Literature review on horizontal collaboration concepts

Ti-htf is the temperature of the HTF entering the generator jacket which has been found from the experimental tests to be in the range between 190 and 200 ^oC. To,htf can be determined by assuming the temperature difference of HTF across the generator to be 6 ^oC. (Po - Pi) and (ηhtf,pump) are the pressure losses in the HTF circuit and the total efficiency of the HTF-pump. These were determined from tests to be 400 kPa and 0.45 respectively. The properties of the HTF can be determined from equations (B.1) to (B.4) in Appendix B.

3.2.1.2 CO2 refrigeration model

The CO2 refrigeration system of the test facility is a volatile-DX system. The basic principle of such system is explained in Section 2.2.1.2.

A simplified diagram of the CO2 system is shown in Figure 3.4. The system comprises an LT CO2 compressor, LT and MT evaporators, a CO2 pump, expansion valve (EXV), regulator valve (RV), liquid receiver, internal heat exchanger (IHX) and a cascade condenser. The condenser is a CO2/brine HX. To provide flexibility to investigate different arrangements for the LT compressor, the discharge line is split into two branches. One branch is connected to the saturated gas line from the liquid receiver and the other feeds directly into the liquid receiver. The figure also shows a bypass circuit from the CO2 pump to the liquid receiver which allows the flow of CO2 refrigerant to the MT evaporator to be varied to enable investigation of the effect of circulation ratio (CR) on the performance of the MT evaporator.

Figure 3.4 Centralised volatile-DX CO2 section

The refrigeration cycle of the system is shown in Figure 3.5. The cycle refers to the schematic diagram in Figure 3.4. It can be seen that the thermodynamic cycle of the system consists of compression of a low pressure, low temperature superheated CO2

vapour to medium pressure level, followed by mixing of the superheated gas from the compression process with saturated CO2 gas from the liquid receiver. The CO2 gas is de-superheated and condensed in the cascade condenser and exits as saturated liquid.

Some of the liquid is pumped to the MT evaporator where evaporation takes place by absorbing heat from air circulating in a refrigerated display cabinet. The CO2 exiting the MT evaporator is a mixture of vapour and liquid. Liquid from the receiver flows through an internal heat exchanger (IHX) before is expanded isenthalpically in the expansion device (EXV) which results in a low temperature vapour-liquid mixture entering the low temperature evaporator. The low temperature two phase mixture evaporates in the LT evaporator removing heat from the LT refrigerated cabinet. The vapour exiting the LT evaporator is superheated in the IHX and then compressed in the LT compressor to complete the cycle.

To simulate the volatile-DX CO2 refrigeration system, some assumptions were made as follows: steady state flow conditions; negligible thermal losses to the environment;

negligible refrigerant pressure drops in the pipes; degree of superheat of the LT evaporator 5 K; CO2 leaving the condenser in saturated liquid; vapour quality at the exit of MT evaporator to be 0.8 (it can be varied depending on the circulation ratio);

isenthalpic expansion in the EXV; negligible refrigerant pressure drop across the

MT

regulator valve (RV). The isentropic efficiency (ηs) of the CO2 pump was assumed to be 0.5 and the overall efficiency (ηo) was predicted from manufacturer’s data to be 0.1. The compression process in the CO2 system was non-isentropic. The isentropic and volumetric efficiency of the compressor can be expressed as a function of pressure ratio (Rp) and can be determined from (Lee et al., 2006).

The electrical motor efficiency of the compressor (ηm) was assumed to be 0.86 (Navarro et al., 2007).

Figure 3.5 P-h diagram of the subcritical CO2 refrigeration cycle

Table 3.2 summarises the main mass and energy balance equations for each component of the CO2 refrigeration system. The properties of CO2 refrigerant were derived from the EES software.

In the integrated system arrangement, the CO2 refrigeration system rejects heat to the secondary loop through the cascade condenser. The temperature of the secondary fluid (brine) should be lower than the temperature of the CO2 refrigerant. This temperature difference is a function of the effectiveness of the heat exchanger. The effectiveness of the cascade condenser and its temperature losses (ΔTcond) can be calculated from:

Pressure (kPa) is absolute pressure

35

The position of each measurement point is shown in Figure 3.4. T3 and T4 are temperatures of CO2 refrigerant entering and leaving the condenser respectively; T35 is the temperature of brine entering the condenser.

Table 3.2 Equations for the CO2 refrigeration system (Numbers of the parameters refer to Figure 3.4)

Components Mass balance Energy balance

LT Compressor m ₁ m₂ m¹H¹ W_comp^._sm²H²

* For simulation purposes mass flow rate of the bypass line was assumed to be 0 kg/s.

The performance of the CO2 refrigeration system can be determined from:

comp

comp

The coefficient of performance (COP) of the integrated system can be calculated from:

int

Wint is the electrical power consumption of the various components of the integrated system which can be expressed as:

pump

Equation (3.19) shows that the integrated system consumes electricity for the LT CO2

compressor (Wcomp), CO2 pump (WCO2,pump), absorption chiller (Wabs), heat transfer fluid pump (Whtf,pump) and brine pump (Wbrine,pump).

The electrical power of the brine pump can be determined from:

 

(Po,brine-Pi,brine) and (ηbrine,pump) are pressure head and total efficiency of the brine pump which were assumed to be 250 kPa and 0.5 respectively. The power consumption of the brine pump can be reduced by minimising the pressure drop of the secondary loop which subsequently reduces the pressure head. This can be achieved by installing the CO2 refrigeration system near the trigeneration plant.