Patent Description:
<CIT> discloses a heat exchanger with the features of the preamble of claim <NUM>.

Patent Literature <NUM> discloses a heat exchanger including a plurality of fins arranged parallel to each other to form a flow passage of gas and heat transfer pipes each passing through the plurality of fins and through which a medium that exchanges heat with the gas flows. The plurality of fins each have a plurality of through-holes and the heat transfer pipes are fitted separately in the plurality of respective through-holes. The plurality of through-holes are provided at equal intervals along a step direction perpendicular to both a direction in which the plurality of fins are arranged and a direction of flow of the gas, and are provided in a plurality of rows along a row direction parallel to the direction of flow of the gas.

The heat exchanger of Patent Literature <NUM> is a part of a refrigeration cycle apparatus such as an air-conditioning apparatus. There has recently been a demand for a reduction in amount of refrigerant charge to reduce the total value of GWP of a refrigeration cycle apparatus. A possible way of reducing the amount of refrigerant charge in a refrigeration cycle apparatus is to reduce the inner capacity of each of the heat transfer pipes of the heat exchanger by reducing the pipe diameter of each of the heat transfer pipes. However, reducing the pipe diameter of each of the heat transfer pipes usually causes a decrease in heat transfer performance of the heat exchanger. For this reason, to maintain the heat transfer performance of the heat exchanger while reducing the pipe diameter of each of the heat transfer pipes, it is necessary to narrow the intervals at which the fins are placed and increase the number of rows of the heat transfer pipes. Meanwhile, narrowing the intervals at which the fins are placed and increasing the number of rows of the heat transfer pipes result in deterioration in ventilation performance of the heat exchanger. That is, there is a trade-off between heat transfer performance and ventilation performance in a heat exchanger whose heat transfer pipes each have a reduced inner capacity. Heat transfer performance and ventilation performance both affect the heat exchanger performance of a heat exchanger. Accordingly, it has been undesirably difficult to improve the heat exchanger performance of a heat exchanger while reducing the inner capacity of each of the heat transfer pipes.

The present disclosure has been made to solve such a problem, and has as an object to provide a heat exchanger that makes it possible to improve the heat exchanger performance of the heat exchanger while reducing the inner capacity of heat transfer pipes and a refrigeration cycle apparatus including the same.

A heat exchanger according to the present invention includes a plurality of fins arranged in parallel to each other and a plurality of heat transfer pipes each extending in a direction intersecting the plurality of fins. In a plane perpendicular to a direction in which the plurality of heat transfer pipes extend, the plurality of heat transfer pipes are placed in a plurality of rows in a row direction that is along a direction of airflow at a row pitch L1. In the plane, the plurality of heat transfer pipes are placed in a plurality of steps in a step direction perpendicular to the row direction at a step pitch L2. Where an outer diameter of each of the plurality of heat transfer pipes is defined as Do, a wall thickness of a portion having a smallest distance between an outer wall surface and an inner wall surface of each of the plurality of heat transfer pipes is defined as tP, an area represented by a numerical expression of L1 × L2 is defined as A, and an area represented by a numerical expression of ((Do - <NUM> × tP)/<NUM>)<NUM> × π is defined as B, a relation of Do < <NUM>, a relation of (<NUM> × tP<NUM> - <NUM> × tP + <NUM>) × Do ^ (-<NUM> × tP<NUM> - <NUM> × tP + <NUM>) ≤ B/A ≤ (<NUM> × tP<NUM> - <NUM> × tP + <NUM>) × In (Do) + (<NUM> × tP<NUM> + <NUM> × tP + <NUM>), and a relation of B/A < <NUM> × tP<NUM> - <NUM> × tP + <NUM> are satisfied.

A refrigeration cycle apparatus according to another embodiment of the present disclosure includes the heat exchanger according to an embodiment of the present disclosure.

The present invention makes it possible to improve the heat exchanger performance of a heat exchanger while reducing the inner capacity of heat transfer pipes.

A heat exchanger according to Embodiment <NUM> is described. <FIG> is a cross-sectional view showing a configuration of some components of a heat exchanger <NUM> according to Embodiment <NUM>. <FIG> shows a configuration of the heat exchanger <NUM> as sectioned along a plane perpendicular to a direction in which the after-mentioned first heat transfer pipes <NUM> extend. The heat exchanger <NUM> is used as a heat source side heat exchanger or a load side heat exchanger of a refrigeration cycle apparatus. The heat exchanger <NUM> is a cross-fin fin-and-tube heat exchanger that allows refrigerant circulating through the heat transfer pipes and air to exchange heat with each other. A usable example of the refrigerant include a hydrofluorocarbon such as R410, R407C, and R32, isobutane, propane, and carbon dioxide. In <FIG>, a thick arrow outline with a blank inside represents a direction of airflow.

As shown in <FIG>, the heat exchanger <NUM> includes, as a plurality of heat exchange units arrayed along the direction of airflow, a first heat exchange unit <NUM> located furthest windward and a second heat exchange unit <NUM> located further leeward than the first heat exchange unit <NUM>.

The first heat exchange unit <NUM> includes a plurality of first fins <NUM> arranged parallel to each other at intervals and a plurality of first heat transfer pipes <NUM> each passing through the plurality of first fins <NUM> and each extending parallel to each other in a direction intersecting the plurality of first fins <NUM>. Each of the plurality of first fins <NUM> has a rectangular flat-plate shape elongated in one direction. Each of the plurality of first fins <NUM> is placed perpendicular to a direction in which the first heat transfer pipes <NUM> extend. The plurality of first fins <NUM> are provided parallel to each other at regular placement pitches in a direction perpendicular to a surface of paper of <FIG>, that is, the direction in which the first heat transfer pipes <NUM> extend. A gap between two first fins <NUM> adjacent to each other serves as air passageway through which air circulates. Note here that a direction that is along the direction of airflow in a plane perpendicular to the direction in which the first heat transfer pipes <NUM> extend is sometimes referred to as "row direction of the heat exchanger <NUM>" or simply as "row direction". Further, a direction perpendicular to the row direction in the plane is sometimes referred to as "step direction of the heat exchanger <NUM>" or simply as "step direction". The step direction of the heat exchanger <NUM> is parallel to, for example, a longitudinal direction of each of the first fins <NUM> and a longitudinal direction of each of the after-mentioned second fins <NUM>.

Each of the plurality of first heat transfer pipes <NUM> extends in the direction perpendicular to the surface of paper of <FIG>. The plurality of first heat transfer pipes <NUM> are arrayed at regular step pitches L2 in one row in the step direction of the heat exchanger <NUM>. Each of the step pitches can be specified by a distance in the step direction between the respective tube axes 12a of two first heat transfer pipes <NUM> adjacent to each other in the step direction. Each of the plurality of first heat transfer pipes <NUM> is a circular pipe having an outer diameter Do. Further, each of the plurality of first heat transfer pipes <NUM> is a circular pipe having a wall thickness tP of a portion having a smallest distance between an outer wall surface and an inner wall surface. The plurality of first heat transfer pipes <NUM> constitute a first row of heat transfer pipes located furthest windward in the heat exchanger <NUM>.

The second heat exchange unit <NUM> includes a plurality of second fins <NUM> arranged parallel to each other at intervals and a plurality of second heat transfer pipes <NUM> each passing through the plurality of second fins <NUM> and each extending parallel to each other in a direction intersecting the plurality of second fins <NUM>. As with the first fins <NUM>, each of the plurality of second fins <NUM> has a rectangular flat-plate shape. Each of the plurality of second fins <NUM> is placed parallel to the first fins <NUM> and perpendicular to a direction in which the second heat transfer pipes <NUM> extend. The plurality of second fins <NUM> are provided parallel to each other at regular placement pitches in the direction perpendicular to the surface of paper of <FIG>, that is, the direction in which the first heat transfer pipes <NUM> extend. Each of the plurality of second fins <NUM> is placed with a displacement of, for example, approximately half a pitch from the corresponding one of the plurality of first fins <NUM>. A gap between two second fins <NUM> adjacent to each other serves as an air passageway. In the present embodiment, each of the first fins <NUM> and each of the second fins <NUM> are separate components. Alternatively, the first fin <NUM> and the second fin <NUM> may be integrally formed. That is, the first heat exchange unit <NUM> and the second heat exchange unit <NUM> may share a plurality of fins with each other.

Each of the plurality of second heat transfer pipes <NUM> extends in a direction parallel to the direction in which the first heat transfer pipes <NUM> extend. The plurality of second heat transfer pipes <NUM> are arrayed at step pitches L2 in one row in the step direction of the heat exchanger <NUM>. Each of the step pitches L2 is equal to a step pitch between first heat transfer pipes <NUM>. Each of the plurality of second heat transfer pipes <NUM> is placed with a displacement of, for example, approximately half a pitch from the corresponding one of the plurality of first heat transfer pipes <NUM>. The plurality of second heat transfer pipes <NUM> constitute a second row of heat transfer pipes as counted from a windward side in the heat exchanger <NUM>. The plurality of first heat transfer pipes <NUM> and the plurality of second heat transfer pipes <NUM> are arrayed at row pitches L1 in the row direction of the heat exchanger <NUM>. Each of the row pitches can be specified by a distance in the row direction between the tube axis 12a of a first heat transfer pipe <NUM> and a tube axis 22a of a second heat transfer pipe <NUM>. A row pitch between first heat transfer pipes <NUM> in the first heat exchange unit <NUM> and a row pitch between second heat transfer pipes <NUM> in the second heat exchange unit <NUM> can both be considered as L1. Each of the plurality of second heat transfer pipes <NUM> is a circular pipe having an outer diameter Do that is equal to the outer diameter of a first heat transfer pipe <NUM>. Further, each of the plurality of second heat transfer pipes <NUM> is a circular pipe having a wall thickness tP that is equal to the wall thickness of a first heat transfer pipe <NUM>.

The heat exchanger <NUM> includes a plurality of refrigerant paths (not illustrated) connected parallel to each other in a flow passage of refrigerant. Each of the plurality of refrigerant paths is formed using one or more first heat transfer pipes <NUM>, one or more second heat transfer pipes <NUM>, or a combination of one or more first heat transfer pipes <NUM> and one or more second heat transfer pipes <NUM>.

<FIG> is a cross-sectional view showing a configuration of some components of a heat exchanger <NUM> according to a modification of Embodiment <NUM>. As with <FIG>, <FIG> shows a configuration of the heat exchanger <NUM> as sectioned along the plane perpendicular to the direction in which the first heat transfer pipes <NUM> extend. As shown in <FIG>, the heat exchanger <NUM> of the present modification differs from the heat exchanger <NUM> shown in <FIG> in that the heat exchanger <NUM> of the present modification includes another second heat exchange unit <NUM> located further leeward than the second heat exchange unit <NUM>.

The second heat exchange unit <NUM> includes a plurality of second fins <NUM> and a plurality of second heat transfer pipes <NUM> each passing through the plurality of second fins <NUM>. As with the first fins <NUM> and the second fins <NUM>, each of the plurality of second fins <NUM> has a rectangular flat-plate shape. Each of the plurality of second fins <NUM> is placed parallel to the first fins <NUM> and the second fins <NUM> and perpendicular to a direction in which the second heat transfer pipes <NUM> extend. The plurality of second fins <NUM> are provided parallel to each other at regular placement pitches in a direction perpendicular to a surface of paper of <FIG>, that is, the direction in which the first heat transfer pipes <NUM> extend. A gap between two second fins <NUM> adjacent to each other serves as an air passageway. In the present embodiment, each of the first fins <NUM>, each of the second fins <NUM>, and each of the second fins <NUM> are separate components. Alternatively, at least two of the first fin <NUM>, the second fin <NUM>, and the second fin <NUM> may be integrally formed.

Each of the plurality of second heat transfer pipes <NUM> extends in the direction parallel to the direction in which the first heat transfer pipes <NUM> extend. The plurality of second heat transfer pipes <NUM> are arrayed at step pitches L2 in one row in the step direction of the heat exchanger <NUM>. Each of the step pitches L2 is equal to a step pitch between first heat transfer pipes <NUM> and a step pitch between second heat transfer pipes <NUM>. The plurality of second heat transfer pipes <NUM> constitute a third row of heat transfer pipes as counted from the windward side in the heat exchanger <NUM>. The plurality of first heat transfer pipes <NUM>, the plurality of second heat transfer pipes <NUM>, and the plurality of second heat transfer pipes <NUM> are arrayed at row pitches L1 in the row direction of the heat exchanger <NUM>. Each of the plurality of second heat transfer pipes <NUM> is a circular pipe having an outer diameter Do that is equal to the outer diameter of a first heat transfer pipe <NUM> and the outer diameter of a second heat transfer pipe <NUM>. Further, each of the plurality of second heat transfer pipes <NUM> is a circular pipe having a wall thickness tP that is equal to the wall thickness of a first heat transfer pipe <NUM> and the wall thickness of a second heat transfer pipe <NUM>.

In the present embodiment, the respective wall thicknesses tP of the first heat transfer pipes <NUM>, the second heat transfer pipes <NUM>, and the second heat transfer pipes <NUM> each range, for example, from <NUM> to <NUM>. Note, however, that the respective wall thicknesses of the first heat transfer pipes <NUM>, the second heat transfer pipes <NUM>, and the second heat transfer pipes <NUM> may be each less than <NUM> or may be each greater than <NUM>.

In a process of manufacturing the heat exchanger <NUM>, the first heat transfer pipes <NUM>, the second heat transfer pipes <NUM>, and the second heat transfer pipes <NUM> may be subjected to pipe expanding. In this case, the respective outer diameters Do of the first heat transfer pipes <NUM>, the second heat transfer pipes <NUM>, and the second heat transfer pipes <NUM> may of course be specified by outer diameters after pipe expanding.

The following describes heat exchanger performance and cost performance in a case in which the outer diameters Do, the row pitches L1, the step pitches L2, and the wall thicknesses tP of the heat transfer pipes of the heat exchanger <NUM> are varied.

Table <NUM> is a table showing effects exerted on the intra-pipe volume V, the extra-pipe heat transfer coefficient αo, the ventilation resistance ΔP, the extra-pipe heat transfer area Ao, and the heat exchanger weight M in a case in which the outer diameters Do, the row pitches L1, the step pitches L2, and the wall thicknesses tP of the heat transfer pipes of the heat exchanger <NUM> according to the present embodiment are varied. It should be noted, in Table <NUM>, when each of the parameters, namely the outer diameters Do, the row pitches L1, the step pitches L2, and the wall thicknesses tP of the heat transfer pipes, are varied, the other parameters are fixed.

The intra-pipe volume V [m<NUM>] is a value obtained by multiplying the cross-sectional area of an interior channel of one heat transfer pipe by the length of the heat transfer pipe. The extra-pipe heat transfer coefficient αo [W/m<NUM>·K] is the proportion of the amount of heat that is transferred between an outer wall surface of a heat transfer pipe and air. The ventilation resistance ΔP [Pa] is a pressure loss of air passing through the heat exchanger <NUM>. The extra-pipe heat transfer area Ao [m<NUM>] is the gross area of the respective outer wall surfaces of the heat transfer pipes of the heat exchanger <NUM>. The heat exchanger weight M [kg] is the weight (core weight) of a heat exchange core unit of the heat exchanger <NUM> and the heat exchange core unit is formed by the heat transfer pipes and the fins.

In a case in which the outer diameter Do is reduced and the step pitch L2 is increased for the purpose of reducing the intra-pipe volume V, that is, the amount of refrigerant charge, the extra-pipe heat transfer coefficient αo decreases, so that energy-saving effectiveness decreases because of lack of heat transfer performance. Accordingly, for improving the heat transfer performance, it is necessary to increase the extra-pipe heat transfer area Ao by increasing the row pitch L1 or to increase the extra-pipe heat transfer coefficient αo by reducing the row pitch L1 and increase the extra-pipe heat transfer area Ao by increasing the number of rows of the heat transfer pipes. However, in either case, the amount of use of the fins or the heat transfer pipes increases, so that there is a possibility that cost performance, that is, the heat exchange performance of the heat exchanger <NUM> per unit weight, may decrease. Further, in a case in which the wall thickness tP of each of the heat transfer pipes is increased for the purpose of reducing the intra-pipe volume V, that is, the amount of refrigerant charge, the amount of use of the heat transfer pipes increases, so that there is a possibility that cost performance may similarly decrease. For these reasons, it is necessary to appropriately set the outer diameters Do, the row pitches L1, the step pitches L2, and the wall thicknesses tP of the heat transfer pipes of the heat exchanger <NUM> to achieve both a reduction in the intra-pipe volume V and an increase in cost performance of the heat exchanger <NUM>.

The following describes the extra-pipe heat exchange performance of the heat exchanger <NUM> per unit weight.

<FIG> each show a relationship between the area ratio of heat transfer pipes to fins and extra-pipe heat exchange performance per unit weight in the heat exchanger <NUM> according to Embodiment <NUM> as a ratio to a maximum value at Do = <NUM> for each outer diameter Do (Do = <NUM>, <NUM>, <NUM>, <NUM>, <NUM>) of the heat transfer pipes.

Note here that the heat transfer pipes may include first heat transfer pipes <NUM>, second heat transfer pipes <NUM>, and second heat transfer pipes <NUM>. The fins may include first fins <NUM>, second fins <NUM>, and second fins <NUM>. The area A is an area represented by the product L1 × L2 of a row pitch L1 and a step pitch L2. The area A is equivalent to the area of each fin per heat transfer pipe. Also, the area B is an area represented by ((Do - <NUM> × tP)/<NUM>)<NUM> × π using the outer diameter Do and wall thickness tP of each of the heat transfer pipes. The area B is equivalent to the cross-sectional area of an interior channel of one heat transfer pipe.

In each of <FIG>, the horizontal axis of the graph represents the area ratio B/A of the area B to the area A. The area ratio B/A represents, as an area ratio, the density at which the heat transfer pipes are placed through the fins. A relationship between the area ratio B/A and the intra-pipe volume V is described here. <FIG> is a graph showing a relationship between the area ratio B/A and the intra-pipe volume V in the heat exchanger <NUM> according to Embodiment <NUM>. <FIG> shows effects of the area ratio B/A on the intra-pipe volume V in cases where Outer Diameter Do = <NUM> and Do = <NUM> and Wall Thickness tP = <NUM>. As shown in <FIG>, the intra-pipe volume V decreases as the area ratio B/A decreases.

In each of <FIG>, the vertical axis of the graph represents the extra-pipe heat transfer performance (Extra-pipe Heat Transfer Performance/Weight) of the heat exchanger <NUM> per unit weight as a ratio to a maximum value at Do = <NUM>. The extra-pipe heat exchange performance is (Extra-pipe Heat Transfer Area Ao × Extra-pipe Heat Transfer Coefficient αo)/ΔP. Extra-pipe Heat Transfer Area Ao × Extra-pipe Heat Transfer Coefficient αo is the extra-pipe heat transfer performance.

A relationship between each of the extra-pipe heat transfer performance, the ventilation resistance ΔP, the heat exchanger weight M, and the extra-pipe heat exchange performance and the area ratio B/A is described here with reference to <FIG>.

<FIG> is a graph showing a relationship between the area ratio B/A and the extra-pipe heat transfer performance (Ao × αo) in the heat exchanger <NUM> according to Embodiment <NUM>. <FIG> shows effects of the area ratio B/A on the extra-pipe heat transfer performance (Extra-pipe Heat Transfer Area Ao × Extra-pipe Heat Transfer Coefficient αo) in cases where Outer Diameter Do = <NUM> and Do = <NUM> and Wall Thickness tP = <NUM>. As the area ratio B/A increases, the heat transfer pipes are located closer to each other and thermal conductivity improves, so that the extra-pipe heat transfer performance (Ao × αo) increases. Further, a comparison made at identical area ratios B/A shows that the extra-pipe heat transfer performance (Ao × αo) increases as the outer diameter Do of each of the heat transfer pipes decreases. A reason for this is that the heat transfer pipes are located closer to each other as the outer diameter Do of each of the heat transfer pipes decreases. For example, as shown in <FIG>, a comparison made at identical area ratios B/A shows that the extra-pipe heat transfer performance (Ao × αo) is higher when Do = <NUM> than when Do = <NUM>. Further, as an example, in a case in which Area Ratio B/A = <NUM>, L1 = L2 = <NUM> when Do = <NUM>, and L1 = L2 = <NUM> when Do = <NUM>. That is, the heat transfer pipes are located closer to each other in a case in which Do = <NUM> than in a case in which Do = <NUM>.

<FIG> is a graph showing a relationship between the area ratio B/A and the ventilation resistance ΔP in the heat exchanger <NUM> according to Embodiment <NUM>. <FIG> shows effects of the area ratio B/A on the ventilation resistance ΔP in cases where Outer Diameter Do = <NUM> and Do = <NUM> and Wall Thickness tP = <NUM>. As the area ratio B/A increases, the heat transfer pipes are located closer to each other and resistance to the flow of air passing through the heat exchanger <NUM> increases, so that the ventilation resistance ΔP increases. In particular, the heat transfer pipes are located to closer to each other as the outer diameter Do of each of the heat transfer pipes decreases, with the same area ratio B/A. For this reason, when the area ratio B/A increases, those heat transfer pipes with a smaller outer diameter Do suffer from earlier closure of air trunks through which air circulates and from a higher rate of increase in the ventilation resistance ΔP than do those heat transfer pipes with a large outer diameter Do.

<FIG> is a graph showing a relationship between the area ratio B/A and the heat exchanger weight M in the heat exchanger <NUM> according to Embodiment <NUM>. <FIG> shows effects of the area ratio B/A on the heat exchanger weight M in cases where Outer Diameter Do = <NUM> and Do = <NUM> and Wall Thickness tP = <NUM>. The value of the weight (core weight) of the heat exchanger <NUM> has a positive correlation with the amount of material of the heat exchanger <NUM> to be used and the manufacturing cost of the heat exchanger <NUM>. For this reason, the value of Extra-pipe Heat Exchange Performance/Weight represented by the vertical axis of the graph in each of <FIG> is equivalent to the cost performance of the heat exchanger <NUM>. As the area ratio B/A decreases, the number of heat transfer pipes that are mounted in the heat exchanger <NUM> decreases, so that the heat exchanger weight M decreases.

<FIG> is a graph showing a relationship between the area ratio B/A and the extra-pipe heat exchange performance in the heat exchanger <NUM> according to Embodiment <NUM>. <FIG> shows effects of the area ratio B/A on the extra-pipe heat exchange performance ((Ao × αo)/ΔP) in cases where Outer Diameter Do = <NUM> and Do = <NUM> and Wall Thickness tP = <NUM>. Further, <FIG> is a graph showing a relationship between the area ratio B/A and the extra-pipe heat exchange performance per unit weight in the heat exchanger <NUM> according to Embodiment <NUM>. <FIG> shows effects of the area ratio B/A on the extra-pipe heat exchange performance per unit weight ((Ao × αo)/ΔP/M) in cases where Outer Diameter Do = <NUM> and Do = <NUM> and Wall Thickness tP = <NUM>. As shown in <FIG>, the characteristic of the extra-pipe heat exchange performance to the area ratio B/A has a maximum value. Further, as shown in <FIG>, the heat exchanger weight M monotonically increases when the area ratio B/A increases. For this reason, as shown in <FIG>, the characteristic of the extra-pipe heat exchange performance per unit weight to the area ratio B/A also has a maximum value. Further, as the area ratio B/A increases, the heat exchanger weight M increases, so that the extra-pipe heat exchange performance per unit weight has a lower gradient in a region in which the area ratio B/A is high. Further, as the outer diameter Do of each of the heat transfer pipes decreases, the rate of change in the ventilation resistance ΔP increases, so that the extra-pipe heat exchange performance per unit weight to the area ratio B/A having a maximum value is lower. Further, as shown in <FIG>, the maximum value of the extra-pipe heat exchange performance ((Ao × αo)/ΔP) increases as the outer diameter Do of each of the heat transfer pipes decreases.

Continued reference is made to <FIG> vary in value of the wall thickness tP from one another. <FIG> is a graph showing a case in which the wall thickness tP is <NUM>. <FIG> is a graph showing a case in which the wall thickness tP is <NUM>. <FIG> is a graph showing a case in which the wall thickness tP is <NUM>. <FIG> is a graph showing a case in which the wall thickness tP is <NUM>. In general, in a case in which a hydrofluorocarbon is used as refrigerant, the wall thickness tP from approximately <NUM> to <NUM> when Do = <NUM> or smaller is often used.

The extra-pipe heat exchange performance of the heat exchanger <NUM> according to the present embodiment per unit weight as shown in <FIG> is calculated by the following method.

In general, the heat transfer coefficient αa [W/m<NUM>·K] between air and the fins is defined by the following equations.

Note here that Nu is a Nusselt number and Re is a Reynolds number. Pr is a Prandtl number, λa is the thermal conductivity of air, and v is the kinematic viscosity of air. At ordinary temperatures and pressures, Pr = <NUM>, λa = <NUM> [W/m·K], and v = <NUM> [m<NUM>/s]. Further, C<NUM> and C<NUM> are constants, and NL is the number of rows of the heat transfer pipes.

The characteristic length De [m] is defined by the following equations.

Note here that Vc [m<NUM>] is a free flow volume, FP [m] is a fin pitch, tF [m] is the thickness of each of the fins, and dc [m] is a fin collar outer diameter.

The wind velocity U [m/s] based on a free passage volume between fins and the front wind velocity Uf [m/s] of the heat exchanger are defined by the following equations.

Note here that Qair [m<NUM>/s] is the flow rate of air flowing into the heat exchanger, EH is the overall height of the heat exchanger in the step direction, and EL is the overall height of the heat exchanger in a direction in which the fins are stacked.

In general, the extra-pipe heat transfer coefficient αo is defined by the following equations.

Note here that η is fin efficiency and αa is an air-side heat transfer coefficient. Ao [m<NUM>] is the air-side total heat transfer area of the heat exchanger, Ap [m<NUM>] is the air-side pipe heat transfer area of the heat exchanger, AF [m<NUM>] is the air-side fin heat transfer area of the heat exchanger, and Acon [m<NUM>] is the area of contact between the heat transfer pipes and the fins. Ao, Ap, AF, and Acon are values that can be calculated once the dimensions dependent on the shape of the heat exchanger, namely the number NL of rows of heat transfer pipes, the number ND of steps of heat transfer pipes, the number NF of fins, the row pitch L1, the step pitch L2, the fin pitch FP, the fin thickness tF, and the outer diameter Do of each of the heat transfer pipes, are determined. The contact heat transfer coefficient αc between the heat transfer pipes and the fins of the heat exchanger is constant.

The fin efficiency η is defined by the following equations.

Note here that dF [m] is a fin equivalent diameter and λF [W/m·K] is the thermal conductivity of the fins.

The ventilation resistance ΔP [Pa] is defined by the following equations.

Note here that f is a coefficient of friction loss, ρ is the density of air, and C<NUM> and C<NUM> are constants.

It should be noted that the constants C<NUM>, C<NUM>, C<NUM>, and C<NUM>, which are used in the Nusselt number Nu and a coefficient of flow loss f, are set to represent the thermal conductivity αa and ventilation resistance ΔP of the fins of a heat exchanger of a commercially widely-distributed common air-conditioning apparatus.

The extra-pipe heat exchange performance of the heat exchanger <NUM> according to the present embodiment per unit weight as shown in <FIG> is calculated under the following conditions.

As a comparative example, a performance calculation is performed under the following calculation conditions. The other parameters are similar to the aforementioned calculation conditions. The calculation conditions of the comparative example are conditions under which the intra-pipe volume is smallest in Patent Literature <NUM> (<CIT><CIT>).

Further, under the calculation conditions of the comparative example, the area ratio B/A is <NUM> in a case in which Wall Thickness tP = <NUM>, is <NUM> in a case in which Wall Thickness tP = <NUM>, is <NUM> in a case in which Wall Thickness tP = <NUM>, and is <NUM> in a case in which Wall Thickness tP = <NUM>.

As shown in <FIG>, there is a region in which the outer diameter Do of each pipe is less than <NUM>, Extra-pipe Heat Exchange Performance/Weight [Ratio] exceeds <NUM>%, and the area ratio B/A can fall below that of the comparative example. That is, if the area ratio B/A falls below that of the comparative example, the intra-pipe volume V can be made smaller than that of the comparative example, and the cost performance of the heat exchanger <NUM> can be made higher than that of the comparative example.

A range of numerical values of the area ratio B/A in which Extra-pipe Heat Exchange Performance/Weight [Ratio] exceeds <NUM>% and the area ratio B/A can fall below that of the comparative example varies with the outer diameter Do and the wall thickness tP. For example, as shown in <FIG>, in a case where Wall Thickness tP = <NUM> and Do = <NUM>, Extra-pipe Heat Exchange Performance/Weight [Ratio] exceeds <NUM>% and the area ratio B/A falls below that of the comparative example, provided <NUM> ≤ B/A ≤ <NUM>. Further, for example, as shown in <FIG>, in a case where Wall Thickness tP = <NUM> and Do = <NUM>, Extra-pipe Heat Exchange Performance/Weight [Ratio] exceeds <NUM>% and the area ratio B/A falls below that of the comparative example, provided <NUM> ≤ B/A ≤ <NUM>. Further, for example, as shown in <FIG>, in a case where Wall Thickness tP = <NUM> and Do = <NUM>, Extra-pipe Heat Exchange Performance/Weight [Ratio] exceeds <NUM>% and the area ratio B/A falls below that of the comparative example, provided <NUM> ≤ B/A ≤ <NUM>.

An upper limit of the range of numerical values of the area ratio B/A in which the outer diameter Do of each pipe is less than <NUM>, Extra-pipe Heat Exchange Performance/Weight [Ratio] exceeds <NUM>%, and the area ratio B/A can fall below that of the comparative example, shown in <FIG>, is expressed by Formula (<NUM>) below as a function of the outer diameter Do and the wall thickness tP. Further, a lower limit of the range of numerical values of the area ratio B/A in which Extra-pipe Heat Exchange Performance/Weight [Ratio] exceeds <NUM>% and the area ratio B/A can fall below that of the comparative example, shown in <FIG>, is expressed by Formula (<NUM>) below as a function of the outer diameter Do and the wall thickness tP.

It should be noted that In is a natural logarithm whose base is e.

Further, the area ratio B/A of the comparative example is expressed by Formula (<NUM>) below as a function of the wall thickness tP.

The upper limit function F (Do, tP) is an approximate expression calculated, for example, by a logarithmic approximation of the method of least squares after obtaining, for each wall thickness tP and each outer diameter Do, an upper limit value of the range of numerical values of the area ratio B/A in which Extra-pipe Heat Exchange Performance/Weight [Ratio] exceeds <NUM>% and the area ratio B/A can fall below that of the comparative example. Further, the lower limit function G (Do, tP) is an approximate expression calculated, for example, by a power approximation of the method of least squares after obtaining, for each wall thickness tP and each outer diameter Do, an upper limit value of the range of numerical values of the area ratio B/A in which Extra-pipe Heat Exchange Performance/Weight [Ratio] exceeds <NUM>% and the area ratio B/A can fall below that of the comparative example. Further, the area ratio function H (tP) of the comparative example is an approximate expression calculated, for example, by a power approximation of the method of least squares after obtaining a value of the area ratio B/A of the comparative example for each wall thickness tP.

With Formulas (<NUM>) to (<NUM>) above, a relationship among the outer diameter Do, the area ratio B/A, and the wall thickness tP in which Extra-pipe Heat Exchange Performance/Weight [Ratio] exceeds <NUM>% and the area ratio B/A can fall below that of the comparative example is expressed by Formula (<NUM>) below.

Specific examples of the range of numerical values identified by Formula (<NUM>) above under the aforementioned calculation conditions are described here with reference to <FIG>.

<FIG> are each a graph showing a relationship between the outer diameter Do of each of the heat transfer pipes and the area ratio B/A in the heat exchanger <NUM> according to Embodiment <NUM>. In each of <FIG>, the vertical axis of the graph represents the area ratio B/A of the area B to the area A. The horizontal axis of the graph represents the outer diameter Do of each of the heat transfer pipes.

In each of <FIG>, the upper limit function F (Do, tP) is shown as "B/A UPPER LIMIT". Further, the lower limit function G (Do, tP) is shown as "B/A LOWER LIMIT". Further, the area ratio function H (tP) of the comparative example is shown as "B/A COMPARATIVE EXAMPLE". <FIG> vary in value of the wall thickness tP from one another. <FIG> is a graph showing a case in which the wall thickness tP is <NUM>. <FIG> is a graph showing a case in which the wall thickness tP is <NUM>. <FIG> is a graph showing a case in which the wall thickness tP is <NUM>. <FIG> is a graph showing a case in which the wall thickness tP is <NUM>.

As shown in <FIG>, at each wall thickness tP, Extra-pipe Heat Exchange Performance/Weight [Ratio] exceeds <NUM>% and the area ratio B/A can fall below that of the comparative example, provided the outer diameter Do and the area ratio B/A fall within the range greater than or equal to "B/A LOWER LIMIT", less than or equal to "B/A UPPER LIMIT", and less than "B/A COMPARATIVE EXAMPLE" and the outer diameter Do falls within the range of Do < <NUM>. That is, the intra-pipe volume V can be made smaller than that of the comparative example, and the cost performance of the heat exchanger <NUM> can be made higher than that of the comparative example.

As noted above, configuring the heat exchanger <NUM> such that when Outer Diameter Do < <NUM>, Lower Limit Function G (Do, tP) ≤ Area Ratio B/A ≤ Upper Limit Function F (Do, tP) and Area Ratio B/A < Area Ratio Function H (tP) of Comparative Example allows the amount of refrigerant charge to fall below that of the comparative example while allowing Extra-pipe Heat Exchange Performance/Weight [Ratio] to exceed <NUM>%. This in turn makes it possible to improve heat exchanger performance while reducing the inner capacity of each of the heat transfer pipes of the heat exchanger <NUM>. Therefore, the heat exchanger <NUM> according to the present embodiment can achieve both improvement in cost performance and a reduction in total value of GWP through a reduction in amount of refrigerant charge. As a result, this makes it possible to reduce the amount of refrigerant charge while improving energy-saving effectiveness in a refrigeration cycle apparatus including the heat exchanger <NUM>.

Further, the foregoing calculation conditions of the heat exchanger <NUM> according to the present embodiment correspond to cooling rated conditions of an air-conditioning apparatus serving as an example of a refrigeration cycle apparatus. This makes it possible to, under the cooling rated conditions of an air-conditioning apparatus, reduce the amount of refrigerant charge while improving energy-saving effectiveness. It should be noted that even under other conditions such as cooling intermediate conditions, heating rated conditions, and heating intermediate conditions of an air-conditioning apparatus serving as an example of a refrigeration cycle apparatus, the heat exchanger <NUM> according to the present embodiment brings about effects that are similar to those brought about under the cooling rated conditions.

A heat exchanger according to Embodiment <NUM> is described. <FIG> is a cross-sectional view showing a configuration of some components of a heat exchanger <NUM> according to the present embodiment. As with <FIG>, <FIG> shows a configuration of the heat exchanger <NUM> as sectioned along the plane perpendicular to the direction in which the first heat transfer pipes <NUM> extend. Constituent elements having the same functions and workings as those of Embodiment <NUM> are given the same reference signs, and a description of such constituent elements is omitted.

In the heat exchanger <NUM> of the present embodiment, as shown in <FIG>, the outer diameter Doa of each of the first heat transfer pipes <NUM> of the first heat exchange unit <NUM> located furthest windward is smaller than the outer diameter Dob of each of the second heat transfer pipes <NUM> of the second heat exchange unit <NUM> (Doa < Dob). A step pitch L2 between first heat transfer pipes <NUM> is identical to a step pitch L2 between second heat transfer pipes <NUM>. Further, each of the plurality of first heat transfer pipes <NUM> is a circular pipe having a wall thickness tP that is equal to the wall thickness of a second heat transfer pipe <NUM>.

In both the first heat exchange unit <NUM> and the second heat exchange unit <NUM>, the relation of Formula (<NUM>), which is described above in Embodiment <NUM>, is satisfied. Further, a value of B/A in the first heat exchange unit <NUM> is smaller than a value of B/A in the second heat exchange unit <NUM>.

<FIG> is a cross-sectional view showing a configuration of some components of a heat exchanger <NUM> according to a modification of the present embodiment. In the heat exchanger <NUM> of the present modification, as shown in <FIG>, a step pitch L2a between first heat transfer pipes <NUM> of the first heat exchange unit <NUM> located furthest windward is greater than a step pitch L2b between second heat transfer pipes <NUM> of the second heat exchange unit <NUM> (L2a > L2b). The outer diameter Do of each of the first heat transfer pipes <NUM> is identical to the outer diameter Do of each of the second heat transfer pipes <NUM>. Further, each of the plurality of first heat transfer pipes <NUM> is a circular pipe having a wall thickness tP that is equal to the wall thickness of a second heat transfer pipe <NUM>.

As described above, the heat exchanger <NUM> according to the present embodiment further includes a plurality of heat exchange units, arrayed along the direction of airflow, each of which has one or more of the plurality of heat transfer pipes. The plurality of heat exchange units include a first heat exchange unit <NUM> located furthest windward and at least one second heat exchange unit <NUM> located further leeward than the first heat exchange unit <NUM>. A value of B/A in the first heat exchange unit <NUM> is smaller than a value of B/A in the at least one second heat exchange unit <NUM>.

In general, in the first heat exchange unit <NUM> located furthest windward, frost easily forms, as a great temperature difference between the first fins <NUM> or the first heat transfer pipes <NUM> and air results in an increased amount of heat that is exchanged. The foregoing configuration makes it possible to make the first heat exchange unit <NUM> lower in heat exchange performance than the second heat exchange unit <NUM>. This makes it possible to inhibit the formation of frost in the first heat exchange unit <NUM> and therefore makes it possible to prevent an air trunk of the first heat exchange unit <NUM> from being closed by an increased amount of frost that is formed. This makes it possible to improve cost performance while reducing deterioration in ventilation performance of the heat exchanger <NUM>.

A refrigeration cycle apparatus according to Embodiment <NUM> is described. <FIG> is a refrigerant circuit diagram showing a configuration of a refrigeration cycle apparatus <NUM> according to Embodiment <NUM>. In the present embodiment, an air-conditioning apparatus is an example of the refrigeration cycle apparatus <NUM>. As shown in <FIG>, the refrigeration cycle apparatus <NUM> includes a refrigeration cycle circuit <NUM> through which refrigerant circulates. The refrigeration cycle circuit <NUM> is configured such that a compressor <NUM>, a four-way valve <NUM>, an outdoor heat exchanger <NUM>, an expansion valve <NUM>, and an indoor heat exchanger <NUM> are connected in a circular pattern via refrigerant pipes. Further, the refrigeration cycle apparatus <NUM> includes an outdoor fan <NUM> configured to supply air to the outdoor heat exchanger <NUM> and an indoor fan <NUM> configured to supply air to the indoor heat exchanger <NUM>. In the refrigeration cycle apparatus <NUM>, the compressor <NUM> is driven so that a refrigeration cycle is executed in which the refrigerant circulates through the refrigeration cycle circuit <NUM> while the refrigerant changes its phase. The outdoor heat exchanger <NUM> allows the air supplied by the outdoor fan <NUM> and the refrigerant, which is an inner fluid, to exchange heat with each other. The indoor heat exchanger <NUM> allows the air supplied by the indoor fan <NUM> and the refrigerant, which is an inner fluid, to exchange heat with each other. As at least either the outdoor heat exchanger <NUM> or the indoor heat exchanger <NUM>, the heat exchanger <NUM> of Embodiment <NUM> or <NUM> is used.

The refrigeration cycle apparatus <NUM> includes an outdoor unit <NUM> and an indoor unit <NUM> as heat exchange units. The outdoor unit <NUM> houses the compressor <NUM>, the four-way valve <NUM>, the outdoor heat exchanger <NUM>, the expansion valve <NUM>, and the outdoor fan <NUM>. The indoor unit <NUM> houses the indoor heat exchanger <NUM> and the indoor fan <NUM>. The outdoor unit <NUM> and the indoor unit <NUM> are connected to each other via a gas pipe <NUM> and a liquid pipe <NUM>, which are some of the refrigerant pipes.

Operation of the refrigeration cycle apparatus <NUM> is described by describing cooling operation as an example. For cooling operation, the four-way valve <NUM> is switched such that refrigerant discharged from the compressor <NUM> flows into the outdoor heat exchanger <NUM>. The high-pressure gas refrigerant discharged from the compressor <NUM> flows into the outdoor heat exchanger <NUM> via the four-way valve <NUM>. During cooling operation, the outdoor heat exchanger <NUM> operates as a condenser. That is, the outdoor heat exchanger <NUM> allows refrigerant circulating through inside and outdoor air supplied by the outdoor fan <NUM> to exchange heat with each other, so that the refrigerant transfers heat of condensation to the outdoor air. This causes the gas refrigerant having flowed into the outdoor heat exchanger <NUM> to condense into high-pressure liquid refrigerant.

The liquid refrigerant having flowed out of the outdoor heat exchanger <NUM> is decompressed by the expansion valve <NUM> into low-pressure two-phase refrigerant. The two-phase refrigerant having flowed out of the expansion valve <NUM> flows into the indoor heat exchanger <NUM> via the liquid pipe <NUM>. During cooling operation, the indoor heat exchanger <NUM> operates as an evaporator. That is, the indoor heat exchanger <NUM> allows refrigerant circulating through inside and indoor air supplied by the indoor fan <NUM> to exchange heat with each other, so that the refrigerant removes heat of evaporation from the indoor air. This causes the two-phase refrigerant having flowed into the indoor heat exchanger <NUM> to evaporate into low-pressure gas refrigerant. The indoor air having passed through the indoor heat exchanger <NUM> is cooled by exchanging heat with the refrigerant. The gas refrigerant having flowed out of the indoor heat exchanger <NUM> is suctioned into the compressor <NUM> via the gas pipe <NUM> and the four-way valve <NUM>. The gas refrigerant suctioned into the compressor <NUM> is compressed into high-pressure gas refrigerant. During cooling operation, the refrigeration cycle described above is continuously and repeatedly executed. Although not described, for heating operation, a direction of refrigerant flow is switched by the four-way valve <NUM> such that the outdoor heat exchanger <NUM> operates as an evaporator and the indoor heat exchanger <NUM> operates as a condenser.

As described above, the refrigeration cycle apparatus <NUM> according to the present embodiment includes the heat exchanger <NUM> of Embodiment <NUM> or <NUM>. This configuration allows the refrigeration cycle apparatus <NUM> to achieve both a reduction in total value of GWP and improvement in energy-saving effectiveness.

Embodiments <NUM> to <NUM> and the modifications described above may be combined with each other.

Claim 1:
A heat exchanger (<NUM>), comprising:
a plurality of fins (<NUM>, <NUM>) arranged in parallel to each other; and
a plurality of heat transfer pipes (<NUM>, <NUM>) each extending in a direction intersecting the plurality of fins (<NUM>, <NUM>),
in a plane perpendicular to a direction in which the plurality of heat transfer pipes (<NUM>, <NUM>) extend, the plurality of heat transfer pipes (<NUM>, <NUM>) being placed in a plurality of rows in a row direction that is along a direction of airflow at a row pitch L1,
in the plane, the plurality of heat transfer pipes (<NUM>, <NUM>) being placed in a plurality of steps in a step direction perpendicular to the row direction at a step pitch L2,
where an outer diameter of each of the plurality of heat transfer pipes (<NUM>, <NUM>) is defined as Do,
a wall thickness of a portion having a smallest distance between an outer wall surface and an inner wall surface of each of the plurality of heat transfer pipes (<NUM>, <NUM>) is defined as tP,
an area represented by a numerical expression of L1 × L2 is defined as A, and
an area represented by a numerical expression of ((Do - <NUM> × tP)/<NUM>)<NUM> × π is defined as B,
a relation of Do < <NUM>, characterized by
a relation of (<NUM> × tP<NUM> - <NUM> × tP + <NUM>) × Do ^ (-<NUM> × tP<NUM> - <NUM> × tP + <NUM>)≤ B/A ≤ (<NUM> × tP<NUM> - <NUM> × tP + <NUM>) × In (Do) + (<NUM> × tP<NUM> + <NUM> × tP + <NUM>), and
a relation of B/A < <NUM> × tP<NUM> - <NUM> × tP + <NUM> being satisfied.