FI64456C - PAO UTNYTTJANDE AV EN VEHICLE PUMP SIG GRUNDANDE FOERFARANDE VIDTILLVARATAGANDE AV VAERME - Google Patents

Description

64456 Heat recovery method based on heat pump utilization

Pa utnyttjande av en värmepump sig grundande förfarande vid tillvaratagande av värme

The invention relates to a method based on the utilization of a heat pump for heat recovery, in which a heating mass or one or more mass streams are connected to evaporators of a heat pump system and a heated mass flow to its condensing devices. with respect to the heated mass flow such that the temperature of the heated mass flow rises when in heat exchange exchange with the circulating media in the condensers of the heat pump circuits, and by which both the heat transfer and heat receiving mass streams are connected to either so that the temperature of the heat transfer mass flow decreases as it passes through the evaporators of the heat pump circuits.

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It is known that a heat pump is a machine that transfers heat energy from a lower temperature to a higher one.

The idea of using a heat pump to heat and utilize waste heat has long been known. With regard to the state of the art bypassing the present invention, reference is generally made to "Possibilities and effects of heat pump use in heating and heat recovery", Aittomäki, Kalema, Lappalainen, Talsio, Wiksten, VTT HVAC Technology Laboratory, Communication 23, Otaniemi, March 1975 and 25 special solutions Refrigeration Engineering ", HJ Maclntire, F.W. Hutchinson, John Wiley 4 Sons Inc., New York, 1950, Chapter VI.

The technical performance of a heat pump is described by the heat factor ε, which is defined as e = transferred heat output / work done, where the transferred heat output is the sum of the heat output from 30 heat sources and the work done. The benefit of a heat pump is thus essentially dependent on the magnitude of the heat factor.

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The larger ε, the more useful the heat pump is. It can also be easily shown that the theoretical heat coefficient eteor of a single-circuit heat pump system can be written as follows: 5 Eteor "rH (1) where t = temperature of the mass flow to be heated after the heat pump T = heating" "" "10 So the heat factor depends on the prevailing temperatures.

The previously known heat pump systems have been largely single-circuit, i.e. only one medium can be used to transfer heat from the heat source to the object to be heated.

Even in the current application cases, the theoretical heat coefficient of such single-circuit heat pump systems remains less than 4 due to the magnitude of the temperatures of the medium to be heated and / or heated. The theoretical heat coefficient of currently commonly used heat pump systems can be calculated according to formula (1) above. In addition, this theoretical heat factor is reduced by the losses associated with both the compressor and other parts of the heat pump system. The main such source of loss is the non-ideal operation of the compressor. Thus, keeping the efficiency of the compressor in the optimum range is essential, as described above, in order to achieve a good thermal coefficient. Therefore, the compressor operating point should be selected optimipainesuh-I (discharge pressure / suction-side pressure) of the area and seek to prevent the compressible gaseous medium in the heat pump system efficiency degradation causes superheating.

30 However, this is not possible even with the special solutions currently known, since the pressure ratio of such systems is determined by the temperature levels required by the application and the properties of the available media. Also, it is not possible to avoid superheating due to efficiency-loss in a single-circuit system.

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The heat coefficient of the currently known single-circuit heat pump systems already varies between 2 and 3, even in the current application cases, due to the reasons mentioned above.

As an example of the prior art for multi-circuit heat pump systems, reference is made to DE-OS 2,637,230 and U.S. Patent 4,124,177.

The main energy used for heat pumps is electricity.

In several countries, such as In Finland, the electricity production method joh-10 support is much more expensive than heat. In Finland, the price ratio of electricity / heat today usually varies between 2.5 and 4, depending on the tariff policy.

In such circumstances, it is natural that the heat pump, despite all 15 energy-saving measures, has not been able to arouse wider interest. Therefore, both in view of the limited energy resources of the whole country and in order to alleviate the economic burden of energy and fuel imports from a non-energy self-sufficient country, it would be important to develop a heat pump system with a heat factor 20 substantially higher than the current electricity / heat ratio.

The object of the present invention is therefore to provide a heat recovery method based on the utilization of a heat pump, in which the heat coefficient achievable is considerably higher than currently known and thus enables heat recovery even in applications where it has not previously been economically possible.

It is a particular object of the invention to provide a method by which an optimal circuit division is achieved in a multi-circuit heat pump.

In order to achieve the above and later objects, the invention is characterized in that the ratio of the evaporation temperatures of the medium of each of two successive heat pump papers is arranged to be substantially constant.

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Theoretically, the invention will be most advantageous when the system according to the method comprises an infinite number of separate heat pump pupils. However, given the increase in heat coefficient ε to be described in more detail as a function of the increase in the number of circuits used according to the invention and the fact that increasing the number of circuits at least with current technology also increases equipment costs, the optimal result is often achieved when 6 to 10 separate systems comprising a heat pump circuit. In this case, it is effectively utilized that the heat coefficient e initially rises relatively sharply as the number of heat pump circuits arranged according to the invention is increased, however, so that the growth rate slows down with larger numbers of circuits.

The invention will now be described in detail with reference to some embodiments of the invention shown in the figures of the accompanying drawing.

Figure 1 schematically shows a six-circuit heat recovery system according to the invention.

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Figure 2 shows the temperature distribution in the different circuits of the system according to Figure 1.

Figure 3 shows the theoretical total heat coefficient obtained by the method according to the invention as a function of the number of separate heat pump circuits at the temperatures according to the example below.

Figure 4 shows in the T-s coordinate system the physical basis of a preferred embodiment of the invention in which the evaporated vaporized medium in the heat pump circuit is kept suitably wet as it enters the compressor of the heat pump circuit.

Figure 5 shows a circuit diagram of a heat pump system according to the invention applied in waste water heat recovery.

Referring to Figures 1-2, an embodiment of the method according to the invention will now be described. The heat recovery-5 intake system according to Fig. 1 comprises six heat pump circuits p ... p ^., The evaporators of which are connected in series in the order of the rising evaporation temperature (k = 1 ... 6) in series with the wastewater flow K. Correspondingly, the condensers L ^ ... L ^ of the heat pump circuits p ^ ... p ^ are connected in series in the order t ascending state of the condensing heat-10 with respect to the water stream m to be heated in the heat exchange connection. Each heat pump circuit p ^ ... p ^ further comprises compressors Κ ^.,. Κ ^ known per se and expansion valves T n ... T,.

Fig. 2 shows the temperature levels of the system according to Fig. 1 and in the different heat pump circuits p ^ ... p ^. The temperature differences of the heat transfer mass flow M are denoted by Tc and the temperatures of the mass flow to be heated are denoted by tc, respectively. According to Figure 2, the temperature decreases in the direction of the heating mass flow M. Correspondingly, the temperature t 20 rises in the flow direction of the mass flow m to be heated.

In the following, the heat coefficient ε improving effect of the method according to the invention will be described with reference to Figs. 1 and 2.

25 Enter the heat capacity flows of waste water and water to be heated (heat capacity flow = mass flow x specific heat, [. C j = [_cJ7 = W / K)

ratio by a (a * = C / c). Let us first assume that the compressor K

isentropic efficiency η. is 1. Then it can be shown that the theoretical thermal coefficient of the system (e ^ ok ^ is 6 64456 ekok -1 * -Γ1- <2) _ ° e - 1

T -T bN

No. 5 where SN * a [\ - "1 <3) k = l N 1 10 where Π = · k ^ '... · A ^ and A ^ = - ^ - (3) k = 1 (1+ a) - a = £ - k-1 N is the number of heat pump circuits.

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It can further be shown that the optimal circuit division prevails (£ kok gets the maximum value) when Τ ^ + · ^ / Τ ^ = T ^ / T ^ _ ^, i.e. T ^ + ^ / T ^ = constant = m.

Then Sjj is reduced to 20 S - 7 [--- vf - T-3 <*) f (l + a) - am]

1 / n

25 where m is determined by the equation m = (T ^ / Tq)

T

1 r N a

It can further be shown that when N - * · «°, then SN - * - ^ (-) - 1 J

. o and then 30 Ekok 1 + “t. r T I (5) —-— - Γ (_ν - il - 1 T -T a LVT * J No o

Example T = 323 K T „= 323 K T = 283 K a = 1 - o N o In this case, the heat factor of an N-stage heat pump system with optimal circuit distribution is 35 7 644S6 ε. 1 - 1 οό'ο -l 11 ^ | [π (6) kok 323 f 1 ^ η

40 * - (2 - =, ^ J

5 where now m = (323/283) 1 / N, and an infinitely many circuit-implemented system has ε. . = 1 + ... -—- - 8.08.

kok «323 r 323 .1 _.

10 40 1-283 ”J

By assigning numerical values to N, a graphical representation according to Fig. 3 is obtained, according to which as N increases, ε ^ ο ^ approaches asymptotically the value calculated above 8.08.

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Thus, in the previously known conventional single-circuit systems, the values according to the example theoretically have the best possible thermal coefficient of only 4.04, while, for example, in the case of 30 circuits it is already as good as 7.81. With the 6-circuit system shown in Figure 1, ε ^ ο ^ = 6.8, i.e. about 70% better than with the single-circuit system.

On the other hand, the multi-circuit system also achieves two important additional practical advantages compared to the single-circuit system, which affect the efficiency of the compressor, and which have been referred to in the past.

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First, each circuit operating at a relatively low temperature difference (condensing - vaporisation) is maintained in the compressor pressure ratio K (discharge pressure / suction-side pressure) for a suitable amount to allow the operation of the compressor optimihyötysuhdealueella. Due to the larger temperature difference, high pressure ratios are required in the single-circuit system, whereby the efficiency of the compressor (η.) Is already substantially reduced by X s.

Second, the multi-circuit system allows the use of different media in the different circuits and allows the medium to be kept suitably wet as it enters the compressor. The inherent limitation of a single circuit system to the use of a particular medium has become a limiting factor in its use, as it is very difficult, often even impossible, to find a medium suitable for a wide temperature range. In a multi-circuit system, this does not present any problem, because the most suitable medium can be selected for each circuit on the basis of the temperature range prevailing in it, independently of the other circuits.

For example, by treating liquids at the temperature of the above-mentioned exemplary case according to Figures 1 and 2 in six circuits, the medium R21 (ρ ^, Ρ2>} ·}) of Figure 1 can be suitably selected for the medium R21 and the hotter circuits (ρ ^, ρ ^, ρ ^) medium R11.

In the heat pump circuits p ^ ... pN, keeping the evaporated medium in the recirculating evaporator suitably wet as it enters the compressor K 15 will be considered in more detail below with reference to Fig. 4.

Figure 4 shows the rotation process of the heat pump circuit p in the T-s coordinate system. It can be shown that the best point for starting the pressing work is the point F, which is determined by the point of intersection of the isentrope passing through the point P and the iso-20 term. This point location is the same for all isentropic efficiency η. values. The humidity of the steam at point F is oi- X s of size x. Then it can be shown that T, r C "(T -T) TC 'T, T, 25 * · [i * -Ηζ - - -Shf - ((1 - ^> 1 <7> where Ah ^ = heat of vaporization at temperature T ^ C '= specific heat of the medium (liquid) 30 Cp "= specific heat of the medium (vapor).

We show by way of example that the choice of the pressing point is essential for the heat coefficient.

35 As an example Consider the case where the medium is water vapor and T ^ = 321.6 and T £ = 369 K. Let η. ^ = 0.7 of the compressor. Known heat pump solutions t 9 64456 ti have a suction space at point F '(Fig. 4), the temperature of which is about 10 ... 30 ° C higher than T ^.

When the compression starts from the point F '(superheating 30 ° C), it can be shown by calculation that ε = 4.7, i.e. substantially weaker than when the compression starts from the point F, where ε * 5.5. In reality, this difference is even greater, as superheating also means that the temperature difference between the heat source and the medium must be dimensioned to be the same, which is not taken into account here.

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The improvement discussed above is due to the fact that when the gas temperature rises sharply due to adiabatic compression, the volume also tends to increase and if no cooling is provided, additional work has to be done to save the pressure corresponding to the desired condensing temperature due to the gas volume increase. By introducing the gas into the wet compression, the evaporating liquid effectively cools the gas, reducing the "unnecessary" additional work required and thus improving the heat coefficient. In practice, the steam content of the suction space can be adjusted to suit e.g. adjusting the temperature of the discharge side of the compressor by means of, for example, K-20 revolutions fan speed.

The following is a detailed implementation example of the system according to Figures 1, 2 and 5: 25 - wastewater volume 1.87 kg / s

- waste water inlet temperature T, 313 K and outlet temperature T = 283 K

o o - heated water flow 1.87 kg / s - initial temperature of heated water t 313 K.

According to the optimal temperature distribution, T ^ * 287.8 K, T ^ * 292.7 K, T3 - 297.6 K, T4 = 302.6 K and Tj = 307.7 K. Choose Τχ - 288 K, T2 = 293 K, T3 = 298 K, T4 = 303 K and Tj = 308 K.

Assume that the isentropic flux ratio of the compressor K is about * 0.65 X s 35 and the isentropic efficiency T 0.3 of the expansion turbine * 0.3.

The appropriate main dimension values for the system are given in tabular form below: 10 6 4 4 5 6 oo o ro »—I oo o> m cn ri γ-ho '« o VO rl * · »# * * # 'm σ \ o oo co »·> ro σ' <t <r ι-h O ro to rl O l- ~ 00 i — i co cn o rc oo m IA *> er> rs« s ρϋ σ> σ > o oo o \ o cn o oo cn cn σ> m cr »<3 · * * cn

r-l O

00 rM o * nr *> »r-ι CN CN <T \ rl 00" V, Λ Λ AM f) PS oo ct o cn oo <tr ^ or ^ cn O Cn <* 00 cn ^ 3 * r *> cn

r-l O

cn

r-l r-l IA

CN CN CM Γ-- r — I Γ- * CN

cn ***% »· *« »»>

Oi Ό Ό O O 00 CO r-l O CO

m σ 'co oo co <r "r cn

r-l O

00 r-l 00 r-l cn cn cn m t — i cn CN V, vO * A a »*> PS cn ct» o tn oo r · * and o r-v cn

* * »Cn cn -T

r-l o cn

H VO

rl «d 0> x ~ o cn • HO CM -if r — 1 <f en rl CO '' '" o * · r «« r> * t • f4 cn o en o ri oo r-ι <t » o cn

*> 00 vT Γ> * f — I -sT

pu Oi I-I esi CO., LTD

»

M D

T3 »C - Λ X Λ! CU 0 3 \ d js a> to »« -i-i o) o r — C O CU 4J e

H CO i-C, d ^ E *: o -H

U 0) 1-1 Äi ^! Q.cO

: θ öo * 4-> seo S Cu CU jj, co ·! —i O x x: crj

S 4JCO ^ ÄOx-1C

m) x u · μ e 4J js -ru

Γ-l ce-HMCU »0) 3 S-J

UJ> O CO CO CO -U 4-1 o a) o) lu co co jxZ t — i r — i * r-i 4-> cn

d d -H 3 CO 3 -rl d 0) CO

• ι-l · Η> 3 0) 4J 4-1 4J -H * J 0)

co CO co> * -i co: o CO -I-I 3 M

• H CU cn CO a -H p, Ο. , ο l> Cu H 3 CO »—I 0 MS H M O Θ

: c0 B cfl · Η O 3 KO SCO 3 3 O

> M £ HfeiCU, rHxJHuJi «ä 64456

Total heat output 286 kW Total compressor power 52 kW Total turbine power 1 kW Total electric power 52 - 1 e 51 kW 5 Heat factor = 286/51 = 5.6

For comparison, the following heat coefficients should also be calculated for this example: 10 - Theoretical upper limit of the heat coefficient (for an infinite circuit system) from equation (5):

Sok - l0- «15 - 6-circuit theoretical upper bound from equations (4) and (2): S„ k '8 · 94 <S6' ° · 10791> - Corresponding 1-circuit theoretical upper bound: 20 ekok * + 3Ϊ3 “5 , 21 313¾- · ° -1186 * 1 <S1 '1' [, 313 * θ '° · ιι86> 25 2 "283

An example of an apparatus for implementing the method according to the invention will now be described with reference to Fig. 5. Figure 5 shows a preferred form of device implementation of the system of Figures 1 and 2.

An electric motor M is placed on the frame, which drives the first expansion turbine group TV ^ »TV2 And ^ 3 ·

The shaft 10 is further adapted, via a clutch, to rotate the screw compressors Κ ^, ^, Κ ^ by means of gears 35. The size of the gears 12,644 56 can be selected differently to provide the desired volume flow.

Correspondingly, the system includes a second compressor group Κ ^, Κ ,. and K 2, 5 which are driven by gears. In Fig. 5, the above-mentioned gear transmissions are schematically shown by reference numerals 11 ^ ..- lg.

The evaporators and condensers L ^ ... L ^ 10 of the heat pump circuits ρ ^.,. Ρ ^ have a conventional structure and are already known, so it is not necessary to describe the details of their implementation in this context.

An embodiment of the invention has been described above, in which the temperature of the heating mass flow (M) decreases as this mass flow passes through the evaporators (^ ^.,. Η ^) of the heat pump circuits (p ^ ... pN). in cases where the heating mass flow (ή) is not necessarily the actual mass flow, but the heat source used is, for example, soil, inland water, seawater 20 or the like.

The invention can also be applied in that several different mass flows are used as the heating mass flow, which are distributed to different heat pump circuits (p ^) according to the temperature of the different mass flows, e.g. so as to obtain a temperature distribution (T ^) similar to Fig. 2. .Τ ^^. For example, if deep seawater is used as the heating mass, then by taking seawater from different depths, where their temperature varies, a temperature distribution Tq similar to Fig. 2 can be obtained ....

As shown in Fig. 5, the same motor M is driven by separate heat pump circuits 0 o p ^ .. «Pjj via its shaft 10 and switches 11 _... 11 ~. The separate heat pump circuits ρ ^.,. Ρ ^ are continuous and the system is controlled by varying the load by changing the speed of all compressor units by -35 miles and not by changing the number of heat pump circuits and switching the circuits off, which has a detrimental effect on heat coefficient according to the above theory.

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The following are claims within the scope of which the various details of the invention may vary within the scope of the inventive idea.

Claims (1)

64456 u Claim A method based on the utilization of a heat pump for heat recovery, wherein the heating mass or one or more mass streams (K) are connected to evaporator devices (H) of the heat pump system and the heated mass stream (m) to its condenser devices (L). ) comprises a plurality of continuous separate heat pump circuits (p ^ ... p ^), and the condensers (L ^ ... L ^) of the heat pump circuits are connected in series with respect to the heated mass flow (m) so that the temperature of the heated mass flow (m) (m) ί £) rises when in heat exchange contact with the circulating media (L ^ ... LN>) in the condensers (L ^ ... LN>) of the heat pump silicon (10). are either the same mass flow or different mass flows, are connected to each other countercurrently via heat pump circuits so that the temperature of the heat dissipating mass flow (M) 15 (Tc) counts as it passes through the evaporators (Η ^.,. Η ^) of the heat pump circuits (ρ ^.,. Ρ ^), characterized in that the ratio of the evaporation temperatures (Τ ^ / Τ ^ _ ^) of the medium of each of two successive heat pump circuits is arranged essentially constant.