Categorization and analysis of heat sources for organic Rankine cycle systems

Renewable and Sustainable Energy Reviews 64 (2016) 790–805

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Renewable and Sustainable Energy Reviews journal homepage: www.elsevier.com/locate/rser

Categorization and analysis of heat sources for organic Rankine cycle systems Huixing Zhai a, Qingsong An b, Lin Shi a,n, Vincent Lemort c, Sylvain Quoilin c a Key Laboratory for Thermal Science and Power Engineering of Ministry of Education, Department of Thermal Engineering, Tsinghua University, Beijing 100084, China b Key Laboratory of Efficient Utilization of Low and Medium Grade Energy of Ministry of Education, Tianjin University, Tianjin 300072, China c Thermodynamics Laboratory, University of Liège, Campus du Sart Tilman, B49, B-4000 Liège, Belgium

art ic l e i nf o

a b s t r a c t

Article history: Received 12 January 2015 Received in revised form 21 April 2016 Accepted 28 June 2016

Organic Rankine cycles (ORC) are an effective way to convert low-medium temperature heat to electricity that cannot be used for conventional high-temperature Rankine cycles. Even though there has been many studies of ORC systems over the past few decades, ORC heat sources have received relatively little attention. The heat sources providing energy to the ORC have different characteristics that significantly impact the theoretical analyses and system designs of ORC systems. This paper gives a theoretical categorization of heat sources to give uniform boundary conditions for further theoretical studies on cycle choices and working fluid screening. The ideal cycles for each heat source type are analyzed. Performance metrics are recommended for different heat source ORC systems. The general characteristics of the different heat sources including waste heat from industrial and power systems and geothermal, solar and biomass sources are given with their influences on the ORC systems. Finally, the market characteristics for ORC systems using different heat sources are reviewed with suggestions for future developments of ORC systems. & 2016 Published by Elsevier Ltd.

Keywords: Heat source characteristic Thermodynamic categorization Performance metrics ORC Organic Rankine cycle

Contents 1. 2.

3.

4. 5.

Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 791 Characteristics and categorization of heat sources. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 791 2.1. Heat source characteristics . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 791 2.1.1. Heat source type . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 791 2.1.2. Heat source temperature . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 792 2.1.3. Heat source capacity . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 792 2.1.4. Heat source dynamics . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 793 2.1.5. Heat source related costs. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 793 2.2. Thermodynamic categorization. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 793 Ideal conversion cycles for various heat sources . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 795 3.1. Ideal cycle for “Type A” sources . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 795 3.2. Ideal cycle for “Type B” sources. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 796 3.3. Ideal cycle for “Type C” sources. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 796 Performance metrics . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 796 Typical heat source analyses and their influence on the ORC system. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 797 5.1. Waste heat from industrial processes. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 797 5.2. Waste heat from power generation systems . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 798 5.2.1. Internal combustion engines . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 798 5.2.2. Gas turbines . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 799

n

Corresponding author. Postal address: Department of Thermal Engineering, Tsinghua University, Haidian District, Beijing 100084, China. E-mail addresses: [email protected] (H. Zhai), [email protected] (Q. An), [email protected] (L. Shi), [email protected] (V. Lemort), [email protected] (S. Quoilin). http://dx.doi.org/10.1016/j.rser.2016.06.076 1364-0321/& 2016 Published by Elsevier Ltd.

H. Zhai et al. / Renewable and Sustainable Energy Reviews 64 (2016) 790–805

5.3. Geothermal . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5.4. Solar . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5.5. Biomass . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5.6. Multiple heat sources . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6. ORC site market economics. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7. Development directions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8. Conclusions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Acknowledgements . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

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799 800 800 801 801 802 802 802 803

1. Introduction

2. Characteristics and categorization of heat sources

As the energy crisis has intensified year after year, all countries are trying to make use of waste heat and renewable energy sources to generate electricity. However, these energy resources are widely scattered, have low temperatures and can be unstable, which complicated their efficient use. Low temperature sources have lower conversion efficiencies and possibility low commercial profitability. Technologies are needed to efficiently turn these kinds of heat sources into electricity. This paper focuses on 80– 350 °C heat sources with many researches focusing on using, organic Ranking cycles (ORC) to exploit these resources. ORC systems are suitable for the recovery of low-temperature heat by using low boiling point organic working fluids [1]. ORC power systems have simple configurations that easily scale down to low-capacity systems, have straightforward maintenance, and have good performance with varying working conditions [2,3]. Moreover, the ORC system effiencies can be improved in various ways, including the use of a recuperator, using working fluid mixtures, or using transcritical working conditions to minimize exergetic losses [4–7]. Various review articles [1,6–9] have given overviews of ORC cycle layouts, working fluid selection and system components. Most studies of ORC power systems have considered the technical constraints linked to the heat source, but only for specific cases [10–14]. Since the performance metrics also varies widely, the results of different studies are difficult to compare. Even though there has been many studies of ORC systems in the past few decades, the effect of the ORC heat sources have received relatively little attention. The heat source is, however, of great significance for ORC working fluid screening, cycle calculations and component design since different heat sources lead to different working conditions and, thus, different working fluids and cycle layouts. The heat source type also affects the choice of the performance metric (first and second law efficiencies, power output, overall efficiency, etc.). Some reviews have focused on ORC system with one kind of heat source such as internal combustion engines, solar energy or biomass which have provided valuable detailed information, but have not indicated low to categorize sources and no general guidelines regarding to the effect of different heat source types [15–18]. This paper focuses on the ORC heat sources to characterize the main heat source types as theorectical basis for ORC system selection. The ideal cycles for each heat source type are analyzed to guide the initial cycle layout and working fluid screening. Performance metrics are then recommended for different ORC applications. The heat source categorization and the recommended performance metrics set reasonable unified boundary conditions for future studies and enable objective comparisons. The effect of the heat sources characteristics on the ORC system are reviewed to provide insights into possible system designs. Finally, the market conditions for ORC systems using different heat sources are reviewed with indications on how the ORC market will develop in the future.

ORC are suitable for the recovery of low-temperature heat source by using low boiling point organic working fluids and where the system need miniaturization for its simple configuration as shown in Fig. 1. The typical low-medium temperature heat sources or miniaturization needed occasion include waste heat from industrial systems, internal combustion engines (ICE) and gas turbines, together with renewable energy resources like geothermal, solar (collector and pond) and biomass sources. Fig. 2 shows how the ORC systems can be coupled to these sources. Heat from the industrial process and ICE and gas turbines mostly comes from fossil fuels and their waste heat can be converted by ORC. Geothermal heat and solar heat are usually converted by ORC systems through direct or indirect heat transfer systems. Biomass can be converted through direct combustion with the heat being directly transferred to the ORC. Other possibilities for biomass conversion include gasification and pyrolysis but these are more expensive [19]. Eventhough the low-temperature heat comes from different sources, it has common parameters that can be analyzed and can be categorized according to these parameters. 2.1. Heat source characteristics The main heat source parameters include the type (open/ closed), the temperature, the medium, the capacity, the dynamic behavior and the cost. 2.1.1. Heat source type The heat source type depends on how the heat source connects with ORC. The schematic diagram of closed open type heat sources are shown in Fig. 3. In some publications, these two types are also referred to as sealed and open types [20], reservoir and stream [21], constant and sensible [22], or recycling and once-through [23] heat sources. Closed type heat sources are possibly include solar energy, biomass combustion and industrial chemical reactions. These heat sources present a high exergy content and are usually connected to the ORC through an intermediate heat transfer loop. The intermediate heat transfer fluid (usually a thermal oil) absorbs the heat from the heat source and delivers it to the ORC system. As shown in Fig. 3, the entire heat flow from the heat source goes into the ORC system if heat losses are neglected. Waste heat from industrial processes, ICE exhaust gases and geothermal water are typical open type sources where the heat source stream is cooled and then exhausted. In theory, the outlet temperatures of open systems are not fixed. The heat source is fully exploited only if its temperature is decreased down to the ambient (or reference) temperature. However, in most cases, the lowest outlet temperature cannot be reached due to pinch point limitations or acid dew point constraints. In such cases, only part of the heat potential of the heat source can be recovered by the

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Greek symbols

Nomenclature area (m2) constant pressure specific heat (kJ/kg/K) exergy (kJ/kg) enthalpy (kJ/kg) solar irradiation (W/m2) lower heating value (kJ/kg) mass (kg) pressure (kPa) heat (kJ) heat per unit mass of the working fluid (kJ/kg) entropy (kJ/kg/K) temperature (K) specific volume (m3/kg) work output (kW) work output per unit mass of the working fluid (kJ/kg)

A cp Ex h I LHV M p Q q s T V W w

ORC. Therefore, closed type sources can be seen as supplying their entire heat flow to the ORC while open type sources supply only part of their heat energy. The distinctions between closed and open type sources is not always straightforward in all cases. For example, geothermal brine could be considered as a closed type heat source if its reinjection temperature influences the ground temperature and, therefore, the extraction temperature. Geothermal brine can also be viewed as an open type heat source if the ground is considered to be an infinite source. Biomass is a closed type heat source if both the boiler and the conductive oil loop are viewed as the heat source but is an open type source when only the gas from the boiler is viewed as the heat source. For both closed and open heat sources, the optimization target is normally to maximize the work output [24] or the exergy efficiency [25], which lead to equivalent results. ⋅

⋅

Wnet = Q ⋅η

(1) ⋅

⋅

where, Wnet is the net work output. Q is the heat supplied by the heat source and η is the cycle efficiency. For closed type heat sources or heat sources with outlet temperature limitations, the

η

efficiency

Subscripts and superscripts cd ex hs in min out p ref th TR wf

condensing exergy heat source inlet minimum outlet pinch reference thermal triangular working fluid

⋅

heat supplied to the ORC system, Q , is fixed. Thus, optimizing the cycle efficiency, η, is equivalent to optimizing the net power output ⋅

Wnet . For open type sources without outlet temperature limitations (or the limitation cannot be practically be reached), the system design should consider not only the cycle efficiency, η, but also the ⋅

heat supplied by the heat source Q . Thus, there is a trade-off between the heat source outlet temperature and the cycle efficiency to achieve the maximum work output. 2.1.2. Heat source temperature Heat source temperatures can vary greatly. This paper focuses on 80–350 °C heat sources. The temperature is indeed the heat source key parameter which directly influences the choice of working fluid. Based on the temperature of the heat source and on the working fluid critical temperature, cycles can be divided into trans-critical cycles and subcritical cycles. This distinction significantly influences both the energy recovery efficiency and the heat exchanger design [26]. Gaseous heat sources include gases generated by industrial processes and the exhaust streams of combustion processes. Liquid phase heat sources include high-temperature water loops and two-phase flows (primarily steam). The heat source flow rate directly impacts the capacity and heat transfer efficiency. The heat transfer coefficients that can be achieved with liquids are much better than with gases, so the pinch temperatures for liquid heat sources are often 5–15 K [24,25] while for gaseous heat source is the pinch temperatures are around 30 K [27] due to the gas/fluid heat exchanger heat transfer rates. Large temperature pinches lead to larger heat transfer losses while low temperature pinches lead to larger heat transfer areas and higher system costs. Thus, each system has an optimal pinch temperature difference. The chemical composition of the medium also influences the system design. For geothermal systems, the outlet temperature of the heat source steam must be high enough to keep the dissolved minerals in solution. Likewise, for combustion by-products, the gas stream outlet temperature must be above the acid dew point. 2.1.3. Heat source capacity Capacity is also an important heat source characteristic. For a specific heat source temperature, the capacity is determined by the temperature change and flow rate. The heat source capacity ⋅

Fig. 1. Simple ORC configuration.

directly influences the system capacity and is referred to as Q hs in

H. Zhai et al. / Renewable and Sustainable Energy Reviews 64 (2016) 790–805

793

Fig. 2. ORC systems using various heat sources [8].

sink temperature (e.g. the cooling fluid temperature). ⋅

⋅

Q hs = Mhs⋅cphs⋅(Ths _ in − Tref )

(4)

For open type heat sources with temperature limitations, the heat source capacity is defined as the heat source content if the flow temperature is reduced to the lowest allowed temperature. ⋅

⋅

Q hs = Mhs⋅cphs⋅(Ths _ in − Ths _

Fig. 3. Schematic diagram of open and closed type heat sources.

the following. The heat own capacity can be expressed in different forms depending on the heat source type.

min

)

(5)

2.1.4. Heat source dynamics The heat source dynamics are another important characteristic. Some heat sources maintain a relatively fixed temperature, like geothermal water sources. In other cases, heat source temperature and capacity are highly variable, such as with industrial waste heat and solar energy. For large but regular changes, such as with solar energy, thermal energy storage is used to smooth the changes. For relatively small, irregular changes (random perturbations), the components should be designed to tolerate off-design conditions to keep the system running efficiently. Slow variations of the heat source conditions, such as seasonal variations, must also be taken into account when sizing the system. In most cases, the system is not sized for the maximum available power, but is down sized to optimize the economic profitability throughout the year.

⋅

In the case of closed type heat sources, Q hs is related to the energy content of the heat source. For solar systems, ⋅

Q hs = A⋅I

(2)

where A is the collector cross-sectional area and I is the solar irradiation (in W/m2). For biomass systems, ⋅

2.1.5. Heat source related costs Different heat sources also have quite different additional costs which may strongly influence the whole system cost. The additional costs include components such as exploration and drilling costs of geothermal ORC systems, the cost of system miniaturization for ICE waste heat recovery systems and the solar collector cost for solar ORC systems. 2.2. Thermodynamic categorization

⋅

Q hs = Mfuel⋅LHVfuel ⋅

(3)

where Mfuel is the biomass mass flow rate to the burner and LHVfuel is the lower heating value. In the case of open type heat sources without temperature limitations, the heat source capacity is defined as the heat source content if the flow temperature is reduced down to the reference

The design of an ORC system based on the heat source conditions requires an analysis of the thermodynamics related to the cycle type and working fluid choices. Heat sources should be categorized in terms of thermodynamics of the cycle layout and the working fluid to make a general design method. Only the heat source temperature and the pinch temperature difference reflect

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Fig. 4. Thermodynamic categorization of heat sources.

in a T-s diagram influence the cycle layout and working fluid choice. The heat source temperature is a key parameter with the largest influence on the cycle efficiency and working fluid selection. The inlet temperature and outlet temperatures determine the optimal cycle layout, for instance whether to use a recuperator or not. The source medium then determines the pinch temperature difference in the evaporator. Heat sources can be divided into three types according to their inlet and outlet temperatures as shown in Fig. 4.

Fig. 4A shows the T-s diagram for a finite capacity heat source without an outlet temperature limitation Tmin labeled “Type A”. “Type A” sources are all open type sources. Fig. 4B shows the T-s diagram for a finite capacity heat source with outlet temperature limitation labeled “Type B”. “Type B” sources can be either open type sources or closed type sources. For an open type source, Tmin is the outlet temperature limitation. For instance, waste heat from industrial processes must be above the gas acid dew point to avoid corrosion. For a closed type source, Tmin is the outlet temperature

Fig. 5. Optimization flow diagram of ORC using different heat source types (Ths_out: heat source outlet temperature; Tlimit: heat source outlet temperature limit).

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795

of the heat transfer cycle and is always set to a relatively high temperature to guarantee a high cycle efficiency. Fig. 4C shows the T-s diagram for an infinite heat source with a constant temperature labeled “Type C”. “Type C” sources are like the constant temperature condensation process that release the latent heat. In thermodynamic analysis, whether the heat source is open or closed type does not influence how the heat source couples with the cycle. However, it will influence the optimization process as shown in Fig. 5. The thermodynamic analysis is only the first step in designing an ORC system. The following steps must consider the other characters of the heat source. For example, the heat source capacity influences the working fluid mass flow rate and the heat exchanger size. The heat source dynamics determine whether energy storage is needed while the additional heat source costs will influence the project cost.

3. Ideal conversion cycles for various heat sources The optimal ORC layouts can be quite different for different heat source types. This section describes how to identify the ideal cycle for each heat source type. The ideal cycle provides the theoretical energy conversion limit for the heat sources. These ideal cycles provide guidance the for thermal cycle design, to decide whether to adopt a supercritical cycle or a subcritical cycle and whether to use a recuperator. The best ORC design that is closest to the ideal cycle that giving best performance. These cycle types can allow for general conclusions like a recuperator is not required if there is no lower limit on the outlet temperature [25]. The heat sink conditions may also be important, but they have less effect. To avoid having a very large number of cases, the heat sink is assumed to be infinite and isothermal. 3.1. Ideal cycle for “Type A” sources “Type A” sources are finite capacity heat sources without outlet temperature limitations. In many publications [9,28–30] the nonisothermal heat source profile on the T-s diagram is shown as a straight line and the ideal cycle for the open heat source is assumed to be triangular cycle as in Fig. 6. The triangular cycle efficiency is then defined as [28,29]:

Fig. 7. Comparison of a real sensible energy heat source with the assumed straightline.

⋅

⋅

ηthTR =

Wnet ⋅

=1−

Q in

Q out ⋅

Q in

=

T1 − T2 T1 + T2

However, for a sensible heat, isobaric heat source, the heat source temperature variation is not a simple straight line on the Ts diagram. The basic thermodynamic functions are

⎛ ∂h ⎞ ⎛ ∂s ⎞ dh = Tds + Vdp ⇒ ⎜ ⎟ = T ⎜ ⎟ ⎝ ∂T ⎠ p ⎝ ∂T ⎠ p

(7)

⎛ ∂h ⎞ and ⎜ ⎟ = cp ⎝ ∂T ⎠ p

(8)

⎛ ∂T ⎞ T T ⇒ Δs = cp ln 2 then, ⎜ ⎟ = ⎝ ∂s ⎠ p cp T1

(9)

Fig. 7 compares these two definitions and highlights the nonlinearity of the sensible energy heat sources. Air from 280 K to 650 K at 0.1 MPa has a significantly different slope than the straight line. If the pinch point is fixed, the average evaporating temperature of the ideal cycle for the real heat source is lower than that for the constant T/cp heat source. Thus, the differences in the heat sources directly influence the cycle performance. This difference is due to the fact that Eq. (6) assumes a linear relation slope between T and s with ⋅ ⋅ ⎧1 ⎫ Q in = Mwf ⋅⎨ [T1 (s1 − s2 ) − T2 (s1 − s2 )] + T2 (s1 − s2 ) ⎬ ⎩2 ⎭ ⋅

(10)

⋅

Q out = Mwf T2 (s1 − s2 )

Fig. 6. Triangular cycle for “Type A” sources.

(6)

(11)

However, this linear relation slope is not exact according to Eq. (9). In addition, cp for most heat sources is not constant, so the variations are not linear. As a consequence, the triangular cycle cannot be considered as the ideal cycle for an open type sensible energy heat source, but only an approximation. A more realistic heat source profile derived from Eq. (9) is shown in Fig. 4A. Even if cp is simplified to a constant since cp does not vary significantly, the gradient of the heat source line decreases as the temperature decreases. Then, the ideal cycle keeps the temperature difference at Tp during the entire evaporating process. Thus the ideal cycle includes an isobaric heating process, an isentropic expansion process and an isothermal cooling process. The heat source and its ideal cycle are shown in Fig. 8. To

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“Lorentz cycle” if the heat sink has a temperature glide as well as the working fluid [31]). The cycle efficiency and the work output calculations are calculated in a similar manner to Eqs. (16) and (17) with the reference temperature being replaced by the minimum heat source temperature: For the ideal cycle, T1 ¼Tin, T2 ¼ Tmin and the work output, Wnet, is ⋅ ⋅ ⋅ ⎛ Tcd T⎞ Wnet = Q ⋅ηth = Mhscphs (Tin − Tout ) ⎜ 1 − ln 1 ⎟ ⎝ T1 − T2 T2 ⎠ ⋅ ⎛ T ⎞ = Mhscphs ⎜ Tin − Tmin − Tcd ln in ⎟ ⎝ Tmin ⎠

(18)

3.3. Ideal cycle for “Type C” sources

Fig. 8. Division of an ideal nearly triangular cycle into an infinite number of micro Carnot cycles.

“Type C” sources are infinite capacity, constant temperature Ths heat sources and are usually closed type sources. Since this is an isothermal heat source, the ideal cycle is a Carnot cycle as shown in Fig. 4C. The system efficiency is given by

ηth = 1 − Tcd/T1 make the best use of the heat source, T2 should be equal to Tcd. The work output and the cycle efficiency of the “nearly triangular cycle”, the cycle is divided into an infinite number of micro Carnot cycles as shown in Fig. 8. The work output of the cycle per unit mass flow rate of the working fluid is

w=

∫T

T1

2

⎛ T ⎞ ⎜ 1 − cd ⎟ Tds ⎝ T ⎠

(12)

From Eq. (9),

Tds = cp dT

(13)

Substituting of Eq. (13) into Eq. (12) and integrating yields

w=

∫T

T1

2

⎡ ⎛ T ⎞ T⎤ ⎜ 1 − cd ⎟ cpwf dT = cpwf ⎢ (T1 − T2 ) − Tcd ln 1 ⎥ ⎝ ⎣ T ⎠ T2 ⎦

(14)

The heat absorbed during the heating process is

q = cpwf (T1 − T2 )

ηth

w Tcd T = =1− ln 1 q T1 − T2 T2

For the ideal cycle, T1 ¼Tin ¼ Ths, so the work output is ⋅ ⋅⎛ ⋅⎛ ⋅ T ⎞ T ⎞ Wnet = Q ⋅ηth = Q ⎜ 1 − cd ⎟ = Q ⎜ 1 − cd ⎟ ⎝ ⎝ T1 ⎠ Ths ⎠

A large number of publications have focused on the first-law cycle efficiency to optimize the cycle and select the working fluid. However, other publications have shown that this approach is biased and does not reflect the real system performance for at least two reasons:

 In the case of open type sources, a higher efficiency can be

 (16)

For the ideal cycle, T1 ¼Tin, T2 ¼Tout ¼Tcd and work output, Wnet,

(20)

4. Performance metrics

(15)

So the cycle efficiency is

(19)

reached by increasing the evaporation temperature, but this results in a lower heat source utilization which may result in a lower output power [22,24]. Optimizing the efficiency alone can lead to unrealistic operating conditions, such as the use of a high critical temperature working fluid for a low temperature heat source, which results in low vapor densities and the need for large components in the system [32].

is ⋅ ⋅ ⋅ ⎛ Tcd T⎞ Wnet = Q ⋅ηth = Mhscphs (Tin − Tout ) ⎜ 1 − ln 1 ⎟ ⎝ T1 − T2 T2 ⎠ ⋅ ⎛ Tin ⎞ = Mhscphs ⎜ Tin − Tcd − Tcd ln ⎟ ⎝ Tcd ⎠

Therefore, alternative performance indicators are required and have been proposed in the literature. The exergy efficiency is the most common indicator among these alternative performance indicators,

(17)

3.2. Ideal cycle for “Type B” sources

⋅

ηex =

Wnet ⋅

Ex

(21)

⋅

“Type B” sources are finite capacity heat sources with outlet temperature limitations. The designs for open type sources with outlet temperature limitations are quite similar to those without outlet temperature limitation, except that Tout is now fixed to the minimum temperature instead of Tcd. The ideal thermal “nearLorentz cycle” is shown in Fig. 4B. As with the “Type A” sources, the ideal cycle dose not have a linear T and s relationship (i.e. a

where Wt is the actual cycle work output. For a closed type heat source, ⋅ ⋅ ⎛ T ⎞ Ex = Q hs ⎜ 1 − ref ⎟ ⎝ Ths ⎠

(22)

For an open type heat source without an outlet temperature limitation,

H. Zhai et al. / Renewable and Sustainable Energy Reviews 64 (2016) 790–805 ⋅

⋅

Ex = Mhs [hin − href − Tref (sin − sref )] ⋅ ⎛ T ⎞ = Mhs cphs ⎜ Tin − Tref − Tref ln in ⎟ ⎝ Tref ⎠

(23)

For an open type heat source with an outlet temperature limitation, ⋅

⋅

Ex = Mhs [hin − hmin − Tref (sin − smin )] ⋅ ⎛ T ⎞ = Mhs cphs ⎜ Tin − Tmin − Tref ln in ⎟ ⎝ Tmin ⎠

(24)

When the condensation temperature is as low as the reference ⋅

temperature, Ex is equal to the work output power of the ideal cycle as shown in part 3. This also shows that the ideal conversion cycle has the lowest condensation temperature (reference temperature) and makes the best use of the heat source. Thus, the exergy efficiency is equivalent to the ratio of the output power divided by the output power of the ideal cycle. Some authors have used the exergy efficiency or the output power as the performance metrics, which addresses the first issue listed in this section but not the second. Other authors have used thermo-economic indicators (e.g. the cost per kW) or included the required total heat exchanger area in the objective function. The various performance metrics that have been used to characterize the ORC systems in the literature are summarized in Table 1. The economic indicators (e.g. the cost per kW, the Net Present Value, the payback time, etc.) can vary greatly between sources due to differences in the assumed conditions and to the different cost functions available in the literature. Therefore, economic indicator results are more suitable for economic evaluations of specific systems rather than for comparing systems. They are particularly useful for system optimizing or for working fluid Table 1 Performance metrics for ORC systems. Performance Metric

Heat source

References

First law efficiency/ Cycle efficiency

Industry Geothermal

[33–37] [38,39]

Daily, monthly or yearly average efficiency

Solar

[40]

Second law efficiency/ Exergetic efficiency

Industry Solar Biomass

[33–37] [41,42] [43–46]

Work output per unit heat source mass flow rate

Industry ICE & GT Geothermal Industry

[33,34] [11] [24,47] [48]

Industry ICE & GT Geothermal

[49] [11] [50,51]

Expander size

Industry Solar

[52] [53]

Energy utilization efficiency/ Heat source utilization efficiency Cost per installed kW Electricity production cost (epc)

Industry Geothermal Industry Industry ICE & GT Geothermal Biomass

[48,49] [54,55] [32] [56] [11] [47] [57]

Combined cycle efficiency Overall or global efficiency

ICE & GT [58–60] Solar [61,62] Biomass (power and heat [46,63,64] efficiency)

Power to heat ratio

Biomass

Work output per unit working fluid mass flow rate Work output per heat transfer area

[65]

797

selection since they include the effects of the thermodynamic conditions on the system design and cost. The power output per heat transfer area has also been used in a number of publications. This indicator relies on the assumption that the major cost in an ORC system is the heat exchangers [66]. The heat exchanger cost is then used in the economic indicator for the cost per installed kW, with the major advantage that this does not require cost functions for the turbine, which are not readily available in the literature and can vary greatly depending on the turbine technology. However, this approach is only suitable when operating costs (e.g fuel costs) are negligible, which is typically the case in waste heat recovery applications. Finally, economic indicators should only be sued to compare heat exchangers of the same type, because there are large differences in the surface efficiencies of different exchanger types. The expander size and cost are strongly influenced by the working fluid selection and by the operating conditions. Thus, the expander cost should be used as a boundary condition to ensure a reasonable fluid selection for the application, but this cannot be used as the sole performance indicator. The exergy efficiency is the most commonly used thermodynamic performance indicators. The output power is useful for open type sources, whereas the first law efficiency is appropriate for closed type heat sources.

5. Typical heat source analyses and their influence on the ORC system Typical low-medium temperature heat sources include waste heat from industrial systems and ICE and gas turbine power system, geothermal hot water sources, solar heat and biomass energy. The characteristics for these typical low-medium heat sources are shown in Table 2. The theoretical categorization of the source types is useful for cycle optimization and working fluid screening. Besides the theoretical heat source type categorization, the source capacity, dynamics and cost should also be taken into consideration when analyzing a specific heat source. 5.1. Waste heat from industrial processes Typical waste heat sources include flow streams of subcooled liquid, hot air, waste gas, high pressure hot water, waste water and steam [80]. Industrial waste heat from industry has a broad range of temperatures. In general, steel, glass, non-ferrous metals and ceramics production processes have 300–400 °C waste heat streams. Heat sources with temperatures around 150 °C can be found in the food, refining and building industries [68]. Typical pinch point temperature in heat exchangers are 5–15 K for liquid heat sources [24,25] and 30 K for gas heat sources [27]. Smaller design pinch point temperatures will give higher energy and exergy efficiencies but larger required heat transfer areas and heat exchanger costs. Thus, the pinch point temperature should take both the system performance and cost into consideration. Some researchers have optimized the pinch point temperature using work output per heat exchanger area as a performance metric [49,82] or by employing a term of thermo-economic optimization [32]. The exhaust gas composition from combustion processes can also be varied, but often includes CO2, NOx and sulfides in addition to nitrogen and water [67]. Table 3 shows an example of the exhaust gas composition from a rotary kiln for producing cement. The SOx content limits the heat source outlet temperature since SO2 is oxidized to SO3 which then combines with H2O to become

Water

Table 3 Exhaust gases from a rotary kiln for producing cement [83].

[46,79–81] 100–1500 kW Fluctuates, uses oil to stabilize

250–550 °C GT

Oil

6 kW–5 MW Varies with daily and monthly

80–100 °C cooling system; 400–900 °C exhaust gas

Component Concentration CO2 NO2 NOx SO2 O2

180–400 g/scm 0.01–0.3 g/scm 0.20–3 g/scm 0.01–3.5 g/scm 100–200 g/scm

sulfuric acid vapor. The dew point for the sulfuric acid vapor is referred to as the acid dew point, which is usually higher than the water dew point. More SOx, O2 and H2O in the flue gas will result in higher acid dew point temperatures. Typical acid dew point temperatures can be as high as 120 °C depending on the composition [25,67]. The heat source outlet temperature must then be somewhat higher than the acid dew point to protect the heat exchanger. In addition to these constraints, the following issues must also be considered:

 In many cases, the waste heat recovery system can not interfere

 

with the original industrial process. The leads to additional requirements such as negligible pressure drop in the flue gases or installation of the heat exchanger after the flue gas treatment system. Some system have more than one heat source in complex industrial processes. Then the ORC must be optimized with the multiple heat sources [84,85]. With direct evaporation, the energy flows directly from the heat source to the working fluid. During transients, the heat source temperature can abruptly fluctuate, which can cause overheating and chemical decomposition of the working fluid [26]. One solution is indirect evaporation with a heat transfer oil acting as a buffler between the fluctuating waste heat stream temperature and the ORC to guarantee smooth system operation. Another solution is the use of a good control system to ensure safe operation with fluctuating boundary conditions [86,87].

With these conditions, industrial waste heat sources are always “Type B-open” (once-through gas) or “Type C” (water stream). There are often a number of heat sources in the system so the waste heat recovery efficiency should evaluate the entire waste heat recovery efficiency for the industrial process.

Around 300 °C B-closed Biomass

80–90 °C A Solar pond

o 300 °C B-closed Solar collector

80–180 °C B-open

A, B-open

5.2. Waste heat from power generation systems

Geothermal

Power waste heat

ICE A, B-open

Water or Air with other gases Air with other gases Water with brine Water or oil

MW system 5000–6000 €/kWe MW system 5000–6000 €/kWe o 30 MW Varies with daily and monthly

1600 €/kWe for medium scale Boiler cost 1000–1250 €/kW 10,000 €/kWe for small scale Cost of straw is around 9– 14c€/kWh e

[26]

[75–78]

[24,73,74] 0.6–27 MW Varies a little

Drilling cost and the cost to pump the brine Collector cost is 2/3 of the entire system cost None

Non

Micro GTþ ORC, 2500–3000 €/kW 1000–4000 €/kWe 50 kW-6.5 MW Fluctuates if not monitored

[55,59,72]

Miniaturization required No data 95 kW–6.5 MW Starting and braking moments

[69–71]

None System total cost 1500–4500 €/kW, Operating cost 5.65c €/kWh 125 kW–3 MW 10 MW for big industry detrimental to the system: stalling or temperature shocks A, B-open, C. 80–500 °C, Mostly 200– 300 °C Industry

Air, water or steam

Heat source cost System Cost Capacity Dynamic behavior Media Source temperature Type Heat source

Table 2 Key heat source characteristics.

[26,67,68]

H. Zhai et al. / Renewable and Sustainable Energy Reviews 64 (2016) 790–805

Main references

798

5.2.1. Internal combustion engines The main characteristics of waste heat from internal combustion engines (ICE) are multiple heat sources, on-vehicle-mounting (thus, space constrained) and variable working conditions during start or braking [71]. The two primary heat sources for ICE systems are the heat from the engine cooling system (80–100 °C) and the exhaust gas (400– 900 °C) [70]. 60–70% of the fuel energy is lost through the exhaust gas and the cooling system [88]. The ORC output can be used as mechanical work or electricity. To directly use the mechanical power, the expander rotation speed in the ORC should be a fixed ratio to the engine speed, which may not correspond to the optimal cycle efficiency. Alternatively, electricity generated by the ORC can be used to power the air conditioning system or hybrid vehicles. However, the efficiency of current automotive alternators is about 50–60%, which reduces the overall system efficiency [69].

H. Zhai et al. / Renewable and Sustainable Energy Reviews 64 (2016) 790–805

The impact of an engine ORC waste heat recovery system on the overall fuel consumption has been studied (almost exclusively theoretically) in various publications. An ORC can improve the efficiency by about 10% without consuming extra fuel [89]. Srinivasan, et al. [90] showed that an ICE with an ORC as a bottoming cycle improves the energy conversion efficiency by 7% as compared with the original ICE cycle with an 18% deduction in the NOx and CO2 exhaust. The energy from the two heat sources can be more efficiently used by using the heat from the engine cooling system to preheat the working fluid or by using a dual loop ORC system [70] with the exhaust heat recovered by a high temperature cycle and the remaining exhaust heat and the cooling system heat recovered by a low temperature cycle. However, systems using two heat sources are very complicated. Vaja and Gambarotta [27] compared different ORC configurations with a simple cycle recovering only energy from the exhaust gas and the other with a cycle recovering heat from both the exhaust gas and the cooling system. Much less heat is available from the engine cooling heat than from the exhaust gas, but leading to significant differences between the systems. There is, therefore, a trade-off between a simple configuration (not using heat from the cooling system) and a higher work output (using heat from the cooling system). ORC systems on vehicles also fact important space and weight constraints. Thus, such systems should use low volume flow rate working fluids (organic working fluids are much more suitable than water in this case), compact heat exchangers and simple configurations. Thus, waste heat from ICE systems are “Type A” (cooling cycle) or “Type B-open” (exhaust gas). The cooling cycle as a heat source is often ignored. An ICE has its own efficiency which can be improved by an ORC bottoming cycle using waste heat from the ICE to maximize the combined cycle efficiency. 5.2.2. Gas turbines Gas turbines are mainly natural gas turbines and biogas turbines with very different exhaust gas conditions. Table 4 lists information for different kinds of gas turbines. The gas turbine exaust is always at a high temperature with considerable energy which can be recovered by conventional gas and steam combined cycles or ORC. Gas and steam combined cycles have good effiencies when the exhaust gas temperature is quite high such as with heavy duty gas turbines [59]. ORC systems are usually used to recover waste heat from small gas turbines below 50 MW [55] or recuperative gas turbines [59] whose waste heat temperatures are lower than the conventional gas turbines. Gas turbine efficiencies can be quite low at part-load conditions [92], but an ORC system can make up for the low efficiency at part load conditions [93]. The earliest record of a combined gas turbine-ORC system is for a 30 kW gas turbine generating 18.9 kW more electricity with the ORC [94]. Leslie et al. [58] showed a 5.5 MW electricity production increase with a waste heat recovery system based on a converted ORC system to recover waste heat Table 4 Typical gas turbine conditions. Type

Power range

Exhaust gas temperature range

Reference

Micro-gas turbines Biogas gas turbines

o 500 kWe 250 kWe– 3 MWe 4–50 MWe

250–300 °C  450 °C

[72] [91]

350–400 °C

[59]

100–400 MWe

550–650 °C

[59]

Larger gas turbines with regeneration Heavy duty gas turbines

799

from a 27 MW gas turbine. Wisniewski and Borsukiewicz-Gozdur [95] showed that superheating did not improve the cycle efficiency without regeneration. Ahmadi et al. [96] showed that a gas turbine-ORC-absorption chiller system had a higher exergetic efficiency than a combined heat and power (CHP) system or a single gas turbine and discharged less CO2. Small micro gas turbine systems are also an important research direction. Micro gas turbine systems have power ranges between 25 and 300 kW and efficiencies of almost 30%. Thermal efficiencies of 40% can only be reached with more complex thermal cycles like combined cycles [97]. ORC systems are suitable for waste heat recovery from micro gas turbines due to their small site and relatively low temperatures of 250–300 °C [72]. Clemente et al. [98] showed that R245fa, isopentane and isobutene are more suitable than siloxanes for waste heat recovery from a 100 kWe gas turbine because they have lower volume ratios while expanding so the system can have a single stage expander. The source is either “Type A” or “Type B-open” depending on the gas turbine fuel. If the gas turbine uses natural gas or purified biogas as the fuel, the exhaust gas has little SOx, the acid dew point is not a problem and the exhaust is a”Type A” source. However, the exhaust is a “Type B-open” source if the gas turbine uses heavy or light oil as the fuel. The recommended performance metrics is also the combined cycle efficiency, the same as for waste heat recovery from an ICE. 5.3. Geothermal Geothermal energy is a renewable resource due to the large amount of heat inside the earth and is one of the renewable sources that can be used year-round. On average, the ground temperature increases 30 K per 1 km into the earth [68]. The average heat flux at the surface is 0.082 W/m2 and the total energy content can reach 4  1013 W, which is several times greater than the current worldwide energy consumption [99]. Geothermal energy sources exist in two forms as shown in Table 5. Hydrothermal reservoirs contain hot water or steam which can directly exchange heat with ORC systems. Enhanced geothermal systems (EGS) use well drilled into the shale with cold water injected into the earth to generate hot water or steam at temperatures less than 650 °C. As a result, both types of geothermal sources provide suitable temperatures for heat recovery by ORC systems. The temperatures, pressures and compositions of the geothermal water can be quite different depending on the geological conditions [101]. Geothermal water vapor usually contains CO2, H2S, HCl, HF, NH3, CH4, and H2 with proportions that vary with the geothermal field, as shown in Table 6. The return stream temperature must then be somewhat higher than the salt saturation temperature given by its composition to prevent salt from precipitating out of the geothermal vapor or water. Geothermal systems are sensitive to the environmental conditions. Since the heat source is already at a low temperature, the heat sink temperature is the key parameter for the system performance. If the condensing (sink) temperature is high, the efficiency will be very low. In Iceland the environmental temperature Table 5 Geothermal sources. Form

Thermal state

Hydrothermal reservoirs

Hot water or steam Hot shale

Enhanced Geothermal Systems (EGS)

Temperature range

Reference

80–180 °C

[24]

100–650 °C

[100]

800

H. Zhai et al. / Renewable and Sustainable Energy Reviews 64 (2016) 790–805

Table 6 Typical chemicals in the steam from some geothermal fields [73]. Constituents g/kg

THE GEYSERS USA

LARDERELLO Italy

MATSUKAWA Japan

WAIRAKEI New Zealand

CERRO PRIETO Mexico

H2O CO2 H2S NH3 CH4 þH2 Others

995.9 3.3 0.2 0.2 0.2 0.2

953.2 45.2 0.8 0.2 0.3 0.3

986.3 12.4 1.2

997.5 2.3 0.1

984.3 14.1 1.5 0.1

0.1

0.1

is less than 10 °C all year-round, so 70 °C geothermal water is hot enough to generate electricity with an ORC. Two 280 kWe ORC systems using R245fa as the working fluid have been constructed to make use of 74 °C geothermal water and have relatively high efficiencies since the condensing temperature is only 5 °C [102,103]. The cost of drilling the wells and the pump work are the main reasons for the high costs of geothermal ORC systems. The pumps can consume 30–50% of the power produced by the system and the well drilling costs can be as much as 70% of the entire system cost [2]. All geothermal sources will be “Type B-open” sources when combined with ORC systems. Geothermal sources are high pressure water which exchange heat with the working fluid in ORC and have outlet temperature limitations to avoid salt corrosion. Since the drilling and pumping costs are very high, the work output per unit heat source mass flow rate is usually used to evaluate the system to make good use of the geothermal water. 5.4. Solar The solar insolation reaching the earth is 1.7  1017 W, more than ten thousand times the current worldwide energy consumption rate [68]. Solar ponds and Concentrated Solar Power (CSP) systems are used to convert the solar power to electricity. One key characteristic of solar energy is its low energy density which requires large collector areas. The temperature ranges for various collector types are shown in Table 7. An ORC is more suitable than a steam cycle when the temperature is lower than 400 °C [75]. The traditional tradeoff with concentrated solar systems is that high temperature collectors provide higher system efficiencies but lower collector efficiencies and also higher costs. Thus, each system design has an optimal collector temperature [104]. In the European POWERSOL project, the collector outlet temperatures can be 100–150 °C, 200–250 °C or 800 °C to explore the economics and system efficiencies for each collector temperature [76]. Some examples of the collector efficiency's impact on the overall system efficiency are summarized in Table 8. Another key characteristic of solar heat is the variation caused by the fluctuating solar flux with both diurnal and seasonal variations. Dynamic or quasi steady-state models are needed to evaluate the effect of these variations on the power generation and the economics. Several studies have used simulations based on local meteorological data [77,110,111] or compared the influence of the meteorological conditions on the system [112]. Thermal energy storage reservoirs are used between the solar collector and the ORC system to keep the system running efficiently and smoothly. Such systems then have three operating modes with the solar only mode during periods of low insolation, solar with energy storage during periods of high insolation and energy releasing from the storage at night [41]. Solar ORC systems can be used to drive steam compression refrigerators to make ice [113] and for reverse osmosis sea water desalination [114–116]. In addition, solar ORC can also bring

Table 7 Solar collector outlet temperatures [8,26,40]. Collector type

Temperature (°C) Conversion technology

Solar pond

80–90

ORC

o 150

Heat or ORC

90–200

Heat or ORC

200–450

ORC or steam cycle ORC or steam cycle steam cycle steam cycle

Simple collectors

Concentrated solar collectors

Flat plate collector Evacuated tube collector Parabolic trough collector Fresnel reflector technology Parabolic discs Solar tower

100–400 750 1000

electricity to remote districts that would otherwise require expensive, long distance transmission line [117,118]. Solar energy from collectors are “Type B-closed” heat sources. Solar energy from solar ponds are “Type A” sources. The solar collector efficiency then significantly influences the system thermal efficiency. Thus, the system thermal efficiency takes the collector influence into account. The solar energy also changes with the time and seasons, so the calculation of the solar system efficiency cannot only use the conditions for one day, but the average efficiency over the whole year or at least one whole month should be used as the indicator. 5.5. Biomass Biomass is the fourth largest energy source in the world which provides 10% of the world's primary energy consumption [119]. The narrowed definition of biomass refers to straw and wood from agriculture and forestry operations, leftover farming materials, forestry and agricultural waste and animal dung from stock farming. When grown in a sustainable way, biomass is a renewable resources with lower CO2 emissions than fossil fuels. Both the external and internal combustion can be used to generate electricity from biomass. External combustion burns the biomass in a boiler and the exhausted heat is transferred to a steam cycle, ORC system or Stirling engine (which can be considered to be “external” combustion engines). When used in an internal combustion engine (e.g. a gas turbine or a vehicle engine), the biomass must be turned into a gas (through pyrolysis and gasification) or into a liquid (biofuels). The main drawback of external combustion is its lower conversion efficiency while internal combustion has serious gas purification issues [120]. Biomass resources are widely distributed and have low energy densities, so they are suitable for low-capacity systems that make use of the heat source locally to reduce transportation costs. Depending on the biomass type, the high temperature exhaust gases may contain large quantities of potassium ions, sulphur ions and chlorine ions, which can cause corrosion in the boiler [57]. The boiler also has the highest exergy loss (up to 55%) [44]. Typical

H. Zhai et al. / Renewable and Sustainable Energy Reviews 64 (2016) 790–805

801

Table 8 Impact of the collector efficiency on overall efficiency of an ORC system. Year

Project

Collector

Collector efficiency

ORC efficiency

Overall efficiency

Reference

2003 2004 2005 2006

EPFL-Lausanne Arizona Public Service STG International SEGS VI

linear Fresnel collectors Solargenix parabolic trough collectors parabolic trough collectors LS-2 collectors provided by Solargenix

57%

7.74%

60%

15% 20.7%

4.2% 12.1% 2–10% 12.1%

[105] [106] [107] [108]

2010

Tianjin University

flat plate solar collectors evacuated solar collectors

55% 71%

3.2% 4.2%

[109]

biomass ORC systems use an intermediate oil loop between the boiler and the ORC. This oil loop lowers the boiler pressure, buffers load changes and makes the system control and operation safer [26]. Thermal stability of the working fluid (typically a silicon oil) limits the ORC temperature to 330 °C [79]. Biomass ORCs are often operated as CHP systems to increase their overall efficiency, so they require local heating demand, such as a district heating system. The ORC condenser temperature is generally 60–120 °C [121]. Low critical temperature working fluids have relatively high pressure at high condensing temperatures and are not suitable for biomass ORCs. Biomass sources for ORC systems are then “Type B-closed” sources. Since most biomass ORC are used in CHP systems, the power to heat ratio (or heat to power ratio) is the recommended indicator for the CHP design. In addition, the overall efficiency for power generation and heating is a useful indicator to evaluate the biomass resource use. 5.6. Multiple heat sources In addition to individual heat sources, researchers have also studied system using multiple heat sources to better match the temperature profiles of the working fluid and the heat sources. In such cases, advanced concepts, such as ORC cascades are required to minimize the exergy destruction and losses. An example of such a system by Kane et al. [105] used solar energy and the exhaust gas heat and cooling water heat from a biomass engine with the temperature diagram shown in Fig. 9. Tempesti and Fiaschi [122] analyzed a system that used both solar and geothermal energy sources so that it could run continually without thermal storage. In general, higher temperature heat sources are used for evaporation and superheating, while lower temperature heat sources are used to preheat the working fluid [123]. Even though multiple heat source systems can have high theoretical efficiencies, the systems can become very complex and difficult to control which may negatively impact the overall cost.

Fig. 9. ORC system cascade using multiple heat sources [105].

6. ORC site market economics The market economics of the ORC heat sources are also important. Statistics for ORC sites with various heat sources and the main manufacturer have been compiled up to year 2014 for Turboden [124], Ormat [125], Tri-o-gen [126], Enertime [127], GMK [128], Bosch [129], Orcan [130], ElectraTherm [131], Rank [132], Enogia [133], Adoratec [134], BEP Europe [135] and Kaishan [136]. The number of installed ORC systems worldwide and the total capacity are plotted in Fig. 10. The shares of each type of heat sources are shown in Fig. 11. ORC sites have been built since 1975 with the number of ORC units increasing rapidly after 2005. The generating capacity and the number of units were both highest in 2014 market the ORC will continue to grow as shown in Fig. 10. However, the number and generating capacity of the ORC system are still very few.

Fig. 10. Generating capacity and number of ORC units of different years.

The statistics in Fig. 11 show that geothermal and biomass ORC systems are the most mature ORC systems. Geothermal systems provide the most power while biomass systems have the largest number of sites. Even though the biomass and geothermal ORC systems are the most mature, they still depend on electricity price subsidies from the government at this stage. Researches showed that the waste heat recovery applications have the lowest costs while the geothermal and biomass CHP plants exhibit higher total

802

H. Zhai et al. / Renewable and Sustainable Energy Reviews 64 (2016) 790–805

Fig. 11. Generating capacity and number of ORC sites of different heat sources.

cost [2]. Thus, there should be big development potential for recovering the waste heat for industry and power system by ORC systems.

7. Development directions ORC systems have many advantages such as equipment simplicity, easy miniaturization, and good efficiencies at part load condition. Thus, ORC systems are an important thermal power conversion technology. However, the economics are bad with low temperature heat sources. Although the market is developing rapidly, there are still very few ORC systems. The ORC market will continue to develop in several areas: 1. Current ORC systems are economically viable 1) when coupled with other economically viable systems, such as distributed energy systems, to improve the commercial value and cascade the energy utilization to improve the overall system efficiency; 2) when used with low temperature cold sinks, such as used in Iceland or when coupled with cryogenic energy storage systems since low temperature sinks improve the ORC system efficiency. 3) ORC systems should be applied where they have unique development potential, such as the biomass ORC and solar ORC systems used in rural areas for domestic power supplies or in cars or ships. 2. Fundamental improvements on the ORC efficiency. The ORC type should be selected based on the heat source characteristics. Working fluid mixtures should be used to better match the system with the heat source and cold source temperature glides. The theoretical fundamentals of subcritical ORC are well known except for the use of mixtures in real systems. However, transcritical ORC systems still have basic issues that need more study, such as the basic thermal properties of the working fluid and the heat transfer correlations in the trans-critical region.

80–350 °C and their influence on the ORC performance. This theoretical categorization of ORC heat source will help to make future ORC research more systematic and targeted. Key major characteristics are the difference between the open type and closed type sources and the source inlet and outlet temperatures which strongly affect the ORC design. Guidances are then given for optimizing ORC systems for different heat source types. The ideal thermal cycles for these different sources are studied with an example showing that, contrary to general belief, the triangular cycle is not the ideal cycle for waste heat recovery. The ideal thermal cycles identified for each source type give the theoretical thermal conversion limits and also provide guidance for the ORC layout choice. ORC system cycle optimization should take the unique characteristics of different heat sources into consideration. For example, solar ORC systems require thermal storage and ICE ORC systems should have simple configurations. Different performance metrics should be used for different heat sources to make the results of different studies comparable. A database of commercial ORC systems is introduced to provide statistics regarding existing heat sources. Geothermal and biomass ORC systems are the most mature ORC systems. Geothermal systems provide more than half of the installed generating capacity while biomass systems are the most common. Waste heat recovery sources also account for a significant share of the current ORC systems. ORC systems for other types of heat sources are still being developed. Finally, this review leads to effected future developments for ORC systems. ORC systems are not yet economical, but this study gives advices on the use of current ORC systems and on how to improve the performance of the ORC system itself. The system costs are expected to significantly decrease as the ORC system develop.

Acknowledgements 8. Conclusions The characterization and understanding of the heat source is a precondition for ORC research. The heat source characteristics impact the design and performance metrics of ORC systems. This paper provides a theoretical categorization of heat sources and reviews the typical heat source characteristics for temperatures of

This work was supported by the State Key Program of the National Natural Science Foundation of China (Grant No. 51236004), the Science Fund for Creative Research Groups (No. 51321002) and the ORC technology research scholarship of the ASME IGTI ORC Power Systems Committee (2013). The authors thank Dr. Ian Bell for his valuable comments of this paper.

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