A turbine based domestic micro ORC system

Transcription

1 Available online at ScienceDirect Energy Procedia 00 (2017) IV International Seminar on ORC Power Systems, ORC September 2017, Milano, Italy A turbine based domestic micro ORC system Piotr Klonowicz *, Łukasz Witanowski, Łukasz Jędrzejewski, Tomasz Suchocki, Piotr Lampart Institute of Fluid-Flow Machinery Polish Academy of Sciences, Fiszera 14 st., Gdańsk , Poland Abstract The paper presents an analysis of a turbine based micro ORC system for a domestic biomass boiler. The assumed nominal boiler capacity is in the range between 15 kw and 20 kw. Such a small thermal output is particularly difficult from the point of view of power generation. One of the major issues is the economic aspect of the system. In general, small systems are relatively more expensive. Another problem is poor electrical efficiency which results from the efficiencies of the individual components such as the expansion unit and the feed pump. Various working fluids are considered and also various methods of evaporation such as thermal oil loop, pressurized water loop and direct evaporation. The system efficiencies are examined. The main focused is laid on the turbine design. Due to low volume flow rates the rotational speeds must be high, even exceeding 100 krpm for some of the designs. Various turbine types are considered, however, the preferred choice is a partially admitted highly loaded impulse stage. Very small volume flow rates lead to blade channels that are difficult to manufacture, throats of the supersonic nozzles can be particularly problematic. Manufacturing difficulties impose worse tolerances which lead to bigger relative clearances and blade edge thicknesses which have a negative impact on the efficiency. High speed require appropriate bearing system which must be reliable and, particularly for such a small system, simple. Gas bearings (e.g. foil bearings) or ceramic bearings seem to fulfill these requirements The Authors. Published by Elsevier Ltd. Peer-review under responsibility of the scientific committee of the IV International Seminar on ORC Power Systems. Keywords: micro CHP, ORC, microturbine, turbine * Corresponding author. Tel.: ; fax: address: pklonowicz@imp.gda.pl The Authors. Published by Elsevier Ltd. Peer-review under responsibility of the scientific committee of the IV International Seminar on ORC Power Systems.

2 2 Piotr Klonowicz et al. / Energy Procedia 00 (2017) Introduction In the face of climatic changes and ever growing energy costs one should explore new and clean energy sources and also aim for more responsible ways of energy conversion and use. The latter can be partly achieved through more efficient power production. Today, still lots of households use solid fuel boilers for heating and despite of more and more popular biomass application the coal is still widely used. Sadly, in Poland the coal covers over 70% of the heat demand in the residential sector. What is more, about one third of the burned fuel is used for heating only without cogeneration which is in total 130 PJ per year [1]. From the exergetic and the economic point of view this is a waste of resources because a high quality fuel is used to rise the temperature only by a few degrees. On the other hand the domestic cogeneration in the range of 15 kw to 20 kw is technologically and economically challenging. One of the promising technologies is the Stirling engine which is powered by the hot fumes produced by a boiler. These devices allow one to create maximally about 16% of electrical energy even for very small units [2 4]. Another very interesting technology is the Peltier module which bases on thermoelectric effect. It has many advantages including scalability and lack of moving parts which results in low maintenance and high durability [5 7]. Both of these technologies require further development and presently their greatest drawback is the high investment cost. Nomenclature P T microturbine output power [kw] P P feed pump driving power [kw] P NET ORC net power [kw] ƞ T microturbine isentropic efficiency [%] ƞ P feed pump isentropic efficiency [%] h enthalpy [kj/kg] p pressure [bar(a)] T temperature [ C] ΔT PPE evaporator pinch temperature difference [K] ΔT PPC condenser pinch temperature difference [K] m working fluid mass flow rate [kg/s] m h heating fluid mass flow rate [kg/s] m c cooling fluid mass flow rate [kg/s] Q IN delivered heat flow rate [kw] ƞ B boiler (burner) efficiency [%] ƞ ORC ORC efficiency [%] ƞ NET net system efficiency [%] N S specific speed [-] V 2 volume flow rate at the turbine outlet [m 3 /s] turbine isentropic enthalpy drop [J/kg] h 02s The indices 1, 2, 3, 4, 1h, 2h, 1c, 2c, 23, 41, 12h and 12c correspond with the control points of the cycle, the indices min and max denote respectively the minimum and maximum allowable values An alternative for these technologies can be the use of micro ORC systems. Recently, these kinds of micro CHP units have been the subjects of study of many research teams which partially confirms that this topic is up-to-date and attractive [8 12]. Designing of such a unit for a domestic boiler is not easy as these systems must fulfill many requirements such as safety, low cost, durability, low noise levels, maintenance-free operation, good dynamic properties, compact size etc. In case of the electrical power of 1 kw to a few kw, due to small volume flow rates of the working fluids, a volumetric expander is usually applied. An efficient solution is the use of, among others, the scroll expanders [13,14]. The advantage of these devices is a good efficiency even for very small volume flow rates, and also low rotational speed. The main drawbacks are high vibration and noise level, relatively large overall

3 Piotr Klonowicz et al. / Energy Procedia 00 (2017) dimensions and the wear of the working scrolls. An interesting option is the use of a microturbine which has excellent properties from the point of view of noise, vibrations and overall dimensions. However, in order to achieve acceptable efficiency levels the rotational speeds must be very high so that the specific speed values are reasonable. This leads to the problem of the electric generator and the rotor support. High speed generators have recently become relatively easily available and with an appropriate design the electric losses caused by high frequencies can be maintained at reasonable levels [15,16]. As many research works show a very effective way of supporting the rotor is to use gas bearings which can operate even at ultra-high rotational speeds [17 19]. The presented work describes some examples of micro ORC units for domestic applications but it focuses on the corresponding turbine selection and design. 2. Examples of micro ORC units One of the most important tasks influencing not only the efficiency but also the economic aspects is the selection of the suitable working fluid. This choice depends largely on the parameters of the ORC system. In order to achieve high efficiencies the working fluid temperature should be as high as possible. On the other hand the domestic boilers usually produce hot water at temperatures below 90 C which in fact is not feasible for cogeneration. To solve this problem it is possible to design a special boiler which uses for example a thermal oil loop. This approach is a common practice in case of larger ORC units. Oil loop, however, is an additional complication for a domestic system. Alternatively, a water loop can be applied but this requires relatively high pressure within the loop which may be problematic. An attractive solution could be the direct evaporation of the working fluid in a fuel burner as it could simplify the system, limit the number of heat exchangers and other supporting devices such as the loop circulation pump. It must be underlined, however, that the main problems associated to direct evaporation within the burner include the thermal stability of the working fluid which could be locally overheated and undergo chemical decomposition as well as the risk of a leakage of the working fluid inside the burner which could potentially lead to an explosion. These considerations are, however, outside of the scope of the paper and the presented results for directly heated ORC unit should be regarded as a theoretical possibility. 3. Thermodynamic cycle analysis The schematic of thermodynamic cycle of the micro ORC unit has been presented in Fig. 1 together with the T-S diagram which shows the thermodynamic process of the system. It is a unit without regeneration because such a system is simpler and more dynamic, although one can consider also a regenerated case. Fig. 1. Schematic of the micro ORC system. According to the presented nomenclature and control points the output power of the turbine, the driving power of the pump and the ORC net power can be calculated as: P T = m (h 1 h 2 ) = m η T (h 1 h 2s ) (1) P P = m (h 4 h 3 ) = m (h 4s h 3 )/η P (2) P NET = P T P P. (3)

4 4 Piotr Klonowicz et al. / Energy Procedia 00 (2017) The heat transfer within the evaporators both for the direct evaporation and the heating loop as well as for the condenser has been presented in Fig. 2. Fig. 2. Heat transfer in the evaporators - (left) direct evaporation, (middle) thermal fluid loop; heat transfer in the condenser (right). In case of both evaporation methods the definition of the pinch temperature will be different. In case of the direct evaporation it is: T PPE = T 2h T 4 while in case of the water or oil loop it is (4) T PPE = T 12h T 41. (5) In both cases the condenser pinch temperature will be T PPC = T 23 T 12c (5) The mass flow rates of the heating, cooling and working fluids can be calculated on the basis of the delivered heat flow rate: Q IN = m h(h 1h h 2h ) = m (h 1 h 4 ) (6) m c(h 2c h 1c ) = m (h 2 h 3 ). (7) It can be assumed that the direct evaporation variant and the thermal oil variant will reveal similar performances as both systems can be designed for high temperature working fluids. In case of water loop a lower temperature fluid must be selected in order to limit the water pressure. The heating water temperature was set to 180 C. The boiling pressure at this temperature is equal to about 10.5 bar(a). The ORC efficiency can be defined as: η ORC = P NET /Q IN (8) and the net system efficiency can be expressed as η NET = η ORC η B. (9) In case of the direct evaporation case the burner efficiency can be calculated on the basis of the parameters of the fumes but for simplicity the efficiency of the burner and the boiler will be assumed as equal to 80%, the temperature of the hot fumes is set to 800 C. The delivered heat flow rate in both cases is equal to 20 kw. For the water cycle, the evaporator pinch temperature difference is equal to 10 K and for both cases the condenser pinch temperature difference is also 10 K. The efficiencies of the turbine and the pump are set to 60% and 40% respectively. Superheating and subcooling in both cases are equal to 5 K. It was assumed that the maximum evaporation pressure does not exceed 15

5 Piotr Klonowicz et al. / Energy Procedia 00 (2017) bar(a), and the minimum condensing pressure is greater than 0.2 bar(a) as it can be very difficult to achieve the values below that in such a small system. The minimum temperature of the cooling water leaving the condenser was set to 55 C so that it could be used for heating purposes. The vapour temperature T 1, which was the decision variable, was selected so that the best efficiency under the given assumptions is achieved. The thermodynamic properties of the fluids were obtained from the NIST Refprop database [20]. The assumptions used for cycle analysis were gathered in Table 1 and Table 2. Some exemplary results for both cases have been presented in Table 3 and Table 4. Table 1. Most important assumptions for the high temperature system. Fluid Q IN [kw] T 1h [ C] T 1c [ C] ΔT PPC [K] ƞ T [%] ƞ P [%] ƞ B [%] T 2c_min [ C] P 3_min [bar] P 1_max [%] MM MDM Table 2. Most important assumptions for the low temperature system. Fluid Q IN [kw] T 1h [ C] T 1c [ C] ΔT PPC [K] ΔT PPE [K] ƞ T [%] ƞ P [%] ƞ B [%] T 2c_min [ C] p 3_min [bar(a)] p 1_max [bar(a)] R acetone cyclopen Table 3. Results for the high temperature system. Fluid m [kg/s] p 1 [bar] T 1 [ C] T 2 [ C] p 3 [bar] T 3 [ C] P T [kw] P P [kw] P NET [kw] ƞ ORC [%] ƞ NET [%] MM MDM Table 4. Results for the low temperature system. Fluid m [kg/s] p1 [bar] T 1 [ C] T 2 [ C] p 3 [bar] T 3 [ C] P T [kw] P P [kw] P NET [kw] ƞ ORC [%] ƞ NET [%] R acetone cyclopentane It can be seen that the efficiencies for the low temperature system are slightly higher. The best system net efficiency was obtained for the acetone 8.5% and the worse for MDM 6.1. However, if a regenerator was added the efficiency of the high temperature variants would be superior. The overall system performance can be more precisely estimated if more detailed expansion unit analysis is performed. 4. Turbine selection This part is devoted to preliminary design of the turbines dedicated to the systems described above. This task is considered particularly problematic due to the low volume flow rate not only because of the high rotational speed required to maintain reasonable values of the specific speed but also because of technological aspects connected with manufacturing of the very small flow channels. Specific speed is a parameter linearly dependent on the rotational speed. It also depends on the volume flow rate and the isentropic enthalpy drop in the stage N S = n V 2 3/4. (10) h 02s

6 6 Piotr Klonowicz et al. / Energy Procedia 00 (2017) Optimal values of this parameter will be between 0.05 and 0.1 depending on the volumetric expansion ratio [21]. It can be argued that for such small system multistage designs are not convenient not only because of the high costs but also due to poor efficiency benefit. In fact, the overall efficiency can be inferior to that of a single stage due to the relatively large stator clearances which will result in significant internal leakages as only the first stator can be designed with zero clearance. The next stages require labyrinth sealing between stationary and rotating elements. Nonetheless, these kinds of designs for very small power applications have been considered in the past [22]. An interesting kind of design is a single stage reaction radial inflow turbine with axial outflow. These types of stages offer high efficiencies and also have been considered for micro ORC systems [19]. The manufacturing process of such a stage requires high precision and is also difficult because of necessity of creating a three-dimensional geometry of the impeller. For the presented application a partially admitted impulse stage seems to be a reasonable choice. Partial admission allows one for creating larger throats and for imposing a rotational speed limit. Below certain values of specific speed the secondary losses become very significant but they can be reduced by partial admission. Unfortunately, in a partially admitted stage additional losses occur so there will be always a certain admission size at which the efficiency is the highest. The unsupplied part of the rotor will produce the pumping loss. The pumping loss can be reduced if the peripheral speed (or velocity ratio) is decreased. Thus, it is possible to find the minimum loss point at certain admission size and velocity ratio value through optimization. The described approach was applied for the presented cases through a mean line design method [23]. To account for the losses within the blade channels the Traupel [24] model has been adopted. It includes the profile, secondary, trailing edge, shock and partial admission losses. The assumed reaction is 10% (impulse stages) so that partial admission is possible in highly reaction stages the expansion losses (another component of the total partial admission loss) are beyond acceptable. In the presented cases shrouded rotors have been assumed. The nozzle angle is set to 15 which is close to optimal for these kinds of stages [25,26]. The blade chords and the trailing edge thicknesses are proportional to the square root of the mean diameter. It means that for the larger stages the aspect ratios get relatively larger and the trailing edges become relatively thinner. In this approach, for a given specific speed, the optimal design point can be found. The examples of turbines designed for MM, acetone and cyclopentane for various rotational speed along with their efficiencies and admission sizes are shown in Fig. 3, 4 and 5. Fig. 3. Turbines designed for MM. Fig. 4. Turbines designed for acetone.

7 Piotr Klonowicz et al. / Energy Procedia 00 (2017) Fig. 5. Turbines designed for cyclopentane. The pressure ratio of the expansion units is an order of magnitude higher for the high temperature systems which influences the turbine design and performance. First of all, the stages with higher enthalpy drop require higher peripheral speeds so they are in general more structurally loaded. Also, the higher the enthalpy drop is the larger the maximum Mach number in the flow is and the higher the losses connected with the shocks are. On the other hand, the systems operating with low condenser pressure, e.g. systems with MM or MDM as working fluids, are characterized by relatively high volumetric flow rates which has a positive impact on the efficiency as the specific speed at given rotational speed is higher. It can be seen that the efficiencies and admission sizes change with rotational speed and the best efficiency is always obtained at full admission for the highest speed. Moreover, the best efficiency is obtained for MM 71% and the worst for acetone 46%. The MM turbine is characterized by a good performance even at relatively small speed. The calculated values were used in the thermodynamic model. The efficiencies of the low temperature cycle were higher under the assumption of unified turbine efficiencies. The results for highest calculated values of turbine efficiency show the net efficiencies of the MM, acetone and cyclopentane cycles are 9.5%, 8.8%, 8.6% respectively. 5. Conclusions The presented work shows examples of domestic micro ORC systems for a few working fluids along with the thermodynamic and microturbine analyses. It must be underlined that the final choice of the working fluid is not only related to the performance but also strongly depends on aspects such as the fluid price, toxicity, explosiveness, availability, environmental impact and other. It was shown that it is feasible to design turbines for very small power that achieve acceptable efficiencies. Examples of such microturbines have been presented for MM, acetone and cyclopentane. For acetone and cyclopentane it was possible to design a turbine with efficiency of 62% at the rotational speed of 200 krpm. For MM the best turbine efficiency was equal to 71% at the rotational speed of 100 krpm. The results show that the analysis of the expansion unit is complementary to the thermodynamic analysis as it allows one for a better estimation of the system performance. Under the assumption of unified turbine efficiency equal to 60% the net system efficiencies were equal to 8.0%, 8.5% and 8.3% for MM, acetone and cyclopentane respectively. The system efficiency for the best turbine efficiencies (71%, 62% and 62%) was equal to 9.5%, 8.8% and 8.6% (MM, acetone and cyclopentane) so the fluid considered previously the worst of these three actually proved to perform the best. Acknowledgements This work has been founded by The Polish Agency for Enterprise Development and from The Smart Growth Operational Programme (European funds) within the project No. POIR /15 carried out jointly with the SARK Company. References [1] J. Kamiński, A. Malik, An analysis of the Polish district heating sector: present situation and key issues for further development, Bull. Miner.

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