Experimental Investigations of Thermal and Flow Characteristics of Condensation

Based on many years of research by the authors, a laboratory test stand was proposed, whose schematic diagram is shown in Figure 18.1.

The test stand shown in Figure 18.1 allows performing experimental measurements of condensation of various homogeneous refrigerants (pure) and their mixtures (including zeotropic mixtures) in single-pipe minichannels and in multipots in the range of the inner diameter d = 0.21-3.30 mm. A detailed description of the methodology and the results of the authors' research in the scope of condensation characteristics in minichannels is provided in the works of Bohdal et al. (2012,2015) and Shin and Kim (2004a,b).

FIGURE 18.1

Schematic diagram of the test stand; 1—measuring section of pipe minichannel, 2—water duct, 3—cooling compressor unit, 4—air-cooled condenser, 5—liquid refrigerant tank, 6—filter-dryer of refrigerant, 7—electromagnetic valve, 8—finned air cooler, 9—expansion valve supplying, 10—heat exchanger for the pickup of the superheat of the refrigerant, 11—liquid refrigerant subcooler, 12—electronic refrigerant flow meter, 13—medium-pressure sensor on the inlet to the measuring section, 14—medium-pressure sensor on the outflow from the measuring section, 15—differential pressure sensor, 16—electronic water flow meter, 17—computer, and 18—data acquisition system.

The schematic diagram of the test stand (Figure 18.1) should be distinguished by two cooling systems with refrigerant, cooperating in parallel, that is installation of a single-stage refrigeration system supplied from a refrigerating compressor unit and cooling installation feeding the measuring section of pipe minichannel 1. The superheated steam of the refrigerant after compression in the piston compressor 3 was passed through the subsequent elements of the installation from position 3 to position 9 in Figure 18.1.

The condensing process of the refrigerant took place in the flow inside the pipe minichannel 1 (Figure 18.2). The superheated vapor of the refrigerant after discharge from the compressor 3 was directed by a control valve both to the basic system and to the supply system of the measuring section 1. Before the inflow of the refrigerant to the inlet section of the measuring section, the water-cooled heat exchanger 10 was installed (for receiving the superheat and for the preparation of the refrigerant condition at the inlet to

FIGURE 18.2

Schematic diagram of the measuring section of the pipe minichannels; 19—copper pipe (diameter 8/10 mm), 20—connector, 21—water-cooling channel, 22—thermocouple for measuring water temperature, 23—thermocouple for measuring minichannel surface temperature, and 24—insulation (other markings as in Figure 18.3).

the minichannel). The pressure of the refrigerant at the inflow and outflow of the measuring section was measured using piezoresistive sensors 13 and 14 with PMP 131-A1401AlW-type transmitters produced by Endress + Hauser; the measuring range of the sensors is 0-2.5 MPa. The pressure drop of the refrigerant over the length of a 1,000-mm pipe minichannel was additionally measured using the differential pressure sensor 15 with a Deltabar SPMP-type transducer. The liquid of refrigerant leaving the measurement section was subcooled in the heat exchanger 11. The flow rate of the refrigerant liquid was measured using a Coriolis flow meter. (The same type of flow meter was used for cooling water.) The liquid of refrigerant after leaving the measuring section was discharged to the installation feeding the air cooler 8 (Charun 2012).

Figure 18.2 shows a schematic diagram of the pipe minichannel section with control and measurement instrumentation. The basic element of the measurement system was a minichannel section with an internal diameter d_w = 0.21-3.30 mm and an overall length 1,000 mm. (The pressure drop of the refrigerant was measured throughout this length.) A section of the pipe minichannel was placed in the water channel 21 (Figure 18.2). It was a channel made of aluminum, with a rectangular section (dimensions 28 x 24 mm) and an internal length of 950 mm (active length of the measuring section). In nine cross sections along the effective length of the minichannel, the temperature of its external surface was measured using К-type thermocouples installed at a distance of 100 mm. In the same sections, thermoelectric sensors were placed to measure the temperature of the cooling water flowing in the water channel. All thermoelectric sensors, before installation on the test section, were sized against a standard glass thermometer with an elemental scale of 0.1°C, performing their individual characteristics.

Indirectly, the vapor quality x and heat flux density q were determined according to a specially developed methodology. To determine the heat flow Q discharged during condensation into cooling water (flowing in the water channel 2 in Figure 18.1), the concept presented in the authors' work was used (Shin and Kim 2004a,b), with air as a cooling factor. Comments included in the authors' works (Del Col et al. 2014, Kandlikar et al. 2013) were concluded, also those regarding the investigations of zeotropic mixtures.

Before performing the basic tests (on the test stand shown in Figure 18.1), testing investigations of the indirect method were carried out on a separate test station. The schematic diagram of this test section is shown in Figure 18.3. In Figure 18.3b is shown the test section made from minichannel with the same internal diameter and an active length of 950 mm. There is the same arrangement of thermocouple sensors to measure the temperature of

FIGURE 18.3

Schematic diagram of the comparative method to determine the heat flux discharged during condensation in the minichannel; (a) basic measurement section according to Figure 18.1 and (b) additional measuring section heated by electricity.

the refrigerant and the cooling water. Figure 18.3a shows the section with a part of the installation from Figures 18.1 and 18.2.

The refrigerant did not flow inside the section (b) of the minichannel (m = 0) during the testing measurements, while the cooling water was flowing outside, with the same flow rate as during the basic tests in section (a). The minichannel section (b) was included in the independent electric heating system, where it works as an electrical resistor. The electric current parameters such as the supplied electrical power Q_el were measured in each case along with the values of the outer wall surface temperature of the minichannel T_wz and the cooling water temperature T_H in each subsequent measurement cross section (in nine subsequent cross sections, for i = 1, 2, ..., 9). The supplied electric power Q_el was transmitted in the form of heat flux Q, (according to Joule's law), causing an increase in the outer surface temperature of the minichannel in each of its cross sections. Knowing the values of Q_el = Q„ the internal diameter d, and the length L, of each part of the measurement section, the heat flux density <7, related to its internal surface was determined:

On the basis of the measured values of the wall temperature T_wzi and the water temperature T_H„ in the /th cross section, thermal characteristics were made in the form of c], =f(T_wzi - T_H) =f(AT), where AT = T_wzi - T_Hi. From the experimental characteristics, the local heat flux density was determined on the inside surface of the minichannel in a given measurement cross section. Figure 18.4 shows an example characteristic =/(ДТ) for one of the cross sections. Such auxiliary experimental characteristics were prepared for all cross sections of the tested pipe minichannels.

Knowing the value of in a given cross section of minichannel, the local value of the heat transfer coefficient a_x was determined:

where T_{w i} is the temperature of the internal surface of the minichannel in a given cross section and T_s is the saturation temperature.

The vapor quality x of the refrigerant at the inflow to the test section of the minichannel was determined by an indirect method from the energy balance equation of the exchanger 10 shown in Figure 18.1. The exchanger was cooled by water flowing countercurrent to the refrigerant in the amount of m_H. The temperature of the cooling water at the inflow T_{H en} and the outflow T_Hex/ the mass flow rate m_R, and the thermal parameters of the refrigerant on the inflow (superheated vapor) and outflow (wet saturated vapor with vapor quality x) were measured. Heat exchange between the refrigerant and cooling water in the exchanger 10 took place in a sensible and latent heat transfer, and the energy balance equation takes the following form:

where ДТ_Н is the increase in cooling water temperature exchanger, T_R the temperature of the superheated refrigerant vapor at the inflow to the exchanger, r the heat of evaporation, c_pH and c_pR the specific heat of the water and the refrigerant, and Дх the change in vapor quality in the flow through the exchanger, from the value x = 1 (saturated dry steam) to the value x on the outflow from the exchanger (and on the inflow to the measuring section of the mini-channel), i.e.,

By adjusting the cooling water flow rate, it was possible to obtain the value of x = 1 on the outflow from the exchanger 10. The local values of the vapor quality x in individual cross section of the minichannel were determined using the dependence:

The presented experimental method allowed determining the local and average values of the heat transfer coefficient with an accuracy of ±10%.

On the basis of the measurements made on the test stand described above, according to the developed methodology, it was possible to implement research programs for thermal and flow characteristics of condensation. Under the thermal condensation characteristics in minichannels should be understood the dependence of the local heat transfer coefficient on vapor quality x in the form a_x = f(x) for G = const and the average heat transfer coefficient on the mass flux density: a„ =/(G), for x_a = const.

Flow condensation characteristics determine the dependence of the local pressure drop per unit of minichannel length (Дp/L)_x or the average pressure drop (Ap/L)_a from local vapor quality x and mass flux density G. Figure 18.5 presents the examples of thermal and flow characteristics of condensation in a minichannel.

Investigations of Two-Phase Flow Structures during Refrigerants Condensation in Minichannels

Knowledge of two-phase flow structures plays an important role in the selection of the adequate correlation for the calculation of the heat transfer coefficient and flow resistance. From the physical point of view, the following two-phase flow structures, characteristic of refrigerants condensation in minichannels, can be distinguished:

• • Bubble flow: The structure of gas phase (intermittent gas phase) in the form of bubbles in a continuous liquid phase.
• • Plug flow: Gas bubbles can reach a size comparable to the internal diameter of the channel and move mainly in the upper part of its cross section.
• • Slug flow: With increasing flow rate, shearing stresses cause an increase in the range of waves forming gas bubbles along the flow channel,
• • Stratified flow: The liquid and gas phases are separated by a smooth phase separation surface. This usually occurs at low speeds of both phases.
• • Wave flow: Along with the increase in the gas phase velocity, there are disturbances at the phase separation boundary, which results in the formation of waves moving in the direction of flow.
• • Annular flow: The gas phase has a high velocity in the gas core, and on the wall surface of the channel, a liquid condensate film is formed whose thickness is generally asymmetrical in relation to the cross section of the channel, with a larger thickness at the bottom of the channel.

Figures 18.6 and 18.7 show examples of two-phase condensation flow structures in vertical and horizontal minichannels. As it can be seen, two-phase flow structures in a vertical minichannel differ slightly from those observed in horizontal channels (Figure 18.7).

FIGURE 18.5

Sample experimental thermal (a) and flow (b) characteristics of R404A refrigerant condensation in a pipe minichannel tube with an internal diameter of 1.4 mm prepared on the basis of measurements results made on the test stand shown in Figure 18.1 [17].

On the test stand presented in Sikora (2017), experimental measurements were carried out in the scope of flow structures during the condensation of HFE-7100 and Novec 649 refrigerants in horizontal minichannels with an internal diameter d = 0.5-2.5 mm. Using the authors' own database of condensation experimental results, the so-called "calculation method" of flow structures identification was used. The idea of this method is to compare the

FIGURE 18.6

Examples of two-phase flow structures in a vertical minichannel (minichannel with an internal diameter of 2.01 mm) according to Chen and Tian (2006): DB—bubble dispersion flow, В—bubble flow, S—slug flow, F—frothy, A—annular, AM—annular-misty.

FIGURE 18.7

Classification of two-phase flow structures in a horizontal minichannel by Coleman and Garimella (2003).

own experimental results with the results of calculations according to the criteria of two-phase flow structure maps of other authors. It was necessary to determine the boundaries of individual flow structures, using a map of two-phase flow structures by Thome (2004-2008) described in Figure 18.8 and Coleman and Garimella (2003) (Figure 18.9) The parameters describing, in this case, the limits of the flow structure boundaries are the vapor quality x and the mass flow density G.

Figure 18.8 presents the results of a computational analysis of flow structures according to Thome's map (Thome 2004-2008) for both the studied refrigerants. Figure 18.9 shows the same kind of comparison, but according to the map of Coleman and Garimella (2003). It was found that in the range of parameters in which own experimental investigations of the described refrigerants condensation in horizontal minichannels were carried out,

FIGURE 18.8

The flow structure map by Thome (2004-2008).

FIGURE 18.9

The flow structure map by Coleman and Garimella (2003).

annular, plug, and slug flow structures are often observed. Unfortunately, the used maps have clear limitations in terms of low mass flux density values. For this reason, most of the results for the Novec 649 condensation are outside the scale of those maps. Hence, there have been difficulties with the clear definition of the flow structure. Another difficulty was the occurrence of the so-called transitional structures, such as annular-wavy structures. Most maps of two-phase flow structures were created for the R134a refrigerant, and therefore, there may be discrepancies during the identification of the flow structures for the investigated low-pressure refrigerants.

FIGURE 18.10

Exemplary flow structures of the HFE-7100 refrigerant during condensation in a minichannel with an internal diameter d = 2.0mm: (a) wavy—T_s = 80°C, G = 1,060 kg/m²s, .r = 0.62; (b) slug—T = 82°C, G = 1,060 kg/m²s, x = 0.40; (c) bubbly—T, = 61°C, G = 1,060 kg/m²s, x = 0.20; (d) annular—T_s = 75°C, G = 1,060 kg/m²s, x = 0.70.

FIGURE 18.11

Exemplary flow structures of Novec 649 refrigerant during condensation in a minichannel with an internal diameter d = 1.0mm: (a) wavy—T_s = 55°C, G = 1,698 kg/m²s, x = 0.68; (b) bubbly— T_s = 44°C, G = 1,521 kg/m²s, x = 0.27.

The experimental visualization of two-phase flow structures during the condensation of HFE-7100 refrigerant in minichannels made of glass was carried out on the experimental test stand. Figure 18.10 shows a picture of a two- phase structure recorded with a high-speed camera. Figure 18.11 presents the visualization results for the Novec 649 refrigerant. The mass flow density G in the experimental research was higher than 1,000 kg/m²s in most cases. A shift of bubble and slug flow occurrence was observed. The most frequent structures in this case were the annular-wavy and wavy structures.