Rising energy costs and legal requirements such as the EU F-Gas Regulation are increasingly bringing the topic of sustainability in refrigeration systems into focus. The use of CO₂ (R744) as a refrigerant, in combination with highly efficient compressors and controllable ejectors, enables high efficiency in both full-load and part-load operation. So-called high-lift applications with ejectors for R744 systems can increase the performance per displacement for peak loads, improve the energy efficiency of the system, and at the same time reduce investment and operating costs.
How to optimally coordinate the control of the individual components is illustrated by a simulated design of a distribution center and a heat pump for a district heating network. In addition, the energy advantage is demonstrated using the distribution center as an example.

The potential distribution center, designed as a system with a high-lift ejector, has an MT refrigeration capacity of 1,000 kW (evaporation temperature −10 °C) and an LT refrigeration capacity of 300 kW (evaporation temperature −35 °C). The ejector acts as a high-pressure control valve and allows operation as either a high-lift or FGB system. Both in the LT stage and downstream of the gas cooler, an internal heat exchanger is used. The usable superheat at the evaporators and the subsequent suction lines is assumed to be 6 K each. Pressure losses in piping and heat exchangers are neglected.

The design is based on full-load and part-load operation for five ambient temperatures to account for seasonal influences, as shown in Table 1. Regarding the load profile, a distinction is made between ‘open’ and ‘closed’.

The current standard solution for the use of CO₂ as a refrigerant is the so-called flash gas bypass system (FGB system). In this system, the pressure in the intermediate pressure receiver is controlled by an FGB valve and is above the required evaporation pressure. For two different temperature levels, for example medium temperature (MT) and low temperature (LT), a booster configuration is used. The concept of parallel compression conveys the flash gas from the intermediate pressure receiver to the high-pressure level with less pressure difference, thereby increasing energy efficiency.

An alternative to FGB and parallel compression systems is the booster system with ejector(s) (Fig. 1). Here, the high-pressure ejectors are installed at the outlet of the gas cooler.
Two operating modes can be distinguished:

• In operation with suction effect, the ejector relieves the MT compressor stage and loads the parallel stage. Driven by the energy of the motive mass flow coming from the gas cooler, the pressure of the suction mass flow from the receiver downstream of the MT evaporators is raised from MT suction pressure to intermediate pressure. In this state, the FGB valve is closed. Due to the pressure lift generated by the ejector stage, the energy advantage of parallel compression comes into play, as more mass flow can be compressed from intermediate to high pressure with a lower pressure ratio and increased suction gas density. In this process, the pressure in the intermediate pressure receiver is controlled by the parallel compression stage.
• If the ejector is operated without suction effect, the control valve upstream of the ejector suction port is closed. The parallel compression stage is therefore out of operation, and the ejector functions as a precise high-pressure control valve. In this operating mode, the pressure in the intermediate pressure receiver is controlled by the FGB valve.

For the simulation, seven controllable BITZER ejectors of the HDV series are used. With a motive nozzle inlet condition of 92 bar(a) and 31 °C, they have motive mass flows ranging from 800 to 9,500 kg/h. Due to the controllability of the ejector, no additional high-pressure control valve is required. For the MT and parallel stage of the distribution center, the newly developed transcritical ECOLINE 8-cylinder CO₂ reciprocating compressors from BITZER are used (8CTE-140K, displacement 99.2 m³/h at 50 Hz mains frequency). Capacity control is possible via frequency inverter or alternatively via mechanical capacity control.
The calculations presented were carried out using the BITZER selection software.

For operating condition (a), a typical summer day, a design is carried out for different pressure lifts (7–12 bar). The influence on the parameters COP, suction gas temperature, and discharge gas temperature of the MT compressor stage is shown in Fig. 2. If the pressure lift of the ejector decreases, the mass flow drawn off by the ejector downstream of the MT evaporator increases. This reduces the MT mass flow that can be mixed with the superheated LT discharge gas. Consequently, the suction gas temperature of the MT compressors and the discharge gas temperature rise. At a pressure lift of 7 bar, the suction gas temperature would be 49 °C and the discharge gas temperature 188 °C. The maximum discharge gas temperature of the compressor is limited to 160 °C — at 7 bar pressure lift, the thermal limit of the compressors would therefore be exceeded. At high ambient temperatures and low pressure lifts, attention must therefore be paid to the suction and discharge gas temperatures of the MT compressor stage. The smaller the mass flow ratio of MT evaporator to LT evaporator, the more likely this phenomenon occurs.
A pressure lift of 9.5 bar is selected for the full-load operating point. This achieves an optimal system solution in terms of COP, number of compressors, and discharge gas temperature. Five 8CTE-140K compressors (two in the MT stage and three in the parallel stage) as well as one 4TME-20K and three 4PME-25K compressors of the subcritical ECOLINE ME series from BITZER are required for the LT stage. The lead compressor of each stage is operated with a frequency inverter. The respective active displacement for the MT and parallel stage is shown in Table 2. Of the five ejectors required, shown in Table 2, the HDV-E65 is throttled. The discharge gas temperature of the MT compressor stage is 154.5 °C and thus within the expected range (see Fig. 2). Since continuous compressor operation at a discharge gas temperature of 154.5 °C is not recommended, an additional air-cooled desuperheater is used downstream of the LT compressor stage. This reduces the discharge gas temperature of the MT stage to 149.5 °C. The COP is 1.69, which is 21.5 percent higher than that of the corresponding FGB system (Fig. 3).

Under operating conditions b, c, d, and e, the system operates subcritically. The potential energy available at the ejector inlet, which is linked to the high-pressure level, decreases as the ambient temperature drops. Compared to an FGB system, this results in a decreasing COP improvement (Fig. 3), as the ejector increasingly fails to relieve the MT stage and load the parallel stage (Table 2). For operating conditions d (winter night) and e (winter day 2), the potential energy available at the motive nozzle inlet is no longer sufficient for high-lift operation—the valve on the suction side of the ejector is closed, and the circuit operates as an FGB system. The ejector functions as a high-pressure control valve.

For operating condition e (winter day 2) with a relative MT load of 65 percent and LT load of 70 percent, the system requires one more MT compressor than in full-load operation for operating condition (a). However, the maximum number of compressors can be reduced by using a ‘swing compressor,’ which can operate either as an MT or as a parallel compressor depending on the operating condition.
Considering both full-load and part-load conditions is essential when designing a system with a high-lift ejector.

For the calculations, operating conditions from a real H1200-AW air-to-water heat pump for district heating in Mejlby (Denmark) are used. The data was provided by Fenagy A/S. The heat pump is designed as a high-lift system and is equipped with eight BITZER 6DTEU-50LK compressors — four in the MT stage and four in the parallel stage. These operate constantly at 50 Hz. For the design, the FENeject is replaced by one HDV-E95 and two HDV-E65 ejectors. The operating conditions to be analyzed (Table 3) are taken from field data of the heat pump. For winter operation, an additional operating condition for an ambient temperature of −5 °C is assumed. Pressure losses in piping and heat exchangers are neglected, and the power consumption of fans, pumps, and controls is not considered.
The results of simulation and actual measurement are shown in Fig. 4. The simulated and measured heating capacity differ by a maximum of 2.7 percent. The simulated heating COP is 8–10 percent higher than the measured value. However, the simulated COP only refers to the power consumption of the compressors, whereas the measured COP also includes the power consumption of fans, pumps, and controls as well as pressure losses.

In the heat pump, the mass flow does not shift from the parallel stage to the NK stage or vice versa. Compared to the booster system for the distribution center, the conditions at the ejector inlet hardly change with the ambient temperature. The state at the motive nozzle inlet and the associated pressure are defined by the return/flow temperature of the water and the pinch point in the gas cooler – these fluctuate little with the season. The required relative ejector nozzle area decreases as the ambient temperature drops (Fig. 4). Depending on the operating point, one of the three available ejectors is throttled or, if necessary (for an ambient temperature of -5 °C), switched off.
The design of the heat pump proves to be significantly simpler than that of a booster system.

Conclusion and outlook

The analysis of the distribution center shows that the COP improvement achieved by using an ejector depends on the ambient temperature:

• The higher the ambient temperature, the higher the pressure and the driving energy available at the ejector inlet – and the greater the energy benefit from parallel compression.
• If the high pressure or the outlet temperature at the gas cooler/condenser decreases, the mass flow that can be drawn off by the ejector also decreases.

A shift of the load from the parallel stage to the LT compressor stage must be taken into account. Depending on ambient temperature and load profile, more LT compressors may be required for part-load operation than for full-load operation. The number of compressors needed can be reduced by using a “swing compressor”, which can operate either as a parallel or LT compressor.
When designing a booster system as a high-lift application, high gas cooler outlet temperatures or pressures can lead to high suction gas temperatures/discharge gas temperatures in the LT stage. The ejector’s pressure lift must be optimized with regard to system efficiency and suction/discharge gas temperatures.
While in the design of the distribution center high pressure and ambient temperature are linked, the air-to-water heat pump must be considered differently. High pressure and gas cooler outlet temperature are linked to the water temperature and the pinch point of the gas cooler. With varying ambient temperature, the system must operate continuously in transcritical mode, and no load is shifted from the parallel stage to the LT stage. Since COP improvement through ejectors is greatest at high gas cooler outlet temperatures/pressures, coupling heat and refrigeration is desirable.

(published: 2023)

Further information:
ECOLINE reciprocating compressors for subcritical CO₂ applications with high standstill pressures
ECOLINE reciprocating compressors for transcritical CO₂ applications
ECOLINE reciprocating compressors with LSPM motor for transcritical CO₂ applications
Application guide for the use of CO₂ / R744