Printed Edition of the Special Issue Published in Energies

Studying the structure of the lip seal is important to minimize coolant flow and maximize the sealing effect. In the wave-shaped lip seal, the web number of the radial seal is the same as that of the conventional lip. In the remaining parts of the wave-shaped lip seal, the same mesh as that used for the conventional lip seal is used.

Figure 1. Schematics of the flow phenomenon around the rim seal and the concept of wave-shaped rim seal: (a) Flow phenomenon near the rim seal; (b) Pressure differences causing the egress and ingress flow

Results and Discussion

In the radial slot, the circumferential velocity (Vcir) is dominant due to the rotation of the rotor. For Type 2 W, the modified pressure field in the radial seal weakens the strength of the outlet and inlet flow. The CCO2 contour of the rotor disc is described in region A (near the common path) and region B (near the radial seal of the rotor) (marked in Figure 11a(i)).

Figure 7. Circumferential velocity contour in wheelspace of Type 1. (a) Locations of Planes 4, 5, 6, and 7

Conclusions

Note on a class of solutions of the Navier-Stokes equations representing steady rotationally symmetric flow.Q. In Proceedings of the Mathematical Proceedings of the Cambridge Philosophical Society; Cambridge University Press: Cambridge, United Kingdom, 1953; Volume 49, pp. Aerodynamic aspects of sealing gas turbine-rotor-stator systems: Part 2: The performance of simple seals in a quasi-asymmetric external flow. Int.

Gas Turbine Cycle with External Combustion Chamber for Prosumer and Distributed

Energy Systems

Introduction

R&D work is carried out in order to optimize the biochemical processes of the bioreactor depending on the type of plant biomass and the use of biogas in the fuel cell [45,46]. In the case of an external combustion chamber, it is possible to vent some of the turbine exhaust air and omit a combustion chamber. In the case of an external combustion chamber, it is thus possible to vent some turbine exhaust air and omit a combustion chamber.

Figure 1. Turbine set operating according to the open cycle with regenerator (Variant 1-V1)

Materials and Methods 1. Nomentclature and Units

For small power stations (from a few kW to a few hundred kW) a maximum temperature of 900◦C in front of the turbine has been assumed and a low efficiency of the components has been assumed. The results of thermodynamic calculations of the cycles with gas turbines using fuels with different calorific values ​​were presented in the paper. The following main relationships have been used in the analysis, including energy balances and definitions for the efficiencies.

Table 3. Assumptions adopted for the design analysis of turbine generator variants [27,51].

Results

The influence of the cycle structure and the calorific value of the fuel on the relative efficiency (with respect to methane) of both variants of turbine aggregates. A direct comparison of the efficiencies of the two cycle variants of the turbine assembly (V1 and V2) is interesting. The influence of the calorific value of the fuel on the efficiency ratio of the cycles with the external combustion chamber and bypass (V2) and after with the regenerator (V1).

Figure 3. Effect of compression ratio

Final Conclusions

The use of fuels with a lower calorific value usually leads to a lower efficiency of the power plant. Simulation of the operation of a spark ignition engine with different fuels and its contribution to technology management. Sustainability. Technical and logistic analysis of the extension of the energy supply system with the cogeneration unit supplied with biogas from the water treatment plant.

Figure 18. Example of an experimental variant of the innovative isothermal turbine.

Computational Simulation of PT6A Gas Turbine Engine Operating with Different Blends of

Biodiesel—A Transient-Response Analysis

The PT6A Engine Model

For this we apply the continuity balance of the fluid for each part of the engine. This is the case with the combustion model, the fuel system or the external loads associated with the propeller. The application of the balances to the parts of the engine results in a 0-dimensional model representing the operation in a transient state.

This feature directs the airflow into the engine's components in the same direction as the aircraft's displacement. The quotient between specific heats of the air is denoted by γ and ηis the polytropic efficiency. Preparatory work for modeling the compressor has to do with the estimation of its geometry.

This change is only related to the tangential velocities of the air given by the velocity triangle analysis. This affects the maximum temperature in the fuel chamber when the engine is started. This means that the expansion ratio of the gas is known for the turbine section.

Mass compatibility conditions are related to maintaining the air mass flow rate.

Figure 1. Cross section and stations of the Pratt-Whitney PT6A-65 engine.

Numerical Results

It is readily understood that the rotational speed of the axial and centrifugal compressor stages is matched to the compressor's turbine via the engine spool. Finally, we address the transient operation of the engine using the fuel mixtures, specifically at engine start-up when the maximum temperature can be reached. We first validate the computational model by considering the standard operation of the PT6A-65 engine reported in the operation manual [29].

The accumulated polytropic indices and irreversible efficiencies for the compression and expansion sections of the PT6A-65 engine are shown in Table 4. Those values ​​are captured in the energy balances and polytropic processes of the compressor and turbine stages. In any case, the maximum pressures are related to the use of the Jet-A1 fuel.

Regardless of fuel type, the temperature never exceeds a value of 1050 K during idle operation. Finally, we evaluate the engine start-up procedure with the fuel types tested in previous scenarios. We aim to evaluate the starting procedure of the PT6A-65 engine at both 100% and 60% throttle.

Then the fuel intake generates a temperature peak in the combustion chamber, which in all scenarios is the maximum temperature reached during the full operation of the engine.

Table 2. On-ground steady operation conditions using Jet-A1 fuel. Extracted from Ref. [29].

Materials and Methods

General specifications of the test facility and engine operating conditions are shown in Table 1. The engine is then cooled down until the EGT is below 120◦C and the main engine components are visually inspected. The baseline experiments without biodiesel content (B0-RPM38, B0-RPM70, B0-RPM80, B0-RPM100) were selected to evaluate the design operating regimes of the engine using pure Jet A1.

The idling mode (38% RPM) is the state of stabilization after starting the engine without acceleration. During the steady state of the engine at each operating mode, OAT, EGT, ωsha f t, T and Pf. The minimum EGT value occurs between 70% RPM and cruise regimes, which is due to the optimal heat dissipation of the engine [21].

Thus, modification in the hydromechanical properties of the fuel mixture changes the FCU operation, which increases the tension of the needle valves and the fuel injection pressure. 19], an increase in the viscosity of the fuel mixture and the clogging of biodiesel in the filtration system were registered during those previous experiments. The experimental methodology and the facility used enabled the measurement of operating parameters of the engine.

At the cruise regime (80% RPM), a maximum reduction of CO and HC emissions was observed with a maximum variation of 25% and 58%, respectively.

Figure 2. (a) General J69 engine experimental facility detailing peripheral components and instrumentation: (1) turbojet engine model J69T-25A, (2) tachometer, (3) fuel pressure manometer, (4) air inlet duct, (5) air inlet temperature probe set, (6) exhaus

Hardware-In-The-Loop” System

• Description of the “Hardware-In-The-Loop” Apparatus
• Dynamic Simulation Model of the Building and Heat Emission System
• Description of the Case Study
• Results
• Conclusions

From this perspective, some works based on dynamic simulation of the heating and ventilation (HVAC) building system have been proposed in the literature [6,12]. The paper focuses in particular on the effects of the supply temperature to the heat emission system and the resulting on-off signal from the typical zone thermostats located inside the building. The emulator produces a fluid enthalpy variation corresponding to the thermal effect evaluated by the dynamic simulation of the simulated building (see Section 3).

Thermal storage (520 L) is connected to the user circuit and can be used to provide thermal energy for the EU in place of the AHP. The experimental analysis based on the HiL system requires an accurate energy model of the building to drive the EU. However, these quantities and their effects on the thermal evolution of the building are included in the dynamic energy model.

The signature level indicates that different values ​​of the coefficient are used according to the fan speed (generally high, medium and low). Under such conditions, the thermal losses of the water circuit are relevant compared to the heat provided by the AHP (32% in MOD#1 and MOD#2 and 15% in MOD#3). The capacity ratio of the device is evaluated according to the average actual thermal output, Q, over that considered.

Evaluation of the control performance of hydronic radiant heating systems based on the emulation using hardware-in-the-loop simulation.

Figure 1. Components and logical scheme of the building emulator and integrated energy system (BE-IES) “hardware-in-the-loop” apparatus.

Hybrid Fuel Cell—Supercritical CO 2 Brayton Cycle for CO 2 Sequestration-Ready Combined Heat

• Theory
• Experimental
• Results and Discussion 1. Fuel Cell Performance
• Conclusions and Future Work

This section establishes the theory of the overall concept and the individual components of FFCTH. The electrical efficiency of the standard sCO2 cycle (ηSSGT) given in Figure 1 is represented in equation (1) where CO2 is the total CO2 mass flow rate in the cycle. Partial oxidation of the fuel (i.e., methane in this study) and the oxygen mixture sent to the fuel-rich combustor results in the generation of syngas (H2+CO).

The efficiency of the sorption/desorption based air separation unit is believed to be within the acceptable range. Also note that the air separation efficiency is considered acceptable in the design. The initial conditions for air (or oxygen) entering the fuel-rich combustion chamber of the FFC are 1 bar and 298 K.

However, the electrical efficiency of FFCTH is generally higher, as will be shown below. It is obvious that the trend in Figure 7 matches the trend in syngas composition in the fuel-rich exhaust. The electrical efficiency of the FFCTH without sequestration is 6.85% higher than the standard sCO2 turbine setup without sequestration.

The P/H of FFCTH with oxygen (ready for sequestration) is much higher than with air (not ready for sequestration at all)Φ.

Figure 1. Schematic of a standard sCO 2 Brayton cycle with recuperation and recompression showing the various state points in the system.

Design Analysis of Micro Gas Turbines in Closed Cycles

Modelling

These connections are illustrated by comparing the examples of microturbines of 50 kW, 35 kW and 10 kW. The flow parts of the compressors (Figure 5) and the turbines (Figure 6) were designed by CFD (Computational Fluid Dynamics) codes while the heat exchangers (regenerators) and the combustion. The flow parts of the turbine and the compressor were calculated with ANSYS software.

Examples of the distribution of the calculated parameters in the flow channels (velocity and pressure lines) are shown in Figures 5 and 6 for the compressor stage and the turbine stage, respectively. The calculations were performed for a temperature T3 before the turbine equal to T3=850◦C (within the typical range for very small power gas turbines). Figure 8 shows the example of the turbine power plant, while the main parameters of the special variants are shown in table 4 and in figure 9, the compressor rotors and Figure 10, the turbine discs.

The results are shown in Table 4 and the interpretations of the cycles are presented in Figure 11. As a result of the calculations, we can conclude that it is enough to apply a highly efficient gas turbine in a closed cycle and reduce the pressure at the compressor inlet to obtain a power plant with a smaller output but the same high efficiency. In Figure 12, the overall efficiency of the cycles with a regenerator and the upper temperature T3=850◦C is represented as the function of the compressor pressure ratio.

The authors presented the results of the project analysis of micro turbines, which prove that it is possible to increase the efficiency of small power gas turbines.

Figure 3. Scheme of a gas turbine closed cycle with regenerator (a) and schema of an open cycle with a regenerator (b); where: I—compressor, II—turbine, III—generator, IV—high temperature exchanger, V—low temperature exchanger, VI and VII—valves, VIII—gas