WO2005071329A1

WO2005071329A1 - Parallel flow evaporator

Info

Publication number: WO2005071329A1
Application number: PCT/EP2004/014664
Authority: WO
Inventors: Sivert Vist; Jostein Pettersen
Original assignee: Norsk Hydro Asa
Priority date: 2004-01-20
Filing date: 2004-12-23
Publication date: 2005-08-04

Abstract

Parallel flow evaporator having a tube-like inlet manifold, and a number of heat exchanging tubes connected to this inlet manifold, where in that the surface of the free cross-section area of the manifold is defined by the equations Gtr = f(Ac, X, M, F) and is the transition between gravitational (stratified-wavy) and inertia (intermittent) dominated flow defined in the flow chart of Kattan. Formula (I) in which C1, is between 0,8 and 1,5 and C2 is between 3,0 and 6,0 and B1* GTr<Gi<B2*Gtr, in which Ac is the free surface cross-section area of the manifold, M is the mass flow rate as defined by the system requirements, X is the vapour fraction of the fluid supplied to the manifold, and F are the characteristics of the fluid supplied to the manifold as far as these are needed to define the Kattan equation.

Description

Parallel flow evaporator

The invention relates to heat exchanger with parallel tubing and manifolds. Such an evaporator is generally known.

In such an evaporator the distribution of the fluid entering the inlet manifold and subsequently the heat exchanger is governed by a number of conditions. Basically the transition between the manifold and a heat exchanger tube is a T- junction, and flow conditions inside the heat exchanger are substantially influenced by that T-junction.

In heat pumps and refrigeration systems a refrigerant which changes phase in the process cycle is commonly used. Two heat exchangers are often used, one to evaporate liquid phase refrigerant in order to cool an external fluid such as air or water, and the other to condense refrigerant vapour by transferring heat to another external fluid.

Compact heat exchangers with flat multiport tubes and folded louvered fins are increasingly used, e.g. in the automotive industry. Because of the small cross- sectional area of each tube, several tubes are placed in parallel to achieve the necessary capacity of the heat exchangers at acceptable pressure drops. In condensers, the refrigerant normally enters in a pure vapor phase, such that maldistribution of the flow has not presented serious problems. However, in evaporators, there is a large challenge to distribute the two-phase flow evenly into the parallel heat exchanger branch tubes. In evaporators used e.g. in automotive air conditioning systems, the flat multiport tubes are oriented vertically to facilitate condensate drainage.

Horizontal manifolds are used to feed parallel vertical branch tubes with refrigerant. Such manifolds, receiving two-phase refrigerant flow, are vulnerable to maldistribution both on a total mass flow level, but also on distribution of the liquid phase among the parallel branch tubes. Pressure gradients along the inlet and outlet manifolds cause gross mass flow maldistribution, while phase stratification in the inlet manifold causes maldistribution of the liquid phase. These problems are universal to evaporators having two-phase flow distribution in horizontal manifolds, regardless of the type of branch tubes used (round tubes, flat multiport tubes or oval tubes).

Refrigerant maldistribution gives reduced performance due to nonuniform feeding of liquid to the evaporator branch tubes. Due to nonuniform heat exchanger temperatures, the air temperature distribution becomes nonuniform, with most cooling at the refrigerant tubes receiving most liquid. This provides uneven frosting patterns and reduced dehumidification properties.

Present (prior art) manifold internal diameters have been determined by branch tube width and maximum manifold mass flux corresponding to average tube mass flux. Inlet and outlet manifold diameter are equal because of ease of production and cost.

It might be expected that the two-phase flow distribution in T-junctions would depend on the momentum of the incoming fluid. Therefore it is useful to consider the flow patterns which occur in two-phase flow in pipes before considering the available T-junction measurement data. One complication in the analysis of horizontal pipe flow compared to vertical flow is that the flow is not symmetrical around the axial centre axis.

The phase distribution in the inlet tube is important for the phase split in the junction, therefore the analysis of flow pattern at the inlet of the junction is essential for the understanding of the flow distribution phenomena. The flow patterns that can be observed in horizontal two-phase flow as follows :

Bubbly flow At low vapour flow rates the vapour is distributed in discrete bubbles in a continuous liquid phase. The bubbles tend to flow in the upper part of the tube due to buoyancy. Plug flow An increase in vapour flow rate cause the bubbles to coalescence into large elongated plug-type bubbles flow in a continuous liquid phase in the upper part of the tube. Slug flow The liquid flow is contained in liquid slugs, separating the successive vapour bubbles. The length of the vapour bubbles can vary considerably and contain liquid droplets. Vapour bubbles may be dispersed in the liquid slug. Stratified flow The liquid is flowing in the lower part of the tube with a relatively smooth interface to the vapour in the upper part. Wavy flow At increasing vapour velocity, the interface between the gas and the liquid becomes wavy. Annular flow At even higher velocities, a liquid film will form a continuous annulus along the tube wall with the vapour flowing in the core. Due to gravity, the film will be thicker at the bottom of the tube ("crescent" liquid interface).

Dispersed mist flow The liquid is transported in droplets in the continuous gas phase.

Flow pattern maps are often used in predicting two-phase flow patterns represented as areas on a graph and the transition lines between the flow patterns. The coordinates of the flow pattern maps are the actual superficial vapour and liquid velocities or generalized parameters containing these velocities.

In an number of articles published in the Journal of Heat Transfer 120 (1998), pages 140-147, pages 148-155 and pages 156-165 respectively and entitled Flow boiling in horizontal tubes, part 1 , part 2 and part 3 respectively by N. Kattan et al. a flow chart has been proposed. An example for such a chart is shown in Figure 1 for HFC-134 a refrigerant in which the following symbols are used : S = Stratified SW = Stratified - Wavy I = Intermittent A = Annuclar MF = Mist

A new design concept, using different and specifically designed manifold diameters for inlet and outlet manifolds, is solving/reducing above problems. The problem is solved in a Parallel flow evaporator having a tube-like inlet manifold, and a number of heat exchanging tubes connected to this inlet manifold, characterised in that the surface of the free cross-section area of the manifold is defined by the equations

Gtr = f(A_c, X, M, F) and is the transition between gravitational (stratified-wavy) and inertia (intermittent) dominated flow defined in the flow chart of Kattan Gι = M/Ac Bι = Cι*148,7 1 +14,89*A_C*10⁵ B₂ = C_?*148.7 1 +14,89*A_C*10⁵ in which Cι is between 0,8 and 1 ,5 and C₂ is between 3,0 and 6,0 and

in which A_c is the free surface cross-section area of the manifold, M is the mass flow rate as defined by the system requirements, X is the vapour fraction of the fluid supplied to the manifold, and F are the characteristics of the fluid supplied to the manifold as far as these are needed to define the Kattan chart.

It has been focused that the gross mass flow maldistribution among the parallel heat exchanger tubes is caused partly by the pressure gradients along the manifolds, and by the nonuniform feeding of the vapor and liquid phases into the inlet manifold. The pressure drop due to friction along the inlet manifold is partly outweighted by the retardation pressure gain caused by fluid velocity reduction. In the outlet manifold, the pressure drop due to friction and acceleration is additive. The problem of phase maldistribution in the inlet manifold can be reduced by reduction of the cross-sectional flow area, such that gravitationally induced phase stratification is avoided. This places a lower bound on mass flux and consequently an upper bound on inlet manifold cross-sectional flow area.

In conclusion, there is a rationale to increase the cross-sectional flow area of the outlet manifold to reduce difference in pressure gradient over the heat exchanger branch tubes, while keeping the inlet manifold cross-sectional area small enough to avoid stratification in the inlet manifold. The current invention defines guidelines for designing the outlet manifold such that gross flow maldistribution is minimized. Design guidelines for the inlet manifold are also given, such that the detrimental effect of gravitational stratification is minimized. These objectives are obtained by using the design parameters as defined by claim 1.

In a preferred embodiment of the invention Cι is between 0,9 and 1 ,5, most preferably between 0,95 and 1,5 whereas C2 is between 3,0 and 5,0 most preferably between 4,2 and 6,0.

Furthermore it is preferred that the outlet manifold has a greater cross-section than the inlet manifold.

Background information about how to apply the design parameters as defined and described here after references being made to the annexed drawings in which Fig. 2 is a schematical drawing of a parallel flow evaporator, and

Fig. 3 is a cross-section of a manifold tube showing the definition of freeflow cross-sectional area.

As shown in Fig. 3 the free flow cross-sectional area is defined by Ac. In fig. 3 there is shown a round tube manifold 10 with a flat multiport tube 11 protruded into the manifold 10 to a defined degree. The free area below (in fig. 3) the end of the tube 11 is defined as the free flow cross-sectional area. The free flow cross sectional area of the outlet manifold should be bounded by a lower limit such that the manifold pressure loss is less than 10 % of the average branch tube pressure loss :

In fig. 2 there is shown an inlet manifold 20, an outlet manifold 21 and a number of multiport tubes 11 extending between the two manifolds. The pressure drop over the whole length of the inlet manifold is defined as ΔP, the pressure drop over the whole length of the outlet manifold is defined as APo and the pressure drop over the length of the multipart tube is defined as APt. This results in the equation :

APo < 0.1 APt

This is obtained by using a large enough cross-sectional area in the outlet manifold. The following lower limit on the cross-sectional area should be used :

Ao > 2 ∑ At

The above restriction in manifold pressure loss would restrict the impact on capacity degradation within 5 % of nominal capacity for evaporators.

Based on experimental investigations, it is found that the mass flux at the inlet of the inlet manifold should be bounded by a lower limit, in order to get an approximate equal phase split in the branch tube junctions. The limits are defined by :

Where G tr is the transition between gravitational (stratified-wavy) and inertia (intermittent) dominated flow defined in the flow chart of Kattan. For R134a at a saturation temperature of 5C, and a vapor mass fraction of 0.1 , Gtr = 240 kg/(m²s) in a tube with inner diameter of 11 mm. Corresponding values at vapor fractions of 0.2 and 0.3 is 180 and 150 kg/(m²s), respectively. For calculating the Gtr, the Equations given by Kattan. (1998) should be used. The empirical parameter B* was found from experimental investigations :

β* = Cx* 148.7 1 + 14.89 Ac x 10

Where Ac is the free flow cross sectional area of the manifold (indicated in Figure 2 for a manifold with a protruded flat multiport tube). Good results can be obtained by selecting Ci between 0,8 and 1 ,5 and C2 between 3,0 and 6,0. The upper limit is obtained by limiting the inlet mass flux by an upper bound in order to avoid excessive frictional pressure loss along the inlet manifold. In order to have optimal results it has been found that the values of G number defined.

Claims

1. Parallel flow evaporator having a tube-like inlet manifold, and a number of heat exchanging tubes connected to this inlet manifold, characterised in that the surface of the free cross-section area of the manifold is defined by the equations

G_tr = f(A_c, X, M, F) and is the transition between gravitational (stratified-wavy) and inertia (intermittent) dominated flow defined in the flow chart of Kattan. Gι = M/Ac Bι = Cι*148,7 1+14,89*A_C*10⁵ B₂ C_?*148,7 1 + 14,89*A_c*10⁵ in which Ci is between 0,8 and 1 ,5 and C₂ is between 3,0 and 6,0 and

in which A_c is the free surface cross-section area of the manifold, M is the mass flow rate as defined by the system requirements, X is the vapour fraction of the fluid supplied to the manifold, and F are the characteristics of the fluid supplied to the manifold as far as these are needed to define the Kattan equation.

2. Evaporator according to claim 1 , characterised in that Ci is between 0,9 and 1 ,5, preferably between 0,95 and 1 ,5

3. Evaporator according to claim 1 or 2, characterised in that C₂ is between 3,0 and 5,0, preferably between 4,2 and 6,0

4. Evaporator according to anyone of the preceding claims, characterised in that the outlet manifold has a greater cross-section than the inlet manifold.