**7. Nomenclature**

114 Heat Exchangers – Basics Design Applications

Figure 12 shows the temperature profiles along the thermosyphons for different loading rates. In the plots of Figure 12 it can be seen that profiles with higher temperatures are of the thermosyphons loaded with rates of 24.81%, 18.91% and 15.06%. This means that these thermosyphons reached higher temperatures, compared to the other two thermosyphons, under the same conditions of heat supply and cooling air flow. Moreover, these plots show that the lowest average temperature differences between condenser and evaporator are presented in the thermosyphon loaded with 10.17% in the range of heat input of 500 W to

Also, it can be noted that for this thermosyphon it was presented a relative increase of temperature at the bottom of the evaporator for each value of the heat supplied, specifically for the heat supply 2000 W, this indicates the start of drying in the area of

In the plots of Figure 12 it can be observed that at the top of the condenser (thermocouple Tc1) there is a higher temperature than in the middle of the condensation zone (low thermocouple Tc2). For this flow of cooling air, the rate of heat extraction is higher compared to the steam generation in the evaporator, so in this region drying occurs causing a rapid

The methodologies for the calculation of three key parameters implemented in the design and manufacture of two-phase thermosyphons were developed. These key parameters are: the relationship of the lengths of evaporation and condensation zones, the operational pressure values and the evaporation rate of the working fluid during the filling process. The developed methodologies were applied in the design and manufacture of several two-phase thermosyphons. Distilled water was used as working fluid and hydrazine hydrate was

Two experimental installations were designed and constructed. One was used to load the thermosyphons, without using a vacuum pump, eliminating the non-condensing gases; the other one is a wind tunnel modified to test the performance of the thermosyphons using electrical resistances as heat source. A series of experiments to investigate the effect of parameters as heat power supply, amount of working fluid and speed of cooling air, on the performance of two-phase thermosyphons were carried out. In each experimental test, the temperature distribution along the external surface of thermosyphons, the heat power supply and the dissipated heat power by the device as well were registered. The results showed that the thermosyphons work isothermally with efficiencies around 90% for a

The results of these investigations can be used to design and construct high efficiency twophase thermosyphons for heat recovery from waste gas with a temperature up to 250 ºC.

The authors wish to express their thanks to CONACyT, COFAA and National Polytechnic

working fluid loading of 20% of the internal volume of the thermosyphon.

800 W.

evaporation.

**5. Conclusion** 

condensation on the top of the thermosyphon.

added to it as corrosion inhibitor.

**6. Acknowledgment** 

Institute of Mexico for their support of this work.

*A* – area, [m2]; *a* – width of the gas-gas heat exchanger, [m]; *cp* – heat capacity at constant pressure, [J/kg K]; *d* – diameter, [m]; *D* – binary diffusion coefficient, [m2/s]; *E* – modulus of elasticity, [Pa]; Gr – Grashof number; g – gravitational force, [m/s2]; *H* – specific enthalpy, [kJ/kg]; *h* – convective heat transfer coefficient, [W/m2 K]; *hm* – mass transfer coefficient, [m/s]; *I* – electric current, [A]; *k* – thermal conductivity, [W/m K]; *Leva* – characteristic length of evaporation, [m]; *l* – length, [m]; *m* – mass flow, [kg/s]; *m* – mass, [kg]; Nu – Nusselt number; Pr – Prandtl number; *p* – pressure, [Pa]; *Q* – heat flow, [W] *R* – electric resistance, [Ω]; Ra – Rayleigh number; *r* – radius, [m]; – stress, [N /m2]; *S* – entropy, [kJ/ºC] *s* – specific entropy, [kJ/kg ºC]; *T* – temperature, [oC, K]; *t* – wall thickness of container, [m]; *U* – voltaje, [V]; *u* – specific internal energy, [kJ/kg]; *V*– volumen, [m3]; *v* – velocity, [m/s]; x – quality; *Ψ*– loading rate; – thermal diffusivity, [m2/s]; *β* – volumetric thermal expansion coefficient, [K-1]; – kinematic viscosity, [m2/s]; – density, [kg/m3]. – specific volume, [kg/m3]; – relative humidity; Subscripts: 0 – normal conditions; *a* – air;

**5** 

Tomasz Bury

*Poland* 

*Silesian University of Technology* 

**Impact of a Medium Flow Maldistribution** 

**on a Cross-Flow Heat Exchanger Performance** 

The plate exchangers (with the mixed current) and the finned cross-flow heat exchangers, which core has the form of a bunch of pipes with flat plate ribs, have the most important meaning among the currently applied heat exchangers with extended surface. These heat exchangers are usually used for heat transfer between a liquid flowing inside the tubes and a gaseous medium flowing outside the tubes, on the ribs side. Small size, low weight and a high efficiency determine the strong position of such devices. Compact ribbed heat exchangers are commonly used in thermal technique, refrigeration, air-conditioning and

A typical thermodynamic analysis of a cross-flow heat exchanger is usually aimed in determination of the heat transfer surface for the desired design and its capacity. There are several simplifying assumptions made during such calculations, for example neglecting of the heat losses to the environment, uniform flow of media through the exchanger, heat transfer coefficients determined for the average temperatures. These assumptions are fulfilled very rarely in reality and of course it affects the analytical results to some degree. The subject of this work is evaluation of the impact of a non-uniform flow of media (or flow maldistribution) on very popular finned cross-flow heat exchangers performance. The reasons for such maldistribution occurring in an exchanger include the layout of the exchanger with respect to other components in the system, effects of manufacturing tolerances, the design of the flow circuits in the exchanger and the design of the inlet and outlet headers. In some instances, the maldistribution could also be induced due to temperature effects. These factors become even more critical when the exchangers are applied in compact designs which involve a tortuous flow path for both the fluid streams. This situation may lead to some losses in the total heat flow rates transferred in the heat exchanger and affects its thermal efficiency. There is therefore the obvious question: to what

One of the most important parameters describing such heat exchangers is the heat transfer coefficient on the gas side. Usually, this coefficient is determined as an average value for the whole heat transfer surface. This is of course another simplification. Beside of these simplifying assumptions, a variety of constructions being applied causes significant

extent inequality of media flows worsens effects of the heat exchanger?

**1. Introduction** 

automotive industry.

**1.1 Characteristics of the problem** 

*eva* – evaporation; *ext* – external; f – working fluid; *g* – hot gases; *int* – internal; *t* – total; *w* – water.

#### **8. References**

