FORM 2
THE PATENTS ACT, 1970
(39 of 1970)
&
THE PATENTS RULES, 2003
COMPLETE SPECIFICATION
[See section 10, Rule 13]
HEAT EXCHANGER AND REFRIGERATION CYCLE APPARATUS;
MITSUBISHI ELECTRIC CORPORATION, A CORPORATION ORGANISED AND
EXISTING UNDER THE LAWS OF JAPAN, WHOSE ADDRESS IS 7-3,
MARUNOUCHI 2-CHOME, CHIYODA-KU, TOKYO 1008310, JAPAN
THE FOLLOWING SPECIFICATION PARTICULARLY DESCRIBES THE INVENTION
AND THE MANNER IN WHICH IT IS TO BE PERFORMED

2
DESCRIPTION
Technical Field
[0001]
5 The present disclosure relates to a heat exchanger including a plurality of fins
and a plurality of heat transfer pipes each extending in a direction intersecting the
plurality of fins and to a refrigeration cycle apparatus including the same.
Background Art
[0002]
10 Patent Literature 1 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
15 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.
Citation List
20 Patent Literature
[0003]
Patent Literature 1: Japanese Unexamined Patent Application Publication No.
2013-92306
Summary of Invention
25 Technical Problem
[0004]
The heat exchanger of Patent Literature 1 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
30 GWP of a refrigeration cycle apparatus. A possible way of reducing the amount of

3
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
5 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
10 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
15 difficult to improve the heat exchanger performance of a heat exchanger while
reducing the inner capacity of each of the heat transfer pipes.
[0005]
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
20 exchanger performance of the heat exchanger while reducing the inner capacity of
heat transfer pipes and a refrigeration cycle apparatus including the same.
Solution to Problem
[0006]
A heat exchanger according to an embodiment of the present disclosure
25 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
30 transfer pipes are placed in a plurality of steps in a step direction perpendicular to the

4
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
5 expression of L1  L2 is defined as A, and an area represented by a numerical
expression of ((Do – 2  tP)/2)2   is defined as B, a relation of Do < 5.5 mm, a
relation of (0.0219  tP2
– 0.0185  tP + 0.0043)  ln (Do) + (1.6950  tP2
+ 1.8455 
tP + 1.5416)  B/A  (0.2076  tP2
– 0.1480  tP + 0.0545)  Do ^ (–0.0021  tP2
–
0.0528  tP + 0.0164), and a relation of B/A < 0.0076  tP2
– 0.0417  tP + 0.0574 are
10 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.
Advantageous Effects of Invention
15 [0007]
An embodiment of the present disclosure makes it possible to improve the heat
exchanger performance of a heat exchanger while reducing the inner capacity of heat
transfer pipes.
Brief Description of Drawings
20 [0008]
[Fig. 1] Fig. 1 is a cross-sectional view showing a configuration of some
components of a heat exchanger 100 according to Embodiment 1.
[Fig. 2] Fig. 2 is a cross-sectional view showing a configuration of some
components of a heat exchanger 100 according to a modification of Embodiment 1.
25 [Fig. 3] Fig. 3 is a graph showing 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 100 according to Embodiment 1 for each outer diameter Do of the
heat transfer pipes.
[Fig. 4] Fig. 4 is a graph showing a relationship between the area ratio of heat
30 transfer pipes to fins and extra-pipe heat exchange performance per unit weight in the

5
heat exchanger 100 according to Embodiment 1 for each outer diameter Do of the
heat transfer pipes.
[Fig. 5] Fig. 5 is a graph showing a relationship between the area ratio of heat
transfer pipes to fins and extra-pipe heat exchange performance per unit weight in the
5 heat exchanger 100 according to Embodiment 1 for each outer diameter Do of the
heat transfer pipes.
[Fig. 6] Fig. 6 is a graph showing 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 100 according to Embodiment 1 for each outer diameter Do of the
10 heat transfer pipes.
[Fig. 7] Fig. 7 is a graph showing a relationship between the area ratio B/A and
the intra-pipe volume V in the heat exchanger 100 according to Embodiment 1.
[Fig. 8] Fig. 8 is a graph showing a relationship between the area ratio B/A and
the extra-pipe heat transfer performance (Ao  o) in the heat exchanger 100
15 according to Embodiment 1.
[Fig. 9] Fig. 9 is a graph showing a relationship between the area ratio B/A and
the ventilation resistance P in the heat exchanger 100 according to Embodiment 1.
[Fig. 10] Fig. 10 is a graph showing a relationship between the area ratio B/A
and the heat exchanger weight M in the heat exchanger 100 according to
20 Embodiment 1.
[Fig. 11] Fig. 11 is a graph showing a relationship between the area ratio B/A
and the extra-pipe heat exchange performance in the heat exchanger 100 according
to Embodiment 1.
[Fig. 12] Fig. 12 is a graph showing a relationship between the area ratio B/A
25 and the extra-pipe heat exchange performance per unit weight in the heat exchanger
100 according to Embodiment 1.
[Fig. 13] Fig. 13 is a graph showing a relationship between the outer diameter
Do of each heat transfer pipe and the area ratio B/A in the heat exchanger 100
according to Embodiment 1.
30 [Fig. 14] Fig. 14 is a graph showing a relationship between the outer diameter

6
Do of each heat transfer pipe and the area ratio B/A in the heat exchanger 100
according to Embodiment 1.
[Fig. 15] Fig. 15 is a graph showing a relationship between the outer diameter
Do of each heat transfer pipe and the area ratio B/A in the heat exchanger 100
5 according to Embodiment 1.
[Fig. 16] Fig. 16 is a graph showing a relationship between the outer diameter
Do of each heat transfer pipe and the area ratio B/A in the heat exchanger 100
according to Embodiment 1.
[Fig. 17] Fig. 17 is a cross-sectional view showing a configuration of some
10 components of a heat exchanger 100 according to Embodiment 2.
[Fig. 18] Fig. 18 is a cross-sectional view showing a configuration of some
components of a heat exchanger 100 according to a modification of Embodiment 2.
[Fig. 19] Fig. 19 is a refrigerant circuit diagram showing a configuration of a
refrigeration cycle apparatus 200 according to Embodiment 3.
15 Description of Embodiments
[0009]
Embodiment 1.
A heat exchanger according to Embodiment 1 is described. Fig. 1 is a crosssectional view showing a configuration of some components of a heat exchanger 100
20 according to Embodiment 1. Fig. 1 shows a configuration of the heat exchanger 100
as sectioned along a plane perpendicular to a direction in which the after-mentioned
first heat transfer pipes 12 extend. The heat exchanger 100 is used as a heat
source side heat exchanger or a load side heat exchanger of a refrigeration cycle
apparatus. The heat exchanger 100 is a cross-fin fin-and-tube heat exchanger that
25 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. 1,
a thick arrow outline with a blank inside represents a direction of airflow.
[0010]
30 As shown in Fig. 1, the heat exchanger 100 includes, as a plurality of heat

7
exchange units arrayed along the direction of airflow, a first heat exchange unit 10
located furthest windward and a second heat exchange unit 20 located further
leeward than the first heat exchange unit 10.
[0011]
5 The first heat exchange unit 10 includes a plurality of first fins 11 arranged
parallel to each other at intervals and a plurality of first heat transfer pipes 12 each
passing through the plurality of first fins 11 and each extending parallel to each other
in a direction intersecting the plurality of first fins 11. Each of the plurality of first fins
11 has a rectangular flat-plate shape elongated in one direction. Each of the plurality
10 of first fins 11 is placed perpendicular to a direction in which the first heat transfer
pipes 12 extend. The plurality of first fins 11 are provided parallel to each other at
regular placement pitches in a direction perpendicular to a surface of paper of Fig. 1,
that is, the direction in which the first heat transfer pipes 12 extend. A gap between
two first fins 11 adjacent to each other serves as air passageway through which air
15 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 12 extend is
sometimes referred to as "row direction of the heat exchanger 100" 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 100" or simply as
20 "step direction". The step direction of the heat exchanger 100 is parallel to, for
example, a longitudinal direction of each of the first fins 11 and a longitudinal direction
of each of the after-mentioned second fins 21.
[0012]
Each of the plurality of first heat transfer pipes 12 extends in the direction
25 perpendicular to the surface of paper of Fig. 1. The plurality of first heat transfer
pipes 12 are arrayed at regular step pitches L2 in one row in the step direction of the
heat exchanger 100. 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 12
adjacent to each other in the step direction. Each of the plurality of first heat transfer
30 pipes 12 is a circular pipe having an outer diameter Do. Further, each of the plurality

8
of first heat transfer pipes 12 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 12 constitute a first row of heat transfer pipes
located furthest windward in the heat exchanger 100.
5 [0013]
The second heat exchange unit 20 includes a plurality of second fins 21
arranged parallel to each other at intervals and a plurality of second heat transfer
pipes 22 each passing through the plurality of second fins 21 and each extending
parallel to each other in a direction intersecting the plurality of second fins 21. As
10 with the first fins 11, each of the plurality of second fins 21 has a rectangular flat-plate
shape. Each of the plurality of second fins 21 is placed parallel to the first fins 11
and perpendicular to a direction in which the second heat transfer pipes 22 extend.
The plurality of second fins 21 are provided parallel to each other at regular
placement pitches in the direction perpendicular to the surface of paper of Fig. 1, that
15 is, the direction in which the first heat transfer pipes 12 extend. Each of the plurality
of second fins 21 is placed with a displacement of, for example, approximately half a
pitch from the corresponding one of the plurality of first fins 11. A gap between two
second fins 21 adjacent to each other serves as an air passageway. In the present
embodiment, each of the first fins 11 and each of the second fins 21 are separate
20 components. Alternatively, the first fin 11 and the second fin 21 may be integrally
formed. That is, the first heat exchange unit 10 and the second heat exchange unit
20 may share a plurality of fins with each other.
[0014]
Each of the plurality of second heat transfer pipes 22 extends in a direction
25 parallel to the direction in which the first heat transfer pipes 12 extend. The plurality
of second heat transfer pipes 22 are arrayed at step pitches L2 in one row in the step
direction of the heat exchanger 100. Each of the step pitches L2 is equal to a step
pitch between first heat transfer pipes 12. Each of the plurality of second heat
transfer pipes 22 is placed with a displacement of, for example, approximately half a
30 pitch from the corresponding one of the plurality of first heat transfer pipes 12. The

9
plurality of second heat transfer pipes 22 constitute a second row of heat transfer
pipes as counted from a windward side in the heat exchanger 100. The plurality of
first heat transfer pipes 12 and the plurality of second heat transfer pipes 22 are
arrayed at row pitches L1 in the row direction of the heat exchanger 100. Each of
5 the row pitches can be specified by a distance in the row direction between the tube
axis 12a of a first heat transfer pipe 12 and a tube axis 22a of a second heat transfer
pipe 22. A row pitch between first heat transfer pipes 12 in the first heat exchange
unit 10 and a row pitch between second heat transfer pipes 22 in the second heat
exchange unit 20 can both be considered as L1. Each of the plurality of second
10 heat transfer pipes 22 is a circular pipe having an outer diameter Do that is equal to
the outer diameter of a first heat transfer pipe 12. Further, each of the plurality of
second heat transfer pipes 22 is a circular pipe having a wall thickness tP that is
equal to the wall thickness of a first heat transfer pipe 12.
[0015]
15 The heat exchanger 100 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 12,
one or more second heat transfer pipes 22, or a combination of one or more first heat
transfer pipes 12 and one or more second heat transfer pipes 22.
20 [0016]
Fig. 2 is a cross-sectional view showing a configuration of some components of
a heat exchanger 100 according to a modification of Embodiment 1. As with Fig. 1,
Fig. 2 shows a configuration of the heat exchanger 100 as sectioned along the plane
perpendicular to the direction in which the first heat transfer pipes 12 extend. As
25 shown in Fig. 2, the heat exchanger 100 of the present modification differs from the
heat exchanger 100 shown in Fig. 1 in that the heat exchanger 100 of the present
modification includes another second heat exchange unit 30 located further leeward
than the second heat exchange unit 20.
[0017]
30 The second heat exchange unit 30 includes a plurality of second fins 31 and a

10
plurality of second heat transfer pipes 32 each passing through the plurality of second
fins 31. As with the first fins 11 and the second fins 21, each of the plurality of
second fins 31 has a rectangular flat-plate shape. Each of the plurality of second
fins 31 is placed parallel to the first fins 11 and the second fins 21 and perpendicular
5 to a direction in which the second heat transfer pipes 32 extend. The plurality of
second fins 31 are provided parallel to each other at regular placement pitches in a
direction perpendicular to a surface of paper of Fig. 2, that is, the direction in which
the first heat transfer pipes 12 extend. A gap between two second fins 31 adjacent
to each other serves as an air passageway. In the present embodiment, each of the
10 first fins 11, each of the second fins 21, and each of the second fins 31 are separate
components. Alternatively, at least two of the first fin 11, the second fin 21, and the
second fin 31 may be integrally formed.
[0018]
Each of the plurality of second heat transfer pipes 32 extends in the direction
15 parallel to the direction in which the first heat transfer pipes 12 extend. The plurality
of second heat transfer pipes 32 are arrayed at step pitches L2 in one row in the step
direction of the heat exchanger 100. Each of the step pitches L2 is equal to a step
pitch between first heat transfer pipes 12 and a step pitch between second heat
transfer pipes 22. The plurality of second heat transfer pipes 32 constitute a third
20 row of heat transfer pipes as counted from the windward side in the heat exchanger
100. The plurality of first heat transfer pipes 12, the plurality of second heat transfer
pipes 22, and the plurality of second heat transfer pipes 32 are arrayed at row pitches
L1 in the row direction of the heat exchanger 100. Each of the plurality of second
heat transfer pipes 32 is a circular pipe having an outer diameter Do that is equal to
25 the outer diameter of a first heat transfer pipe 12 and the outer diameter of a second
heat transfer pipe 22. Further, each of the plurality of second heat transfer pipes 32
is a circular pipe having a wall thickness tP that is equal to the wall thickness of a first
heat transfer pipe 12 and the wall thickness of a second heat transfer pipe 22.
[0019]
30 In the present embodiment, the respective wall thicknesses tP of the first heat

11
transfer pipes 12, the second heat transfer pipes 22, and the second heat transfer
pipes 32 each range, for example, from 0.1 to 0.4 mm. Note, however, that the
respective wall thicknesses of the first heat transfer pipes 12, the second heat
transfer pipes 22, and the second heat transfer pipes 32 may be each less than 0.1
5 mm or may be each greater than 0.4 mm.
[0020]
In a process of manufacturing the heat exchanger 100, the first heat transfer
pipes 12, the second heat transfer pipes 22, and the second heat transfer pipes 32
may be subjected to pipe expanding. In this case, the respective outer diameters Do
10 of the first heat transfer pipes 12, the second heat transfer pipes 22, and the second
heat transfer pipes 32 may of course be specified by outer diameters after pipe
expanding.
[0021]
The following describes heat exchanger performance and cost performance in
15 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 100 are
varied.
[0022]
Table 1 is a table showing effects exerted on the intra-pipe volume V, the extra20 pipe heat transfer coefficient o, the ventilation resistance P, 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 100 according to the present
embodiment are varied. It should be noted, in Table 1, when each of the
25 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.
[0023]

12
[Table 1]
Direction of change V o P Ao M
Do Increase + + + – +
Decrease – – – + –
L1
(Row)
Increase Unchanged – + + +
Decrease Unchanged + – – –
L2
(Step)
Increase – – – + –
Decrease + + + – +
tP Increase – Unchanged Unchanged Unchanged +
Decrease + Unchanged Unchanged Unchanged –
[0024]
The intra-pipe volume V [m3
] is a value obtained by multiplying the cross5 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 o [W/m2
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 P [Pa] is a pressure loss of air
passing through the heat exchanger 100. The extra-pipe heat transfer area Ao [m2
]
10 is the gross area of the respective outer wall surfaces of the heat transfer pipes of the
heat exchanger 100. The heat exchanger weight M [kg] is the weight (core weight)
of a heat exchange core unit of the heat exchanger 100 and the heat exchange core
unit is formed by the heat transfer pipes and the fins.
[0025]
15 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 o 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
20 the extra-pipe heat transfer area Ao by increasing the row pitch L1 or to increase the
extra-pipe heat transfer coefficient o 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

13
exchange performance of the heat exchanger 100 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
5 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 100 to achieve both a reduction in the intra-pipe volume V and an
increase in cost performance of the heat exchanger 100.
10 [0026]
The following describes the extra-pipe heat exchange performance of the heat
exchanger 100 per unit weight.
[0027]
Figs. 3 to 6 each show a relationship between the area ratio of heat transfer
15 pipes to fins and extra-pipe heat exchange performance per unit weight in the heat
exchanger 100 according to Embodiment 1 as a ratio to a maximum value at Do = 5.5
mm for each outer diameter Do (Do = 2.0 mm, 3.0 mm, 4.0 mm, 5.0 mm, 5.5 mm) of
the heat transfer pipes.
[0028]
20 Note here that the heat transfer pipes may include first heat transfer pipes 12,
second heat transfer pipes 22, and second heat transfer pipes 32. The fins may
include first fins 11, second fins 21, and second fins 31. 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 – 2  tP)/2)2 25   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.
[0029]
In each of Figs. 3 to 6, the horizontal axis of the graph represents the area ratio
30 B/A of the area B to the area A. The area ratio B/A represents, as an area ratio, the

14
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. 7 is a
graph showing a relationship between the area ratio B/A and the intra-pipe volume V
in the heat exchanger 100 according to Embodiment 1. Fig. 7 shows effects of the
5 area ratio B/A on the intra-pipe volume V in cases where Outer Diameter Do = 3.0
mm and Do = 5.5 mm and Wall Thickness tP = 0.2 mm. As shown in Fig. 7, the
intra-pipe volume V decreases as the area ratio B/A decreases.
[0030]
In each of Figs. 3 to 6, the vertical axis of the graph represents the extra-pipe
10 heat transfer performance (Extra-pipe Heat Transfer Performance/Weight) of the heat
exchanger 100 per unit weight as a ratio to a maximum value at Do = 5.5 mm. The
extra-pipe heat exchange performance is (Extra-pipe Heat Transfer Area Ao  Extrapipe Heat Transfer Coefficient o)/P. Extra-pipe Heat Transfer Area Ao  Extrapipe Heat Transfer Coefficient o is the extra-pipe heat transfer performance.
15 [0031]
A relationship between each of the extra-pipe heat transfer performance, the
ventilation resistance P, the heat exchanger weight M, and the extra-pipe heat
exchange performance and the area ratio B/A is described here with reference to
Figs. 8 to 12.
20 [0032]
Fig. 8 is a graph showing a relationship between the area ratio B/A and the
extra-pipe heat transfer performance (Ao  o) in the heat exchanger 100 according
to Embodiment 1. Fig. 8 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
25 Coefficient o) in cases where Outer Diameter Do = 3.0 mm and Do = 5.5 mm and
Wall Thickness tP = 0.2 mm. As the area ratio B/A increases, the heat transfer pipes
are located closer to each other and thermal conductivity improves, so that the extrapipe heat transfer performance (Ao  o) increases. Further, a comparison made at
identical area ratios B/A shows that the extra-pipe heat transfer performance (Ao 
30 o) increases as the outer diameter Do of each of the heat transfer pipes decreases.

15
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. 8, a comparison made at identical area ratios B/A shows that the extrapipe heat transfer performance (Ao  o) is higher when Do = 3.0 mm than when Do
5 = 5.5 mm. Further, as an example, in a case in which Area Ratio B/A = 0.06, L1 = L2
= 21.7 mm when Do = 3.0 mm, and L1 = L2 = 39.8 mm when Do = 5.5. That is, the
heat transfer pipes are located closer to each other in a case in which Do = 3.0 mm
than in a case in which Do = 5.5 mm.
[0033]
10 Fig. 9 is a graph showing a relationship between the area ratio B/A and the
ventilation resistance P in the heat exchanger 100 according to Embodiment 1.
Fig. 9 shows effects of the area ratio B/A on the ventilation resistance P in cases
where Outer Diameter Do = 3.0 mm and Do = 5.5 mm and Wall Thickness tP = 0.2
mm. As the area ratio B/A increases, the heat transfer pipes are located closer to
15 each other and resistance to the flow of air passing through the heat exchanger 100
increases, so that the ventilation resistance P 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
20 diameter Do suffer from earlier closure of air trunks through which air circulates and
from a higher rate of increase in the ventilation resistance P than do those heat
transfer pipes with a large outer diameter Do.
[0034]
Fig. 10 is a graph showing a relationship between the area ratio B/A and the
25 heat exchanger weight M in the heat exchanger 100 according to Embodiment 1.
Fig. 10 shows effects of the area ratio B/A on the heat exchanger weight M in cases
where Outer Diameter Do = 3.0 mm and Do = 5.5 mm and Wall Thickness tP = 0.2
mm. The value of the weight (core weight) of the heat exchanger 100 has a positive
correlation with the amount of material of the heat exchanger 100 to be used and the
30 manufacturing cost of the heat exchanger 100. For this reason, the value of Extra-

16
pipe Heat Exchange Performance/Weight represented by the vertical axis of the
graph in each of Figs. 3 to 6 is equivalent to the cost performance of the heat
exchanger 100. As the area ratio B/A decreases, the number of heat transfer pipes
that are mounted in the heat exchanger 100 decreases, so that the heat exchanger
5 weight M decreases.
[0035]
Fig. 11 is a graph showing a relationship between the area ratio B/A and the
extra-pipe heat exchange performance in the heat exchanger 100 according to
Embodiment 1. Fig. 11 shows effects of the area ratio B/A on the extra-pipe heat
10 exchange performance ((Ao  o)/P) in cases where Outer Diameter Do = 3.0 mm
and Do = 5.5 mm and Wall Thickness tP = 0.2 mm. Further, Fig. 12 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 100 according to Embodiment 1.
Fig. 12 shows effects of the area ratio B/A on the extra-pipe heat exchange
15 performance per unit weight ((Ao  o)/P/M) in cases where Outer Diameter Do =
3.0 mm and Do = 5.5 mm and Wall Thickness tP = 0.2 mm. As shown in Fig. 11, the
characteristic of the extra-pipe heat exchange performance to the area ratio B/A has a
maximum value. Further, as shown in Fig. 10, the heat exchanger weight M
monotonically increases when the area ratio B/A increases. For this reason, as
20 shown in Fig. 12, 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
25 transfer pipes decreases, the rate of change in the ventilation resistance P
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. 11, the
maximum value of the extra-pipe heat exchange performance ((Ao  o)/P)
increases as the outer diameter Do of each of the heat transfer pipes decreases.
30 [0036]

17
Continued reference is made to Figs. 3 to 6. Figs. 3 to 6 vary in value of the
wall thickness tP from one another. Fig. 3 is a graph showing a case in which the
wall thickness tP is 0.1 mm. Fig. 4 is a graph showing a case in which the wall
thickness tP is 0.2 mm. Fig. 5 is a graph showing a case in which the wall thickness
5 tP is 0.3 mm. Fig. 6 is a graph showing a case in which the wall thickness tP is 0.4
mm. In general, in a case in which a hydrofluorocarbon is used as refrigerant, the
wall thickness tP from approximately 0.15 to 0.2 mm when Do = 5.5 or smaller is
often used.
[0037]
10 The extra-pipe heat exchange performance of the heat exchanger 100
according to the present embodiment per unit weight as shown in Figs. 3 to 6 is
calculated by the following method.
[0038]
In general, the heat transfer coefficient a [W/m2
K] between air and the fins is
15 defined by the following equations.
[0039]
[Math. 1]
[0040]
20 Note here that Nu is a Nusselt number and Re is a Reynolds number. Pr is a
Prandtl number, a is the thermal conductivity of air, and  is the kinematic viscosity of
air. At ordinary temperatures and pressures, Pr = 0.72, a = 0.0261 [W/mK], and  =
0.000016 [m2
/s]. Further, C1 and C2 are constants, and NL is the number of rows of
the heat transfer pipes.
25 [0041]

18
The characteristic length De [m] is defined by the following equations.
[0042]
[Math. 2]
5 [0043]
Note here that Vc [m3
] 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.
[0044]
The wind velocity U [m/s] based on a free passage volume between fins and
10 the front wind velocity Uf [m/s] of the heat exchanger are defined by the following
equations.
[0045]
[Math. 3]
15 [0046]
Note here that Qair [m3
/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.
[0047]
20 In general, the extra-pipe heat transfer coefficient o is defined by the following

19
equations.
[0048]
[Math. 4]
5 [0049]
Note here that  is fin efficiency and a is an air-side heat transfer coefficient.
Ao [m2
] is the air-side total heat transfer area of the heat exchanger, Ap [m2
] is the airside pipe heat transfer area of the heat exchanger, AF [m2
] is the air-side fin heat
transfer area of the heat exchanger, and Acon [m2
] is the area of contact between the
10 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,
15 are determined. The contact heat transfer coefficient c between the heat transfer
pipes and the fins of the heat exchanger is constant.
[0050]
The fin efficiency  is defined by the following equations.
[0051]
20 [Math. 5]
[0052]
Note here that dF [m] is a fin equivalent diameter and F [W/mK] is the thermal

20
conductivity of the fins.
[0053]
The ventilation resistance P [Pa] is defined by the following equations.
[0054]
5 [Math. 6]
[0055]
Note here that f is a coefficient of friction loss,  is the density of air, and C3 and
C4 are constants.
10 [0056]
It should be noted that the constants C1, C2, C3, and C4, which are used in the
Nusselt number Nu and a coefficient of flow loss f, are set to represent the thermal
conductivity a and ventilation resistance P of the fins of a heat exchanger of a
commercially widely-distributed common air-conditioning apparatus.
15 [0057]
The extra-pipe heat exchange performance of the heat exchanger 100
according to the present embodiment per unit weight as shown in Figs. 3 to 6 is
calculated under the following conditions.
[Calculation Conditions]
20 Dry-bulb temperature of air flowing into heat exchanger 100: 35 degrees
Celsius
Wet-bulb temperature of air flowing into heat exchanger 100: 24 degrees
Celsius
Wind velocity at front of heat exchanger 100 of air flowing into heat exchanger
25 100: 1.2 m/sec
Refrigerant: R32

21
Outer diameter Do of heat transfer pipe: 2.0 mm to 5.5 mm
Wall thickness tP of heat transfer pipe: 0.1 mm to 0.4 mm
Material of heat transfer pipe: copper
Row pitch L1: 11 mm to 22 mm
5 Step pitch L2: 5 mm to 42 mm
Thickness of fin: 0.10 mm
Fin pitch FP: 1.50 mm
Material of fin: aluminum
Shape of fin: flat fin
10 [0058]
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
15 Patent Literature 1 (Japanese Unexamined Patent Application Publication No. 2013-
92306).
Outer diameter Do of heat transfer pipe: 5.5
Row pitch L1: 20.35 mm
Step pitch L2: 20.35 mm
20 Fin pitch FP: 1.50 mm
[0059]
Further, under the calculation conditions of the comparative example, the area
ratio B/A is 0.053 in a case in which Wall Thickness tP = 0.1 mm, is 0.049 in a case in
which Wall Thickness tP = 0.2 mm, is 0.046 in a case in which Wall Thickness tP =
25 0.3 mm, and is 0.042 in a case in which Wall Thickness tP = 0.4 mm.
[0060]
As shown in Figs. 3 to 6, there is a region in which the outer diameter Do of
each pipe is less than 5.5 mm, Extra-pipe Heat Exchange Performance/Weight [Ratio]
exceeds 100%, and the area ratio B/A can fall below that of the comparative example.
30 That is, if the area ratio B/A falls below that of the comparative example, the intra-pipe

22
volume V can be made smaller than that of the comparative example, and the cost
performance of the heat exchanger 100 can be made higher than that of the
comparative example.
[0061]
5 A range of numerical values of the area ratio B/A in which Extra-pipe Heat
Exchange Performance/Weight [Ratio] exceeds 100% 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. 4, in a case where Wall Thickness tP =
0.2 mm and Do = 3.0, Extra-pipe Heat Exchange Performance/Weight [Ratio]
10 exceeds 100% and the area ratio B/A falls below that of the comparative example,
provided 0.013  B/A  0.043. Further, for example, as shown in Fig. 4, in a case
where Wall Thickness tP = 0.2 mm and Do = 4.0, Extra-pipe Heat Exchange
Performance/Weight [Ratio] exceeds 100% and the area ratio B/A falls below that of
the comparative example, provided 0.023  B/A  0.049. Further, for example, as
15 shown in Fig. 5, in a case where Wall Thickness tP = 0.3 mm and Do = 3.0, Extrapipe Heat Exchange Performance/Weight [Ratio] exceeds 100% and the area ratio
B/A falls below that of the comparative example, provided 0.009  B/A  0.033.
[0062]
An upper limit of the range of numerical values of the area ratio B/A in which
20 the outer diameter Do of each pipe is less than 5.5 mm, Extra-pipe Heat Exchange
Performance/Weight [Ratio] exceeds 100%, and the area ratio B/A can fall below that
of the comparative example, shown in Figs. 3 to 6, is expressed by Formula (1) 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
25 Exchange Performance/Weight [Ratio] exceeds 100% and the area ratio B/A can fall
below that of the comparative example, shown in Figs. 3 to 6, is expressed by
Formula (2) below as a function of the outer diameter Do and the wall thickness tP.
[0063]
Formula (1): Upper Limit Function
F (Do, tP) = (0.0219  tP2
– 0.0185  tP + 0.0043)  ln (Do) + (1.6950  tP2 30 +

23
1.8455  tP + 1.5416)
It should be noted that ln is a natural logarithm whose base is e.
[0064]
Formula (2): Lower Limit Function
G (Do, tP) = (0.2076  tP2
– 0.1480  tP + 0.0545)  Do ^ (–0.0021  tP2 5 –
0.0528  tP + 0.0164)
[0065]
Further, the area ratio B/A of the comparative example is expressed by Formula
(3) below as a function of the wall thickness tP.
10 [0066]
Formula (3): Area Ratio Function of Comparative Example
H (tP) = 0.0076  tP2
– 0.0417  tP + 0.0574
[0067]
The upper limit function F (Do, tP) is an approximate expression calculated, for
15 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 100% and the area ratio B/A can fall
below that of the comparative example. Further, the lower limit function G (Do, tP) is
20 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 100%
and the area ratio B/A can fall below that of the comparative example. Further, the
25 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.
[0068]
30 With Formulas (1) to (3) above, a relationship among the outer diameter Do,

24
the area ratio B/A, and the wall thickness tP in which Extra-pipe Heat Exchange
Performance/Weight [Ratio] exceeds 100% and the area ratio B/A can fall below that
of the comparative example is expressed by Formula (4) below.
[0069]
5 Formula (4)
Do < 5.5 mm,
(0.0219  tP2
– 0.0185  tP + 0.0043)  ln (Do) + (1.6950  tP2
+ 1.8455  tP +
1.5416)  B/A  (0.2076  tP2
– 0.1480  tP + 0.0545)  Do ^ (–0.0021  tP2
– 0.0528
 tP + 0.0164), and
B/A < 0.0076  tP2 10 – 0.0417  tP + 0.0574
[0070]
Specific examples of the range of numerical values identified by Formula (4)
above under the aforementioned calculation conditions are described here with
reference to Figs. 13 to 16.
15 [0071]
Figs. 13 to 16 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 100 according to Embodiment 1. In each of Figs. 13 to 16, the vertical
axis of the graph represents the area ratio B/A of the area B to the area A. The
20 horizontal axis of the graph represents the outer diameter Do of each of the heat
transfer pipes.
[0072]
In each of Figs. 13 to 16, 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
25 LIMIT". Further, the area ratio function H (tP) of the comparative example is shown
as "B/A COMPARATIVE EXAMPLE". Figs. 13 to 16 vary in value of the wall
thickness tP from one another. Fig. 13 is a graph showing a case in which the wall
thickness tP is 0.1 mm. Fig. 14 is a graph showing a case in which the wall
thickness tP is 0.2 mm. Fig. 15 is a graph showing a case in which the wall
30 thickness tP is 0.3 mm. Fig. 16 is a graph showing a case in which the wall

25
thickness tP is 0.4 mm.
[0073]
As shown in Figs. 13 to 16, at each wall thickness tP, Extra-pipe Heat
Exchange Performance/Weight [Ratio] exceeds 100% and the area ratio B/A can fall
5 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 < 5.5 mm. That is, the intra-pipe
volume V can be made smaller than that of the comparative example, and the cost
10 performance of the heat exchanger 100 can be made higher than that of the
comparative example.
[0074]
As noted above, configuring the heat exchanger 100 such that when Outer
Diameter Do < 5.5 mm, Lower Limit Function G (Do, tP)  Area Ratio B/A  Upper
15 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 100%. This in turn makes it possible to improve heat exchanger
performance while reducing the inner capacity of each of the heat transfer pipes of
20 the heat exchanger 100. Therefore, the heat exchanger 100 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
25 heat exchanger 100.
[0075]
Further, the foregoing calculation conditions of the heat exchanger 100
according to the present embodiment correspond to cooling rated conditions of an airconditioning apparatus serving as an example of a refrigeration cycle apparatus.
30 This makes it possible to, under the cooling rated conditions of an air-conditioning

26
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
5 apparatus, the heat exchanger 100 according to the present embodiment brings
about effects that are similar to those brought about under the cooling rated
conditions.
[0076]
Embodiment 2.
10 A heat exchanger according to Embodiment 2 is described. Fig. 17 is a crosssectional view showing a configuration of some components of a heat exchanger 100
according to the present embodiment. As with Fig. 1, Fig. 17 shows a configuration
of the heat exchanger 100 as sectioned along the plane perpendicular to the direction
in which the first heat transfer pipes 12 extend. Constituent elements having the
15 same functions and workings as those of Embodiment 1 are given the same
reference signs, and a description of such constituent elements is omitted.
[0077]
In the heat exchanger 100 of the present embodiment, as shown in Fig. 17, the
outer diameter Doa of each of the first heat transfer pipes 12 of the first heat
20 exchange unit 10 located furthest windward is smaller than the outer diameter Dob of
each of the second heat transfer pipes 22 of the second heat exchange unit 20 (Doa
< Dob). A step pitch L2 between first heat transfer pipes 12 is identical to a step
pitch L2 between second heat transfer pipes 22. Further, each of the plurality of first
heat transfer pipes 12 is a circular pipe having a wall thickness tP that is equal to the
25 wall thickness of a second heat transfer pipe 22.
[0078]
In both the first heat exchange unit 10 and the second heat exchange unit 20,
the relation of Formula (4), which is described above in Embodiment 1, is satisfied.
Further, a value of B/A in the first heat exchange unit 10 is smaller than a value of B/A
30 in the second heat exchange unit 20.

27
[0079]
Fig. 18 is a cross-sectional view showing a configuration of some components
of a heat exchanger 100 according to a modification of the present embodiment. In
the heat exchanger 100 of the present modification, as shown in Fig. 18, a step pitch
5 L2a between first heat transfer pipes 12 of the first heat exchange unit 10 located
furthest windward is greater than a step pitch L2b between second heat transfer pipes
22 of the second heat exchange unit 20 (L2a > L2b). The outer diameter Do of each
of the first heat transfer pipes 12 is identical to the outer diameter Do of each of the
second heat transfer pipes 22. Further, each of the plurality of first heat transfer
10 pipes 12 is a circular pipe having a wall thickness tP that is equal to the wall thickness
of a second heat transfer pipe 22.
[0080]
In both the first heat exchange unit 10 and the second heat exchange unit 20,
the relation of Formula (4), which is described above in Embodiment 1, is satisfied.
15 Further, a value of B/A in the first heat exchange unit 10 is smaller than a value of B/A
in the second heat exchange unit 20.
[0081]
As described above, the heat exchanger 100 according to the present
embodiment further includes a plurality of heat exchange units, arrayed along the
20 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 10
located furthest windward and at least one second heat exchange unit 20 located
further leeward than the first heat exchange unit 10. A value of B/A in the first heat
exchange unit 10 is smaller than a value of B/A in the at least one second heat
25 exchange unit 20.
[0082]
In general, in the first heat exchange unit 10 located furthest windward, frost
easily forms, as a great temperature difference between the first fins 11 or the first
heat transfer pipes 12 and air results in an increased amount of heat that is
30 exchanged. The foregoing configuration makes it possible to make the first heat

28
exchange unit 10 lower in heat exchange performance than the second heat
exchange unit 20. This makes it possible to inhibit the formation of frost in the first
heat exchange unit 10 and therefore makes it possible to prevent an air trunk of the
first heat exchange unit 10 from being closed by an increased amount of frost that is
5 formed. This makes it possible to improve cost performance while reducing
deterioration in ventilation performance of the heat exchanger 100.
[0083]
Embodiment 3.
A refrigeration cycle apparatus according to Embodiment 3 is described. Fig.
10 19 is a refrigerant circuit diagram showing a configuration of a refrigeration cycle
apparatus 200 according to Embodiment 3. In the present embodiment, an airconditioning apparatus is an example of the refrigeration cycle apparatus 200. As
shown in Fig. 19, the refrigeration cycle apparatus 200 includes a refrigeration cycle
circuit 50 through which refrigerant circulates. The refrigeration cycle circuit 50 is
15 configured such that a compressor 51, a four-way valve 52, an outdoor heat
exchanger 53, an expansion valve 54, and an indoor heat exchanger 55 are
connected in a circular pattern via refrigerant pipes. Further, the refrigeration cycle
apparatus 200 includes an outdoor fan 56 configured to supply air to the outdoor heat
exchanger 53 and an indoor fan 57 configured to supply air to the indoor heat
20 exchanger 55. In the refrigeration cycle apparatus 200, the compressor 51 is driven
so that a refrigeration cycle is executed in which the refrigerant circulates through the
refrigeration cycle circuit 50 while the refrigerant changes its phase. The outdoor
heat exchanger 53 allows the air supplied by the outdoor fan 56 and the refrigerant,
which is an inner fluid, to exchange heat with each other. The indoor heat
25 exchanger 55 allows the air supplied by the indoor fan 57 and the refrigerant, which is
an inner fluid, to exchange heat with each other. As at least either the outdoor heat
exchanger 53 or the indoor heat exchanger 55, the heat exchanger 100 of
Embodiment 1 or 2 is used.
[0084]
30 The refrigeration cycle apparatus 200 includes an outdoor unit 110 and an

29
indoor unit 120 as heat exchange units. The outdoor unit 110 houses the
compressor 51, the four-way valve 52, the outdoor heat exchanger 53, the expansion
valve 54, and the outdoor fan 56. The indoor unit 120 houses the indoor heat
exchanger 55 and the indoor fan 57. The outdoor unit 110 and the indoor unit 120
5 are connected to each other via a gas pipe 130 and a liquid pipe 140, which are some
of the refrigerant pipes.
[0085]
Operation of the refrigeration cycle apparatus 200 is described by describing
cooling operation as an example. For cooling operation, the four-way valve 52 is
10 switched such that refrigerant discharged from the compressor 51 flows into the
outdoor heat exchanger 53. The high-pressure gas refrigerant discharged from the
compressor 51 flows into the outdoor heat exchanger 53 via the four-way valve 52.
During cooling operation, the outdoor heat exchanger 53 operates as a condenser.
That is, the outdoor heat exchanger 53 allows refrigerant circulating through inside
15 and outdoor air supplied by the outdoor fan 56 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 53 to condense into
high-pressure liquid refrigerant.
[0086]
20 The liquid refrigerant having flowed out of the outdoor heat exchanger 53 is
decompressed by the expansion valve 54 into low-pressure two-phase refrigerant.
The two-phase refrigerant having flowed out of the expansion valve 54 flows into the
indoor heat exchanger 55 via the liquid pipe 140. During cooling operation, the
indoor heat exchanger 55 operates as an evaporator. That is, the indoor heat
25 exchanger 55 allows refrigerant circulating through inside and indoor air supplied by
the indoor fan 57 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 55 to evaporate into low-pressure gas
refrigerant. The indoor air having passed through the indoor heat exchanger 55 is
30 cooled by exchanging heat with the refrigerant. The gas refrigerant having flowed

30
out of the indoor heat exchanger 55 is suctioned into the compressor 51 via the gas
pipe 130 and the four-way valve 52. The gas refrigerant suctioned into the
compressor 51 is compressed into high-pressure gas refrigerant. During cooling
operation, the refrigeration cycle described above is continuously and repeatedly
5 executed. Although not described, for heating operation, a direction of refrigerant
flow is switched by the four-way valve 52 such that the outdoor heat exchanger 53
operates as an evaporator and the indoor heat exchanger 55 operates as a
condenser.
[0087]
10 As described above, the refrigeration cycle apparatus 200 according to the
present embodiment includes the heat exchanger 100 of Embodiment 1 or 2. This
configuration allows the refrigeration cycle apparatus 200 to achieve both a reduction
in total value of GWP and improvement in energy-saving effectiveness.
[0088]
15 Embodiments 1 to 3 and the modifications described above may be combined
with each other.
Reference Signs List
[0089]
10: first heat exchange unit, 11: first fin, 12: first heat transfer pipe, 12a: tube
20 axis, 20: second heat exchange unit, 21: second fin, 22: second heat transfer pipe,
22a: tube axis, 30: second heat exchange unit, 31: second fin, 32: second heat
transfer pipe, 50: refrigeration cycle circuit, 51: compressor, 52: four-way valve, 53:
outdoor heat exchanger, 54: expansion valve, 55: indoor heat exchanger, 56: outdoor
fan, 57: indoor fan, 100: heat exchanger, 110: outdoor unit, 120: indoor unit, 130: gas
25 pipe, 140: liquid pipe, 200: refrigeration cycle apparatus, Do: outer diameter, Doa:
outer diameter, Dob: outer diameter, L1: row pitch, L2: step pitch, L2a: step pitch,
L2b: step pitch, tP: wall thickness

We Claim:
[Claim 1]
A heat exchanger, comprising:
a plurality of fins arranged in parallel to each other; and
5 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 being placed in a plurality of rows in a
row direction that is along a direction of airflow at a row pitch L1,
10 in the plane, the plurality of heat transfer pipes 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 is defined
as Do,
a wall thickness of a portion having a smallest distance between an outer wall
15 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 – 2  tP)/2)2   is
defined as B,
20 a relation of Do < 5.5 mm,
a relation of (0.0219  tP2
– 0.0185  tP + 0.0043)  ln (Do) + (1.6950  tP2
+
1.8455  tP + 1.5416)  B/A  (0.2076  tP2
– 0.1480  tP + 0.0545)  Do ^ (–0.0021
 tP2
– 0.0528  tP + 0.0164), and
a relation of B/A < 0.0076  tP2
– 0.0417  tP + 0.0574
25 being satisfied.
[Claim 2]
The heat exchanger of claim 1, further comprising 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,
30 wherein the plurality of heat exchange units include a first heat exchange unit

32
located furthest windward and at least one second heat exchange unit located further
leeward than the first heat exchange unit, and
a value of B/A in the first heat exchange unit is smaller than a value of B/A in
the at least one second heat exchange unit.
[Claim 3]
A refrigeration cycle apparatus, comprising the heat exchanger of claim 1 or 2.