Category Archives: Particle Image Velocimetry (PIV)

Structural synthesis method

The synthesis is based on Multibody Systems Method (MBS) according to which a mechanical system is defined as a collection of bodies with large translational and rotational motions, linked by simple or composite joints [5]. The functional design process at structural level consists in the following stages:

♦ Identification of all possible graphs on basis of the following input data:

— spatiality of the multibody system;

— type of the geometrical constraints gc (simple or/and compound);

— number of bodies nb;

— the mobility of the multibody system M.

♦ Selection, from multitude of the identified graphs, of the graphs that are admitting supplementary conditions imposed by the specific field of utilisation.

♦ Successive transformation of the selected graphs into mechanisms by:

— mentioning the fixed body and the role of the other bodies (ex.1-fixed body, 2-input body, 3-output body etc.);

— identification of distinct graphs versions based on the preceding particularisation;

— transformation of these graphs versions into mechanisms by mentioning the types of constraints gc (rotation, translation etc.) [2], [7].

The graphs of the multibody system are defined as a features based on the modules introduced in the next figure and are considering the number of bodies and the relationships between them.

The identification of all possible graphs starts with definition of the types of the geometrical restrictions between the bodies considering the chosen space S (gc, min= 1, gc, max = S-1) [9]. In the Fig.4 the notations “R” and “T” represent restriction rotation type and respective translation restriction type. All the other notations represent composite joints as combinations of the ones mentioned before.

Fig.4 Restriction types

For example, in the planar space (S = 3), all the possible graphs can be designed using the restrictions types from Fig.4, where gc= 1 (Fig.4.a), gc = 1+1 (Fig.4.b), gc = 2 (Fig.4.c), considering the correlations between the number of bodies nb, the mobility M and the sum of the geometrical constraints Igc.

The relation between M, S, nb, Zgc is [8.]:

M=S(nb-1)-Zgc (1)

In respect with the relative motions of the sun on the sky dome the mobility of the mechanisms that orients the receiver of the conversion system deals with a degree of freedom equal with 2.

The Site

The performance of the system strongly depends on the site chosen. The performance analysis in this study is performed for a site close to Seville (Spain). Time series for the ambient temperature and the direct normal irradiation (DNI) have been derived from satellite data. The yearly sum of the DNI for the site is 2012 kWh/m2. The distribution of the mean monthly DNI in W/m2is displayed in figure 3.





800,00-1000,00 □ 600,00-800,00

□ 400,00-600,00

□ 200,00-400,00 0,00-2 00,00



Hour [-]

Figure 4 displays the distribution of the mean monthly ambient temperature. Since a dry cooling condenser is chosen for the power plant the ambient temperature affects the condenser temperature of the power plant and thus its efficiency. Accordingly the power block efficiency will decrease in summer due to the higher ambient temperature.

The parabolic troughs used for the power plant cannot use the complete DNI but only

the DNI multiplied by the cosine of the incident angle. Figure 5 displays the corrected DNI sorted according to the number of hours they occur. For the simulation it is assumed that the power plant is only operating at a corrected DNI higher than 250 W/m2. According to figure 5 this limit is exceeded for 2770 hours per year. Thus the useful solar energy is reduced to 1726 kWh/m2a compared to 2012 kWh/m2a. Plant operation for values lower than 250

W/m2 is not profitable due to thermal losses in the solar field and high parasitic losses in the whole plant.

Experimental set-up

The thermal solar collector was of a direct-flow design; which consisted of tubes with a co­axial heat transfer conduit connected to a manifold with a parallel inlet/outlet pipe configuration. Twenty evacuated tubes were installed equating to an absorber area of 2.046 m2, with vacuums in the order of 10-5 mbar. Filtered water was the heat transfer fluid used throughout this study, with no anti-freeze component being added to the system. All data was recorded using Thermomax’s in-house solar simulator as shown in Figure 1. Thirty-six Tungsten Halogen lamps (3500 K) were used to produce 18 kW of power to irradiate the collector surface. This system was capable of generating average irradiances in the range 200 to 1500 Wm-2. A correction factor for the spectral discrepancy between the solar simulator and natural sunlight was applied using the effective transmittance — absorptance product method4 and the cool sky was used to minimise the influence of
thermal irradiance on the collector. The solar simulator rig (i. e. solar cradle and cool sky) was capable of a 90° rotation between the vertical and horizontal planes. Acquisition of the data was recorded on in-house software program designed using National Instruments LabVIEW 7, where the deviation of the measured parameters was consistent with EN12975-2.

The Progress of the AndaSol projects in Spain

Rainer Kistner, Milenio Solar S. A., Klaus Grethe, Milenio Solar S. A., Michael Geyer, Flagsol GmbH, Henner Gladen, Solar Millennium AG, Jose Alfonso Nebrera, Cobra S. A

The general scope of the AndaSol projects is the erection and operation of the first large-scale parabolic trough power plants in Europe. Each AndaSol project represents a fully dispatchable capacity of 50 MWe with a highly efficient steam power cycle, combined with a solar field consisting of concentrating parabolic trough collectors as the power source and a 6 hour thermal storage system.

The AndaSol projects are developed and promoted by the Solar Millennium Group and its Spanish subsidiaries, Milenio Solar S. A. and AndaSol-2 S. A., in close cooperation with the Spanish ACS-Cobra Group. Since the publication of the Spanish Royal Decree 2818, that regulates the purchase and compensation of renewable electricity installations, in December 1998 the Solar Millennium group has screened high potential sites in Southern Spain and has identified the Marquesado de Zenete in the Province of Granada as one of the most promising sites for the implementation of the first commercial solar thermal power plants with parabolic trough collectors in Europe. With the financial support of the German Ministry for Environment, a three-year’s solar radiation measurement campaign has been conducted since March 2000, indicating an excellent annual direct normal radiation level of approximately 2.200 kWh/m2/year.

The conceptual design of the AndaSol plant was fully developed by Solar Millennium AG and its subsidiary Flagsol GmbH from 2000 until 2002 and served as basis for the subsequent preparation of the permitting applications for the administrative authorization, land use permit, construction license, environmental impact declaration and grid access application.


Anica Trp, Kristian Lenic, Bernard Frankovic
Faculty of Engineering University of Rijeka, Vukovarska 58, HR-51000 Rijeka, Croatia
Phone: + 385 51 651 514, Fax: + 385 51 675 801, E-mail: anica. trp@riteh. hr

Latent thermal energy storage system of the shell-and-tube type with the phase change material (PCM) filling the shell side, which is used as a heat storage device in solar heating applications, has been analysed numerically and experimentally in this paper. The heat transfer in this type of thermal energy storage system is a conjugate problem of transient forced convective heat transfer between the heat transfer fluid (HTF) and the wall, heat conduction through the wall and solid-liquid phase-change of the PCM. Since phase-change heat transfer is non-linear due to the moving phase change boundary, analytical solutions are only known for a few problems with simple geometry and simple boundary conditions. Numerical methods provide a more accurate approach, so various techniques have been developed. Many authors [1] — [5] have used the enthalpy formulation, in which the energy equation for PCM is written in terms of enthalpy. A transient heat transfer phenomenon in a shell-and-tube latent thermal energy storage system has been studied numerically by Bellecci and Conti in [6] — [8], and numerically and experimentally by Lacroix in [9] and [10]. These authors have used an enthalpy method for solving phase-change heat transfer and employed standard empirical correlations to calculate convective heat transfer coefficient. Cao and Faghri in [11] and [12] have simulated numerically, using the temperature transforming model for phase-change heat transfer, the transient behaviour of the shell-and-tube thermal energy storage system employing a low Prandtl number HTF. Ismail and Abugderah in [13] have modelled numerically a phase change thermal energy storage system of the same type. In both papers, the transient HTF momentum and energy equations were solved simultaneously with the tube wall and the PCM energy equations, as one domain, in order to avoid the errors due to the use of empirical correlations. Zhang and Faghri in [14] have semi-analytically studied a shell-and-tube latent thermal energy storage system employing a moderate Prandtl number HTF. They concluded that the laminar forced convective heat transfer inside the tube never reaches a thermally developed state and must be solved simultaneously with the phase-change of the PCM, so the application of CFD methods is required. In [15] a transient phase-change heat transfer with conjugate forced convection in the shell — and-tube latent heat storage unit, with water as HTF and calcium chloride hexahydrate as PCM has been analysed numerically. In this paper, a transient heat transfer phenomenon during charging and discharging of the shell-and-tube latent thermal storage unit has been analysed numerically and experimentally. The mathematical model has been formulated. The enthalpy method for modelling phase-change heat transfer has been used. The dimensionless conservation equations for HTF, wall and PCM, with initial and boundary conditions, have been discretised by fully implicit control volume approach, that has been implemented in developed fOrtRAn computer code, and solved simultaneously using an iterative procedure. Numerical model has been validated with experimental data. Series of numerical calculations have been performed in order to analyse transient phase — change heat transfer during charging and discharging of the latent thermal energy storage unit.


dWX dWX dWX dP 1

дт x dX R dR dX Re

The governing differential equations of the HTF, the tube wall and the PCM, with initial and boundary conditions, have been solved as one domain. The computational domain has been discretised by the control volume approach and the SIMPLER algorithm. The formulation has been fully implicit in time and the convection-diffusion terms have been treated using power-law scheme. Algorithm has been implemented in developed FORTRAN computer code and the resulting discretisation equations have been solved simultaneously using an iterative procedure. Time-wise temperature distributions of HTF, tube wall and PCM have been obtained by numerical calculations and transient phase — change heat transfer behaviour during charging and discharging of the shell-and-tube thermal energy storage unit, i. e. melting and solidification of the PCM has been simulated.

Solar-Hybrid Gas Turbine Technology

Solar gas turbine systems use concentrated solar power to heat the pressurized air in a gas turbine before entering the combustion chamber (Fig. 1). The solar heat can therefore be converted with the high thermal efficiency of a modern recuperated or combined gas turbine cycle. The combustion chamber closes the temperature gap between the receiver outlet temperature (800-1100°C at design point) and the turbine inlet temperature (950- 1300°C) and provides constant turbine inlet conditions despite fluctuating solar input. The solar power tower technology is used with concentration ratios up to 1000 suns to achieve the high receiver temperatures.

Status of Solar-Hybrid Gas Turbine Technology

A pressurized volumetric air receiver with secondary concentrator has been developed and successfully tested as the so-called REFOS receiver technology in the scope of several German national and international R&D projects [2], [3]. In 2002, three receiver modules have been coupled in series to a 240 kWe gas turbine and successfully operated at receiver temperatures of up to 800°C [4]. More detailed information about the receiver development and recent test results with receiver temperatures up to 960°C can be found in [3]. This section gives an overview over the recent results.

To make use of this promising technology a high temperature volumetric pressurized receiver had been developed in the past by DLR and successfully tested up to a temperature of 800°C and 15 bars [5]. Within the SOLGATE1 project a test system was operated, consisting of a solarized gas turbine connected to a pressurized receiver. The system was set up at the solar tower test facility of the Plataforma Solar de Almeria (PSA), Spain, at 60m height. The test setup consists of 3 receiver modules, a helicopter engine with modified combustor and a generator. In the receiver cluster the air from the compressor of the turbine is heated from 290°C to 1000°C with solar energy. A bypass allows to reduce the temperature at the combustor inlet to 800°C, due to current limitations of the combustor design. The combustor can be fed also with gas. The two-shaft turbine is driving the compressor and the generator which is connected to the grid.

Power Conversion Unit

The Power Conversion

Unit included the gas turbine, a converted helicopter engine and the balance of plant systems.

The gas turbine was installed on a stationary skid and integrated with a generator and auxiliary systems (Fig. 3). A separate oil cooling system and a new electric and control system were added as

Co-funded by the EC under contract no. ENK5-CT-2000-00333.

well as a more powerful starter. The new super alloy combustor included larger diameter inlet flanges to allow for air to enter at over 800°C compared to 300°C in the original combustor. The injector and the igniter were purposely designed to allow continuous operation at elevated air inlet temperatures. Furthermore, the injector had to function at a wide range of air-to-fuel ratios without flameout. A fuel shut off valve and a metering valve were added for fuel flow control. The solar test loop introduced the receiver array between the compressor and the combustor, the dead volume was increased from a few litres to about 3 m3 and the pressure drop was also significantly higher. This had to be taken into account in the control logic.


The receiver is designed for the actual available parabolic trough collectors, e. g. the EuroTrough, US-Trough or LS-3. Therefore the dimensions of the absorber are predetermined by roughly 4 m length and 70 mm diameter. Vacuum insulation and a highly selective absorber coating are used to achieve high efficiency. The most significant innovations are the design of the bellows, the glass-to-metal seal and the anti reflective coating.

Bellow Design

To reduce shading of the absorber tube by the bellows, a new design for the components at the tube ends was developed. Compared to the existing design, where bellow and glass-to-metal seal are placed one after another here the components are placed on top of each other. The bellow is embedded between the glass tube and the absorber tube, the length of the whole component is significantly reduced. An active length of the receiver of more than 96% is achieved, which is at least 2% more compared to existing designs.

Bellow design of the SCHOTT Receiver

Another advantage of the design is the protection of the glass-to-metal seal (GMS) from concentrated solar radiation. Even when beeing shielded with an aluminium envelop from the outside, radiation can hit the GMS at the northern end of the receiver particularly in winter time from the inside. This radiation can cause high temperature gradients at the GMS. To correct this problem, a mirror has been placed at the end of the bellow.

Glass-to-metal seal (GMS)

To ensure vacuum stability over the hole lifetime of the product, a glass-to-metal seal with high mechanical strength and temperature resistance is needed. This implies a stress free bonding between the metal part and the glass part of the GMS. Using materials with different coefficients of thermal expansion (CTE) can produce stress "frozen” in the glass during manufacturing. Thermal stress can also evolve from temperature changes during operation. Breakage of the glass-to-metal sealing is the main cause for damages in existing power plants [3].

To solve this problem SCHOTT developed a material combination for the glass-to-metal seal with adjusted CTE’s of both partners in the range of 5 x 10-6 K -1. The materials used are a standard nickel alloy and a new developed borosilicate glass with extremely high optical transmittance.

In this way a reliable sealing of the glass envelope with the metallic bellow is ensured even under extreme temperatur changes. The low thermal stress allows to plug a 1 mm thick metal tongue in to the glass thus giving the glass-to-metal seal high mechanical strength.

Structural synthesis of sun-tracking mechanisms

General criteria: M = 2, nb = 3 or 4, S = 6.

Specific criteria:

— rotational type restriction perpendicular to the equatorial plane, between the base and the output body (PV panel or collector absorber);

— direct link between base and the input/output body is rotational.

The specific criteria lead us to reduce the synthesis of the system to a mechanism with M=1, because one of the driving constraints is defined by a specific criterion. Consequently, the mechanism will be analysed for only one motion which covers the variation of the sun altitude generated by the Earth’s precession motion.

In this case, the synthesis is reduced to a planar space S = 3, the mobility will be considered as M=1 and the number of bodies decreases to nb = 2 or 3.

This planar space is defined as a plane perpendicular on the equatorial plane and determines the mechanism to orient the panel in concordance with the variable altitude of the sun along the year.

Starting from the fundamental structures with nb = 2, M = 1or nb = 3, M = 1(Fig.5) possible graphs are identified.

Fig.5. Possible graphs for mechanism’s configuration

In respect to the specific criteria, the proper graphs were selected and presented with the associated mechanisms in Fig.6.


Fig.6. Selected graphs and associated mechanisms

In agreement with the structural schemes obtained after the utilisation of the synthesis method, several technological solutions are presented bellow.

Fig11. Technological solution for the structure from Fig.6.4

Thermodynamic Analysis of the Power Plant

For the thermodynamic analysis of the steam cycle the commercial simulation program IPSEpro® is used. IPSEpro® has a model library containing all components necessary for the simulation of a conventional power plant. This library has been extended by DLR by all relevant solar components.

Conceptual Design of the Power Plants

Collector fields for the superheated steam driven power cycle and the saturated steam cycle are presented and investigated. The design calculations are performed for the site in Seville for the 21st of June at noon with a direct normal irradiation of 850 W/m2.

Power Plant with Superheated Steam

The schematic diagram of the power plant with superheated steam is displayed in figure 6. Seven collector loops are connected in parallel. Due to the symmetry of the collector field only 4 loops are displayed and calculated. The mass flux is multiplied by 1.75. The collector field is operated in recirculation mode. In the superheating section there are two collectors connected in series, while in the evaporation section eight collectors are connected in series. The two sections are subdivided by a separation drum. One injection nozzle per superheater row is used to control the outlet temperature of each loop. A fraction of the feed water is fed in front of the recirculation pump in order to cool down the recirculated water to avoid cavitation. The mass flux feeded in front of the recirculation pump is controlled to maintain a temperature of approx. 30 K below the according saturation temperature. The steam quality at the separator inlet is set to 0.85. This is enough to guarantee a sufficient cooling of the absorber tubes during steady state operation. During cloudy periods where frequent transients are expected the steam quality should be significantly lower. The steam temperature at the collector outlet (not displayed in figure 6) is set to 400°C. The mechanical efficiency of all pumps is set to 0.97 and the pump efficiency to 0.62. The electrical and mechanical efficiency of all driving motors is set

to 0.98. The isentropic efficiency of the high pressure turbine stage is 0.71 and that of the low pressure stage 0.78. The generator efficiency is 0.96.

Evaporator [Superheater

Power Plant with Saturated Steam

The schematic diagram of the saturated steam power plant is displayed in figure 7. In this case nine parallel collector loops with eight collectors in series are needed. Again due to the symmetry of the collector field only five loops are displayed and calculated. In the end the mass flux is multiplied by the factor 1.8.

The outlets of the parallel rows collectors are connected to a main header entering a separator drum. Again the steam quality at the separator inlet is set to 0.85. The saturated water from the separator is recirculated to the collector inlet whereas the saturated steam is fed to the steam turbine. For the simulation the same component efficiencies have been used as for the superheated steam process.

Design Condition

The design calculations are performed for both options for the specified site for the 21st of June at noon and a DNI of 850 W/m2. The incidence angle for that time and site is 14°. The most important design efficiencies for the two investigated options are listed in table 2.

Table 2: Design performance of the saturated and superheated steam processes.

Superheated Steam

Saturated Steam

Solar Field Efficiency [%]



Power Block Net Efficiency [%]



Total Net Efficiency [%]



The thermal efficiency of the solar field is the thermal output of the solar field divided by the DNI on its total aperture area. In case of the saturated steam process the highest fluid temperature in the collector field is the saturation temperature (285°C at 70 bar). Since the collectors of the superheating section of the superheated steam process are operated at
temperatures above the saturation temperature the thermal losses of the superheated steam collector field are higher. Accordingly the efficiency of the saturated steam collector field is 1.8 % higher.


The second row displays the net efficiency of the power block that is defined as the net electrical output of the power plant divided by the thermal output of the solar field. Due to the higher life steam parameters of the superheated steam process the net efficiency of the power block is 1 % higher. The low values of the power block efficiencies are mainly caused by their simplicity. Finally the net efficiency of the superheated steam power plant is 0.2 % higher compared to the saturated steam process. For a final assessment of both options their part-load performance will be investigated in the next section.

Experimental results

The study of the standard PV system and the hybrid PVT/AIR systems includes outdoor tests for the determination of the steady state electrical efficiency pel of the corresponding PV modules of all systems and the thermal efficiency pth of the PVT/AIR models. The electrical efficiency pel of PV modules depends on their temperature (TPV) and the incoming solar radiation and it is calculated by the measured data as: pel=ImVm/GAa, where Im and Vm are the current and the voltage of Pv module operating at maximum power and G the irradiance on the system aperture plane. Test results showed that the pel for PV, PVT/UNGL and PVT/TFMS type models was: r|el=0.1659-0.00094Tpv, while for PVT/GL type models was: pel=0.1457-0.00094TPV. The value of TPV for the standard module was calculated by the equation: TPV=30+0.0175(G-300)+1.14(Ta-25) (Lasnier and Ang, 1990) that correlates TPV with the parameters G and Ta. The TPV of the corresponding modules of PV+REF, PV-TILT and of all PVT/AIR systems was based on modified formulas of the above equation, which give approximately the PV module temperature and are presented in Table 1. These modified formulas were experimentally validated and correspond to the increase of PV operating temperature due to the reduced heat losses to the ambient.

The thermal efficiency nth of the PVT/AIR models depends on the incoming solar radiation (G), the input air temperature (Ti) and the ambient temperature (Ta). During tests for the determination of system thermal efficiency, the PV modules were connected with a load to simulate real system operation and to avoid PV module overheating by the solar radiation that is converted into heat instead of electricity. The steady state thermal efficiency i"|th of the tested hybrid PVT/AIR solar energy systems is calculated for the operating conditions with the lowest thermal losses (T=Ta) by the equation: i"|th = m Cp(T0-Ti)/GAa, where m is the fluid mass flow rate, Cp the fluid specific heat, Ti and To the input and output fluid temperatures and Aa the aperture area of the PV/T model. The results are presented in Table 1 for all studied systems. The experimental results for the thermal efficiency of the PV-TILT and PVT/AIR-TILT type systems were extracted from the tests where an additional thermal insulation sheet was mounted on the back of these systems, to simulate the tilted roof. In the calculation of the electrical and thermal output of the compound systems PV+REF and PVT/AIR+REF, we considered the net solar radiation on the aperture surface of PV modules and not the additional on the reflector, in order to have direct comparative results with the standard installation mode of the systems. The calculation of thermal efficiency i"|th (for T=Ta) of PVT+REF systems varies from a minimum value for December (CR=1.0) up to the maximum value for June (CR=1.3).