing other thermodynamic heat to power cycles than the conventional power cycle. Various thermodynamic cycles for converting low-grade heat into power have been developed, in which the major ones are: zeotropic vapor cycle, organic Rankine cycle, Trilateral Flash cycle, supercritical Rankine cycle. This study serves the state-of-the-art review of thermodynamic vapor cycles for converting the low-grade heat into electrical power and present recommendations for future research and development to advance the conversion of low- to medium-grade heat to power.

Keywords: Thermodynamic vapor cycle, Low-grade heat, Heat to power

1. Introduction

As more importance is focused on sourcing sustainable energy and increasing energy efficiency by industries, the amount of usage of low-grade heat will increase. Generating electricity from renewable energy sources is becoming an inevitable necessity [1]. Thermoelectric devices and heat engines are two methods for converting thermal energy to electricity [2]. Thermoelectric devices are small with low environmental impact; however, they have lower efficiency and also more costly in comparison with heat engines [3]. Thermoelectrics are more limited to very special applications such as vehicle heat recovery.

Mechanical heat engines have been used for conversion of thermal energy to power since many decades. A heat engine is a device which converts the thermal energy into mechanical work. The

Bahram Saadatfar, Reza Fakhrai and Torsten Fransson, JMES Vol 2 Issue 1 2014

energy conversion is driven by the temperature gradient between hot and cold reservoir. There are many well-known mechanical heat engines: vapor cycles, Otto cycle, Bryton cycle, Diesel cycle, and Stirling. The effort is to introduce advanced power vapor cycle configurations to reach higher cycle efficiencies. Vapor cycles can be utilized low and medium temperature energy resources such as geothermal [4], solar thermal [5], biomass [6] and industrial waste heat (main characteristics of typical heat source and heat sink are listed in Table 1 [7,8]). Vapor cycles can also be used in a combined cycle to reach higher efficiency [9,10].

2. Vapor power cycles for finite thermal sources

This section starts with a short review on a simple steam Rankine cycle as a conventional and practical vapor power cycle. Then it is given a review of advanced vapor power cycles such as the organic Rankine cycle (ORC), the Kalina cycle, and flash cycle (FC). These advanced cycles are often used for conversion of low to the medium heat source to power. These heat sources include solar thermal, geothermal energy, waste heat sources from industrial process and exhaust gases.

2.1 Steam Rankine cycle

The traditional approach to generate electricity from heat is utilizing water as a working fluid in Rankine cycle and steam turbines [11]. The most part of the world’s electricity is generated by steam Rankine cycle (SRC). A schematic diagram of the steam Rankine power plant and the temperature vs. entropy diagram are shown in Fig. 1. It consists of four main components: boiler, steam turbine, condenser, and feed water pump. In an ideal cycle, saturated water pumped to a high pressure, then superheated steam at state 3. The superheated steam expands through the turbine and generates power and after that condensed to saturated water. The efficiency of such a system improves by reheating and utilizing feedwater components. Large steam Rankine plants operates near 550°C with efficiencies around 30% [12].

Steam Rankine cycle is also employed in combined cycles to recover waste heat from high temperature gas turbine exhaust. This is done by the Heat Recovery Steam Generator (HSRG) unit. Depending on the gas inlet temperature and the amount of recoverable heat, using single or multiple-pressure HSRG is feasible [13].

Fig. 1. (a) Simple diagram of steam Rankine cycle (b) T-s diagram of the simple steam Rankine cycle

Bahram Saadatfar, Reza Fakhrai and Torsten Fransson, JMES Vol 2 Issue 1 2014

There are some disadvantages that limit the SRC. As an example, in condenser, the pressure of steam needs to be brought down to less than atmospheric pressure, which means demanding the desecrators unit to protect the air leak into a system in the last stage of turbine and condenser. The low-pressure steam turbine due to a large volumetric flow ratio may be inefficient [14]. Utilizing molecularly heavier fluid as a working fluid can result in higher isentropic efficiencies around 60%. Moreover, heavier organic fluid has vapor pressure above the atmospheric [15]. Moisture droplet formation in the last stage of steam turbine results in superheat working condition or reinforcing blade material in wet turbine to avoid erosion [16]. Capital cost will be greater for superheating system, due to the lower vapor heat transfer coefficients, and mostly necessities of multi-stage reheating and expansion.

For the temperature range of available renewable thermal energy resources, vapor cycles using water shows to be a poor working fluid for the conversion system [17], thus several potential working fluids so called organic fluids.

2.2 Organic Rankine cycle

The principle of the organic Rankine cycle (ORC) is the same as steam Rankine cycle, but instead using organic working fluids with low boiling points for recovering heat from low-temperature heat source. Ringler [18] studied water, ethanol and toluene, by comparing some physical properties such as latent heat, thermal stability, and freezing temperature as well as environment aspects. The study shows that water better suits for the high temperature waste heat recovery and organic fluids are superior for the recovery of low-grade waste heat recovery.

Pure organic fluids such as HCFC123 [19], HFC-254fa [20], isobutene [21], n-pentane [22] and aromatic hydrocarbons [23] have been examined by authors [24,25]. Tchanche et al. [26] examined working fluids for a low temperature solar ORC below 90°C and shows that R134a and R152a are the most suitable fluids. Chen et al. [27], and Tchache et al. [28], and Hung et al. [15], also have published reviews on ORCs for a number of different low temperature energy sources. Mixtures for ORC are also had been studied in some literatures [24,29]. Papadopoulos et al. [17,30] proposed a computer-aided-molecular-design technique to generate the general molecular structure for optimal ORC working fluids. Comparison of fluid's properties in SRC and ORC is summarized in Table 2. A list of common working fluids in commercial installations is listed in Table 3 [31], and a short list of installed ORC waste heat recovery plants is available in Table 4 [32,33]. Schematic configuration of an ORC system using pure fluid, and its T-s diagram is presented in Fig. 2.

The organic working fluids have many different physical properties from water. The slope of the saturation curve in T-s diagram ( dT/ ds ) is one of the main point in categorizing working fluids. The T-s slope for water is positive ( ds/0 dT ). However, for many organic fluids this diagram would be negative like R22 (  0 ) or infinite like R11 (  0 ).

Bahram Saadatfar, Reza Fakhrai and Torsten Fransson, JMES Vol 2 Issue 1 2014

Fig. 2 (a) schematic diagram of ORC cycle (b) T-s diagram for ORC cycle

Ideally, an “isentropic” working fluid would be the best because of not requiring reinforcing material for the turbine blades as well as not having substantial superheat at the turbine exit. Wet fluids usually need to heat up to superheat state. Tabor and Bronicki [34] determined the theoretical relationship for the slope of the saturated vapor curve on the temperature-entropy (T-S) diagram using Maxwell relations and the Clausius-Clapayron equation with Eq. (1), and showed that more complex molecules show a more “drying” behavior with a less positively sloped saturated vapor curve on a T-s diagram. Where, s is entropy, C p is constant pressure specific heat,  is constant volume, H is vaporization heat, V is vaporization volume change, and the subscript  represents the saturated vapor curve.

1 T  (1)  s CP V H   TTTVP 

The turbine built for ORCs normally requires a single-stage expander, which means a simpler, lower capital costs and less maintenance [35,36]. Some important advantages of the ORC over conventional steam power plant are:  Less heat for the evaporation process.  The evaporation process occurs at lower pressure and temperature.  The expansion process after the expander ends in the vapor region. Hence there is no need to superheating to prevent the risk of blades erosion.  The smaller pressure drop and ratio will be much smaller and thus simple single stage turbines can be used.

The thermal energy source and working fluid have temperature difference. This temperature difference inherently results in irreversibility [37]. If the thermal source is a single phase with a near linear temperature change profile along the heat exchanger, a temperature mismatching occurs. Temperature mismatching reduces heat exchanger effectiveness, and destroys exergy and results in pinch point. In Fig. 3 (a) [38,39], shows the temperature profile of an ORC using pure working fluid. It is shown that for a vapor power cycle using a pure working fluid, the temperature profile increases linear, then phase change occurred in constant temperature and again, there is a linear increase.

Bahram Saadatfar, Reza Fakhrai and Torsten Fransson, JMES Vol 2 Issue 1 2014

The area in the plot between thermal source and working fluid can be qualitatively shown the exergy destruction. Guo et al. [36] analyzed the thermal efficiency as well as exergy efficiency of the organic Rankine cycle based on the locations of heat transfer pinch point in the evaporator. Li et al. [40] studied Influence of coupled pinch point temperature difference on performance of ORC. Although some have proposed doing the heat addition process in multi pressure steps, but installing additional heat exchangers is not economical for low grade heat sources [41].

Fig. 3. Stream temperatures variation during heat addition for (a) single component working fluid ORC, (b) zeotropic cycle [38]

A number of solutions such as utilizing unconventional working fluids [42,43] and modifying cycle configuration [44] have been proposed to minimize temperature mismatching. Several cycle configurations have also been studied for ORC applications. Different types of ORC with subcooling or superheating are studied by Liu et al. [45] and showed that superheating and subcooling had negative ORC efficiency. Desideri and Bidini [46] deliberate different types of Rankine cycles for geothermal heat recovery. They results show that regenerated Rankine cycle is a promising solution for low temperature liquid geothermal source. Roy et al. [47] investigated the performance of an ORC with superheating, for, R123, R12, R717, and R134a. The efficiency of a basic ORC with a regenerative ORC was compared by Dai et al. [48] for 10 fluids including R236ea and R245ca.

2.2.1 Improvement of the organic Rankine cycle

The ORC cycle can be improved by using a regenerator in case of using dry fluids. The expander exit state of the dry working fluid is not reached the two phase state, and temperature at this point is higher than the condensing temperature. This higher temperature fluid can be utilized to preheat the liquid before it enters the evaporator Fig. 4 [50]. Heat transferred between the expander outlet and the evaporator inlet in a counter-flow heat exchanger. The results would be higher efficiency. Drescher and Brüggemann [49] examined simple and recuperated cycles with different source and sink temperatures.

Besides the effect of the working fluid, another key aspect affecting the design of best ORC system is the thermodynamic cycle arrangement. The basic cycle can be upgraded, from the simple ORC concept, in order to increase the heat recovery potential. Heat recuperation, superheating, two-pressure level and supercritical cycles are some possible methods [50]. A few of these strategies are already implemented in commercial machines, while others are under

Bahram Saadatfar, Reza Fakhrai and Torsten Fransson, JMES Vol 2 Issue 1 2014

investigation. Bao et al. [51] studied autocascade cycles with exergy analysis. Regenerative and simple cycle is compared by Mago et al. [52]. Moreover, dual loop cycle performance of ORC are studied by Zhang et al. [53].

Fig. 4 Different ORC layout [50]

2.3 Zeotropic vapor cycles

Zeotropic mixture has a property named “temperature glide” in isobaric evaporation and condensation process. Because of the temperature difference they can reduce exergy destruction rate [54]. Another advantage of using zeotropic mixtures is that the power cycle can be operated at a relatively low pressure. Zeotropic mixtures are characterized by non- isothermal phase transitions at a constant pressure. It can be seen from Fig. 3(b) that zeotropic working fluid’s temperature glide follows close to the temperature profile of the thermal source compared with the ORC’s in Fig. 3 (a). This shows a reduction in the heat addition irreversibility [55,56].

Even though the temperature matching can be improved with zeotropic mixtures in vapor cycles, the disadvantage is that the mixtures always have a lower convective heat transfer coefficient in comparison with pure components for a given temperature difference because of the “diffusion resistance” to heat transfer for mixtures [57].

Bahram Saadatfar, Reza Fakhrai and Torsten Fransson, JMES Vol 2 Issue 1 2014

A few authors have examined mixtures of natural hydrocarbons and fluorinated hydrocarbons as a working fluid in ORC. Demuth [58] studied geothermal power plant with two-component mixtures of natural hydrocarbons for working at 137 °C and 182 °C geothermal water source. It is reported that subcritical cycle efficiency increases up to 14% in comparison with pure fluid propane. For a low temperature solar, Wang and Zhao [59] studied turbine design with three mass fractions of R245fa/R152a at fixed evaporation and condensation temperature. Comparing to pure R245fa the mixtures showed a lower volume flow rate at the inlet of the expander; in addition, a lower volume flow ratio for the expansion can be seen. These mean a reduction of turbine dimensions and costs [60]. Chys et al. [61] examined the effect of using mixtures as working fluids in ORC with a simulation model of the cycle including all elements of an actual installation. In the next sections, a brief review of three other zeotropic vapor cycles is described.

2.3.1 The Kalina cycle

In early1980s, Kalina introduced a new thermodynamic power cycle using an ammonia–water mixture as a working fluid and was named Kalina cycle [62]. The Kalina cycle is a modified Rankine cycle consisting proprietary designs for using the ammonia-water working fluid. The scheme of a typical Kalina cycle is shown in Fig. 5. Intermediate concentration of ammonia in state 1 is pumped to state 2, and then heated in a recuperator and heat up by brine in evaporator. In state 5, the liquid and vapor part is split. The vapor phase in state 6, send to the expander and mixed with liquid part. This is cooled in recuperator and condenser.

Theoretically, the Kalina cycle can convert around 45% heat to power, and up to 52% in gas turbine combined cycle plant; this compares with 35% and 45% for steam cycle [63]. The Kalina cycle cycles show about 32% more power in the industrial waste heat application comparison with an ORC. Nonetheless, the Kalina cycle in small biomass direct-fired based cogeneration do not have better performance than ORC. The adoption of the Kalina cycle to heat source sink is one of the advantages over the ORC cycle, as the ammonia–water composition and the system high and low pressure levels can be adjusted [64].

Fig. 5. (a) Simple diagram (b) temperature-heat diagram of the Kalina cycle

Bahram Saadatfar, Reza Fakhrai and Torsten Fransson, JMES Vol 2 Issue 1 2014

The more application of the Kalina cycle is limited to medium–low temperatures heat sources (300–400 °C in heat recovery, and 100–120 °C in the binary geothermal plants) and to small power conversion systems [65–68]. The results for evaluating the impact of an ammonia–water environment on the life expectancy of traditional power plant materials in the Canoga Park demonstration plant indicate that for turbine throttle operating temperatures of up to 540 °C, traditional materials of construction are acceptable. Nevertheless, at more than 400 °C using ammonia–water mixture is corrosive and not advisable [69,70]. Whittaker studied the corrosion effect at the Kalina cycle geothermal power plant in Husavik, Iceland. The analysis states that mild steel and aluminum seem to be unsuitable materials but several stainless steels (304, 316) are appropriate for Kalina Cycle Systems [69,71].

Numerous derivatives of the Kalina cycle have been designed for specific applications such as geothermal energy, solar thermal energy, waste heat recovery, and combined cycle systems. The Kalina cycle requires precision for the mixing and separation processes, and small pinch temperatures, which mean larger heat exchangers. Presently, the Kalina cycle is second only to the ORC in terms of popularity in real application. Some installed units are listed in Table 5 [72].

2.3.2 The Maloney-Robertson cycle

The ammonia–water mixture as a working fluid for power production was first presented by Maloney and Robertson [73] and then enhanced by Kalina. The schematic diagram of the system is shown in Fig. 6. The rich ammonia from the heater sent to the super heater (state 7). The superheated vapor expanded in a turbine, and then mixed with the weak solution from distillation unit (state 11) and is used to absorb the rich vapor in ammonia to regenerate the base solution (state 1). The Maloney-Robertson cycle was less complex than the Kalina cycle; however, it used the mixture composition in the volatile component prior to heat rejection. Ibrahim and Klein [74] and Park and Sonntag [75] studied the Kalina and showed that since during the condensation process in the Kalina cycle the heat exchanges with the environment, there is a limitation on the working fluid temperature exiting the turbine. An absorption condensation process can be employed to avoid this limitation.

Fig. 6 Schematic diagram of the Maloney Robertson cycle

Bahram Saadatfar, Reza Fakhrai and Torsten Fransson, JMES Vol 2 Issue 1 2014

2.3.3 Combined power and cooling cycles

A modified ammonia–water mixture combined power and cooling cycle was presented by Goswami [76] and is shown in Fig. 7 [77]. Vijayaraghavan and Goswami [78] proposed modified equations for both thermal and exergy efficiency based on a cascade analysis for cooling and power production. Also, Martin and Goswami [79] investigated a theoretical and experimental study of cooling production of cycle. The results indicated that it is possible to have cooling but high turbine efficiencies should be reached.

Fig. 7 Schematic diagram of combined cooling and power cycle. a) Internal cooling source. b) external cooling source [77].

2.3.4 Uehara cycle

Different cycle for ocean thermal energy conversion (OTEC) studied by Uehara et al. [80,81]. They investigated the main components of an OTEC plant. The simulation results was done with 26C hot water and 4C cold water for 100MW OTEC system, and reported that R717 as one of the appropriate working fluid. The cycle is an improved Kalina cycle with adding second turbine, a heater and after condenser. A simple diagram of the Uehara cycle is shown in Fig. 8. The working fluid in Uehara cycle is ammonia-water mixture. The warm seawater heat up the mixture and turn into vapor-liquid mixture, then the vapor separated from liquid and sends to first turbine. After passing the heater the vapor directed to second turbine. The mixed vapor from the second turbine is absorbed with ammonia water in the absorber. The unabsorbed part is condensed to liquid by the cold seawater.

Bahram Saadatfar, Reza Fakhrai and Torsten Fransson, JMES Vol 2 Issue 1 2014

Fig. 8 Simple diagram of Uehara cycle

There are number of theoretical investigations on OTEC system [82,83]. Ganic and Moeller [84] analyzed the ratio of the total heat transfer area to the net power output of the OTEC system for a 1-MW OTEC with R717 operated in a simple closed Rankine cycle. Faizal and ahmed [85] examined an experimental study on a new designed closed-cycle with R134-a as a working fluid and achieved a maximum efficiency of about 1.5% in the system. Sun et al. [86] optimized the design for an ORC in OTEC with R717 and R134a by numerical work.

2.4 Transcritical and supercritical cycles

The Kalina cycle practically is complex and needs more maintenance [87]. Another method to avoid the temperature mismatching in the constant temperature phase change region, could be transferring heat to the working fluid at pressures above the critical pressure. A cycle that operates partly under supercritical conditions is known as a transcritical cycle, whereas a cycle operating completely in the supercritical regime is known as a supercritical cycle. Carbon dioxide, helium, refrigerants, and alkanes are some suggested working fluid for these cycles. Compared with a ORC, the working fluid in a Transcritical Rankine cycle (TRC) can be heated to the supercritical state, which results in reducing the exergy destruction in the heating process due to better thermal match [88]. The improvement of temperature matching for a TRC is illustrated in Fig. 9 (a) [38].

Fig. 9 Stream temperatures variation during heat addition for (a) transcritical cycle (b) organic flash cycle [38]

Bahram Saadatfar, Reza Fakhrai and Torsten Fransson, JMES Vol 2 Issue 1 2014

Some research has been conducted on TRCs as an effective method to recover low-grade heat. Gu and Sato [89] studied a supercritical power cycle with the aim of maximizing thermal efficiency by optimization of temperature or pressure in condenser. Vetter et al. [90] compared the sub-critical and supercritical processes with different refrigerants and showed that propane or R143a can have specific net power output more than 40% at a 150°C geothermal source. Zhang et al. [91] investigated the highest thermal efficiency and exergy efficiency of R123 and R125 in subcritical and transcritical ORC, and showed that R125 has a higher thermal and exergy efficiency compared with R123. Mikielewicz [92] showed that transcritical power cycles results around 5% higher efficiency than of subcritical ORC cycles but with a bigger vapor generator size. Gawlik [93] made an analysis for low- to moderate-temperature binary geothermal power plants and transcritical power cycles. Karellas et al. [88] examined the performance of transcritical and subcritical Rankine cycles for R227ea and R134a at turbine inlet from 105°C to 135 °C. Schuster et al. [94] studied the work output, thermal efficiency, and heat transfer capacity of water, R236fa, R245fa, and isobutene, in transcritical and subcritical Rankine cycles with a heat source at 210 °C. Khennich and Galanis [95] investigated the performance of a subcritical Rankine cycle in comparison with superheating, 100 °C heat source and a 10 °C heat sink, for five working fluids. Their results show two optimum evaporation temperatures. The first minimizes the total thermal conductance of two heat exchangers, and the other maximizes the net power output.

The pump efficiency is a key parameter in transcritical cycles. Quoilin [96] reported a 22% efficiency on a diaphragm pump with RHFC-245fa as a working fluid. Bala et al. [97] examined the overall efficiency of sliding-vane refrigerant pumps with different working fluids. They reported about 20% as the highest reported efficiency. As a theoretical point of view the transcritical cycle has merit; however, the design of a supercritical turbine for fluids other than water is still in the developmental phases [98]. Until now, Carbon dioxide (CO2) is the most suggested working fluid for supercritical cycle utilizing finite thermal energy sources.

2.4.1 CO2 transcritical cycle

Carbon dioxide as an environmentally-friendly natural working fluid with zero ozone depletion potential (ODP), non-flammable, non-toxic has been widely studied as the working fluid in the supercritical or transcritical Rankine cycles. However, a disadvantage of CO2 is its low critical point of 31.18 °C and its potential effect on the condensation process [99]. In addition, high critical pressure of CO2, which is as high as 7.38 MPa, is another disadvantage. The basic solar- driven carbon dioxide transcritical power system consists of four main components: a solar collector, an expansion machine, a condenser and a pump is shown in Fig. 10(a) [100]. Fig. 10 (b) [100], represents the corresponding cycle for the carbon dioxide transcritical power system consisting these four processes.

Bahram Saadatfar, Reza Fakhrai and Torsten Fransson, JMES Vol 2 Issue 1 2014

Fig. 10 (a) Solar transcritical CO2 power system. (b) T-s chart for CO2 transcritical power cycle [100]

Vélez et al. [101] considered a CO2 transcritical power cycle and presented an increase up to 25% for the exergy efficiency, and up to 300% for the energy efficiency with rising the turbine inlet temperature from 60 to 150 °C. Sarkar et al. [102] evaluated the different operating conditions and cycle performance on the supercritical CO2 recompression cycle. Bıyıkoğlu and Yalçınkaya [103] conducted a parameter study on a geothermal supercritical condition power cycle using CO2 under different reservoir conditions. Song et al. [100] examined a solar transcritical CO2 power cycle driven and disclosed that there is an optimum turbine inlet pressure where the system efficiency and the net power output both have maximum values. They [104] also made the dynamic simulation of a solar-driven CO2 transcritical power system by simulating both daily and yearly performances of the system. Some experiments have been done to examine the numerical simulation and heat transfer characteristics of supercritical CO2 in solar collectors. Cayer et al. [105] studied a thermodynamic analysis of a CO2 transcritical low-grade heat power cycle and assessed the effect of the pressure on the cycle performance. Also, Cayer et al. [106] examined the a parametric optimization using six performance indicators: thermal efficiency, specific net output, exergetic efficiency, total UA and heat exchangers’ surface as well as cost of the system. Baik et al. [107] compared the power production for R125 and CO2 transcritical cycles for a low grade heat source. Lakew et al. [108] introduced a new approach to improve the performance of carbon dioxide Rankine cycle by using thermal driven pump. Chacartegui et al. [109] proposed supercritical and transcritical CO2 cycles for solar energy applications and used a stand-alone closed cycle and a topping CO2 gas turbine and a bottoming ORC.

Although the CO2 transcritical cycle has been promising potential, the design, construction, and implementation of an inexpensive and reliable turbine will ultimately limit its practical value.

2.5 Trilateral cycle

The trilateral cycle (TLC) proposed by Smith et al. [110] is another offered cycle to improve temperature matching (Fig. 11 (b)) operating at reasonable pressures, faces similar challenges to that of the transcritical cycle. Their examined the R134a as the working fluid and a screw expander. The saturated liquid is flashed into two phase, then, the resulted vapor–liquid mixture send to condenser and after that the liquid is pumped to high pressures and heated up

Bahram Saadatfar, Reza Fakhrai and Torsten Fransson, JMES Vol 2 Issue 1 2014

to the saturation temperature [111]. The schematic configuration of the TLC and its T-s diagram is shown in Fig. 11 [112].

The advantage of the TLC is the more efficient heat transfer between the heat source and working fluid. Löffler [113] reported that TLC - systems shows efficiencies around 50%–100% higher than those of ORC - systems. A model for comparison of TLC - and ORC - systems have been introduced by Zamfirescu and Dincer [111]. They used this model with an inlet heat source of 150 °C and a cooling air of 25 °C for ammonia-water mixtures and for the ORC working fluids were pure R141b, R123, R245ca, and R21. In another work, Fischer [114] reported that exergy efficiency for power production is higher by 14%–29% for the TLC with the two-phase expander utilization compare with the ORC.

Similar to the transcritical cycle, designing of an efficient and reliable two-phase expander is under research; however recent progresses are being made for screw expander.

Fig. 11 (a) Configuration of a TLC- system (b) The Trilateral cycle in the T-s diagram [112]

2.6 Organic Flash Cycle (OFC)

To avoid the requirement of a two-phase expander, the liquid and vapor components of the mixture can be separated, and the vapor sends to expander. In the trilateral flash cycle, heat adds to the working fluid in a single-phase liquid, thus there is no isothermal phase change as shown in Fig. 9(b). This cycle is similar to the flash steam cycle that utilizes high temperature and pressures liquid state geofluid extracted from a geothermal well. The extracted liquid geofluid is throttled to a lower pressure or flashed to produce a liquid–vapor mixture [38,115]. The important disadvantage of the steam flash cycle is that since the water is a ”wet” fluid, the steam leaving expander contains a significant amount of moisture [59]. Isentropic expansion of a “wet” fluid from its saturated vapor state creates a two-phase mixture with liquid droplets forming. Even though large steam turbines regularly have isentropic efficiencies of around 80% to 90%, saturated steam cycles still require wet steam turbines. Wet steam turbines are expensive due to need reinforcing materials for blades to protect from erosion make by the liquid droplets [115]. Turbine cost is reduced for the OFC because there is no need for blade reinforcing. The OFC layout and its T-s diagram are shown in Fig. 12 [116]. The effectiveness of

Bahram Saadatfar, Reza Fakhrai and Torsten Fransson, JMES Vol 2 Issue 1 2014

the OFC on waste thermal energy recovery at 300 °C by using isentropic and dry fluid is studied by Ho et al. [38].

Fig. 12 (a) OFC layout (b) T–s diagram of the OFC [116]

There are different modifications to enhance the efficiency of the OFC systems. The double flash OFC, split the flash evaporation cycle into two steps, and then more fluid is vaporized results in more power production [116]. Presently reported available isentropic efficiency for radial inflow turbine around 70% in two phase regime [117], allows potentially replace throttling valve, and achieving higher efficiency with designing two-phase OFC. A twin-screw machine for two-phase expansion with isentropic efficiencies of about 70% is tested by Smith et al [118].

It was presented that OFC showed higher heat addition exergetic efficiency comparing the other vapor cycles, but on the other hand the throttling process showed a drawback. The OFC also potentially offers exploiting thermal energy sources at lower temperatures.

3. Future directions in Low-Temperature waste heat recovery

Directly transfer heat from the waste heat source to the evaporator without an intermediate heat transfer loop is one of the current research areas. This will reduce the cost around 15% and thermal mass of the system. It can reach the 20%-40% increase in ORC efficiency [119]. However some issues surrounding the selection of an ORC working fluid such as safety and flammability. In addition the constraints imposed on the thermodynamic analysis and heat exchanger design also important.

One of the waste heat sources is exhaust gases. However, low temperature recovery from gas streams containing chemicals becomes increasingly challenging. These waste heat sources will have limitations for cooling flue gases to low temperatures. Hence, developing of heat exchangers for working in such corrosive environments is under research and development. Another area for research is heat exchangers with larger heat transfer coefficients which help to reduce the heat transfer area. Some examples of available heat technology are ceramic inserts used in radiant heating tubes and finned tubes.

Bahram Saadatfar, Reza Fakhrai and Torsten Fransson, JMES Vol 2 Issue 1 2014

The most crucial selection for any vapor cycle is the working fluid with which it operates. Considerable development is into examining such aspects as favorable working fluid selection criteria, the properties of fluid mixtures and the predictive modeling of fluid behavior. Some researchers focus on turbine optimization, including real-gas effects, particularly close to the critical point, and developing accurate equations of states [120,121]. Another research area focused on the best design for expander to be used in vapor power cycle with the attention on correct matching between the working fluid employed, expander and the operating conditions in terms of inlet temperature and available thermal power. In addition two-phase expanders are very much in the research and development phase.

Conclusions

The major vapor power cycle and various derivative cycles were reviewed for low-grade heat conversion into power. Organic Rankine cycle has not a good thermal match with the heat source. The Kalina cycle was developed in order to replace Rankine Cycle as a bottoming cycle for a combined-cycle energy system as well as for generating electricity using low-temperature heat resources. The Kalina cycle has a family of configurations. The Maloney-Robertson cycle is less complex than the Kalina cycle. Another suggested method is operating the Rankine cycle partially in the supercritical regime with CO2 as a working fluid named transcritical CO2 Rankine cycle. However, there are obstacles, especially CO2 turbine, which is still under development. The organic flash cycle is based on the design principle of increasing utilization efficiency by increasing temperature matching and reducing exergy losses during the heat addition process. There is a significant potential to improve the organic vapor cycle for the possible applications to lower temperature thermal energy sources.

Acknowledgements

The authors would like to acknowledge the support of the KIC InnoEnergy from the European Institute of Innovation and Technology.

Abbreviations FC Flash Cycle GHG Greenhouse Gas IEA International Energy Agency ODP Ozone Depletion Potential ORC Organic Rankine Cycle RC Recuperator Cycle REG Regenerative Cycle S Entropy [J/kg] SH Superheat SRC Steam Rankine Cycle T Temperature [K] TLC The trilateral cycle

Bahram Saadatfar, Reza Fakhrai and Torsten Fransson, JMES Vol 2 Issue 1 2014

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Apendix 1.

profile Heat source Heat sink Waste Hear Recovery [122] Fluid (300°C ) Air (10°C)

CHP [8] Thermal oil (250°C - Water (70°C) 300°C)

Geothermal [7] Brine (100 °C) Air (10°C)

Table 1. Main characteristics of typical heat source and heat sink.

Bahram Saadatfar, Reza Fakhrai and Torsten Fransson, JMES Vol 2 Issue 1 2014

Water in steam Organic fluid in ORC Rankine cycle Critical pressure High Low

Critical temperature High Low

Specific heat High Low

Viscosity Low Relatively high

Flammability No Yes, depending on fluid

Toxicity No Yes

Environmental impact No High, depending on fluid

Availability Cheap Expensive

Table 2. Summary of working fluids properties comparison in steam and organic Rankine cycles

Bahram Saadatfar, Reza Fakhrai and Torsten Fransson, JMES Vol 2 Issue 1 2014

Working fluid Application HFC-134a Used in geothermal power plants or in very low temperature waste heat recovery.

HFC-245fa Low temperature working fluid, mainly used in waste heat recovery. n-pentane Used in the only commercial solar ORC power plant in Nevada.

Solkatherm Waste heat recovery

OMTS Biomass-CHP power plants

Table 3 Common working fluids in commercial ORC installations [31].

Bahram Saadatfar, Reza Fakhrai and Torsten Fransson, JMES Vol 2 Issue 1 2014

Location Waste heat Heat source Capacity Technology Start up generating system Mirom Roeselare, Waste Hot water (180 C) 3 MW Refrigerant/turboden 2008 (Belgium) incinerator plant

Oxon Italia SPA, Pavia (I), 8.3 MW MAN Exhaust gases 0.5 MW Siloxane/turboden 2008 Italy diesel engine

RHI, Radenthein (A), magnesite Hot exhaust gas 0.8 MW Siloxane/turboden 2009 Austria production

Nieuweroord, Netherlands engines Exhaust 150 kW ORC/triogen 2 × 835 kW Jenbacher biogas engines

Savona, BC, Canada Simple cycle Exhaust 4.5 MW Pentane 2008 gas turbine

Arizona [33] CSP Solar Solar 1 MW 2006

Hawaii Solar Solar 100 KWe Electratherm 2009

Table 4 Short list of ORC waste heat recovery plants [32,33]

Bahram Saadatfar, Reza Fakhrai and Torsten Fransson, JMES Vol 2 Issue 1 2014

Location Heat source Power output (MW) Cangoa Park (USA) Exhaust gas of gas turbine 3/6.5 (515°C), solar gas turbine

Fukuoka City (Japan) Waste heat from 5 incineration plant

Kashima steel works (Japan) Waste hot water (98°C) 3.1

Husavik (Iceland) Geothermal heat (124°C) 2

Unterhaching (Germany) Geothermal 3.4

Table 5 Kalina based power plants [72]