GB2049901A

GB2049901A - Heat Pump Apparatus and Method of Recovering Heat Utilizing the Same

Info

Publication number: GB2049901A
Application number: GB8011166A
Authority: GB
Original assignee: Valmet Oy
Current assignee: Valmet Oy
Priority date: 1979-04-02
Filing date: 1980-04-02
Publication date: 1980-12-31
Also published as: FR2453373A1; JPS55134254A; NO800960L; SE8002494L; SU925256A3; GB2049901B; FI791079A; DE3012670A1

Abstract

A heat pump apparatus comprises a plurality of separate heat pump circuits (P) wherein the condensers (L) of the heat pump circuits are connected in series with respect to the mass current being heated (m) and in a manner such that the temperature of the mass current being heated increases while being in heat exchange relationship with the fluid circulating in the condensers of the separate heat pump circuits. Similarly, a mass current (m) may be cooled on passing in heat exchange relationship with the evaporators of the heat pump circuits. The mass currents may be air or liquid. The apparatus may be used for drying timber (Fig. 6). <IMAGE>

Description

SPECIFICATION Heat Pump Apparatus and Method of Recovering Heat Utilizing the Same This invention relates to a heat pump apparatus and to a method of recovering heat utilizing the same, more particularly to a heat pump apparatus and its method of operation in which the heating mass or at least one mass flow is connected to the evaporators of the heat pump system and wherein the mass flow to be heated is in heat exchange relationship with the condensors thereof.

Heat pump apparatuses are well known and are generally defined as apparatus which cause a flow of heat or thermal energy from a lower-temperature body to a higher-temperature body.

The concept of utilizing heat pump apparatus for heating enclosed environments, such as residences, as well as for exploiting waste heat has been known for many years. In general, heat pumps and their various applications are set forth in the publication: "The Possibilities and Effects of the Use of a Heat Pump in Heating and Heat Storing", Aittomaki, Kalema, Lappalainen, Talsio, Wiksten, VTT (National Technical Research Centre); The HV 8 C-Laboratory, Item 23, Otaniemi, March 1975; and "Refrigeration Engineering", H. J. Maclntire, F. W. Hutchinson, John Wiley 8 Sons, Inc., New York, 1950, Chapter VI.

The technical efficiency of heat pump apparatus is generally designated by the thermal coefficient E, which can be defined as the released thermal energy/the performed work. The "released thermal energy" comprises the sum of the thermal energy extracted from the source of heat and the work performed. Thus, the efficiency of a heat pump apparatus is determined by the value of the thermal coefficient. The higher the value E, the greater the usefulness of the heat pump apparatus. It can also easily be shown that the theoretical thermal coefficient Meteor of a single-circuit heat pump system can be designated as follows Eteor t-T in which t is the temperature of the mass current to be heated subsequent to the heat pump, and T is the temperature of the heating mass current subsequent to the heat pump.

It is seen from the above that the theoretical thermal coefficient of a single-circuit heat pump is influenced by the prevailing temperatures.

Prior art conventional heat pump systems have been of the single circuit type, i.e. only a single fluid is used for transferring heat from the source of heat to the object to be heated.

The theroretical thermal coefficient of such conventional heat pump systems utilized in present day applications is generally lower than 4 due to the temperature level of the fluid to be heated and/or the temperature of the heating fluid. Such a theoretical thermal coefficient of conventional heat pump systems can be calculated according the above-mentioned formula. This theoretical thermal coefficient is usually further reduced by the mechanical losses inherent in the operation of the compressor and other mechanical elements of the heat pump system. However, the most significant losses result from the less than ideal operation of the compressor. It follows that it is essential to maintain the efficiency of the compressor within an optimum range in order to obtain a satisfactory thermal coefficient for the heat pump system.In this connection, the system should be such that the compressor operates within an optimum pressure ratio range (pressure at pressure side/pressure at suction side) and, additionally, attempts to prevent superheating of the heat pump fluid should be made since such superheating will generally result in a decrease of the efficiency of the heat pump system for a gaseous fluid to be compressed.

Unfortunately, in conventional applications of heat pump systems, the above is not always possible since the pressure ratio of such conventional heat pump systems is determined by the temperature levels required for the particular application as well as by the properties of the available fluids. Thus, it is often not possible to avoid super heating in single-circuit heat pump systems so that a decrease in efficiency is usually unavoidable.

For all of the above reasons, the thermal coefficient of present day heat pump systems is generally in the range of between 2 and 3.

Conventional heat pump systems are in the main operated by electricity. Unfortunately, in most countries, e.g. in Finland, owing to its manner of generation, electricity is considerably more expensive than its equivalent in heat energy. For example, in Finland, the ratio of the price of electricity to the equivalent heat energy value varies today between 2.5 and 4 depending upon tariff policies, etc. In such circumstances, it is not surprising that conventional heat pump systems have not generated wide interest and acceptance in spite of the fact that such heat pump systems can be adapted to have reduced energy requirements.

In view of the world-wide problems in connection with limited energy resources and the ever increasing economic burdens suffered by countries which are not self-supporting in energy and therefore must import fuel, it would be extremely desirable to provide a heat pump system having a thermal coefficient which is significantly higher than conventionally obtained thermal coefficients whereby the electricity-heat price ratio obtained by such heat pump system can be reduced.

The present invention in one aspect provides heat pump apparatus, comprising: a plurality of separate heat pump circuits, each of the said circuits being adapted to have a heat transferring fluid circulate therethrough and each including respective evaporator means and condensor means, and means for directing a mass flow to be heated into heat exchange relationship with each of the said condensor means in series, whereby the temperature of the mass flow to be heated rises when in heat exchange relationship with the fluids circulating through the condensor means of the respective heat pump circuits.

The invention in another aspect provides a method of heat recovery utilizing a plurality of separate heat pump circuits, each of the heat pump circuits being adapted to have a heat transferring fluid circulate therethrough and each including a respective evaporator means and condensor means, the method comprising directing a mass flow to be heated into heat exchange relationship with each of the said condensor means in series in a manner such that the temperature of the mass flow to be heated rises when in heat exchange relationship with the fluids circulating through the condensor means of the respective heat pump circuits.

Thus, in accordance with the present invention, there is provided a heat pump apparatus comprising a plurality of separate heat pump circuits wherein the condensors of the heat pump circuits are in serial heat exchange relationship with the mass current to be heated in a manner such that the temperature of the mass current to be heated increases while in heat exchange relationship with the fluid circulating in the condensors of the separate heat pump circuits. In this manner, the thermal coefficient of the heat pump apparatus according to the present invention is significantly improved relative to conventional heat pump apparatuses.

Theoretically, the heat pump apparatus of the present invention in its most preferred embodiment comprises an infinite number of separate heat pump circuits. However, as discussed below in greater detail, the increase in the thermal coefficient of the heat pump apparatus of the present invention with an increasing number of separate heat pump circuits decreases and, additionally, it is clear that increasing the number of separate heat pump circuits will also increase the cost of the apparatus at least to some extent. As explained below, it has been found that the optimum result taking into account the above mentioned factors is often obtained when the heat pump apparatus constructed according to the present invention includes six to ten separate heat pump circuits.In this connection, the thermal coefficient increases relatively abruptly as the number of separate heat pump circuits initially increases but the rate of increase decelerates as the number of circuits further increases.

The invention will be further described, by way of example only, with reference to the accompanying drawings, in which: Figure 1 is a schematic diagrammatic view of a six-circuit heat pump apparatus according to one embodiment of the present invention which operates in accordance with the method of the present invention; Figure 2 is a graph illustrating a typical temperature distribution in the various circuits of the heat pump apparatus illustrated in Figure 1; Figure 3 is a graph illustrating the variation of the theoretical total thermal coefficient kook obtained utilizing the method of the present invention shown as a function of the number of separate heat pump circuits utilizing temperatures set forth in the illustrative example set forth below;; Figure 4 is a graph of temperature (T)-entropy (s) coordinates depicting the physical characteristics of an ideal application of the present invention in which the fluid circulated in the heat pump circuit and which is vaporized in the evaporator is maintained in a suitable wet state or condition as it enters the compressor of the heat pump circuit; Figure 5 is a schematic flow diagram of a heat pump apparatus according to the present invention utilized in connection with waste water heat recovery; Figure 6 is a front elevation view in schematic form of a typical application of a muiti-circuit heat pump apparatus in accordance with the present invention in connection with a timber drying plant or, in general, a condensate drying plant; Figure 7 is a schematic flow diagram of the heat pump system of the condensate drying plant illustrated in Figure 6;; Figure 8 is a front elevation view in schematic form of a six-circuit heat pump apparatus according to the present invention utilized in a plant and illustrating the associated compressors and expansion valves; and Figure 9 is a sectional view taken along the line IX-IX of Figure 8.

Referring to Figure 1, the heat pump apparatus comprises six heat pump circuits P1... P6, whose evaporators H1... H8 are connected in heat exchange relationship in an order according to ascending evaporizing temperatures Tk~1 (k=1 . . . 6), successively in series to a current M of waste water. In a corresponding manner, the condensors L1 .. .L6 of the respective heat pump circuits P1... P6 are connected in heat exchange relationship in an order according to ascending condensing temperature tk, successively in series to the current m of the water to be heated. Each heat pump circuit P1... P6 also includes conventional compressors K1 ... K6 and expansion valves Tvi ... Tv6, respectively.

Figure 2 illustrates the temperature levels Tk and tk of the heat pump apparatus illustrated in Figure 1 in each heat pump circuit P1 . . .P. The temperatures of the mass current M of the heat releasing waste water current are designated Tc and, correspondingly, the temperatures of the mass current m to be heated are designated tc. As seen in Figure 2, the temperature Tc decreases in the direction in which the heating mass current M flows and, correspondingly, the temperature tc increases in the flow direction of the mass current m to be heated.

A mathematical analysis of the improvement in the thermal coefficient E utilizing the heat pump apparatus according to the present invention as illustrated in Figure 1 and 2 is set forth below, The ratio of the heat capacity flows (heat capacity flow=mass flowx nominal heat, [C]=[c]=W/K) of the waste water and of the water to be heated is designated a (a=C/c). It is assumed that the isentropic efficiency nis of the compressor K is 1. It can be shown that the theoretical thermal coefficient (kok) of the system is in which in which

wherein N is the number of the heat pump circuits.

It can be further shown that an optimum circuit division is obtained (kok has its maximum value), when Tk+1/Tk=Tk/Tk-1' in other words Tk+,/Tk=a constant=m. Now 5N can be reduced to the formula 1 1 SN= ----[-------------- -- 1] (4) a [(1+a)-am]N in which m is defined by the equation m=(TN/TO)'XN. It can further be shown that as No, and now

In an example wherein to=323 K, TN=323 K, TO=283 K, and a=1 , the thermal coefficient of the heat pump system with an N-step optimum circuit division is

in which m=(323/283)1/N and in the system with an infinite number of circuits (i.e., N=oo), the thermal coefficient can be calculated::

By assigning N numerical values, the graphic representation illustrated in Figure 3 is obtained, according to which kook approaches asymptotically the value 8.08, calculated above, as N increases.

If the same values utilized in accordance with the above-recited example are applied in connection with a conventional single-circuit heat pump apparatus, it will be seen that the maximum possible thermal coefficient is only 4.04. However, utilizing a heat pump apparatus in accordance with the present invention having 30 circuits, the thermal coefficient will be 7.81. In the six-circuit system illustrated in Figure 1, Ekok will be of the order of 6.8 or some 70% higher than the single circuit arrangement of the prior art.

The multi-circuit apparatus of the present invention results in two additional important practical advantages relative to the single-circuit apparatus of the prior art, both of which affect the compressor efficiency.

Firstly, since each circuit of the multi-circuit system of the present invention operates at a relatively low temperature differential, i.e. the difference in temperature between the condensing and evaporizing stages, the pressure ratio of the compressor K(pressure at pressure side/pressure at suction side) remains at a suitably low level thereby resulting in a compressor operating with its optimum efficiency range. In direct contradistinction, due to the greater temperature difference existing in the single-circuit system of the prior art, the compressor necessarily operates at relatively high pressure ratios so that the compressor efficiency (nits) is substantially lower.

Secondly, the multi-circuit apparatus of the present invention enables the use of different fluids in the various circuits thereby providing the capability of maintaining the fluid in each circuit in a suitably moist or wet state as it enters the compressor. The necessary limitation of the conventional singlecircuit system to a particular refrigerant fluid is a factor which restricts its use since it is extremely difficult and often even impossible to find a fluid which can be employed within an extensive temperature range. However, in a multi-circuit apparatus according to the present invention, this problem does not arise since an appropriate fluid for each circuit can be chosen on the basis of the temperature range prevailing in that particular circuit regardless of the other circuits.

In this connection, for example, if the fluids operating in the six circuits in the apparatus illustrated in Figures 1 and 2 are at the particular temperatures noted in the above example, the refrigerant fluid "R2 1 " can be beneficially used in circuits P1, P2, P3 of the cold end of the apparatus illustrated in Figure 1 while the refrigerant fluid "R1 1" can be beneficially utilized for the circuits P4, P5, P6 of the hotter end.

Turning now to Figure 4, as mentioned above, the fluid circulated in the heat pump circuits P1.. .PN and which are vaporized in the evaporator of each circuit can be now maintained in a suitably wet state as the same enters the compressor by virtue of the present invention. Thus, Figure 4 illustrates a circulation process of a heat pump circuit P on T(temperature)-s(entropy) coordinates. It can be shown that the most appropriate state of the fluid for which initiation of the compressing is accomplished in the pointed designated F which is defined by the intersection point of the isentropy passing via the point P and of the isotherm T,.The location of this point is the same at all values of the isentropic efficiency nl5. Designating the vapor humidity at the point F by x, it can be shown that

in which Ah1 vaporizing heat at the temperature T, Cp=nominal heat of the fluid (liquid) C'p=nominal heat of the fluid (vapor).

It is shown below by means of an example that the choice of the compression point has a substantial effect on the thermal coefficient.

Thus, for example, assuming that the fluid is water vapor and T=32 .6 and T2=369 K, the ni5 of the compressor=0.7. In conventional heat pump apparatus, there is a suction area at the point F' (Figure 4), whose temperature is about 10--300 higher than T,.

Upon the compression being initiated at the point F' (super-heating 300C), it can be shown mathematically that E=4.7 which is substantially less than at the beginning of the compression at the point F, at which time E=5.5. In actuality, this difference is even greater than appears above since superheating also means that the measured temperature difference between the sources of heat and the fluid should be increased by the same amount, which factor has not been taken into consideration in the above calculations.

The improvement in efficiency noted above is due to the fact that as the gas temperature abruptly rises owing to the adiabatic compression, the volume of fluid will increase and, where cooling is not provided, a significant amount of additional work has to be expended in order to reach the pressure corresponding to the desirable condensing temperature due to the gas volume increase. However, if the gas is delivered to the compressor in the wet state, the vaporizing liquid efficiently cools the gas thereby eliminating unnecessary additional work and thereby improving the thermal coefficient. In practice, the vapor content of the suction area can be suitably adjusted, for example by controlling the temperature on the pressure side, e.g. by means of the number of revolutions per minute of the compressor K.

A detailed example of the application of the heat pump apparatus in accordance with the present invention and as illustrated in Figures 1,2 and 5 will now be presented.

(a) amount of waste water=1.87 kg/s (b) inlet temperature T6 of the waste water=313 K and outlet temperature To=283 K (c) water current to be heated=1.87 kg/s (d) initial temperature to of the water to be heated=313 K.

According to the optimum temperature division T1=287.8K, T2=292.7K,T3=297.6K, T4=302.6 K and T6=307.7 K. Assume T=288 K, T2=293 K, T3=298 K, T4=303 K and To=308 K.

Assume that the isentropic efficiency nl5 of the compressor K=0.65 and the isentropic efficiency n of the expansion turbine To=0.3.

The table below lists the suitable characteristic measuring values of the system: Fluid R21 R21 R21 RI1 R11 R11 Suction pressure/ 1.08 bar/ 1.3/288 1.56/293 1.08/ 1.29/ 1.515/ 298/ 303/ 308/ Temperature/x 283/0.96 /0.96 /0.96 0.99 0.99 0.99 Mass flow, kg/s 0.20 0.21 0.21 0.27 0.27 0.28 Volume flow, I/s 41 35 30 42 38 33 Compressor power, kW 8.4 8.5 8.7 8.9 9.0 9.1 Post-compression temperature OC 71 77 83 84 90 95 temperaturetk, K 319 325 331 337 343 349 Turbine power, kW 0.13 0.14 0.15 0.17 0.18 0.19 Released thermal 47.4 47.4 47.7 47.8 48.8 48.0 power, kW Compressor pressure 3.3 3.3 3.2 3.4 3.5 3.6 ratio Total thermal power 286 kW Total compressor power 52 kW Total turbine power 1 kW Total electricity power 52-1=51 kW Thermal coefficient=286/51 =5.6.

For purposes of comparison, the following thermal.coefficients are calculated for the above example as follows: The theoretical upper limit of the thermal coefficient (for an infinite-circuit system) from the equation (5): EhOk=10.43 The theoretical upper limit for a six-circuit system from the equation (4) and (2): Ekok=8.94 (S6=0.1 0791) The theoretical upper limit of the corresponding single-circuit system:

Referring now to Figures 5-9, examples of actual plant applications of the heat pump apparatus according to the present invention are illustrated. Thus, Figures 5, 8 and 9 illustrate a practical form of a plant application of the apparatus of the present invention illustrated in Figures 1 and 2.

An electric motor M is mounted on a frame 14 having a rotary output shaft 10 which drives a first expansion turbine assembly Ty1, TV2 and Ty3 through a coupling 110. The shaft 10 also drives a tooth wheel or gear 1 2a through the coupling 111, the wheel 1 2a itself driving tooth wheels or gears 131, 132, 1 33 located about its circumference. The tooth wheels 131, 132, 1 33 themselves rotate the shafts of compressors K1, K2, K;, respectively.The tooth wheels 131, 132, 133 can be of different diameters in order to ensure the desired volume flow rate.

Further, the rotary shaft 10 of the motor M drives a tooth wheel 1 2b which drives tooth wheels or gears 134, 135 and 1 3s in a corresponding manner to the structure described above. The tooth wheels 134, 135 and 136 drive respective compressors K4, 6 and K6. Figure 5 illustrates the above mentioned tooth wheels or gears which are designated by the reference numerals 111... 11 6 The structure of the evaporators H . . . H6 and of the condensors .... . L6 of the heat pump circuits P1... P6, respectively, is entirely conventional and well known and will therefore not be described in detail.

Referring now to Figures 6 and 7, another example of an application of the heat pump apparatus of the present invention is illustrated in which the heating mass current and the mass current to be heated are one and the same. Thus, Figure 6 is a schematic cross-section of a condensate drying plant for timer packages while Figure 7 is a schematic view of the multi-circuit heat pump apparatus in accordance with the present invention for use at the drying plant illustrated in Figure 6. The drying plant comprises a building 20 in which the timber packages are dried and which is eauipped with conventional rails on which a carriage 21 having wheels 22 is accommodated. The drying carriage 21 is loaded with a package P of damp timber. e drying air stream is circulated through the timber package by way of channels 30 and 31.The lower portions of channels 30, 31 are defined by inclined walls 24a, 24b, respectively, and the distance between the lower edges of the inclined walls corresponds to the width of the timber package P. A blower 26 is located above the timber package and is operated by a motor 25 so that the blower produces the air circulation through the timber package P. The upwardly directed air stream flow which has already passed through the timber package is designated by the arrow F. This air stream is divided into two branches F3 and Fln. Referring to Figure 7, the air stream branch Fln passes through the evaporators H1... H6 of the heat pump circuits P1... P6, which evaporators are designated by reference numeral 27 in Figure 6.In the evaporators H the air stream branch Fln is cooled und dried in a manner such that its humidity x6 is greater than . The same air stream is conducted through the condensors .... . L6 of the respective heat pump circuits p1... p6 whereupon the air flow is warm.

The above described system comprises six separate heat pump circuits p1... p6, to which are connected the compressors K, . K6 in the manner described above as well as the expansion turbines ..... TV6 in accordance with Figure 7 which are equipped with power transmitting mechanisms such as are illustrated in Figures 8 and 9. In the application illustrated in Figure 6, the heat transferred from the evaporators H (27) to the condensors L (28) by means of the heat pump apparatus is designated by arrows W.A hot, dry air stream Fout exits from the heat pump apparatus, i.e. from condensor L8, this air stream then being mixed with the damp and cold air stream F3 (Figure 6) in a manner such that a mixture having a considerable temperature and humidity (xO) results. This mixture is then directed through the timber package P and is designated by the current F2 in Figure 6.

In each of the above described applications of the invention, the temperature of the heating mass current (M) decreases as the mass current passes through the evaporators H, ... Hn of the heat pump circuits P1... Pn. However, there may be provided a multi-circuit heat pump apparatus wherein the heating mass current (M) is not necessarily the actual mass current but, rather, other sources of heat may be utilized such as, for example, soil, inland water, or sea water.

The present invention can also be utilized in a manner such that several different mass currents are used as the mass current, which several currents are divided for the various heat pump circuits according to the temperature of each mass current, i.e. in a manner such that the distribution .... . TN 1 of temperatures (Tk) in accordance with Figure 2 results. For example, if deep sea water is utilized as the heating mass, a temperature distribution .... . TNl as presented in Figure 2 can be achieved by taking the sea water at different depths, at which time their temperature varies.

Claims

1. Heat pump apparatus, comprising: a plurality of separate heat pump circuits, each of the said circuits being adapted to have a heat transferring fluid circulate therethrough and each including respective evaporator means and condensor means, and means for directing a mass flow to be heated into heat exchange relationship with each of the said condensor means in series, whereby the temperature of the mass flow to be heated rises when in heat exchange relationship with the fluids circulating through the condensor means of the respective heat pump circuits.

2. Apparatus as claimed in Claim 1, which comprises a plurality of continuously operating, separate heat pump circuits the compressors of which are provided with a common drive.

3. Apparatus as claimed in Claim 1 or 2, further including means for directing a heat releasing mass flow into heat exchange relationship with each of the said evaporator means in counterflow relationship with the mass flow to be heated whereby the temperature of the heat releasing mass flow decreases through the evaporator means of the heat pump circuits.

4. Apparatus as claimed in Claim 3, wherein the ratio of the evaporating temperatures of the fluid of each two successive heat pump circuits is a constant.

5. Apparatus as claimed in Claim 3 or 4, wherein the mass flow to be heated and the heat releasing mass flow comprise different mass currents and the said two mass flow circuits are directed in counterflow relation with respect to each other.

6. Apparatus as claimed in Claim 3 or 4, wherein the mass flow to be heated and the heat releasing mass flow comprise the same mass current, the said mass current being directed in a first current path in heat exchange relationship with the said condensor means and in a second current path in counter current flow to the said first current path and in heat exchange relationship with the said evaporator means.

7. Apparatus as claimed in any of Claims 1 to 6, wherein the fluid which is adapted to circulate in each of the respective heat pump circuits when in its gaseous state subsequent to passing through the evaporator means of the respective heat pump circuit is maintained in a relatively wet state as it enters the compressor means of the respective heat pump circuit.

8. Apparatus as claimed in any of Claims 1 to 7, wherein each of the said plurality of separate heat pump circuits has a different heat transferring fluid circulating therethrough.

9. Apparatus as claimed in any of Claims 1 to 8, wherein the said plurality of separate heat pump circuits comprises from 6 to

10 heat pump circuits.

1 0. Heat pump apparatus according to Claim 1, substantially as herein described with reference to, and as shown in, the accompanying drawings.

11. A method of heat recovery utilixing a plurality of separate heat pump circuits, each of the heat pump circuits being adapted to have a heat transferring fluid circulating therethrough and each including a respective evaporator means and a condensor means, the method comprising directing a mass flow to be heated into heat exchange relationship with each of the said condensor means in series in a manner such that the temperature of the mass flow to be heated rises when in heat exchange relationship with the fluids circulating through the condensor means of the respective heat pump circuits.

1 2. A method as claimed in Claim 11, further comprising directing a heat releasing mass flow into heat exchange relationship with each of the said evaporator means in counterflow relationship to the flow of the mass flow to be heated whereby the temperature of the heat releasing mass flow decreases through the evaporator means of the heat pump circuits.

13. A method according to Claim 11, substantially as herein described with reference to the accompanying drawings.