Title:
Exhaust gas heat exchanger
Kind Code:
A1


Abstract:
An exhaust gas heat exchanger has a tube which is made of a stainless steel and in which exhaust gas flows, and an inner fin which is made of a stainless steel and arranged in the tube to improve a heat exchange between the exhaust gas and cooling water. The cooling water flows at an outer side of the tube. The fin pitch fp of the inner fin is substantially in the range of 2 mm<fp≦12 mm, and the fin height fh of the inner fin is substantially in the range of 3.5 mm<fh≦12 mm.



Inventors:
Oofune, Yuu (Anjo-city, JP)
Hayashi, Takayuki (Nagoya-city, JP)
Application Number:
11/827409
Publication Date:
01/17/2008
Filing Date:
07/11/2007
Assignee:
DENSO Corporation (Kariya-city, JP)
Primary Class:
Other Classes:
165/183
International Classes:
F28F1/36
View Patent Images:
Related US Applications:



Other References:
11/827,409 Equation Sheet, 1 page.
Primary Examiner:
ROSATI, BRANDON MICHAEL
Attorney, Agent or Firm:
HARNESS, DICKEY & PIERCE, P.L.C. (P.O. BOX 828, BLOOMFIELD HILLS, MI, 48303, US)
Claims:
What is claimed is:

1. An exhaust gas heat exchanger in which exhaust gas generated due to combustion is heat-exchanged with cooling fluid, comprising: a tube in which the exhaust gas flows and outside which the cooling fluid flows; and an inner fin which is arranged in the tube to improve a heat exchange between the exhaust gas and the cooling fluid, wherein: the inner fin has a cross section which has a corrugated shape to include convex portions positioned at crests and troughs of the corrugated shape, and is constructed of an offset fin having lanced segments which are partially lanced and arrayed substantially in a flowing direction of the exhaust gas, the crests and the troughs being alternately arranged, and the cross section being substantially perpendicularly to the flowing direction of the exhaust gas; and a fin pitch fp and a fin height fh of the inner fin are substantially defined by following formulas
3.5 mm<fh≦12 mm,
2 mm<fp≦12 mm, wherein the fin pitch fp is a distance between central lines of the adjacent convex portions positioned at a side of one of the crest and the trough in the cross section of the inner fin, and the fin height fh is a distance between the convex portions which are respectively positioned at the side of the crest and the side of the trough in the cross section of the inner fin.

2. An exhaust gas heat exchanger in which exhaust gas generated due to combustion is heat-exchanged with cooling fluid, comprising: a tube in which the exhaust gas flows and outside which the cooling fluid flows; and an inner fin which is arranged in the tube to improve a heat exchange between the exhaust gas and the cooling fluid, wherein: the inner fin has a cross section which has a corrugated shape to include convex portions positioned at crests and troughs of the corrugated shape, and is constructed of an offset fin having lanced segments which are partially lanced and arrayed substantially in a flowing direction of the exhaust gas, the crests and the troughs being alternately arranged, and the cross section being substantially perpendicularly to the flowing direction of the exhaust gas; and an equivalent circle diameter de is substantially defined by following formulas
when 0<L<5 mm, 1.2 mm≦de≦6.1 mm,
when 5 mm≦L≦15 mm, 1.0 mm≦de≦4.3 mm, wherein L is a length of the lanced segment in the flowing direction of the exhaust gas, and the equivalent circle diameter de is a diameter of an equivalent circle of a field C which is surrounded by the inner fin and the tube and positioned between the adjacent convex portions at a side of one of the crest and the trough in the cross section of the inner fin.

3. The exhaust gas heat exchanger according to claim 2, wherein the equivalent circle diameter de is substantially defined by following formulas
when 0<L<5 mm, 1.3 mm≦de≦5.3 mm,
when 5 mm<L<15 mm, 1.1 mm≦de≦4.0 mm.

4. The exhaust gas heat exchanger according to claim 2, wherein the equivalent circle diameter de is substantially defined by following formulas
when 0<L<5 mm, 1.5 mm≦de≦4.5 mm,
when 5 mm≦L≦15 mm, 1.3 mm≦de≦3.5 mm.

5. An exhaust gas heat exchanger in which exhaust gas generated due to combustion is heat-exchanged with cooling fluid, comprising: a tube in which the exhaust gas flows and outside which the cooling fluid flows; and an inner fin which is arranged in the tube to improve a heat exchange between the exhaust gas and the cooling fluid, wherein: the inner fin has a cross section which has a corrugated shape to include convex portions positioned at crests and troughs of the corrugated shape, and is constructed of an offset fin having lanced segments which are partially lanced and arrayed substantially in a flowing direction of the exhaust gas, the crests and the troughs being alternately arranged, and the cross section being substantially perpendicularly to the flowing direction of the exhaust gas; and a length L of the lanced segment is substantially defined by following formulas
when fh<7 mm and fp≦5 mm, 0.5 mm<L≦65 mm,
when fh<7 mm and fp>5 mm, 0.5 mm<L≦20 mm,
when fh≧7 mm and fp≦5 mm, 0.5 mm<L≦50 mm,
when fh≧7 mm and fp>5 mm, 0.5 mm<L≦15 mm, wherein the length L is a dimension in the flowing direction of the exhaust gas, fp is a fin pitch which is a distance between central lines of the adjacent convex portions positioned at a side of one of the crest and the trough in the cross section of the inner fin, and fh is a fin height which is a distance between the convex portions which are respectively positioned at the side of the crest and the side of the trough in the cross section of the inner fin.

6. The exhaust gas heat exchanger according to claim 5, wherein the length L of the lanced segment is substantially defined by following formulas
when fh<7 mm and fp≦5 mm, 0.5 mm<L≦25 mm,
when fh<7 mm and fp>5 mm, 0.5 mm<L≦8 mm,
when fh≧7 mm and fp≧5 mm, 0.5 mm<L≦18 mm,
when fh≧7 mm and fp>5 mm, 0.5 mm<L≦6 mm.

7. The exhaust gas heat exchanger according to claim 5, wherein the length L of the lanced segment is substantially defined by following formulas
when fh<7 mm and fp≦5 mm, 0.5 mm<L≦7 mm,
when fh<7 mm and fp>5 mm, 0.5 mm<L≦1 mm,
when fh≧7 mm and fp≦5 mm, 0.5 mm<L≦4.5 mm,
when fh≧7 mm and fp>5 mm, 0.5 mm<L≦1.5 mm.

8. An exhaust gas heat exchanger in which exhaust gas generated due to combustion is heat-exchanged with cooling fluid, comprising: a tube in which the exhaust gas flows and outside which the cooling fluid flows; and an inner fin which is arranged in the tube to improve a heat exchange between the exhaust gas and the cooling fluid, wherein: the inner fin has a cross section which has a corrugated shape to include convex portions positioned at crests and troughs of the corrugated shape, and is constructed of an offset fin having lanced segments which are partially lanced and arrayed substantially in a flowing direction of the exhaust gas, the crests and the troughs being alternately arranged, and the cross section being substantially perpendicularly to the flowing direction of the exhaust gas; and a fin pitch fp of the inner fin and a length L of the lanced segment are substantially defined by following formulas
2 mm<fp≦12 mm,
1.1 mm≦X≦4.3 mm, wherein X=de×L0.14/fh0.18, wherein the fin pitch fp is a distance between central lines of the adjacent convex portions positioned at a side of one of the crest and the trough in the cross section of the inner fin, the length L is a dimension in the flowing direction of the exhaust gas, fh is a fin height which is a distance between the convex portions respectively positioned at a side of the crest and a side of the trough in the cross section of the inner fin, de is an equivalent circle diameter which is a diameter of an equivalent circle of a field C, and the field D which is defined in the cross section of the inner fin is positioned between the adjacent convex portions of the side of one of the crest and the trough and surrounded by the inner fin and the tube, in which the inner fin is arranged.

9. The exhaust gas heat exchanger according to claim 8, wherein the length L of the lanced segment is substantially defined by a following formula
1.2 mm≦X≦3.9 mm, wherein X=de×L0.14/fh0.18.

10. The exhaust gas heat exchanger according to claim 8, wherein the length L of the lanced segment is substantially defined by a following formula
1.3 mm≦X≦3.5 mm, wherein X=de×L0.14/fh0.08.

11. The exhaust gas heat exchanger according to claim 1, wherein in a cross section of the inner fin, a ratio of an offset area T to a area of a field C is substantially in a range from 25% to 40%, the cross section being substantially perpendicular to the exhaust gas flowing direction, the field D being positioned between the adjacent convex portions of the side of one of the crest and the trough and surrounded by the inner fin and the tube, in which the inner fin is arranged, the offset area T being an area of a part, which is defined in the cross section of the inner fin and surrounded by the two lanced segments which are adjacent to each other in the exhaust gas flowing direction and offset from each other in the longitudinal direction of the inner fin.

12. The exhaust gas heat exchanger according to claim 1, wherein the lanced segments which are adjacent to each other in the flowing direction of the exhaust gas deviate from each other at an offset amount s in a longitudinal direction of the inner fin, and the offset amount s is larger than about 0.5 mm.

13. The exhaust gas heat exchanger according to claim 1, wherein the tube and the inner fin are arranged at a halfway portion of an exhaust gas recirculation passage through which the exhaust gas of an diesel engine having passed a diesel particulate filter is returned to a suction side of the diesel engine.

14. The exhaust gas heat exchanger according to claim 2, wherein the tube and the inner fin are arranged at a halfway portion of an exhaust gas recirculation passage through which the exhaust gas of an diesel engine having passed a diesel particulate filter is returned to a suction side of the diesel engine.

15. The exhaust gas heat exchanger according to claim 5, wherein the tube and the inner fin are arranged at a halfway portion of an exhaust gas recirculation passage through which the exhaust gas of an diesel engine having passed a diesel particulate filter is returned to a suction side of the diesel engine.

16. The exhaust gas heat exchanger according to claim 8, wherein the tube and the inner fin are arranged at a halfway portion of an exhaust gas recirculation passage through which the exhaust gas of an diesel engine having passed a diesel particulate filter is returned to a suction side of the diesel engine.

17. The exhaust gas heat exchanger according to claim 1, wherein: each of the tube and the inner fin is made of a stainless steel; and the cooling fluid is cooling water.

18. The exhaust gas heat exchanger according to claim 2, wherein: each of the tube and the inner fin is made of a stainless steel; and the cooling fluid is cooling water.

19. The exhaust gas heat exchanger according to claim 5, wherein: each of the tube and the inner fin is made of a stainless steel; and the cooling fluid is cooling water.

20. The exhaust gas heat exchanger according to claim 8, wherein: each of the tube and the inner fin is made of a stainless steel; and the cooling fluid is cooling water.

Description:

CROSS REFERENCE TO RELATED APPLICATION

This application is based on a Japanese Patent Application No. 2006-190428 filed on Jul. 11, 2006, the disclosure of which is incorporated herein by reference.

FIELD OF THE INVENTION

The present invention relates to an exhaust gas heat exchanger. For example, the exhaust gas heat exchanger can be suitably used for an exhaust gas recirculation cooler (EGR cooler), which is provided in an exhaust gas recirculation device (EGR) to cool exhaust gas.

BACKGROUND OF THE INVENTION

Generally, an exhaust gas recirculation cooler (EGR cooler) is used in a diesel type engine or the like as an exhaust gas heat exchanger. For example, with reference to JP-2004-77024A, the general EGR cooler is arranged at a halfway position of an exhaust gas recirculation pipe for partially refluxing exhaust gas of the engine directly to the suction side of the engine.

In this case, the EGR cooler is provided with multiple tubes which are stacked and in each of which an inner fin is arranged. Exhaust gas flowing in the tube is heat-exchanged with cooling water flowing at the outer side of the tube, so that exhaust gas is cooled. In this case, the inner fin is constructed of a straight fin.

In addition to the straight fin or a wave fin, the inner fin can be also constructed of an offset fin which is generally used in an inter cooler or the like to have a different use from the EGR cooler, for example, with reference to JP-3766914.

The offset fin is susceptible to being clogged, although having a higher heat-exchanging capacity than the straight fin. Because there is lot of coal in exhaust gas flowing through the EGR cooler so that the offset fin is susceptible to be clogged, it is difficult to use the offset fin as the inner fin of the EGR cooler.

Moreover, because the cooling method, the required performance, the specifications environment and the like of the EGR cooler are different from those of the inter cooler, specifications (such as fin pitch fp, fin height fh, segment length L and the like) of the offset fin used in the inter cooler can not be directly (without being changed) used in the EGR cooler.

For example, the cooling method of the inter cooler is different from that of the EGR cooler. That is, the inter cooler is generally an air cooling type, while the EGR cooler is generally a water cooling type. Thus, the contribution degree of the inner fin to the heat exchanging capacity in the inter cooler is different from that in the EGR cooler.

Moreover, the temperature (e.g., 170° C.) of the cooling object gas of the inter cooler is different from that (e.g., 400° C.) of the EGR cooler.

Moreover, the inter cooler is made of a different material from that of the EGR cooler. The inter cooler is generally made of aluminum. On the other hand, the EGR cooler is to be made of a stainless steel to maintain a corrosion resistance, because the EGR cooler is exposed to a corrosion environment due to high-temperature oxidation and condensation water.

The specifications of the offset fin are set in such a manner that the heat exchanging capacity (related to cooling method, temperature of cooling object gas, material of inner fin and the like) of the EGR cooler has a maximum value. However, in the case where the specifications of the offset fin for the inter cooler is simply used as the specifications of the offset fin for the EGR cooler, the heat exchanging capacity of the EGR cooler will be lowered.

Moreover, in an exhaust gas recirculation device where the EGR cooler is used, in order to maintain the flow amount in the case of the high load, it is necessary for the pressure loss in the EGR cooler to be small. However, for example, in the case where the specifications (fin pitch fp=2 mm) of the offset fin are defined as disclosed in JP-3766914, the pressure loss in the tube will become excessively large.

The above described disadvantages will occur in not only the EGR cooler but also other sort of exhaust gas heat exchanger which is a water-cooling type and made of the stainless steel.

SUMMARY OF THE INVENTION

In view of the above-described disadvantages, it is an object of the present invention to provide an exhaust gas heat exchanger having an improved performance in the case where an offset fin is used as an inner fin.

According to a first aspect of the present invention, an exhaust gas heat exchanger in which exhaust gas generated due to combustion is heat-exchanged with cooling fluid includes a tube in which the exhaust gas flows and outside which the cooling fluid flows, and an inner fin which is arranged in the tube to improve a heat exchange between the exhaust gas and the cooling fluid. The inner fin has a cross section which has a corrugated shape to include convex portions positioned at crests and troughs of the corrugated shape, and is constructed of an offset fin having lanced segments which are partially lanced and arrayed substantially in a flowing direction of the exhaust gas. The crests and the troughs are alternately arranged, and the cross section is substantially perpendicularly to the flowing direction of the exhaust gas. A fin pitch fp and a fin height fh of the inner fin (32) are defined by 3.5 mm<fh≦12 mm, and 2 mm<fp≦12 mm, wherein the fin pitch fp is a distance between central lines of the adjacent convex portions positioned at a side of one of the crest and the trough in the cross section of the inner fin, and the fin height fh is a distance between the convex portions which are respectively positioned at the side of the crest and the side of the trough in the cross section of the inner fin.

Thus, the pressure loss of the exhaust gas flowing in the tube and the hydraulic resistance of the cooling fluid (such as cooling water) can be restricted. Therefore, the tube can be restricted from being clogged, and can be provided with a higher heat-radiating capacity.

According to a second aspect of the present invention, an exhaust gas heat exchanger in which exhaust gas generated due to combustion is heat-exchanged with cooling fluid is provided with a tube in which the exhaust gas flows and outside which the cooling fluid flows, and an inner fin which is arranged in the tube to improve a heat exchange between the exhaust gas and the cooling fluid. The inner fin has a cross section which has a corrugated shape to include convex portions positioned at crests and troughs of the corrugated shape, and is constructed of an offset fin having lanced segments which are partially lanced and arrayed substantially in a flowing direction of the exhaust gas. The crests and the troughs are alternately arranged, and the cross section is substantially perpendicularly to the flowing direction of the exhaust gas. An equivalent circle diameter de is defined by following formulas


when 0<L<5 mm, 1.2 mm≦de≦6.1 mm,


when 5 mm≦L≦15 mm, 1.0 mm≦de≦4.3 mm,

wherein L is a length of the lanced segment in the flowing direction of the exhaust gas, and the equivalent circle diameter de is a diameter of an equivalent circle of a field C which is surrounded by the inner fin and the tube and positioned between the adjacent convex portions at a side of one of the crest and the trough of the corrugated shape in the cross section of the inner fin.

Thus, the gas density which is a factor considering both the cooling capacity and the pressure loss will be larger than or equal to 93%, so that the exhaust gas heat exchanger has an improved performance can be provided.

According to a third aspect of the present invention, an exhaust gas heat exchanger in which exhaust gas generated due to combustion is heat-exchanged with cooling fluid is provided with a tube in which the exhaust gas flows and outside which the cooling fluid flows, and an inner fin which is arranged in the tube to improve a heat exchange between the exhaust gas and the cooling fluid. The inner fin has a cross section which has a corrugated shape to include convex portions positioned at crests and troughs of the corrugated shape, and is constructed of an offset fin having lanced segments which are partially lanced and arrayed substantially in a flowing direction of the exhaust gas. The crests and the troughs are alternately arranged, and the cross section is substantially perpendicularly to the flowing direction of the exhaust gas. A length L of the lanced segment is defined by following formulas


when fh<7 mm and fp≦5 mm, 0.5 mm<L≦65 mm,


when fh<7 mm and fp>5 mm, 0.5 mm<L≦20 mm,


when fh≧7 mm and fp≦5 mm, 0.5 mm<L≦50 mm,

when fh≧7 mm and fp>5 mm, 0.5 mm<L≦15 mm,

wherein the length L is a dimension in the flowing direction of the exhaust gas, fp is a fin pitch which is a distance between central lines of the adjacent convex portions positioned at a side of one of the crest and the trough in the cross section of the inner fin, and fh is a fin height which is a distance between the convex portions which are respectively positioned at the side of the crest and the side of the trough in the cross section of the inner fin.

Therefore, the gas density can be larger than or equal to 97%. Thus, the exhaust gas heat exchanger has the further improved performance can be provided.

According to a fourth aspect of the present invention, an exhaust gas heat exchanger in which exhaust gas generated due to combustion is heat-exchanged with cooling fluid is provided with a tube in which the exhaust gas flows and outside which the cooling fluid flows, and an inner fin which is arranged in the tube to improve a heat exchange between the exhaust gas and the cooling fluid. The inner fin has a cross section which has a corrugated shape to include convex portions positioned at crests and troughs of the corrugated shape, and is constructed of an offset fin having lanced segments which are partially lanced and arrayed substantially in a flowing direction of the exhaust gas. The crests and the troughs are alternately arranged, and the cross section is substantially perpendicularly to the flowing direction of the exhaust gas. A fin pitch fp and a length L of the lanced segment are substantially defined by following formulas


2 mm<X≦12 mm


1.1 mm≦X≦4.3 mm, wherein X=de×L0.14/fh0.18

wherein the length L is a dimension in the flowing direction of the exhaust gas, fh is a fin height which is a distance between the convex portions respectively positioned at a side of the crest and a side of the trough in the cross section of the inner fin, de is an equivalent circle diameter which is a diameter of an equivalent circle of a field C surrounded by the inner fin and the tube and positioned between the adjacent convex portions of the side of one of the crest and the trough in the cross section of the inner fin, and the fin pitch fp is a distance between central lines of the adjacent convex portions positioned at a side of one of the crest and the trough in the cross section of the inner fin.

Thus, the gas density can be larger than or equal to 93%, so that the exhaust gas heat exchanger has an improved performance can be provided.

BRIEF DESCRIPTION OF THE DRAWINGS

Other objects, features and advantages of the present invention will become more apparent from the following detailed description made with reference to the accompanying drawings, in which:

FIG. 1 is a schematic view showing an exhaust gas recirculation device where an exhaust gas heat exchanger is used according to a first embodiment of the present disclosure;

FIG. 2 is a schematic side view showing an EGR cooler as the exhaust gas heat exchanger according to the first embodiment;

FIG. 3 is a schematic sectional view taken along the line III-III in FIG. 2;

FIG. 4 is a schematic sectional view taken along the line IV-IV in FIG. 3;

FIG. 5 is a schematic perspective view showing the EGR cooler according to the first embodiment;

FIG. 6 is a schematic sectional view of an inner fin of the EGR cooler which is taken along a direction substantially perpendicular to an exhaust gas flowing direction according to the first embodiment;

FIG. 7 is a graph showing a relation between a fin height of an offset fin and a pressure loss ratio according to the first embodiment;

FIG. 8 is a graph showing a relation between the fin height and a hydraulic resistance according to the first embodiment;

FIG. 9 is a schematic sectional view showing an inner fin of an EGR cooler which is taken along a direction substantially perpendicular to an exhaust gas flowing direction according to a second embodiment of the present disclosure;

FIG. 10 is a graph showing a relation between an equivalent circle diameter of an offset fin and an EGR gas density ratio according to the second embodiment;

FIG. 11 is a graph showing a relation between a segment length of an offset fin and an EGR gas density ratio according to a third embodiment of the present disclosure;

FIG. 12 is a graph showing a relation between an EGR gas density ratio and a function using an equivalent circle diameter, a segment length and a fin height according to a fourth embodiment of the present disclosure;

FIG. 13A is a graph showing a variation of a PM advantage sedimentation thickness at the offset fin with respect to time, and FIG. 13B is a schematic view showing a sedimentation of PM at the offset fin; and

FIG. 14 is a graph showing a relation between a heat radiating performance of the EGR cooler and a fin pitch of the offset fin.

DETAILED DESCRIPTION OF THE EXAMPLED EMBODIMENTS

First Embodiment

An exhaust gas heat exchanger according to a first embodiment of the present invention will be described with reference to FIGS. 1-8. The exhaust gas heat exchanger can be suitably used as an exhaust gas recirculation cooler 10 (EGR cooler), for example.

As shown in FIG. 1, the EGR cooler 10 can be provided for an exhaust gas recirculation device. The exhaust gas recirculation device has, for example, an air cleaner 3, a variable tube actuator 4, an inter cooler 5 and an intake manifold 6 which are arranged at a halfway portion of an air suction passage 2 of an engine 1.

The tube actuator 4 and a DPF 8 (diesel particulate filter) are arranged at a half portion of an exhaust passage 7 of the engine 1. A first exhaust gas recirculation pipe 9 is connected with a downstream side of exhaust gas of the DPF 8 and an upstream side of suction air of the tube actuator 4. The EGR cooler 10 and an exhaust gas recirculation valve 11 (EGR valve) are arranged at a halfway portion of the first exhaust gas recirculation pipe 9, which is a pipe for refluxing a part of exhaust gas having passed the DPF 8 to the suction side of the engine.

The exhaust gas recirculation device further has a second exhaust gas recirculation pipe 12 and an exhaust gas recirculation valve 13 (EGR valve) which is arranged at a halfway portion of the second exhaust gas recirculation pipe 12. A part of exhaust gas of the engine is refluxed through the second exhaust gas recirculation pipe 12 directly to the suction side of the engine, immediately before passing the DPF 8. The pressure of exhaust gas flowing through the first exhaust gas recirculation pipe 9 can be lower than that of exhaust gas flowing through the second exhaust gas recirculation pipe 12. In this case, the exhaust gas recirculation can be operated even when the engine 1 has a high load.

In this case, when exhaust gas generated due to combustion in the engine 1 is recycled to the engine 1, the EGR cooler 10 cools exhaust gas by coolant of the engine 1, which is cooling liquid (for example, cooling water) in this embodiment. As shown in FIGS. 2-4, the EGR cooler 10 has multiple tubes 21, multiple inner fins 22, water side tanks 23 and gas side tanks 24, which can be made of a stainless steel and integrated with each other by brazing, welding or the like.

As shown in FIGS. 3 and 4, the tube 21 defines therein an exhaust passage 21a in which exhaust gas flows. Cooling water flows at the outer side of the tube 21, and exhaust gas is heat-exchanged with cooling water through the tube 21.

Specifically, as shown in FIG. 3, the tube 21 having a long side 21c and a short side 21d is provided with a flat-shaped cross section when being viewed from the exhaust gas flowing direction. The multiple tubes 21 are stacked in a stacking direction (for example, up-down-direction in FIG. 3) which is perpendicular to the longitudinal direction (i.e., extension direction of long side 21c) of the tube 21. Moreover, as shown in FIGS. 3 and 4, the outer wall surfaces of the tubes 21 which are adjacent to each other defines therebetween a cooling water passage 21b through which cooling water flows between the adjacent tubes 21.

Cooling water having flowed into the EGR cooler 10 is distributed and supplied for the tubes 21 by the one water side tank 23. Cooling water having flowed through the cooling water passage 21b between the tubes 21 are collected and retrieved by the other water side tank 23. The water side tanks 23 are arranged around the tubes 21 which are stacked, in the vicinity of the two ends (of exhaust gas flowing direction) of the tube 21. Each of the water side tanks 23 is provided with a cooling water port 23a (as cooling water outlet or inlet).

The gas side tanks 24 are respectively arranged at the two ends (of exhaust gas flowing direction) of the tube 21. The gas side tanks 24 are connected with the first exhaust gas recirculation pipe 9. Exhaust gas is distributed and supplied for the tubes 21, by the one gas side tank 24. The exhaust gas having been heat-exchanged is colleted and retrieved from the tubes 21, by the other gas side tank 24.

The inner fins 22 are respectively arranged in the tubes 21, to improve the heat exchange between exhaust gas and cooling water. The inner fin 22 can be fixed to the inner wall surface of the tube 21.

With reference to FIGS. 5 and 6, the inner fin 22, being constructed of the offset fin, has a cross section (taken along a direction which is substantially perpendicular to exhaust gas flowing direction), which has a corrugated shape extending in the longitudinal direction of the tube 21. That is, this cross section of the inner fin 22 has convex portions 31 which are respectively arranged at crest positions and trough positions of the corrugated shape which are alternately arranged. The convex portion 31 of the inner fin 22 is arranged to contact the inner wall surface of the tube 21.

The inner fin 22 (offset fin) is partially lanced (cut and raised) to have multiple lanced segments 32. The lanced segments 32 are arrayed in the exhaust gas flowing direction, in such a manner that the adjacent lanced segments 32 offset from each other in the longitudinal direction of the tube 21 (i.e., longitudinal direction of inner fin 22). In this case, the inner fin 22 can be provided with multiple rows (substantially in exhaust gas flowing direction) of the lanced segments 32.

As shown in FIG. 3, by providing the inner fin 22 in the tube 21, the interior of the tube 21 is divided into multiple passages which are substantially parallel to each other with respect to the longitudinal direction (extension direction of long side 21a) of the tube 21.

That is, as shown in FIG. 5, the wall portions 33 of the lanced segments 32 which define therein the passage are arranged staggeringly in the longitudinal direction of the inner fin 22. In this case, as shown in FIG. 6, it is desirable for an offset amount s to be substantially equal to a half of the passage height u, so that the heat transfer coefficient can become high and the gas resistance can become small. The offset amount s and the passage height u are dimensions in the longitudinal direction of the longitudinal direction of fin 22. In this case, the lanced segments 32 which are adjacent to each other in the flowing direction of the exhaust gas deviate from each other at the offset amount s in the longitudinal direction (which is substantially perpendicular to flowing direction of exhaust gas) of the fin 22.

The inner fin 22 can be shaped in such a manner that the convex portion 31 includes a linear portion or does not include a linear portion in the cross section (taken along a direction substantially perpendicular to exhaust gas flowing direction) of the inner fin 22.

In this case, with reference to FIG. 6 which is a cross section (substantially perpendicular to exhaust gas flowing direction) of the inner fin 22, it is desirable for the ratio of an offset area T to the area of a field C (which is dotted to be indicated) in this cross section of the inner fin 22 is substantially in a range from 25% to 40%, considering that the pressure loss will increase when the ratio of the offset area to the area of the field C is smaller than 25%.

The dotted field C in this cross section of the inner fin 22 is positioned between the convex portions 31 which are arranged at the crest positions (or trough positions) and adjacent to each other in the longitudinal direction of inner fin 22, and surrounded by the inner fin 22 and the tube 21. That is, the dotted field C is positioned between the wall portions 33 (facing each other) of the two lanced segments 32 which are adjacent to each other in the longitudinal direction of the inner fin 22, and surrounded by the inner fin 22 and the tube 21. The offset area T is an area of a part, which is defined in this cross section and surrounded by the wall portions 33 of the two lanced segments 32 which are adjacent to each other in the exhaust gas flowing direction and offset from each other in the longitudinal direction of inner fin 22.

The inner fin 22 can be manufactured by a flat plate which is bent to have a corrugated shape by pressing and further lanced by pressing to form the segment 32.

The lancing of the segment 32 can be performed in such a manner that slits are beforehand formed before the corrugated shape is provided and thereafter the raising is performed. Thus, the inner fin 22 has the cross section with the corrugated shape is formed. Alternatively, the lancing of segment 32 can be also performed in such a manner that the two surfaces of the flat plate are pressed by a press machine so that the cutting and raising are simultaneously performed. Moreover, the inner fin 22 can be also manufactured by rolling, or by a combination of rolling and pressing.

The performance of the EGR cooler 10 is related to the specifications of the inner fin 22 such as a fin pitch fp, a fin height fh and the like. The fin pitch fp is a distance between central lines of the two convex portions 31 (which adjacent to each other) of one of a crest side and a trough side, in the corrugated cross section (taken along substantially perpendicular to exhaust gas flowing direction) of the inner fin 22. The fin height fh is a distance between the tops of the two convex portions 31 which are respectively positioned at the crest side and the trough side in this corrugated cross section.

The optimum specifications of the inner fin 22 are investigated in this embodiment. In this case, experiments are performed for the EGR coolers 10 which are respectively provided with the various fin pitches fp and fin heights fh, to evaluate the pressure loss of the exhaust gas flowing in the tube 21, the hydraulic resistance of cooling water flowing at the outer side of the tube 21, the clogged degree of the tube 21, and the heat radiating performance of the each EGR cooler 10 when exhaust gas and cooling water flow under a predetermined condition. Thus, the optimum specifications of the inner fin 22 can be determined. The predetermined condition is set in such a manner that the temperature Tg1 at the exhaust gas inlet is equal to 400° C., the exhaust gas flow amount is equal to 30 g/s, the exhaust gas inlet pressure Pg1 is equal to 50 kPa, the temperature Tw1 at the cooling water inlet is equal to 80° C. and the flow amount of cooling water is equal to 10 L/min.

FIG. 7 shows the relation between the fin pitch height fh and a pressure loss ratio (ΔPg ratio). The pressure loss is a difference between the exhaust gas pressure Pg1 at the exhaust gas inlet of the water side tank 14 and the exhaust gas pressure Pg2 at the exhaust gas outlet of the water side tank 14. The pressure loss ratio (ΔPg ratio) is a ratio (percentage) when the maximum value of the pressure loss at the various conditions is set as 100.

In this case, the offset fin 22 is provided with the plate thickness of about 0.2 mm, the fin pitch fp of about 5 mm or 7 mm, the length L (which is dimension in exhaust gas flowing direction and named segment length L later) of the lanced segment 32 of about 1 mm or 5 mm, and the curvature radius R (of convex portion 31) of about 0.2 mm.

The curves A-C shown in FIG. 7, indicating the relation between ΔPg and fh, are obtained in such a manner that the build of the EGR cooler 10 has a fixed value (that is, size of water side tank 23 and that of gas side tank 24 are fixed) and the fin pitch fp and the segment length L are provided with different values.

The curve A is obtained in such a manner that the fin pitch fp is equal to about 5 mm and the segment length L is equal to about 1 mm. The curve B is obtained in such a manner that the fin pitch fp is equal to about 5 mm and the segment length L is equal to about 5 mm. The curve B is obtained in such a manner that the fin pitch fp is equal to about 7 mm and the segment length L is equal to about 5 mm.

With reference to the curve A shown in FIG. 7, the ascent variation ratio of the pressure loss when the fin height fh is smaller than or equal to 3.5 mm is larger that when the fin height fh is larger than 3.5 mm. There are inflection points at the curves A-C when the fin height fh is equal to about 3.5 mm. That is, the ascent variation ratio of the pressure loss has different values at the two sides of the in height fh of 3.5 mm.

Thus, in the case where the built of the cooler has the fixed value and the fin pitch fp and the segment length L are substantially equal to each other, the pressure loss when fh is smaller than or equal to 3.5 mm is relatively large and the pressure loss when fh is larger than 3.5 mm is relatively small. Therefore, it is desirable for the fin height fh to be larger than 3.5 mm.

FIG. 8 shows the relation between the fin height fh and a hydraulic resistance ΔPw which is a difference between a water pressure at the cooling water inlet 23a of the water side tank 23 and the cooling water outlet 23a thereof. The relation shown in FIG. 8 is obtained with the inner fin 22 being provided with a same condition as that of FIG. 7.

As shown in FIG. 8, when the fin height fh becomes large and the build of the EGR cooler 10 has a fixed value, the hydraulic resistance ΔPw tends to increase. Thus, a water pump having a high performance becomes necessary to maintain the flowing amount of cooling water (in order to maintain cooling performance) when the hydraulic resistance ΔPw becomes lager than or equal to 3 kPa. For example, the hydraulic resistance ΔPw is substantially equal to 3.2 kPa in the case where the fin height fh is set as 12 mm. Thus, the cost will become high. Therefore, it is desirable for the fin height fh to be smaller than or equal to 10 mm.

Furthermore, when the fin pitch fp becomes small, the offset amount s will become small. In the case where the fin plate thickness t is smaller than or equal to about 0.2 mm, the offset amount s will become excessively small when the fin pitch fp is smaller than or equal to about 2 mm. Thus, the inner fin 22 will be susceptible to being clogged by coal in exhaust gas. Therefore, it is desirable for the fin pitch fp to be larger than 2 mm.

The offset amount s can be set to be larger than 0.5 mm, considering that the advantage sedimentation thickness of the PM (particulate matter) at the surface of the single lanced segment 32 is about 0.25 mm when about 8 hours has elapsed, as shown in FIGS. 13A and 13B. Thus, the clogging can be restricted.

Moreover, the heat-radiating capacity of the inner fin 22 can be heightened, by shortening the segment length L. In this case, the relation between the fin pitch fp and the heat-radiating capacity of the inner fin 22 in the case where the segment length L is provided with a minimum value is investigated. As a result, when the fin pitch fp is larger than about 16 mm, it is difficult for the EGR cooler 10 to be provided with the necessary heat-radiating capacity. Accordingly, it is desirable for the fin pitch fp to be smaller than or equal to about 16 mm. Moreover, it is desirable that the fin pitch fp is smaller than or equal to 12 mm, which is an approximate maximum fin pitch for satisfying the performance required by the exhaust gas regulation, as shown in FIG. 14. In FIG. 14, Q represents the heat radiating amount of the EGR cooler 10, and V represents the capacity of the core (which contributes to heat-exchanging and includes exhaust gas passage and cooling water passage) of the EGR cooler 10. In this case, the relation between Q/V and fp (fin pitch) is determined with respectively setting the fin height fh as 12 mm (fh12) and 3.6 mm (fh3.6) and setting the segment length L as 1 mm (L1) and 10 mm (L10).

According to the above-described investigations, it is desirable for the fin pitch fp and the fin height fh to be in the range defined by the following formula (I).


3.5 mm<fh≦12 mm


2 mm<fp≦12 mm (1)

Thus, the pressure loss of exhaust gas flowing in the tube 21 and the hydraulic resistance ΔPw of cooling water flowing at the outer side of the tube 21 can be restricted, so that the tube 21 can be restricted from being clogged and the heat radiating capacity can be improved.

Second Embodiment

According to a second embodiment of the present invention, the optimum specifications of the inner fin 22 are determined according to different criterions and parameters from those of the above-described first embodiment.

In the second embodiment, the optimum specifications of the inner fin 22 are determined based on the relation between an equivalent circle diameter de and an EGR gas density ratio ρ.

In this case, as shown in FIG. 6, the equivalent circle diameter de means a diameter of an equivalent circle into which the field C in the cross section (substantially perpendicular to exhaust gas flowing direction) of the inner fin 22 is converted. The field C is positioned between the convex portions 31 which are arranged at the crest positions (or trough positions) and adjacent to each other, and surrounded by the inner fin 22 and the tube 21. The equivalent circle diameter de can be calculated by the following formula (2).


de=4×S/W (2)

S represents an area (which corresponds to the cross section area of the circle and is calculated by ΠD2/4 wherein the circle diameter is represented by D) of the cross section of the exhaust gas passage. W represents a length of a wetted perimeter corresponding to a circumference calculated by ΠD wherein the circle diameter is represented by D. The length W is a length (that is, length of the part where the inner wall surface contacts exhaust gas) of the inner wall surface of the single gas passage defined by the inner fin 22 and the tube 21.

Next, the calculation of the equivalent circle diameter de will be described. FIG. 9 is a schematic sectional view of the inner fin 22 which is taken long the direction perpendicular to the exhaust gas flowing direction.

As shown in FIG. 9, the half of the wetted perimeter length W/2 (corresponding to right half of the dotted field C shown in FIG. 6, for example) is indicated by five parts w1-w5. The half of the wetted perimeter length W/2, which is a sum of w1-w5 (that is, W/2=w1+w2+w3+w4+w5), can be calculated based on the fin pitch fp, the fin height fh, the plate thickness t and the curvature radius R of the bent portion of the inner fin 22 according to the following formulas (3)-(7) when the linear length of the part w3 is larger than or equal to zero.


w1=fp/2−(fp/2−(2R+t))/2 (3)


w2=√(R+t)/2 (4)


w3=fh−2(R+t) (5)


w4=ΠR/2 (6)


w5=(fp/2−(2R+t))/2 (7)

The half of the cross section area S/2 of the gas passage (corresponding to right half of the dotted field C shown in FIG. 6, for example) is indicated by four parts a-d. The half of the cross section area S/2, which is a sum of a-d (that is, S/2=a+b+c+d), can be calculated based on the fin pitch fp, the fin height fh, the plate thickness t and the curvature radius R of the bent portion according to the following formulas (8)-(11).


a=(fh−t)(fp/2−(2R+t))/2 (8)


b=(fh−(R+t))R (9)


c=ΠR2/4 (10)


d=(R+t)2−Π(R+t)2/4 (11)

Therefore, the equivalent circle diameter de can be determined according to the fin pitch fp, the fin height fh, the plate thickness t and the curvature radius R of the bent portion.

On the other hand, the EGR gas density ρ (having a unit of kg/m3, for example) is a factor considering both the cooling capacity of the EGR cooler 10 and the pressure loss, and can be calculated according to the following formula (12). The filling factor of the EGR gas will become high when the EGR gas density ρ becomes large. Thus, the EGR rate can be increased.


ρ=Pg2/(R·Tg2) (12)

Pg2 represents an absolute pressure (Pa) of the gas outlet. R represents a gas constant 287.05 J/kg·K. Tg2 represents a temperature (K) of the gas outlet.

FIG. 10 shows a relation between the equivalent circle diameter de and the EGR gas density ratio (ρ ratio), which is a ratio when the maximum value of the EGR gas density ρ is set as 100%. The relations shown in FIG. 10 are obtained with the gas inlet temperature Tg1 of about 400° C., the gas flowing amount of about 30 g/s, the gas inlet pressure Pg1 of about 50 kPa, the cooing water inlet temperature Tw1 of about 80° C., the cooling water flowing amount of about 10 L/min, the fin plate thickness t of about 0.2 mm, the fin height fh of about 9 mm and the curvature radius of about 0.2 mm.

The curve D shown in FIG. 10 is measured when the segment length L is equal to about 1 mm, and the curve E shown in FIG. 10 is measured when the segment length L is equal to about 5 mm. When the segment length L is in the range of about 0<L<5, the relation between the equivalent circle diameter de and the EGR gas density ratio can be indicated by a curve similar to the curve D. When the segment length L is in the range of about 5≦L≦15, the relation can be indicated by a curve similar to the curve E.

With reference to the curve D in FIG. 10, in the case of about 0<L<5, the ρ ratio can become larger than or equal to about 93% by setting the equivalent circle diameter de in the range of about 1.2≦de≦6.1, the ρ ratio can become larger than or equal to about 95% by setting the equivalent circle diameter de in the range of about 1.3≦de≦5.3, and the ρ ratio can become larger than or equal to about 97% by setting the equivalent circle diameter de in the range of about 1.5≦de≦4.5.

With reference to the curve E in FIG. 10, in the case of about 5≦L≦15, the ρ ratio can become larger than or equal to about 93% by setting the equivalent circle diameter de in the range of about 1.0≦de≦4.3, the ρ ratio can become larger than or equal to about 95% by setting the equivalent circle diameter de in the range of about 1.1≦de≦4.0, and the ρ ratio can become larger than or equal to about 97% by setting the equivalent circle diameter de in the range of about 1.3≦de≦3.5.

In this case, the segment length L and the equivalent circle diameter de and the like are provided with the unit of mm.

The relation shown in FIG. 10 is measured when the plate thickness t and the curvature radius R of the fin are equal to 0.2 mm. This relation can be indicated by curves similar to the curves D and E, even when the plate thickness t and the curvature radius R are changed in the range which can be embodied. For example, this relation can be indicated by curves similar to the curves D and E′ when the plate thickness t and the curvature radius R are respectively changed in the range from 0.1 mm to 0.2 mm.

About the construction of the EGR cooler 10, what has not described in the second embodiment is the same as the first embodiment.

Third Embodiment

According to a third embodiment of the present invention, the optimum specifications of the inner fin 22 are determined according to different criterions and parameters from those of the above-described embodiments.

In the third embodiment, the optimum specifications of the inner fin 22 are determined based on the relation between the segment length L and the EGR gas density ratio (ρ ratio).

FIG. 11 shows the relation between the segment length L and the EGR gas density ratio (ρ ratio), which is a ratio when the maximum value of the EGR gas density ρ is set as 100%. The relation shown in FIG. 11 is obtained with the same condition as that of FIG. 10, excepting the fin height fh and the segment length L.

The curve F in FIG. 11 is calculated when fh<7 and fp≦5, for example, when fh is equal to 4.6 and fp is equal to 4.5. Thus, when the segment length L is in the range of 0.5<L≦65, the EGR gas density ratio (ρ ratio) can be larger than or equal to about 95%. When the segment length L is in the range of 0.5<L≦25, the ρ ratio can be larger than or equal to about 97%. When the segment length L is set in the range of 0.5<L≦7, the ρ ratio can be larger than or equal to about 99%.

The curve G in FIG. 11 is calculated when fh<7 and fp>5, for example, when fh is equal to about 4.6 and fp is equal to about 5.5. Thus, when the segment length L is in the range of 0.5<L≦20, the EGR gas density ratio (ρ ratio) can be larger than or equal to about 95%. When the segment length L is in the range of 0.5<L≦8, the ρ ratio can be larger than or equal to about 97%. When the segment length L is in the range of 0.5<L≦1, the ρ ratio can be larger than or equal to about 99%.

The curve H in FIG. 11 is calculated when fh≧7 and fp≦5, for example, when fh is equal to about 9 and fp is equal to about 4.5. Thus, when the segment length L is in the range of 0.5<L≦50, the EGR gas density ratio (ρ ratio) can be larger than or equal to about 95%. When the segment length L is in the range of 0.5<L≦15, the ρ ratio can be larger than or equal to about 97%. When the segment length L is set in the range of 0.5<L≦4.5, the ρ ratio can be larger than or equal to about 99%.

The curve I in FIG. 11 is calculated when fh≧7 and fp>5, for example, when fh is equal to about 9 and fp is equal to about 5.5. Thus, when the segment length L is in the range of 0.5<L≦15, the EGR gas density ratio (ρ ratio) can be larger than or equal to about 95%. When the segment length L is in the range of 0.5<L≦6, the ρ ratio can be larger than or equal to about 97%. When the segment length L is in the range of 0.5<L≦1.5, the ρ ratio can be larger than or equal to about 99%.

In this case, the fin pitch fp, the fin height fh, the segment length L and the like are provided with the unit of mm. The relation shown in FIG. 11 is obtained when the plate thickness t and the curvature radius R of the inner fin 22 are equal to about 0.2 mm. This relation can be indicated by curves similar to the curves F-I, even when the plate thickness t and the curvature radius R are changed in the range which can be embodied. For example, this relation can be indicated by curves similar to the curves F-I when the plate thickness t and the curvature radius R are respectively changed in the range from 0.1 mm to 0.2 mm.

About the construction of the EGR cooler 10, what has not described in the third embodiment is the same as the first embodiment.

Fourth Embodiment

According to a fourth embodiment of the present invention, the optimum specifications of the inner fin 22 are determined according to different criterions and parameters from those of the above-described embodiments.

In the fourth embodiment, the optimum specifications of the inner fin 22 are determined based on the relation between the EGR gas density ratio (ρ ratio) and a function X using the equivalent circle diameter de, the segment length L and the fin height fh.

FIG. 12 shows the relation between the EGR gas density ratio (ρ ratio) and the function X which can be indicated by the following formula (13).


X=de×L0.14/fh0.18 (13)

Moreover, FIG. 12 shows the calculation result a of the EGR gas density ratio (ρ ratio) in the case where the fin pitch fp, the fin height fh and the segment length L are respectively provided with various values.

The curves in FIG. 10 are obtained in the case where the fin pitch fp has an arbitrary value while the segment length L and the fin height fh are provided with a fixed value.

Specifically, the fin pitch fp is provided with a value in the substantial range from 1.5 mm to 14 mm, while the fin height fh is substantially equal to one of 3.6 mm, 4.6 mm, 5.6 mm, 7 mm, 9 mm and 12 mm and the segment length L is substantially equal to one of 1 mm and 10 mm. Other measurement conditions of FIG. 12 are same as those of FIGS. 10 and 11.

As shown in FIG. 12, the curves indicating the relation between the EGR gas density ratio (ρ ratio) and the function X show a similar tendency under different conditions. Thus, when the segment length L and the equivalent circle diameter de are set so that the function X has a value in the substantial range of 1.1≦X≦4.3, the EGR gas density ratio (ρ ratio) can be larger than or equal to about 93%. When the segment length L and the equivalent circle diameter de are set so that the function X has a value in the substantial range of 1.2≦X≦3.9, the ρ ratio can be larger than or equal to about 95%.

The segment length L and the equivalent circle diameter de can be set so that the function X has a value in the substantial range of 1.3≦X≦3.5. Thus, the ρ ratio can be larger than or equal to about 97%. Furthermore, the size of the core of the exhaust gas heat exchanger can be reduced.

In this case, the function X and the like is provided with the unit of mm. The relations shown in FIG. 12 are obtained when the plate thickness t and the curvature radius R of the fin are equal to about 0.2 mm. This relation can be indicated similarly to what is shown in FIG. 12, even when the plate thickness t and the curvature radius R are changed in the range which can be embodied. For example, this relation can be indicated similarly when the plate thickness t and the curvature radius R are respectively changed in the range from 0.1 mm to 0.2 mm.

About the construction of the EGR cooler 10, what has not described in the fourth embodiment is the same as the first embodiment.

Other Embodiment

Although the present invention has been fully described in connection with the preferred embodiments thereof with reference to the accompanying drawings, it is to be noted that various changes and modifications will become apparent to those skilled in the art.

The exhaust gas heat exchanger according to the present invention can be also suitably used as an EGR cooler which is arranged at a halfway portion of the second exhaust gas recirculation pipe 12 through which a part of the exhaust gas of the engine 1 is returned directly to the suction side of the engine 1 before flowing through the DPF 8.

Moreover, the present invention can be also suitably used for the other exhaust gas heat exchanger made of a stainless steel or the like, other than the EGR cooler. The present invention can be suitably used for the exhaust gas heat exchanger through which cooling water is heat-exchanged with exhaust gas discharged to the ambient air to be heated.

Such changes and modifications are to be understood as being in the scope of the present invention as defined by the appended claims.





 
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